EP4547944A2 - New process for isothermal compression and expansion of gases and some devices for its application - Google Patents

New process for isothermal compression and expansion of gases and some devices for its application

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Publication number
EP4547944A2
EP4547944A2 EP23751123.3A EP23751123A EP4547944A2 EP 4547944 A2 EP4547944 A2 EP 4547944A2 EP 23751123 A EP23751123 A EP 23751123A EP 4547944 A2 EP4547944 A2 EP 4547944A2
Authority
EP
European Patent Office
Prior art keywords
gas
piston
liquid
isothermalizer
thermal
Prior art date
Legal status (The legal status is an assumption and is not a legal conclusion. Google has not performed a legal analysis and makes no representation as to the accuracy of the status listed.)
Pending
Application number
EP23751123.3A
Other languages
German (de)
French (fr)
Inventor
Arpad Torok
Current Assignee (The listed assignees may be inaccurate. Google has not performed a legal analysis and makes no representation or warranty as to the accuracy of the list.)
Individual
Original Assignee
Individual
Priority date (The priority date is an assumption and is not a legal conclusion. Google has not performed a legal analysis and makes no representation as to the accuracy of the date listed.)
Filing date
Publication date
Priority claimed from PCT/RO2022/000007 external-priority patent/WO2022271046A1/en
Application filed by Individual filed Critical Individual
Publication of EP4547944A2 publication Critical patent/EP4547944A2/en
Pending legal-status Critical Current

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Classifications

    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F02—COMBUSTION ENGINES; HOT-GAS OR COMBUSTION-PRODUCT ENGINE PLANTS
    • F02G—HOT GAS OR COMBUSTION-PRODUCT POSITIVE-DISPLACEMENT ENGINE PLANTS; USE OF WASTE HEAT OF COMBUSTION ENGINES; NOT OTHERWISE PROVIDED FOR
    • F02G1/00—Hot gas positive-displacement engine plants
    • F02G1/04—Hot gas positive-displacement engine plants of closed-cycle type
    • F02G1/043—Hot gas positive-displacement engine plants of closed-cycle type the engine being operated by expansion and contraction of a mass of working gas which is heated and cooled in one of a plurality of constantly communicating expansible chambers, e.g. Stirling cycle type engines

Definitions

  • the invention refers to a compression process and a similar expansion process of gas and vapors, processes which leads to a progressive increase (respectively decrease) of the gas pressure in the working enclosure, from a p, starting value to a target p f value without significantly affecting its average T m temperature.
  • the methods revealed in the invention for the implementation of the proposed procedures use existing installations (or parts thereof) in the prior art, but also new installations, proposed by this invention. Through the rigorous application of this process, when they are used using only techniques and devices experienced in the prior art, a significant increase in the energy efficiency of these devices is obtained, therefore, a significant reduction in the energy consumed for the gas compression, respectively an increase in the energy supplied as a result of the expansion.
  • the proposed invention does not stop at these results, but proposes a series of new devices, using which the exergetic performance of the compression and detention processes increases even more strongly, by increasing the performance of all technologies in which the compression and/or expansion of gas and vapors have an important share: transport, storage and liquefaction of gases, production of mechanical work using classical and, especially, renewable and waste heat sources, efficient storage in various types of tanks of thermal and mechanical energy from these heat sources, installations for heating living spaces and for providing domestic hot water, refrigerators and heat pumps, installations for treatment and conditioning of air, etc.
  • the invention also contains the description of some installations in these technological fields, to which the application of the described procedure involves a series of structural changes, through which new technologies are obtained, with superior results to those in the prior art.
  • thermodynamic processes in compressors and expanders of the current state of the art are carried out with a polytropic index, between the isothermal and the isentropic index, and differ from one type of apparatus to another, but also vary during the process.
  • polytropic transformations are resorted to even when isothermal or isentropic ones are preferable, because of considerations of minimum cost and technical limitations of the respective configuration.
  • isentropic index for the complete conversion, without energy changes with the environment, of thermal energy into mechanical energy and vice versa
  • isothermal index in which the internal energy of of the gas not to change).
  • the speed at which this energy is discharged depends on the size of the difference between the two temperatures, on the physicochemical characteristics of the gas and the materials from which the compressor is made (which contributes to the definition of an overall heat transfer coefficient C GT ), on the size of the contact surfaces between the gas and its environment (which contributes to the definition of a total heat transfer area A GT ) and on the distribution of temperatures within the gas and by its convective movements.
  • the size of the contact surfaces and how they vary during the compression/expansion process, as well as how temperatures are progressively distributed inside the apparatus, are constructive features. Maintaining a constant average temperature of the gas during the compression process is perfectly possible and can be achieved by maintaining the equality between the mechanical energy ceded by the piston to the gas (dependent on its velocity) and the thermal energy ceded by the gas to heat-absorbing surfaces in its environment. This equality can be achieved by the appropriate modification, through rigorously directed actions, of the speed of heat transfer between the gas and its environment.
  • T iz is made according to the characteristics of each particular case and is a compromise between the amount of energy consumed in addition to the ideal compression and the duration of the compression cycle (therefore with the power of the compression plant).
  • any compressor with positive displacement in the state of the art can behave perfectly isothermally, if the trajectory of the moving component, the one that determines the variation of the gas volume in the enclosure, is properly established (respects the isothermal trajectory), and the methods of increasing the global heat transfer coefficient contribute to increasing the power density of the device.
  • the duration of the isothermal compression cycle can be considerably reduced, if the contact area between the gas and the ambient is large throughout the compression cycle, and especially when the piston approaches the top dead center TDC.
  • a first objective of the invention is to propose isothermal compression and expansion apparatus with full control over the deviations from the ideal trajectory, respecting a sequence of operations through which the proposed objective can be achieved, simultaneously with achieving the best compromise between energy efficiency and power density of the installation.
  • the process described in the present invention and the devices proposed for its implementation have as their objective the consistent use of the sequence of the 3 phases, previously listed, of compression and expansion with maximum energy efficiency, in order to obtain constant temperature differences AT between the gas and its ambient environment.
  • the proposed devices are designed in such a way that the instantaneous heat transfer values from the gas to/from the constitutive elements of the isothermalizer and from them to the cold/hot source are always as high as possible, thus ensuring maximum efficiency for the technological installations that use them.
  • a second objective of the invention is that the devices proposed in the invention benefit from a sufficiently high power density, so that the high energy efficiency is not paid for by diminishing other performances of these devices. Since in the case of a perfectly isothermal transformation, the instantaneous thermal energy taken by the gas from the mechanical energy of the moving components is equal to the instantaneous heat flow between the gas and its environment, we can achieve this goal by methods of sharp reduction of the polytropic coefficient of the transformation.
  • the instantaneous velocities for the entire isothermal trajectory viz(t) (for any compression/expansion ratio) must be as high as possible, which will allow the circulation, in the same time interval, of higher gas flows, keeping the same gap of temperature AT, so an increase in both power and energy efficiency of the device.
  • a third objective of the invention is to propose new complex installations, made by incorporating the types of densifiers and rarifiers described above.
  • - fig. 6 izotermalizer with thermal sponge made of elastic cords and horizontal metal plates
  • - fig. 7 izotermalizer with thermal sponge from peripherally mounted coil springs and metal plates
  • isothermalizer with liquid piston composed of elementary isothermalizers that have the thermal sponge composed of vertical telescopic fins compressible by moving a plate - fig. 24A: schematic diagram of a liquid piston isothermalizer, composed of elementary thermal sponge isothermalizers composed of telescopic vertical fins whose compression is done by moving the solid piston of another elementary isothermalizer
  • gas piston isothermalizer in which the inner tank contains a thermal sponge of closely spaced vertical plates, and the coolant, coming from the main (outer) tank, is sprayed through a sprinkler system, or moves exclusively under the action of its own weight
  • Fig.40 isothermalizer with gas piston, where the inner tank contains a thermal sponge consisting of a metal conveyor belt, and the coolant, coming from the main (external) tank, is sprayed through a sprinkler system - fig. 40A: alternative mounting configuration of the roller belt to the densifier in Fig.40
  • - fig. 43 isothermalizer with gas piston, where the second stage is a tank with a thermal sponge made of metal foam and another of metal inserts, mounted on top, and the gas is cooled by a sprinkler system with variable flow
  • thermodynamic system for energy recovery and thermal storage from residual liquid agents
  • ambient space heating installation consisting of a working gas circuit under pressure, operating according to a Brayton cycle
  • ambient space heating installation consisting of a pressurized working gas circuit, operating after a Brayton cycle and in which the liquid heat transfer agent is replaced by a working gas under pressure
  • thermodynamic sterilization of air composed of an adiabatic compressor and expander, between which a 4-way valve with large passage sections is mounted
  • gas liquefaction system which works according to a Siemens cycle, the condenser being cooled with a heat pump
  • the invention proposes a different solution: instead of determining from the design phase, with an inherent margin of error, how the instantaneous power of the moving device must vary, we introduced an automatic device to the system adjustment, which, based on information collected in real time by a series of transducers, determines, regardless of environmental variations and other disturbances that may occur, the meaning and amount by which the quantity must change at that moment of the force acting on those moving parts that influence the average temperature of the working gas.
  • the procedure in order to apply it not only in the case of piston devices, where the main regulated quantity is the displacement speed of the piston, but in the case of all devices with positive displacement, by introducing automatic systems for regulating the angular speed of the rotor and/or the inlet and outlet flow rates of the cooling fluids (transferring even the piston function to these fluids).
  • the invention also describes a gas piston isothermalizer, which isothermally compresses gases in a closed enclosure with constant volume, by controlling the gas flow rate introduced, simultaneously with the control of the input and output flow rates of a heat transfer fluid.
  • rotary isothermalizers in order to achieve an isothermal transformation, it is necessary to adjust several parameters simultaneously.
  • a single working enclosure for example, compressors with a single vane in the rotor
  • maintaining the temperature between two close limits is done by varying the angular speed of the rotor during each rotation of it , and in the case of those with several enclosures with different volumes, each in different phases of compression (lobe compressors, gear compressors, screw compressors, scroll compressors, etc.), the speed can be kept constant, but in each enclosure, the coolant/heater flow rate and the piston fluid flow rate are varied.
  • the isentropic compression step can be replaced, for a small AT difference, with a polytropic compression.
  • This compression can take place inside the enclosure, through commands given by the controller to the actuator that changes the volume of the enclosure, imposing such a rapid change that the isothermal temperature is reached from the first moments of the process.
  • the polytropic compression is indicated to be done outside the enclosure, in a compressor with a simple configuration (thus with a low cost price), for example in a blower, or other rotating device (which also involves the use of a small intermediate tank), which ensures the rapid attainment of the isothermal temperature, at a higher power density, simultaneously with a higher flow rate and a higher degree of use of the isothermalizer enclosure. In this way, the rapid displacements of the piston, those at high speed in the initial phase, are also avoided.
  • the determination of the isothermal v iz (t) of the actuator is as accurate as possible (which implies, in the case of the theoretical approach, an exact determination of all the quantities involved in the differential equation that describes the phenomenon and of all the correlations between these quantities), and the actuation mechanisms intended to achieve this speed should be robust, have a response time as short as possible and a sufficiently large adjustment range.
  • the appropriate modification of the piston speed can be done by making appropriate kinematic chains, actuated by motor assemblies with variable speed (preferably, direct current motors, linear motors, stepper motors, or with high-performance hydraulic devices for energy transmission).
  • variable speed preferably, direct current motors, linear motors, stepper motors, or with high-performance hydraulic devices for energy transmission.
  • the variation of the working liquid flow rate can be done with the help of nozzles with variable section (pressure-swirl nozzles).
  • temperature sensors mounted in certain regions of the apparatus can transmit signals that, after processing, cause to be sent commands to the flow regulators on the coolant agent pipes. In this way, an isothermal evolution can be achieved, regardless of the temperature of the external environment, the temperature of the components of the compressor, of the thermal sponge or of the cooling agents. Moreover, the information collected can be used to adjust the lubricant flow, the spray coolant flow, etc.
  • Fig.2 shows the scheme of principle for this type of installation.
  • the processor (controller) 12.4 (DC) compares the pressure measured by the pressure sensor 12.5, with the corresponding one in an isothermal transformation, the working temperature Tiz and the position Li of the piston at that time.
  • DC sends the appropriate control to the drive system 12.3 (here, based on a linear motor), moving the piston 12.2 which moves in cylinder 12.1.
  • the excess heat of the gas is absorbed by the thermal sponge 12.6 and the other components of the densifier.
  • Fig.2 is represented the principle diagram of such type of installation, made on the configuration of a solid piston compressor from the state of the art.
  • the figure was represented with a continuous line the circuit of the heat transfer fluid, with a broken line the signals transmitted by the transducers, and with a dotted line (ACAD 10W100) the commands transmitted by the controller to the executive bodies.
  • the system is managed by the processor 12.4, DC, which constantly compares the pressure measured by the pressure sensor 12.5, with that corresponding to the working temperature T, z and the volume / at each moment.
  • the calculation of the instantaneous volume is based on the position at that moment of the piston, provided by the position transducer L, from which the volume occupied by the liquid (calculated on the basis of the signals provided by the two flowmeters) is subtracted.
  • the DC transmits the appropriate command to the drive system 12.3 (here, based on a linear motor), setting in motion the piston 12.2, which moves in the cylinder 12.1.
  • the linear rotating field is created by a cylindrical set of electric coils and a set of permanent magnets (or a second set of coils) positioned on the piston rod, which moves along the axis of the cylinder.
  • the change in the heat flow rate taken by the liquid is made, simultaneously with the modification of the volume of the enclosure, through controller commands to the two 7v servovalves and possibly to the pump drive motor. Operation at maximum efficiency is ensured by the correlation of the two processes of isothermal temperature regulation.
  • the system is tested on the test bench, being able to experiment for a series of different values of T iz , so for different average piston speeds and different values of energy efficiency.
  • the optimal trajectory and the signal to be transmitted to the drive system are chosen.
  • the decision can be made to increase the absorbent surface of the thermal sponge, or to increase the heat transfer agent flow rate.
  • Fig.2a shows a solid piston isothermalizer, whose piston is actuated by means of a disk 12.7 through whose profiled channel the bearings 12.8 run. The disks are driven by means of the axis 12.9 driven by the alternating current motor 12.10.
  • Fig.2A is represented, as a function of time, the allure of a curve viz(t), and in Fig.2B, the way in which this curve is transformed into a closed curve (the profiled channel) whose trajectory depends on the angle of rotation of the engine.
  • the mounting variant with a single disk, with the execution of the profiled channel on one side of it and the coupling system to the motor shaft executed on the opposite side, represented in Fig. 2C is the variant that occupies the smallest volume, compared to the variant with two disks from Fig. 2a, with the version with a profiled cam, or with the connecting rod-crank drive system.
  • a gear wheel is mounted on its axis, next to each device, which drives the rod of the respective device through the teeth on the periphery of each profiled disc.
  • the system is very versatile, allowing these discs to be driven at different speeds, with different trajectories, or to couple different types of devices to the motor shaft.
  • Other modes of motion transmission can also be used, for example through belts.
  • This technical process for making isothermalizers can be applied to any type of compressor/expander in the state of the art, including those which, being driven by a high-speed engine, through a connecting rod-crank system, have a sinusoidal variation of the piston speed and a trajectory close to the adiabatic one, and the average polytropic coefficient of the transformation is high.
  • the prototype of this compressor will be equipped with a complete real-time speed control system, then the AIA (adiabatic-isothermal-adiabatic) curve of the transformation is raised, in which T jz is chosen in such a way that the compression cycle time is the same.
  • T jz is chosen in such a way that the compression cycle time is the same.
  • laboratory devices can also be built to study how the average temperature varies in a closed enclosure with variable volume, and to obtain the curve of variation as a function of time, as well as as a function of the instantaneous position of the moving body, of the average temperature of the gas in the respective enclosure.
  • a series of pressure and translation transducers are mounted in this enclosure to determine the instantaneous position of the mobile organ that determines the change in the volume of the enclosure. Based on the pair of data obtained with the help of these transducers, the time variation, for different input signals, of the average temperatures of the gas in the enclosure is subsequently determined, with the help of gas laws.
  • Another possibility offered by the solution proposed in the invention is that of the global change, during operation, of the average speed of the drive motor (change valid for longer periods of time, of the order of multiples of the duration of a cycle), change that can be obtained by the variation of the supply voltage/current of the direct current motors, or of the frequency of the current with which the alternating current motor is supplied (single-phase, or three-phase, synchronous, or asynchronous) with the help of frequency converters.
  • These speed changes are required in a wide range of applications:
  • the heat transfer agent is only used to equalize the temperature of the thermal sponge and the walls of the device, keeping the temperature difference AT, and changing its temperature by the same value as the temperature of the thermal sponge increases .
  • the temperature difference AT under the conditions of increasing the temperature of the thermal sponge entails the need to increase the power contribution of the piston - in the case of energy storage systems from renewable sources, the energy to be stored can show important variations over time. Similarly, during the supply of stored energy, load changes occur frequently. These variations can be compensated for by changing the speed of the piston drives
  • the hot source, the cold source, or both, involved in the operation of heat engines, refrigerators, or heat pumps are finite sources, or with random variation, the operation of the installation leads to changes in the temperature of these sources.
  • the need to maintain the temperature difference AT implies a change in power, which can be achieved by changing the speed
  • the temperature of the gas in the enclosure T iz is deliberately changed, continuously, depending on the temperature of the sponge, without changing and the temperature difference AT, thus keeping the shape of the isothermal trajectory unchanged and changing only the temperature and pressure level at which the transformation takes place, thus the nominal working power.
  • the lateral surface of the cam must contain a portion of great length (approximately equal to the maximum radius of the disc) which makes as small an angle as possible with the plane passing through the center of rotation and through the tip of the cam.
  • This configuration requires the guide roller located at the end of the piston rod to have as small a radius as possible.
  • the final adiabatic expansion can be done, after the end of the compression phase, inside the enclosure (by reversing the direction of travel of the piston and strongly accelerating it, followed by a strong deceleration to a stop, immediately followed by a new change of direction to carry out the evacuation of the gas from the enclosure. Throughout these stages, it is necessary to achieve very high accelerations and decelerations. For this reason, it is preferable that the return of the gas temperature to the T amb value is done in an external expander.
  • the isothermalizer is provided with an additional enclosure 12.15, which communicates through a section as large as possible with the main enclosure.
  • the additional enclosure 12.15 is constructed in such a way that, together with the surface S 1 ; through which it communicates with the isothermalizer enclosure, it constitutes a compressor. It is equipped with the piston 12.19 which, at the top dead center (TDC), perfectly restores the surface cut from the wall 12.1 of the main compressor, allowing its piston 12.2 to move without obstacles.
  • the axis of the isothermalizer is vertical, allowing a layer of liquid 12.22 to be deposited on the upper surface of the piston 12.1 , which has the role of occupying the entire dead volume, when this piston is at the top dead center. At the same time, this layer of liquid ensures the sealing of the enclosure, against possible liquid leaks past the gaskets of pistons 12.2 and 12.19.
  • the piston 12.19 of the secondary compressor 12.15 moves from TDC to BDC, allowing the gas sucked in through the valve 12.12 to fill both enclosures.
  • This expander is equipped with the piston 12.21 , which at this moment is in position Ti and the discharge valve 12.23, mounted at the highest elevation, through which the expander communicates with the storage tank under constant pressure. Therefore, transvasation has the effect of replacing the compressed air in the isothermalizer with an equal amount of liquid of the same pressure.
  • the position Ti of the piston 12.21 is calculated so that in this position, the volume of the expander is equal to the volume of the compressed gas in the respective cycle. With piston 12.2 in this position, discharge valve 21.13 closes and intake valve 12.12 opens, allowing new intake and compression phases to begin in the main chamber.
  • the pistons of the two adiabatic devices are actuated by elastic springs, which have the advantage of obtaining high accelerations, of the possibility of storing the mechanical energy consumed to compress the spring and its quick release. They allow the easy installation of simple mechanical devices, through which the BDC positions of the adiabatic devices can be easily changed during operation, changing the volumes of these devices and, consequently, the value of the temperature T iz . Also, for systems in which the sponge has the role of storing thermal energy, automatic regulation systems can be designed in which the volume of the two adiabatic devices is adjusted according to the instantaneous temperature of the thermal sponge, or of the coolant.
  • Another improvement of the isothermal transformation process can be obtained by using, whenever possible, double-acting piston isothermalizers.
  • These isothermalizers are composed of two independent enclosures, separated by a common piston: when the compression and discharge phases take place in one of the enclosures, the intake phase takes place in the other enclosure.
  • cycle time with increased compressed air flow
  • a considerable simplification of the profiled cam that guides the movement of the piston by eliminating the steeply sloped portions, which makes it possible to permanently drive the piston with a power constant, eliminating shocks and vibrations.
  • the master piston 13.2 slides inside the master cylinder 13.1.
  • the first part of the piston stroke is divided, by the telescopic construction attached to the piston rod 13.3, into 4 segments of equal length and a segment of variable length, but both the number of segments and their length are at the discretion of the designer.
  • the portion 13.3 of the rod is rigidly attached to piston 13.2 in its center, and an outer ring 13.4 is attached to the opposite end, larger than the diameter of the rod.
  • Sections 13.3i (13.3a, 13.3b, 13.3c and 13.3d), are ring cylinders, which at the top have attached an inner ring 13.5i, and at the bottom have attached an outer ring 13.4i.
  • cylinders have an inner diameter equal to the outer diameter of the outer ring of the previous segment, and their outer diameter is equal to the inner diameter of the inner ring of the next segment.
  • the outer rings of each segment slide on the inner surface of the next segment, and the inner rings slide on the outer surface of the previous segment, the seals 13.6 providing the sealing.
  • the space between the bottom surface of the piston and the upper surfaces of the inner rings, as well as that between the outer surfaces of a segment and the inner surfaces of the next segment are vacuumed. Configurations may also be made, in which this space is occupied by a liquid or gaseous fluid at atmospheric pressure or a different one, if to this fluid is assigned an external reservoir and a series of flexible and fixed pipes for its proper circulation.
  • a suitable range of piston speed trajectories can also be obtained from the combination of the motion of a solid piston driven by a mechanical device, with the additional movement of extending its rod, movement due to the hydraulic power of a liquid agent.
  • the upward jump of the active surface of the piston is repeated each time an outer ring of a segment steps on the lower surface of the inner ring of the next segment.
  • the active surface becomes equal to that of the piston and its movement continues, without jumps, at a decreasing speed, until the piston power equals that required to compress the gas to the desired pressure.
  • the engine power may be exceeded if the telescoping continues in the same way, with ring segments with the inner surface of the outer ring larger than the diameter of the piston (and the densifier cylinder), adding an additional cylinder with the corresponding diameter.
  • the power of an isothermal process depends, in addition to the initial pressure of the working gas and the compression ratio, on the isothermal temperature (established initially, depending on the desired energy efficiency) and on the speed with which it is removed /absorbed excess/necessary heat of the working gas, rate which depends on the device configuration.
  • the second objective of the invention is to propose a series of design techniques for isothermalizer configurations, which will ensure them competitive power densities.
  • the isothermalizers proposed in this invention use to a greater or lesser extent, according to the concrete applications served and the constructive variant chosen, one or more techniques for reducing the polytropic compression/expansion index, from the current state of the art, used as such , or with innovative modifications that increase the performances obtained.
  • One of these techniques is to increase the surfaces of the compressor components that come into contact with the working gas: by changing the ratio between the diameter and the length of the device, by making the outer walls in a bellows-type configuration, or harmonic, which allows changing the length of the cylinder as a result of the displacement of the piston, without reducing the surface that is in contact with the working gas, by adding cooling fins on the inner faces of the piston and/or the cap, fins which, at displacement of the piston, they can interpenetrate, the introduction of deformable metal inserts inside solid piston devices, the introduction of non-deformable metal inserts inside liquid piston or gas piston devices (inserts that do not prevent the movement of this type of piston).
  • the deformable sponge is made up of one or more solid components (in many configurations a liquid component is also introduced, in a fixed amount) with variable volume and/or position.
  • the solid components of the thermal sponge have the total surface Si, which is in direct contact with the working gas, approximately the same throughout the compression, and their degree of deformation is constantly controlled by the position of the piston, with each position of the piston corresponding to a different shape of the sponge, a property ensured by the elasticity of some of its component elements, or by kinematic devices controlled by (or in correlation with) the movement of the piston.
  • the liquid components of a thermal sponge mounted in alternative devices can also play the role of a transport agent of excess thermal energy, if during the discharge and intake phases, they are replaced by cooled components, or they can take over the role of a liquid piston, if during thermodynamic transformation, the amount of liquid introduced is different from that discharged.
  • thermal sponges made of solid elements are more difficult to make, but liquid elements of the sponge can be introduced and discharged, both from the inlet phase and during the thermodynamic transformation, in the form of a jet, in the form of drops, in powder form, in foam form, etc.
  • they can combine the role of liquid piston with that of coolant/heater.
  • a compressor equipped with a solid elastic thermal sponge, in the form of a helical spring, having inside a channel through which a liquid flow circulates for the evacuation of excess thermal energy is disclosed in WO2014005229 - Temperature management in gas compression and expansion (US20140007569).
  • This type of sponge meets some of the listed characteristics, but presents a number of disadvantages: the complexity of its technical realization, the overall low heat transfer coefficient and the low power density.
  • devices with constructive characteristics similar to the thermal sponges described previously are used, their temperature is variable in an oscillatory manner, they most often have the role of transfer heat from one area of the appliance served to another, and back.
  • the easiest to implement is a sponge in the form of a multi- alveolar system, with cells that communicate with each other (foam obtained from elastic, metallic or non- metallic compounds, from natural rubber, synthetic rubber, from elastomers, from elastic polymers, from isoprene, etc. , deformable sheets, with large, regular or irregular surfaces).
  • the actual volume of the sponge V B composed of the total volume of its solid components and the volume of the gas in the closed alveoli (disregarding the volume of gas in the open alveoli) should not change significantly, for any of the piston positions, not even after a large number of compression processes. Some small variations are acceptable, however, if they are oscillations around an average value.
  • the mechanical energy received from the outside by the piston is stored both as potential energy stored in the compressed gas reservoir and in the thermal sponge as internal energy.
  • the energy stored in the sponge is equivalent to the potential energy stored in a reservoir of appreciable dimensions, containing compressed gas at an appreciable pressure.
  • Figures 4 - 20 show some procedures for mounting a solid thermal sponge in solid piston isothermalizers (they can also use fixed amounts of liquid, used for lubrication, for cooling the sponge during compression and for the evacuation of the compressed gas remaining in the cylinder when the piston reaches TDC (having no role in gas compression). The procedures are identical for newly designed isothermalizers as well as for those obtained by modifying the configuration of prior art compressors.
  • Fig.3a The principle diagram of the isothermalizer equipped with a deformable thermal sponge is shown in Fig.3a. It consists of the housing 12.1 (composed of the cover with the check valves 12.12 and 12.13 and the side walls), the piston 12.2 and the thermal sponge 12.6.
  • Check valves one-way valves are only one of the solutions for the intake and discharge of working gases.
  • the complexity of the installation requires that, in many configurations, the circulation of gases, lubricants and cooling fluids be directed by a greater variety of taps, dampers, mechanically or electrically operated valves, etc., but in most drawings these will be represented only schematically, by the same symbol.
  • the operation of the densifier is the same as that of a sponge-free compressor: the gas suction is via the inlet valve, by moving the piston from top dead center TDC to the bottom dead center BDC, with the exhaust valve closed and compression to the desired pressure p f , by moving it from the bottom dead center BDC to the T point, with both valves closed, during which time the heat transfer from the gas to the sponge takes place.
  • the exhaust valve opens so that the gas with pressure p f is exhausted to the desired destination, by moving the piston from T point to top dead center TDC.
  • the isothermalizer in the figure is cylindrical, and the thermal sponge will have the same geometric shape, but with a smaller diameter (calculated so that when its compression is maximum, its outer diameter reaches almost equal to the inner diameter of the cylinder of the isothermalizer, and with a height smaller (to achieve, in case of compression, a first polytropic phase and to allow an easier and faster movement of the piston at the beginning of the compression phase).
  • the sponge is compressed to the non-destructive limit allowed, and a small volume of gas remains in its alveoli (the dead volume), gas that expands, before the intake valve opens, when the piston moves to BDC.
  • the cooling of the thermal sponge is done intermittently, when, after a fixed number of cycles, the measured temperature of the housing reaches a predetermined value, by replacing the gas in the enclosure with a cooling liquid (the ratio of the number of gas/air cycles can reach 1 :1000).
  • a continuous cooling can be done, at the expense of the total volume of admitted gas, by introducing into the cylinder a quantity of liquid, the volume of which must be equal to the volume of the gas with the final pressure, liquid that needs to be recirculated and cooled, by any of the processes mentioned in the invention.
  • the one-piece elastic sponge can be replaced by a batch of elastic elements of much smaller sizes (the cells formed between these elements always communicate with each other) inserted in an elastic deformable bag, made of materials with high tear resistance, having a series of holes of reduced diameter, through which the gas inside communicates with the outside (the deformable bag can be replaced with a network with very small meshes).
  • Fig.4 and Fig.4A show a configuration of this isothermalizer, in which the cooling is permanent. Compared to the previous configuration, it additionally contains an additional thermal sponge 12.24, which may be non-deformable, made of a metal mesh (or foam), fixed to the cover of the enclosure, and at its lower part, a grill 12.25, or a small mesh net.
  • the lower, deformable sponge (up to the limit where the volume of gas in the alveoli is very small) is provided with a series of vertical channels 12.27, with a small diameter, the gas in these channels being in communication with the neighboring alveoli of the sponge. In the upper part, these channels are closed by covers 12.26, which open and close simultaneously with the inlet valve 12.12.
  • These channels can be helical elastic springs, hollow inside, made of a harder material, the sections of the coils being circular or ellipsoidal, embedded in the structure of the sponge since the manufacturing phase.
  • the height of the springs is equal to the height of the thermal sponge, and in the maximum compression phase (in which the coil sections are flattened ellipses) they have a height equal to that of the fully compressed sponge.
  • the walls of the helical channel thus made are also provided with gas communication perforations.
  • the axis of the cylinder is vertical, and a fixed quantity of liquid 12.22 is arranged on the upper face of the piston.
  • the compressed gas in the alveoli, as well as that in the peripheral zone and that in the vertical channels (forced into the neighboring alveoli) is moved to the upper sponge, with a large absorption surface.
  • the piston reaches the T position, the gas reaches the desired pressure.
  • the alveoli of the lower sponge may contain a mixture of gas and liquid, or they may contain only liquid, and part of the liquid may pass through the grate into the upper sponge.
  • the discharge valve is opened, the enclosure is flooded, from the constant pressure tank (open system) with liquid of the same pressure, all the amount of gas in the cylinder, regardless of its location, being directed to the storage tank under constant pressure. In this configuration, the dead volume is zero.
  • the piston continues its stroke until TDC (Fig. 4A), the position in which the lower sponge reaches the maximum degree of compression allowed and pushes the amount of liquid remaining in the sponge towards the discharge pipe.
  • TDC Fig. 4A
  • the inlet valve and the covers of the vertical channels are opened, all the liquid from the upper part drains to the lower one, and the working gas enters the enclosure.
  • the non-deformable thermal sponge is reduced to a ring with the outer diameter equal to the inner diameter of the cylinder and with the inner diameter equal to the outer diameter of the deformable sponge, the intermediate grill being no longer necessary but, to regulate the degree of compression of the elastic sponge, on the underside of the upper cover of the cylinder, between it and the upper surface of the deformable sponge, a plate with the same horizontal section as that of the sponge is mounted.
  • the piston moves towards the TDC position, which is at the lower limit of the ring represented by the non-deformable sponge, the liquid on its upper face is progressively absorbed by the deformable sponge, so that, in this position, all the amount of gas and liquid is in the non-deformable sponge.
  • the opening of the discharge valve leads to the flooding of the annular sponge by the liquid from the pipe connected to the storage tank (under constant pressure), the compressed gas being pumped into the tank.
  • the movement of the piston causes the entire amount of liquid to flow out on the upper face, where an amount equal to that coming from the tank is removed.
  • One of the possibilities to eliminate this quantity is to create a small reservoir in the side wall, immediately above the upper face of the piston, with a volume equal to that of the gas being pumped, a quantity that remains here, until the side surface of the piston passes this cavity, after which it drains into the crankcase.
  • the liquid used for cooling must have a low adhesion to the material of the thermal sponge, for which purpose a series of additives can be added to the liquid and/or to the structure of the thermal sponge.
  • a series of additives can be added to the liquid and/or to the structure of the thermal sponge.
  • the coolant has a high specific gravity (e.g. mercury), has the highest surface heat transfer coefficient, the highest thermal conductivity, and the contact angle between the liquid and the material used to make the sponge should be less than 90°.
  • the isothermalizer in Fig.5 illustrates a configuration in which: the thermal sponge 5.4 of this isothermalizer is a helical elastic metal spring with the rectangular turn section, one end of the spring being fixed to cover 5.1 , the other to piston 5.2.
  • the spring 5.4 With the piston in the BDC , the spring 5.4 is in a detensioned state (or slightly pretensioned). In the illustration, the piston is in an intermediate position.
  • the dead volume of the isothermalizer which includes a central cylindrical space with a diameter equal to the inside diameter of the spring) is quite large.
  • FIG.6 Another constructive variant is represented by the isothermalizer in Fig.6 and the one in Fig.7, where the main components of the thermal sponge are horizontal parallel plates (preferably metal) 5.1 1.
  • These plates shall be attached to a vertical elastic rope system 5.7a, inserted into the elastic bellows 5.7b, respectively to a vertically mounted helical spring system 5.12, with an outer diameter smaller than that of the isothermalizer springs in Fig.5 (where they were the main component of the thermal sponge).
  • the space inside the elastic helical springs 5.12 (circular, rectangular, etc.
  • a guide rod system 5.7c attached to the piston is proposed, which pierce the cover through holes fitted with sealing gaskets 5.8 (or vice versa).
  • tubes are installed instead of rods, through which cooling fluids circulate, the problem of removing the excess heat can be solved completely or partially.
  • These tubes must be provided also with sprinklers to spray droplets of coolant, or aqueous foam.
  • a process that achieves the same objective is the use of thermal tubes with vapor at the saturation limit. These rods also eliminate the possibility of side movements of the thermal sponge and the possibility of it coming into contact with the cylinder walls.
  • the windows that provide the largest absorbent surface of the board are circular, with a diameter equal to the maximum diameter of the support spring.
  • Fig. 7a the fastening system is illustrated when the springs are located at the edge of the cylinder, and in Fig. 7b
  • the number of plates mounted in these systems can be very large, thus ensuring a large heat transfer surface and, consequently, a high piston speed, or a very small difference between the temperature of the gas and that of the thermal sponge.
  • the horizontal plates can be very close together, even touching each other (Fig. 7c). Even when the horizontal plates come into contact with each other, many spaces remain filled
  • FIG. 8 shows an isothermalizer configuration in which, starting from the configuration in Figure 7, the surface area through which heat absorption, by the thermal sponge, from the gas being compressed is considerably increased by the installation of fins, or other vertical elements 5.10.
  • the density, the arrangement on the support plate, the thickness of the plates, etc. may differ from one horizontal plate to another. There is a very high degree of freedom in the shape of these fins, their dimensions (a large thickness ensures a slower increase in the temperature of the sponge, a smaller distance between the fins ensures better cooling of the gas, a larger width of these reduces the required
  • Configurations may be made in which vertical fins are walls separating laterally distinct areas of the cylinder (in a horizontal section, they are a sequence of concentric circles, or rectangles with increasingly smaller sides, or other geometric figures placed in each other, or side by side). In this case,
  • the horizontal plates have the raised edges along the entire contour (5.1 1c, figure 8), like trays. This allows a liquid layer to accumulate permanently or periodically on the horizontal plates (5.11 d, figure 8b). The height of the tray edges will determine the size of the liquid fraction of the sponge 5.1 1 a.
  • the vertical movement of the piston causes the vertical fins to move, penetrate the liquid layers on the horizontal plates and discharge liquid into the liquid layer formed on the top surface of the piston. In this way, the
  • the thermal sponge is made without the elastic springs between the horizontal plates. In this way, the action of the piston is transmitted from one horizontal plate to another successively, not simultaneously.
  • mechanical or electromechanical locking-unlocking mechanisms of horizontal plate movement can be installed in the
  • the isothermalizers in Fig.9 are isothermalizers provided with horizontal plates 5.11 , fixed on the helical springs 5.12, inside which the springs 5.13 are mounted, with progressively decreasing diameters, and the space inside the spring with the smallest diameter may be partially occupied by a full cylinder.
  • Fig. 9A The isothermalizer in Fig. 9A is similar, but the fixing of the horizontal plates is done on deformable supports of the bellows type (harmonica) 5.22, 5.24.
  • Fig.10 shows a configuration of an isothermalizer whose thermal sponge is made with telescopic
  • FIG. 10a an embodiment of the vertical telescopic fins is exemplified.
  • Fig. A is represented, in an intermediate position of the compression phase, a vertical section (1 -1 ) through one of these fins for the case where the horizontal section through the "cylinder" of the densifier is rectangular.
  • Fig.C is represented a vertical section through an intermediate section of the fin, and in Fig.B, a horizontal section through the fin. It is noted that each section 5.10 of the wing (minus the lower one, which is a simple plate
  • each section (except the first, where it is not necessary) has in the upper part some lateral talons 5.10b, which slide through a series of corresponding channels 5.10a, practiced in the inner walls of the upper section, having a length equal to the height of the section, less the stopper
  • Fig.11 shows a diaphragm densifier consisting of the upper housing 5.1 , the lower housing 5.1 b and the elastic
  • the densifier is operated directly by the piston 5.2, but can also be operated by means of a volume of hydraulic oil, in which case the housing 5.1 b has perforations for the oil circulation.
  • the shape of the two housings is modified, the enclosure between them having a shape close to that of a rectangular parallelepiped, with "softening" the edges, which allows aspiration of a larger volume of gas and offers more choice for the type of thermal sponge.
  • Fig.1 1 we chose a
  • the thermal sponge composed mainly of flat metal plates 5.1 1 , supported on supports mounted on the carry-supports of harmonic type, composed of flashplates 5.22, 5.23 and 5.24.
  • the thermal sponge can also have a permanent liquid component, with the role of avoiding the formation of a dead volume, and an itinerant liquid component, that with the help of 6.9b sprinklers, cools the gas, subject to compression.
  • This component can also be used as a liquid piston, with flow rate adjusted in such a way as to obtain an110 isothermal speed for the compression.
  • the diaphragm 5.33 is mounted between two metal plates, 5.30 and 5.32, and is rigidly fixed to the two plates, by means of the plates 5.34, along a median axis, the outer edges of the the diaphragm being rigidly fixed between the two halves of the housing.
  • the free part of the elastic diaphragm is extensible under the action of the piston,115 and the part between the two plates can slide on some rollers 5.31 .
  • the isothermalizers in Fig. 12, Fig.13 and Fig.14 also have in their composition thermal sponges made of elastic and inelastic metallic components which, when the piston is in TDC, occupy almost entirely the internal volume of the isothermalizer.
  • the one in Fig. 12 is constructed from horizontal plates 5.1 1 , which rest on a system of lamellas 5.14a and 5.14b or elastic half-plates 5.14. When fully tensioned,120 two paired half-plates 5.14 almost entirely cover the surface of a horizontal plate, as does the group of plates 5.14a and 5.14b installed between two horizontal plates.
  • Fig.12A it is noted that part of the blades are fixed (for example, by welding 5.14c) to the horizontal plate below it at the periphery of the plate, while the other blades, having reverse curves, are arranged alternately with the first blades and are fixed near the midline of the plate.
  • half-plates they can be welded in any of the two listed125 positions.
  • the piston is at TDC, the space not occupied by the thermal sponge of the isothermalizer enclosure, in the case of precision machining of plates, lamellae and halfplates can be made very small. This space is occupied by the dead volume of the compressed gas. or by the coolant 12.22 arranged on the upper face of the piston.
  • the one in Fig. 13 is made by alternating flat plates 5.1 1 , which slide on an equal number of130 arched plates 5.14, so dimensioned that when the piston is at top dead center, all these plates overlap one another and occupy as large a part as possible from the cylinder of the device.
  • the whole assembly is stabilized by a rod 5.7c, which has one end fixed to the cylinder cover, and the other end pierces the piston through a hole made in the piston and sealed with the gasket 5.8.
  • the rod 5.7c in all configurations in which it is used (eg Fig.7, Fig.9) has a non-circular section and is located in the axis of135 the cylinder, preventing the rotation of the stabilized plates.
  • the cover of the device is provided with a series of holes distributed on the contour of the cover, holes to which the supply pipes with lubricant 5.2c are connected, and the sponge plates are provided with a series of holes 5.1 1 o and 5.14o, distributed in such140 a way that the oil is distributed to all regions where friction occurs.
  • the lubricating liquid drains on the plates and accumulates on the upper face of the piston, from where it is evacuated with the help of a pump, through the pipes 5.2c and is sent to the cooling bath.
  • the compressed gas is collected in the space between the horizontal plates, as well as in an inner parallelepiped space 5.2b.
  • the dimensions of the collection space are set by the width of the arched strips 5.15, and the height of the inner parallelepiped collection space is adjusted by the dimensions of the piece 5.2a, fixed on the moving piston.
  • section 1 -1 an inner, top view of the system is shown.
  • the isothermalizer in Fig.14a is similar to that of Fig.13, but arched plates with a different number of curves with different radii of curvature, are mounted between the horizontal plates. As in previous configurations, the highest absorption power is obtained when all the plates have the same surface, close to the section surface through the cylinder, and in the TDC they perfectly overlap.
  • This type of thermal sponge can be used to reduce the energy consumption of state-of-the-art160 compressors having a superunitary polytropic index, compressors for which the main objective is not to achieve an isothermal compression, but to obtain a large volume of compressed gas in as short time as possible.
  • This objective can be achieved in a more economical way than at the state of the art (where the desired compression ratio is obtained by staged compression, intercaling some heat exchangers between these stages), by inserting a thermal sponge with a maximum absorption surface into the compressor,165 obtained with heat-accumulator elements having a minimum volume, associated with a continuous flow lubrication system, which also takes over the sponge cooling function and reduces the dead volume as much as possible when the piston is in the TDC.
  • a piston actuator system which (at a compression cycle time equal to that of a conventional compressor) introduces a variable piston speed, higher in the exhaust, in the suction and in the first part of the active piston stroke, and170 smaller toward the end of the compression process, it further reduces energy consumption and also makes the cooling system more efficient.
  • the energy consumption necessary to obtain a given compression ratio may be reduced, in a given time, if a properly sized thermal sponge is inserted inside its cylinder. Most of the time, this involves some constructive changes to175 the original compressor (for example, shortening the stroke of the piston, or lengthening the useful part of the cylinder, with a G b value, equal to the thickness of the sponge in a fully compressed state, as well as the adequacy of the lubrication system to the new requirements). In the conditions of the evolution of energy prices and the objectives of reducing thermal pollution and noxious pollution, the expenses necessary for these adaptations will be rewarded. 1 180 For isothermalizers can also be made configurations without elastic components.
  • the isothermalizer in Figure 15 (horizontal section through a vertical cylinder with a rectangular section) consists of a thermal sponge made of metal plates 5.11 , made with a thickness as small as possible (if a high power of the isothermalizer is desired), but large enough for the plates not to be subjected (due to their own weight, or too sudden movements) to some residual deformations. To ensure that the plates
  • section 1 -1 is a vertical section through the cylinder, executed in the area where the carry-supports are mounted.
  • 190 support is made in the form of blades, or narrow rods, on whose inner side (facing the inside of the cylinder) are mounted, (by welding, riveting, embossing, etc.) supports 5.20 of the plates 5.1 1 , made of sheet metal, wire, pieces processed by machining, etc.
  • On each port-support is mounted, at different levels (usually equally spaced), a number of supports equal to the total number of plates, or equal to the number of plates in a set, if the interlaced plate technique is used.
  • One end of the carry-support shall be
  • a fishplate 5.18 rigidly attached to the piston.
  • a short swivel arm is attached, also via a movable joint, which has a guide roller 5.16 attached, which can run on a rail, or in a channel 5.17 of the cylinder cover.
  • the horizontal plates 5.11 are rectangular, occupying almost the entire horizontal section area, but they have practiced in the corners a series of cuts to avoid collision with the carry-supports and
  • 1200 supports on the neighboring levels, as well as to create the consoles 5.21 that are laying on the supports on that level.
  • the carry-supports make the minimum angle (almost 0°) with the vertical axis, and the distance between the plates is maximum.
  • the angle made by the longitudinal axis of carry-supports whith the vertical axis increases, and the distance between the
  • the carry-supports make the maximum angle (almost 90°) with the vertical axis, and the distance between the plates is minimal.
  • the plates can perfectly overlap without intermediate spaces, ensuring a small dead volume and easy circulation for the fluid intended to replace this gas.
  • Fig.16 shows a horizontal section through the cylinder of an isothermalizer which also consists of
  • the carrysupports are made of a sequence of pairs of fishplates 5.23 and 5.24, placed superimposed in the same vertical plane. Both fishplates of these pair have a centrally located hole through which a pin passes, around which both fishplates can rotate. At the same time, this pin, having a corresponding length, can be the support for one of the horizontal plates 5.11. In another configuration, the length of the pin is
  • the shape and dimensions of the thermal sponge components must lead to the achieving of the most regular shapes of the volumes in which the compressed gas is found, when the piston is in position T and a minimum dead volume when the piston is in the TDC position.
  • the gas is now exhausted by moving the piston from position T to the TDC position, where the dead volume reaches the lowest value.
  • a total elimination of the dead volume can be achieved, as in the principle diagram in Fig.17a, by introducing into cylinder 5.1 a, as early as the initial phase, a liquid phase 5.3b of the 5.3a thermal sponge, consisting of an appropriate amount of lubricant, or a heat transfer liquid, the volume of which equals the dead volume.
  • a liquid phase 5.3b of the 5.3a thermal sponge consisting of an appropriate amount of lubricant, or a heat transfer liquid, the volume of which equals the dead volume.
  • the 5.5a output valve opens (due to the valve adjustment, or due to a command received from the control system), located at the
  • FIG. 17b Another possible configuration for the exhaust of compressed gas is shown in Figure 17b, in which 5.1 a is a small densifier, whose inlet window 5.6a is at the same time the discharge window for a
  • the compression conditions in the two densifiers being different, will be different throughout the compression period, also the temperatures of the gas and of thermal sponges they contain.
  • the liquid fraction of the sponge in the densifier 5.1 can be chosen so that when piston 5.2 reaches the TDC , it completely occupies the volume of the cylinder not occupied by the solid fraction, without entering in cylinder 5.1 a at all. At this point, all
  • the initial volume of gas in the two cylinders is transferred to cylinder 5.1 a, and its pressure reaches the final p f value.
  • the gas can be subjected to a new isothermal compression stage, or it can be exhausted into the storage tank by moving the 5.2a piston from the BDC to the TDC.
  • the thermal sponge of the densifier 5.1 a shall be carried out only with a solid fraction, so that the dead volume is as low as possible.
  • the gas can be subjected to a new isothermal compression stage, or it can be exhausted into the storage tank by moving the 5.2a piston from the BDC to the TDC .
  • the thermal sponge of the densifier 5.1 a shall be carried out only with a solid fraction, so that the dead volume is as low as possible.
  • the window 5.6r and the pipe 5.6c which connects directly to the storage tank located at a higher elevation (or with another useful destination) and which contains, at its base, liquid from associated hydraulic circuit, with pressure p f .
  • the 5.6a window serves only for gas suction. In this case, when the piston, in his movement to the TDC reaches the point T, opening of the window 5.6r allows liquid from the 5.6c pipe to enter the cylinder and replace the compressed gas at the
  • the window width may be equal to the thickness of the sponge when the piston is in the TDC, and its length can be equal to the width of the wall), which allows for rapid circulation with reduced losses of exergy.
  • Fig.18A is shown such a process applied to the isothermalizer in Fig .5.
  • the thermal sponge of this isothermalizer is composed of a solid fraction (a rectangular section helical spring) and a liquid fraction that completely eliminates the dead volume of the cylinder when the piston is in the TDC. Gas suction and exhaust are
  • the temperature of the gas (and implicitly, the mechanical work required to compress it during a cycle) and of the thermal sponge increase progressively with the number N of compression cycles. The heat the gas receives is accumulated, along with the mechanical energy received, in the compressed gas storage
  • the densifier sponge can be extracted entirely from the densifier, stored in an isolated enclosure and replaced with an identical sponge having the temperature of T amb . This is possible, for example, if the cylinder has a rectangular section, the 5.6c side windows and the caps closing them have the width equal to the compressed sponge and the length equal to the side wall and if, at the
  • the side caps 5.6c and the edge plates of the sponge 5.4a shall be mechanically coupled together so that they can be translated, sliding on the surface of the piston then on the outer rails (for example, pushed by a 5.2d piston, or by towing).
  • Another way to extract the “overheated” sponge is to extract the piston fully, coupled with the sponge, through the casing of the device. After extraction, the piston-sponge assembly is cooled in a fast
  • thermal sponge cooling is one of the most important problems of building high-performance densifiers.
  • the thermal energy absorbed by the thermal sponge can be eliminated, depending on the characteristics of the densifier, by any of the prior art
  • a lubrication system combined with a lubricant cooling system leads, after a transitional regime, as in polytropic compressors, to a stabilization of the sponge temperature at a T b value, higher or lower depending on the coolant flow and the coolant temperature.
  • the temperature value of T fc can be maintained at a low value if, e.g., the system operation is supervised by an controller which, at predetermined intervals (or dictated by a
  • this agent can be included in a cooling circuit, continuously or intermittently and can also take over the cooling function of the thermal sponge.
  • the elimination of excess heat is done by replacing the compressed gas with colder liquid during the exhaust operation (the warmer liquid being subject to upward forces).
  • the stopping time of the liquid agent in the compressor can be prolonged (periodically or every cycle) by the commands sent to the piston by the controller.
  • Another possibility of replacing this fraction, increasing the flow rate of the gas, is to remove the remaining fluid in the cylinder, during the fresh gas suction
  • a tubular helical spring can be mounted, coupled by flexible tubes to an external cooling circuit with heat exchanger, through which, under the effect of a hydraulic pump, a coolant circulates.
  • the hollow spring is fitted with horizontal spray nozzles. It introduces suspended particles between the metal plates, which considerably increases the cooling speed of the gas and of the thermal sponge. Excess fluid accumulated above the piston is eliminated in continuous flow. This system can also be applied to densifiers in Fig.6,
  • this type of isothermalizer can be cooled by inserting the coolant through the top of the apparatus, at a pressure equal to that of the apparatus.
  • This type of densifier is very suitable for cooling with aqueous foam.
  • Foam regeneration can be done by introducing at a pressure slightly higher than instantaneous gas pressure in trays 5.11 c.
  • the additional coolant introduction system may be mounted on the skeleton supporting the horizontal plate system, inside it, or on similar independent structures.
  • Fig.19 shows an isothermalizer whose cylinder consists of two segments with different diameters.
  • deformable sponge is mounted, whose volume in a fully compressed state is only a fraction smaller than the volume of the cylinder corresponding to this position of the piston.
  • the total volume, the configuration of this thermal sponge and its total surface are chosen according to the speed of the piston, the desired efficiency of the device and the compression ratio.
  • a thermal sponge composed of horizontal metal plates 5.1 1 is represented, which rests on the large helical springs 5.12 and
  • the isothermalizer contains one or more vertical bellows
  • each bellows there is a spring 12.30, whose length, in an untensioned state, is equal to the height of the enclosure when the piston is in the BDC position, and in a
  • the inner spring keeps the liquid bellows in a tensioned state, by means of an unlocking mechanism.
  • the mechanism unlocks, the height of the bellows suddenly reaches its maximum length, and as a result, the gas in the enclosure is suddenly compressed, almost adiabatically.
  • Valve 12.12a being open, the heat transfer liquid fills the
  • the pressure of the liquid is maintained above the pressure in the enclosure, so that it also enters the isothermalizer enclosure, being distributed by the sprinklers in the spaces between the horizontal plates.
  • the flow rate of the agent is regulated by the main regulator, so that, together with the displacement of the piston, it ensures the isothermal trajectory. If the liquid acts only as a heat transfer agent, the height of the liquid layer accumulated on the upper face of the piston is
  • the measurement of the gas pressure at the end of the intake phase indicates exactly (regardless of the temperature distribution in the working gas and in its ambient environment), if its average temperature is equal to that prescribed by the optimal trajectory and, consequently, if it needs to be cooled or additionally heated, as well as how much the piston must be accelerated to reach the optimal isothermal speed.
  • the isothermalizers are also equipped with a cooling/heating system, which introduces in their enclosure, during compression, through devices provided for this purpose (for example, through sprinklers), a flow of heat transfer liquid, which is evacuated through another pipe and transported, with the help of a pump, to a heat exchanger. If the flow rate of liquid discharged is equal to that supplied, the liquid agent is only the heat transfer liquid, but in
  • the piston compressor is a liquid agent supplied by a hydraulic motor, the liquid having remarkable properties: lubrication and sealing, minimile of dead volume and a two-way transmission agent of heat and mechanical energy.
  • the 1475 disadvantage of these systems is that, if thermal sponges made up of solid elements only are inserted into the cylinder, at some point they will be covered by the liquid whose level increases, so the heatabsorbing surface of the sponge shrinks (just when its temperature rises faster).
  • the plates are arranged in such a way that after the introduction of liquid into the cylinder, all the plates are covered with a layer of liquid, this becoming a liquid component of the thermal sponge, as it takes over from the upper surface of the solid plate, the role of absorbing the energy excess heat.
  • low-height stoppers can be mounted to the periphery of the horizontal plates.
  • the isothermalizer is constructed like a prior art liquid compressor, with a cylindrical casing 7.1 and a liquid piston 7I actuated by the solid piston 7.5, actuated, in turn, by a speed control system, the imperative of which is to provide the isothermal speed v iz
  • the solid piston is used in combination with an electric, variable speed drive assembly. In the case of a hydraulic drive with variable liquid flow, cylinder 7.2 is connected directly to the liquid pipe.
  • the thermal sponge has a solid and a liquid component.
  • the solid component consists of the horizontal plates 7.3a provided, on most of the circular sector, with the peripheral skirts 7.3b, intended for the division of the liquid piston.
  • the liquid component consists of the liquid layers that form on each of these solid horizontal plates.
  • 1500 sponge has mainly the role of a piston and acts simultaneously in the N elementary compressors formed by splitting the main compressor.
  • the gas contained in each of these elementary compressors yields heat, mainly to two circular surfaces 7.3a, which have a diameter almost equal to the diameter of the master cylinder. If the horizontal plates 7.3a were missing, cylinder 7.1 , cap 7.2 and the liquid piston would constitute a liquid piston compressor, with an initial volume approximately
  • Fig .21 shows a cross section of a cylindrical densifier (in many applications a rectangular section is more
  • Fig.21A shows a horizontal flat section of it, at the level of an elementary compressor.
  • the horizontal plates 7.3 and 7.4 separate the compressor from the constant pressure tank 7g and the liquid piston 7I, respectively.
  • the liquid piston consists of a fixed volume of liquid agent (the same type of liquid as the one in the tank 7g), equal to the free volume of the compressor (the volume of gas in the cylinder, immediately after the suction phase).
  • the first phase the
  • densifier absorbs a volume of gas through the valve 7a (located in the upper elementary compressor), when the piston 7.5 moves from the TDC to the BDC. At the same time, an identical volume of liquid, located in the densifier, is transferred into the tank 7L.
  • the compression phase follows, in which, after closing the suction valve, the liquid from the tank 7L enters the densifier cylinder and, according to the law of the communicating vessels, is distributed into the N elementary compressors.
  • Opening windows 7.6a is done by moving a movable cap (piston 7.7) that is running tight (by means of seals) on wall 7.6 and is controlled by a differential pressure switch 7p, when the piston 7.5 is in position T and the pressure p f of the liquid in the densifier is equal to the pressure in the tank 7g.
  • the gas pressure in each elementary compressor is equal to the pressure p f , to which is added the pressure given by the liquid column between the measuring point and the elevation of
  • piston 7.7 causes the entire amount of compressed gas in the densifier to be replaced by liquid agent in the tank 7g (in this way, the entire volume bordered by horizontal plates 7.3 and 7.4 is occupied by the liquid) and causes the level of the liquid in this tank to lower. If piston 7.5 continues to move to the TDC, the amount of fluid between the level T and the TDC level (equal to the total volume of compressed gas during a stroke of the piston) is
  • the tank 7g and the piston 7.7 may be missing, the wall 7.3 becomes the outer wall and the windows 7.6a are replaced, each of them with a check valve.
  • the tank 7s is replaced by a simple pipe in the wall of which valves are mounted to each elementary compressor, the lower end
  • the liquid and, indirectly, the plates in the densifier can be cooled by keeping the liquid in the
  • the densifier in Fig.22 is built on the same principle of overlapping a large number of elementary liquid piston compressors, made by interspersing their upper and lower walls, 7.3s and 7.3I, respectively. Compared to the previous configuration, two types of liquid piston mini-densifiers appear in this type of
  • 1565 densifier a mini-densifier 7c between the upper and lower walls, with higher height, cooled by liquid spray and a mini-densifier 7d between the lower and upper walls, with a lower height, without spraying.
  • the liquid piston is inserted into the elementary compressors 7d directly, through the windows 7.3f.
  • a vertical skirt 7.3g is inserted to separate a layer of gas into each densifier.
  • the introduction of the liquid piston into the elementary compressors 7c is done by a distributor of agent 7.1 1 , from which the liquid agent is
  • a horizontal wall 7.3 separates the densifier area from the tank 7g, which is in direct communication with a tank 7s located, this time, in the center of the densifier, having a cylindrical shape and being separated from the densifier by the cylindrical wall 7.6, in which the windows 7.6a are executed, on each basic compressor. To each of these windows corresponds a similar window, located at the same level, in the
  • a sector 7.8 in the side wall, with the height equal to that of the densifier, has a series of windows 7.8a practiced at each mini- densifier. During the compression operation, these windows are closed by a piston 7.8b of the appropriate size and shape, fitted with suitable seals and a horizontal displacement system. Through these windows the liquid agent is removed from the elementary compressors (after the compressed gas
  • a movable flat ring 7.3s 1595 a movable flat ring 7.3s, with outer diameter greater than the inner diameter of the peripheral ring 7.3sa and inner diameter smaller than the outer diameter of the inner peripheral ring 7.3sb.
  • the movable ringsections shall all be fixed on one or more rods 7.9 in a position below the corresponding plate 7.3s in such a way that, through the seals mounted on the edges of the upper surface, air and liquid are not allowed to flow to the lower compressor when the rods 7.9 are in the “closed” position.
  • the liquid agent is inserted through a gate 7a
  • the fluid pressure at the periphery of the master cylinder is equal to that of the gas and liquid in the densifiers 7d, and that of the gas and liquid in the densifiers 7d is slightly lower, this causing the valves 7.10 and 7.1 1 to open, depending on their adjustment. If the flow of liquid entering the sprinklers is lower than that introduced through the windows 7a, the gas pressure increases, which also causes the 1620 gas inlet from the densifiers 7d to open into the densifiers 7c, causing the gas pressure in these densifiers to increase. When passing through the liquid layer, this gas undergoes additional cooling.
  • the inlet valve is a sealing plug directed by a pre-loaded spring, and the sprinkler can be a simple disc-shaped cap, with a horizontal spray holes 7.11 on the side. This valve opens when the difference between the dispenser fluid
  • all the mini-densifiers of type 7d transfer the compressed gas to the mini-densifiers of type 7c ,
  • the compressed gas is evacuated to the tank 7g.
  • liquid piston densifiers can be made, with one or more solid-piston densifiers as the main subassembly. Any of the solid piston densifiers described above, or made on the same constructive principles, may be used. In these configurations, the functionality of the system and its
  • the solid piston densifier becomes a liquid piston densifier.
  • Fig.23 shows the densifier in Fig.16, in which vertical perforations 5. 11 o are made in the components of the solid thermal sponge (horizontal metal plates 5.1 1 ), in such a way that when the plates overlap under compression, these holes also overlap and form continuous channels.
  • the solid thermal sponge horizontal metal plates 5.1 1
  • thermal sponge 5gs is inserted in the upper part of the densifier, made of metal foam, metallic fabrics, other metal inserts with a large absorption surface (in Fig.23 it is made of woven blankets made of metallic wire, superimposed, without separation intervals, mounted on a horizontal support system, of bars, rods, perforated plates, etc.), which absorbs a large amount of heat energy in all phases of compression.
  • the liquid piston penetrates into the holes and grooves executed in the horizontal plates, as
  • Fig.24 shows a constructive variant of the liquid piston isothermalizer with vertical telescopic fins. It is composed of a series of identical elementary isothermalizers, separated from each other by horizontal walls 7.17, and from the enclosure for introducing the active liquid, by a common vertical wall
  • each elementary isothermalizer 1665 7.18, provided for each elementary isothermalizer with a slot, located at its base.
  • each elementary isothermalizer there is a horizontal plate 5.1 1 , with a slightly smaller surface area than the lower wall, which rests on a plate 7.21 , at some distance from the base of the mini-isothermalizer (because in the first phase of compression, the thermal energy to be absorbed is less). All these plates are fixed on a vertical rod 7.20 which penetrates all the base plates and its lower end is fixed to the displacement
  • the rod 7.20 on which the plates 7.21 are fixed passes through the lower wall of the enclosure, and its end is connected to an actuation mechanism, with the role of simultaneously moving the inner plates and the horizontal plates with which they are in contact and with this to perform the compression of all telescopic sections.
  • the method of driving the horizontal plates from the outside can be replaced by a method by which the horizontal plates are moved by the liquid piston itself.
  • a method by which the horizontal plates are moved by the liquid piston itself instead of the plates and the rod used in the previously described configuration, to move the fins of the telescopic sections,
  • 1710 enclosures communicate with each other through flexible pipes, and one of them is connected to an external isothermalizer, which introduces compressed gas into these containers and also compresses the gas inside the containers, together with which it forms an gas piston isothermalizer, independent from the one with a liquid piston, having different inlet and discharge pressures. Also, the pumping of the compressed gas is done independently, by replacing it with liquid from its own storage tank.
  • the gas isothermalizer
  • piston acts in tandem with the liquid piston, in such a way that the intake and discharge of the gas from the containers is done almost simultaneously with that of the gas in the compartments of the liquid piston isothermalizer. If the penetration of liquid into the gas piston isothermal vessels for discharge occurs before the time when the pressure in the liquid piston isothermal reaches the final pressure, the liquid displaced by sinking, as an effect of the additional weight, of the gas piston isothermal vessels , acts like
  • FIG.24C A similar isothermalizer is shown in Fig.24C, in which the compression of the fins is done by the solid piston (which, at the same time, is its upper wall), of an elementary compressor mounted on the lower wall of the main elementary compressor.
  • the liquid piston, entering the cylinder 7.2, is distributed through the slots 7.18f in each elementary isothermalizer 7.18i and compresses the gas inside.
  • the isothermalizer 7.18s is a solid piston isothermalizer.
  • the gas here is compressed, simultaneously with the displacement of the piston 7.18p and the compression of the telescopic fins 5.10. Compression in this isothermalizer ceases when the piston reaches its maximum height, corresponding to full compression of
  • the fins at a gas pressure lower than that in the lower isothermalizer.
  • the liquid piston continues to enter the isothermalizer until the gas pressure reaches the desired value, that of the gas in the compressed gas tank 7r, under constant pressure, in which, along with the gas 7g, the liquid 7I is found.
  • the valves 7.22.2 on the vertical pipes 7.23 are opened successively (due to the differences between the hydrostatic pressures of the isothermalizers, located at different
  • valves 7.18v are opened and the liquid from the main column enters each of the isothermals 7.18s, isothermals that become liquid piston.
  • the compression also continues in these isothermalizers, until the pressure in the tank is reached, when the valves 7.22.1 are opened, through which the gas is discharged and the entire device is filled with liquid.
  • the gas inlet valve 12.13 is opened and the way to empty the liquid from the enclosure is opened.
  • the isothermalizer in Fig.25 is a double-effect liquid piston isothermalizer resulting from joining
  • the isothermalizer in the figure is composed of two identical liquid piston isothermalizers, with a parallelepiped body 7.1 , the fixed vertical walls 7.6 provided with gas outlet holes for each elementary isothermalizer 7c, the movable vertical walls 7.7, attached to them but which can slide laterally, provided with similar exhaust holes, located in such a
  • a solid thermal sponge formed by the horizontal plates 7.3a provided to the outside with the peripheral skirts 7.3a, intended to divide the liquid piston, and towards the inside, in contact to the vertical wall 6.6, with skirts oriented towards the upper part.
  • the liquid piston 7I is actuated by the hydraulic motor 7p, with variable flow due to the actuation of a regulating flap, commanded by a regulation system, whose imperative is to provide the liquid piston with
  • Composite systems may also be made by simultaneously or successively accumulating the effects of a liquid piston and a set of solid pistons.
  • Such configurations are described in Fig.26, Fig.26a and Fig.26b and can have as a model any of the solid piston densifiers described above, if their volume
  • the densifiers 7g in Fig.26 are equipped with a thermal sponge composed of elastic plates 7.3, mounted on a support 7.3s. They are placed in hermetic bags 7.14, made of elastic materials, or other tear-resistant materials, with high heat transfer coefficient, but slightly deformable, even at low pressures.These bags are fastened to a metal plate 7.2s fitted with holes 7.2o, that make the connection between the gas in these densifiers to the gas layer 7gs, located above the
  • each bag 7g is reduced, with the advance of the liquid piston, according to the reduction in the volume of the thermal sponge it contains and may reach a minimum value when the deformation of the sponge is naturally blocked (when all the inner plates overlap), or by externally controlled devices.
  • the so-created enclosures communicate with the top gas layer 7gs in the compressor
  • valves 7.2ps can be separated from it by the valves 7.2ps.
  • the metal plate 7.2s separate, in this way, all the gaseous regions 7s containing thermal sponges with large absorbency area, from the rest of the enclosure 7i.
  • various pipes that create additional communication paths can be inserted into the enclosure, such as pipe 7.16, which makes the equalization of pressures between layers 7i and 7s by means of an intermediate liquid piston.
  • the device is equipped with a series of valves for the
  • valves 7.2a for the initial gas intake, valves 7.2e for the discharge of compressed gas, valves 7a for introducing the liquid piston into the enclosure, for the circulation of the cooling liquid (same as the piston liquid) and for liquid evacuation, simultaneously with the entry of gas into the lower chamber 7i, valves 7.2i which control the entry and exit of the liquid piston into densifiers 7g, valves 7.2ps which can block communication between densifiers 7g and the layer 7gs, valves 7.2 pm
  • the liquid piston enters the enclosure 7i.
  • the initial gas pressures be as high as possible.
  • the volume of the inflatable bags has a maximum value (a value as close as possible to the total volume of the lower enclosure 7i is recommended), and the elastic
  • the inflatable bags return to their rest form, the one with the elastic elements not tensioned, the mechanical energy previously accumulated in the elastic elements of the thermal sponge diminishing the mechanical energy required by the liquid piston to achieve this compression.
  • Part of the thermal energy produced by compressing the gas in the 1805 chamber 7i is given to the walls 7.14 (inflatable bags) and the walls of the chamber 7i, with heat transfer surfaces that decrease with the advance of the liquid piston and with the increase of the gas pressure in all compartments, and a part of the thermal energy resulting from the compression of the gas in the enclosures 7s and 7gs is transferred to the thermal sponges 7g and 7gs, respectively.
  • the gas pressure in the enclosures 7s can increase, at a rate dependent on the Young's modulus of the plates 7.3, until the
  • the mechanical energy stored in the elastic elements can be recovered during the pumping phase of the liquid from the enclosure 7i, by directing it to a hydraulic motor.
  • the thermal sponge of the densifier in Fig.26a is made of plates, or elastic metal strips 5.14,
  • Hermetic bags can have the form of mattresses, with the width I equal to one side of the device enclosure and the length equal to a multiple of the other side L. These mattresses are placed in overlapping layers, along the entire height
  • the mattresses communicate with each other through rigid tubes.
  • an appropriate number of mattresses are used, with a suitable thickness (depending on the type of thermal sponge) and with an area equal to the horizontal section of the cylinder, the mattresses communicating with each other, through more rigid tubes.
  • tubes of length I or L are used (both sizes arranged in alternating layers can be used). All these tubes (and mattresses) communicate with each other, through rigid, metallic or deformable tubes, forming a single enclosure.
  • the isothermalizer contains layers of cylindrical bags or rectangular mattresses, but the thermal sponge is made of elastic metal sheets rolled into more or less helical rolls. In all cases, although the bags can be
  • the bags are inflated, at an initial pressure which may be different from the atmospheric one.
  • the liquid is introduced and discharged into/out of the enclosures, through gates 7a, with hydraulic pumps, the ratio between the inlet and outlet flow being variable, but always unitary (the
  • the upper thermal sponge is a reticulated mesh 7gs, under which the lower thermal sponge is mounted,
  • the liquid piston (which is also the coolant and is part of a cooling circuit, equipped with a hydraulic pump and a heat exchanger), is introduced through the valve 7a from one end of the pipe, in the space between the walls of the pipe.
  • spray nozzles are mounted, both to the inside of the pipe and to its outside, which spray liquid in all areas occupied by gas, compressing it.
  • the gas is maintained by adjusting the ratio between the flow of liquid in and out. At some point, the entire lower part of the cylinder is occupied by liquid, the gas accumulating in layer 7gs. Compression continues with a low liquid flow until the final pressure p F is reached, when valve 7.12e opens, the liquid flow increases rapidly and the compressed gas is fully discharged to the tank/consumer. After compression, the liquid is discharged by suddenly withdrawing the lower cylinder head 7.4 and draining the liquid into a
  • the compressor in Fig.27a is a liquid piston densifier, without solid moving parts, with a cylinder 7.1 of circular section, in which the mass of the thermal sponge is distributed in such a way that the thermal energy from the transformation of the mechanical energy of the piston is absorbed into - a way as uniform as possible. Since, according to the invention, the piston moves with the isothermal speed v iz , the
  • the cover 7.2 of the cylinder is equipped with an exhaust valve 7.12r (the gas is discharged by replacing it with a liquid agent), and the lower wall 7.4 with an inlet valve 7.12a and with valves 7.4e for exhausting the liquid from the cylinder, at the end of compression.
  • thermal sponge is made of vertical cylinders (their cross-section may not necessarily be circular) concentric 7.3v, arranged at greater distances in the central part of the densifier, but increasingly closer towards its periphery. Moreover, the peripheral cylinders are provided with elements to enhance heat absorption (in the figure, horizontal fins 7.3f). Another component of the thermal sponge is the 7gs reticulated metal network, located in the upper part of the cylinder. The pressure in the cylinder is kept
  • Such holes can also be made at lower altitudes, to control the paths taken by the liquid agent, to increase the absorption power of the excess thermal energy and to accentuate the upward convective currents.
  • Another advantage of this configuration is the creation of the possibility that, after the evacuation of the compressed gas from the densifier, the cylinder 7.1 and its cover 7.2 can be raised for a short
  • the isothermal speed can be obtained by continuously changing the angular speed of the rotor, in such a way as to maintain at all times the equality between the instantaneous work delivered to the gas by the piston (in this case, the sliding blade in the rotor) and the instantaneous
  • the isothermalizer described in Fig.28 is a variant of the blade compressor, compressor described in detail in the patent application RO128041 (A2). It is characterized by the fact that it uses only one blade in the rotor. It consists of a stator (empty cylinder 6.2), inside which it rotates around its center
  • the rotor 6.1 In this configuration, the rotor cylinder is empty and its diameter is larger than the radius of the stator.
  • a pocket usually parallelepiped 6.4, obtained by installing side walls along the entire length parallel to the plane formed by the rotor diameters, on one side and on the other side of it, equally spaced from it.
  • this pocket housing in which the parallelepiped blade 6.3 is inserted, the length of which is equal to the inner length of the housing in which it is inserted (so of the
  • stator whose height is equal to the depth of the pocket and the thickness is equal to the inside thickness of the pocket (the four side surfaces of the blade slip tightly onto the inside surfaces of the housing).
  • the length of this housing is equal to the inner length of the stator (the surfaces of the base of the blade also slip tightly onto the inner surfaces of the stator bases).
  • the rotor is tangent to the inner surface of the stator wall. In the configuration in Fig.28, the radius of curvature of the stator wall is modified, over
  • the contact portion between the rotor and the stator is no longer limited to a straight segment, but extends to a curved surface with the desired width.
  • the blade can slide along the entire height of the notch, and when its tip touches the stator wall, it divides it into two chambers, sealed between them. This extreme position of the blade is ensured by the centrifugal force generated by the rotation of the blade, as well as
  • the height of the rotor can be equal to the inside height of the stator, in which case the surfaces of the two bases of the rotor slide over the surfaces of the two stator bases.
  • this height is higher and the sliding movement between the stator bases and the rotor walls is provided by bearings 6.91 and segments or seals, etc.
  • the rotor of the machine is mechanically coupled with an engine (electric or mechanical), and in the case of a expander, with a generator or other mechanical load.
  • On both sides of the tangent surface there are two rectangular slots (6.6d and 6.7d in Fig.28), connected to pipes 6.6 and 6.7 respectively, for the siction and for the exhaust of the working
  • the inlet 6.6d can be free, and on the discharge line 6.7 a valve 6.7a is mounted, automatic or operated by a coil 6.7b (Fig.28). If the machine acts as a rarifier, the suction is through a valve or drawer and the exhaust is usually free. In the configurations in which the axis of the stator is vertical, the suction and the exhaust of the working gas is made by cut-outs executed in the two circular plates that constitute the bases of the stator: a cut for intake in the 6.6v area of the lower
  • the described device is a rotating polytropic compressor that can achieve good performance in certain specific applications. Like any polytropic compressor, it can perform isothermal compression operations when its angular velocity is equal to the isothermal angular velocity u)(t) over the entire duration of a rotation, but even for large temperature differences AT this
  • the central device shall be programed in such a way as to ensure the equality between the mechanical power given to the gas and the thermal power given by the gas to its environment.
  • the isothermalizer in Fig.30 cumulates a series of changes that can be made to the isotermalizer with a blade, changes that can be applied separately, or cumulating several of them, depending on the
  • stator 6.2 changing the shape of stator 6.2 so that a section parallel to the bases is no longer circular, but the new section allows a continuous and watertight slide of the blade, and leads to a favorable change in the isothermal angular speed cuff
  • the stator shape and the curve direction of the blade determine different isothermal speeds. From this point of view, the construction of the densifier may differ from that of the
  • these cavities are designed to make, through valve 6.65, a passage between the uncompressed gas tank and the low
  • the inner cylinder 6.1 is held fixed and the outer cylinder 6.2, together with the tank 6.10 and the cooling system SR, mounted on one of the caps, rotate around it.
  • stator existing on the machine in Fig.28, is replaced by a rolling motion of the rotor on the inside walls of the stator.
  • the rotor moves along a circular path on the flywheel 6.81 (Fig.31 , section 1 -1 ) around a shaft perpendicular to the flywheel.
  • the flywheel rotates around the axis of the stator, at a distance equal to the length difference between the two rays (at the apparatus shown in the figure, where the radius of the stator is
  • Cooling of the gas during compression is done by the sprinklers 6.9b mounted in the wall of the stator, or by itscovers 6.22. As with the compressor in Fig.28, coolant circulation can also be made through the inside of the rotor 6.1 , if it is not used for other purposes. At this ratio of 2:1 between the two diameters, the rotor blade cannot be executed in one piece, and a telescopic blade consisting of two sections 6.31 and 6.32 respectively is
  • stator covers 6.2 (the two bases) must be movable in relation to the walls 6.21 : they rotate through bearings 6.91 mounted on the stator walls and through bearings 6.92 mounted on the rotor walls.
  • Fig. 32 shows how a solid sponge can be implemented, composed of almost parallel plates in
  • the plates are cylindrical metal sheets 6.12, each with a notch along a generator, with an opening slightly larger than the width of the blade, with unequal diameters, with values between stator and rotor diameter, mounted between these two cylinders, so that, compared to the assembly in Fig.28, the central axis of the rotor is moved towards the central axis of the stator, in the plane containing them, with a distance equal to the total thickness of all these plates, without leaving gaps for gas leaks.
  • the cylindrical plates of the thermal sponge are engaged in a rotational motion in which the peripheral points of contact with the rotor and those of contact between successive plates move at the same speed, which would lead to different angular velocities of the plates and at pressures exerted on the rotor blade. If these plates are light enough and elastic enough, they can be driven by the rotor blade in a rotational motion synchronized with that of the rotor. Another way to
  • gear teeth 6.1 m are mounted from place to place, on the outer surface of the rotor cylinder, for example in the shape of triangular prisms (see also Fig.32A). Each such tooth corresponds, in the same plane perpendicular to the central axis, on each plate of the thermal sponge, a hollow 6.12m (obtained, for example, by punching), or a hole, slightly longer than the rotor tooth.
  • the 2060 holes are made in such a way that the holes near the dead center overlap over the corresponding rotor tooth (so the distance between the holes increases as the diameter of the cylindrical plate increases). In this way, as it rotates, the rotor engages with the first plate, this with the next, and so on, equalizing their angular velocities.
  • the rotor is empty and serves to convey the gas with the inlet
  • the coolant 6.4I is conveyed through the rotor blade 6.3, through the pipes 6.9a and through the sprinklers 6.9b, from where it is injected between the sponge plates, the liquid in the stator being discharged, by means of a pump, through a hole made in the lower cover of the stator. .
  • the vane 6.1 12 that separates the different pressure zones is operated from the outside, using the spring 6.113 and performs back and forth movements in the cylinder 6.1 11 , along a fixed axis.
  • Rotor 6.1 performs a rolling motion inside the stator 6.2.
  • the internal volume of the stator is divided into several regions whose volume undergoes successive increases and decreases, depending on the rotation angle of the rotor.
  • the extended regions communicate with each other through the stator wall in a 6.6a portion open to the environment (if
  • the exhaust valve 6.7a actuated by a solenoid valve 6.7b, is located in the region where the volume of gas between two successive blades reaches the minimum value. In the case of rarefiers, the direction of rotation of the rotor and the role of the valves are reversed.
  • a more efficient use of the internal volume of the stator 6.2 is achieved by a simplified construction of the rotor 6.1 , keeping only its central axis, which can be full (Fig.35), or empty (Fig.34), on which they are mounted pockets 6.4, in which the blades 6.3 slide. Between the surface at the base of these pockets and blades are inserted elastic springs and/or lubricating fluid 6.4I. On the isothermalizer rotor in Fig.34, a cylindrical tank 6.4r is also mounted with a smaller radius than in the case of the device
  • the flat plates (preferably metal) 6.1 p are mounted radially, which forms the solid thermal sponge, and the radial pipes 6.11, at the end of which the sprinklers 6.9b are mounted.
  • the thermal sponge On the isothermal rotor, can be mounted solid thermal sponges which can have also other configurations.
  • the thermal sponge is made of metal wires 6.3b, which can occupy all the space that in Fig.33 the rotor of the device occupies.
  • the suction of the working gas is made as in the case of the isothermalizer in Fig.33, through a wide opening 6.6a in the stator wall, but for exhaust a 6.7 wide opening is used, with width, measured on the circumference of the circle, slightly larger than the distance between two successive blades.
  • the exhaust opening continues with a pipe mounted at the highest elevation of the stator, if its axis is horizontal. At the stator end of this pipe, a layer of liquid is permanently maintained,
  • a second, polytropic compression step can be introduced, followed by a cooling of the gas as it passes through the liquid layer. Further rotation of the rotor leads to the emptying of the liquid from the inter-blades space into the tank 6.10 and the intake of the working gas. If the stator axis is vertical, the inlet pipe is mounted on its lower base, through a 6.6v opening, and the discharge pipe is mounted on its upper base, through a 6.7v opening, of the shape and positioning
  • Quasi-isothermal compressions can be obtained just as easily, starting from state-of-the-art liquid ring compressors, with obtaining, at the same discharge temperature, higher compressed gas flow rates, if between the rotor blades of this type of compressor efficient thermal sponges are introduced, similar to those described in Fig.34 and Fig.35.
  • the construction of the isothermalizer in Fig.36 brings together the characteristics of several types of isothermal described above. It is a solid double-acting piston device consisting of a cylinder 5.1 (not necessarily circular in section) and two covers: one upper 5.1 s and one lower 5.11. Together, they delimit a closed enclosure, divided into two compartments by the piston 5.2, the sealing between them being made with elastic gaskets, segments, etc. The piston is moved between a bottom dead center BDC
  • the downward displacement of the piston is also determined by the profile of the profiled cam 6.14, with the speed v2, z , which may be different from v1 iz , due to the constructive differences between the two compartments (these differences can also cause differences in volume, or pressure, which may require the use of different tanks for the storage of compressed gas).
  • the telescopic rods in the two compartments also serve as carry-supports for the horizontal plates 5.1 1 of the two thermal sponges and are made according to the model of the carry-supports in the isothermalizers in Fig.16. These rods also serve as supports for the 6.9b sprinklers and the 6.9 pipes that feed them (these pipes can be placed right inside the splints that make up the telescopic rods. In this way a continuous cooling of the thermal sponges can be achieved.
  • the cooling circuit is composed of the
  • Fig.37A and Fig.37B show some of the modifications by which other state-of-the-art devices, the
  • the 6.9b sprinklers are mounted in the housing (they can also be mounted in the rotor body 6.14, 6.15, respectively 6.16, 6.17) which inject coolant into the space between the gear teeth.
  • This liquid is the liquid piston of the compressors, the flow rate through each sprinkler being controlled by means of 6.9c valves (for devices with larger volumes, these can be adjustable valves with servomotors),
  • Fig.38 contains some proposals for the implementation of thermal sponges in the configuration of some types of scroll compressors and some peristaltic compressors of the prior art, in order to bring as close as possible the polytropic coefficient of the transformations that take place in these devices to the unit value.
  • the proposed objective can be achieved if, in addition to this process, are applied procedures for the angular velocity modification and for the
  • Fig.38A shows a cross section through a compressor, with the two spiral volutes 6.18 and 6.19 (here, Archimedean spirals) interspersed.
  • the two spiral volutes 6.18 and 6.19 here, Archimedean spirals
  • one of the volutes is fixed, the other performing an eccentric orbital motion, without rotating, but there are also compressors in which, to ensure a safer seal
  • the two volutes rotate simultaneously, in the same direction, but with different centers of rotation.
  • the sealing between the compartments with different pressures is achieved by using 13.6 spiral-shaped gaskets, mounted on grooves made on the ridges of the two spirals.
  • the thermal sponge is composed of thin elastic metal plates, having the same spiral shape as the
  • 2200 main spirals about the same length, the same height and the same step, but with a smaller thickness g. They are located between the two main spirals, the distance between them being a multiple of a whole fraction of the distance between the loops of a single volute. For example, if this distance is b, the distance between two successive spirals of the sponge is b/N, where N is the number of plates that make up the sponge. In this way, between every two loops of the spiral considered mobile, are found a fixed
  • thermal sponge spirals will be longer, exceeding both extremities the respective end of the main spiral, as in Fig.38A, and will be perforated, through the resulting holes being inserted a rod 6.22, fixed to the respective main wing, which
  • n is the number of plates between the two spirals, plates that have the thickness g, on both sides of one of the main spirals (the fixed one, if only one is movable) a thickness reduction with depth Ng is performed along the entire length of this spiral, over the entire height t of this volute. At the points where the distance between the two main spirals is minimal, this distance is equal to Ng and is filled, in its entirety, by the spirals of the thermal sponge. If the two covers 6.18 and 6.19 are arranged in a
  • the sponge coils are supported with their lower edge on the lower cover of the compressor, their upper edge being in contact with the upper cover. All these contacts must be tight, at least when the distance between the two main spirals is minimal.
  • the thermal sponge is mounted by temporarily fixing the N thin spirals on each of the faces of the fixed spiral, its thickness thus becoming equal to the thickness of the movable spiral, then the introduction of the movable spiral, followed by the
  • the thickness reduction of the fixed spiral is performed only on a portion of its side faces, which leads to the formation of two channels 6.23 (visible in cross section 1 -1 ), one on each side, along the entire length of the spiral.
  • the depth of this channel has the value 2-N-g, if the thickness of the elastic
  • the 2240 plates is equal to that of the spirals of the thermal sponge.
  • the height of the thermal sponge spirals is equal to the width of this channel, and the spirals are arranged in such a way that when the distance between the two main spirals is minimal, the sponge spirals, together with the spacer plates penetrate these channels and occupy the entire volume, and in the parts in which the main spirals are spaced at a distance L, the sponge spirals are spaced from each other, with a fraction L/N, the same for all
  • the sponge plates always have at least 3 support points in that channel, so that they will never come into direct contact with either the lower or the upper cover.
  • the sealing method proposed in Fig.38A, section 1 - 1 is the mounting, on the bottom and on the edges of the respective channels, of some membranes made of elastic materials, slightly deformable.
  • the volume formed between this membrane and the channel walls is filled with a fluid and is tightly divided, by deformable walls, into regions with a width not greater than necessary to cover the contact surface between the main spirals.
  • the spirals entering the channel press the membrane at the bottom of the channel and push the fluid between the membrane and the bottom of the channel into the regions between the side walls and the membrane, obstructing possible gas leaks, as can be seen in “magnifying glass” 2, which shows an enlarged image of the region in question, when the movable spiral 6.18 steps on the fixed spiral 6.19.
  • Fig.38D shows one of the possibilities to transform a peristaltic compressor into an isothermalizer, by implementing a thermal sponge.
  • a thermal sponge (Fig.38E shows a cross section through its unfolded shape) consisting of metal plates 5.14, elastic, corrugated, similar to those described in Fig.10a and metal plate 5.11 , flat, rigid, occupying a central position, between two sets of corrugated boards. All these plates have the same dimensions when fully tensioned. In order to prevent uncontrolled
  • the shaft of trough 6.2 is a section of an arc of a circle, having its center having the center in the center of rotation of the arm 6.27, the arm on which is mounted, by means of the bearing 6.32, the pressing roller 6.26.
  • the ends of the trough are curved and have a path, towards the downstream and upstream device, outside the range of the pivoting arm.
  • a thermal sponge similar to those in Fig.38E, or Fig.38F,
  • a lamella is mounted with a width equal to the inner width of the trough. Continuous channels are drilled on the side edges of the lamella, in which a sealing gasket is mounted, transforming the lamella into a real piston.
  • the material from which the lamella is made (metal, or plastics similar to those from which peristaltic 2285 tubes are made) must be sufficiently malleable, so that, under the action of the force exerted by the roller, combined with that exerted by the elastic elements of the sponge, to follow a smooth route, with acceptable radii of curvature, but be hard enough to withstand the stresses to which it is subjected for a long time.
  • the end of the lamella in the inlet area of the pressure roller must be shaped in such a way as to allow gradual entry of the roller and must be fixed in relation to the trough so as to prevent movement
  • the other end of the blade must be shaped in such a way as to allow easy and complete evacuation of the compressed gas. If the material of which it is made does not have sufficient longitudinal elasticity, this end may be allowed to slide freely on the bottom of the trough.
  • thermal sponges can be mounted inside the tubes.
  • two peristaltic tubes 6.25 which have the shape of segments from a circular ring and are fixed on a metal bed 6.24, so as to allow the penetration into the circular space between them, a pressing roller 6.28.
  • This roller is mounted on a movable arm 6.27, which is continuously rotated by a drive motor 6.26.
  • a higher power density can be obtained if several pressing rollers are mounted on the same arm, each such roller performing the action of compression on two circular peristaltic tubes, with
  • Each of the peristaltic tubes is provided, at the end of the support on which it is mounted, with a check valve (in some applications, the inlet valves can be dispensed with).
  • a single pressing roller acts on each peristaltic tube. When this roller reaches the peristaltic tube, both valves are closed and the tube is filled with gas at the initial pressure. As soon as the roller passes the first end of the tube, the inlet valve opens so that another portion of gas
  • a pressing roller is mounted, at the same distance from the center of the device, on separate arms, which make equal angles between them.
  • peristaltic tubes can be extended to other types of compressors.
  • peristaltic tubes with thermal sponge can be mounted, the tube being pressed and the gas being compressed by the rotational movement of the vanes.
  • pressure of the gas in the peristaltic tube is usually different from that of the gas inside the compressor and has a different destination, but by careful sizing, the two compressors can become two stages of compression of the same process.
  • the space inside the compressor will be occupied, almost entirely, by the peristaltic tubes. Note that in this configuration, the distance between the two covers can be increased, so that the tips of the two spirals no longer touch the lid of the opposite spiral.
  • the orbital movement of the movable spiral, or the rotation in the same direction of both spirals, has the effect of a peristaltic pressure and leads to the compression of the gas in these tubes.
  • Rotary compressors are especially useful for high gas flow rates, for low compression ratios. To obtain higher pressures can be used the pressure step method. They are ideal for supplying precompressed working gas, at temperature T iz , for densifiers with high compression ratios.
  • Another type of compressor that has multiple gas compression/expansion chambers is the rotary screw .
  • An isothermal compression for this type of compressor can be achieved by using rotors with variable pitch, decreasing towards the exit (in this way, the ratio between the surface of the heat absorbing elements and the volume of the respective enclosure increases, as it does approach the
  • the sprinklers for dispersing the heat transfer liquid are mounted in the housing of the device and/or in the metal body of the rotors, fed from a channel that passes through their axis.
  • a new type of isothermalizer capable of performing energetically efficient isothermal transformations, is the gas piston isothermalizer, whose principle diagram is represented in Fig.39.
  • this isothermalizer has a first stage 8.1 , which includes one or more compressors and solid or
  • liquid piston isothermalizers connected in series or in parallel, which discharge into a tank 8.2i, which constitutes the second stage and whose volume is significantly larger than that of the devices in stage 8.1 , both stages being equipped with heat transfer control systems between the gas and its ambient environment, the pressure and temperature of the gas in each stage being controlled by changing of the speeds of the actuation devices and of the flow rates of the heat transfer agent, by an automatic
  • the first stage compressors In a long compression cycle (the time required to obtain the final pressure in the second stage), the first stage compressors must ensure, at its output, in each short cycle (we will call a short cycle, a
  • 2375 compressor C1 can be used, whose polytropic coefficient is close to the adiabatic one.
  • a quasi-adiabatic compressor is preferable (eg a rotary blower, which can provide a sufficiently high flow rate at the T iz temperature, with reduced dimensional characteristics than a reciprocating adiabatic compressor).
  • the isothermal compressor (isocompressor) C2 raises the gas pressure from the value p 1 to
  • a variable value p r equal to the gas pressure of the second stage, keeping the gas temperature constant. Since, in the adiabatic compressor C1 , the working gas is compressed in a short time, and the isothermal compression that follows requires a much longer period, a high-speed adiabatic compressor with a much smaller internal volume than the isothermal compressor is indicated, which flows into an intermediate tank R, in which the pressure p 1 is kept constant, by equalizing, through the play of the valves, or by changing
  • the pressure p 1 is the inlet pressure in the isocompressor (densifier) C2, and the discharge pressure (and implicitly, the compression ratio) is variable, being equal to the pressure p r of the second stage, a pressure whose value changes very little during a cycle short of the C2 isocompressor.
  • the pressure p r entering the second stage, varies during a total cycle (long cycle), between the pressure p, and the final pressure p f . If the final pressure p f has a
  • the C2 isocompressor can be made by connecting in series several densifiers.
  • the C2 densifier work with a double-effect piston, or with several densifiers in parallel.
  • the piston actually performs a slight compression of the gas in the 8.2i tank, it having to evolve on a new isothermal trajectory (the new isothermal compression process will be exerted on the compressed gas pumped out by C2, to which is added all the gas existing in tank 8.2, also taking into account the heat given up in this tank, to the heat transfer liquid).
  • tank 8.2i is filled with a hydraulic agent
  • the densifier 02 works with a minimum discharge pressure p2, corresponding to an economic regime . It supplies in the second stage gas with the constant pressure p 2 , gradually replacing the hydraulic liquid here (the mechanical energy consumed for pumping being recovered in a hydraulic motor/pump 8.6M, Fig.39A) , until the liquid in the tank is exhausted, when close the discharge valve. Then the compression continues with the tank 8.2i closed, with the gradual and slow
  • this first phase in which the densifier works with a constant compression ratio, pumping the compressed gas into the tank with constant pressure p 2 , has a limited duration, after which the pumping operation of each tranche of gas compressed, from the first stage, to be accompanied by the removal from the reservoir 8.2 of a smaller
  • tank 8.7 is no longer needed, but an additional tank 8.2 is added to the system.
  • the discharge of the first stage switches to the second tank 8.2 and begins the compression of a new tranche of gas, while the gas from the first tank is delivered to the upstream device, through the transvasation of the liquid in the RL tank, using the 8.6M1 pump.
  • two densifiers C2 are used, in parallel, to double the flow of compressed gas and, through their alternative operation, to reduce the periods when, in the second stage, no gas is introduced.
  • the two densifiers can switch to series operation.
  • the fluid level in the tank 8.2 is kept constant by the 8.6M pump.
  • the liquid in the tank 8.2 is cooled by its inclusion in a system that also contains the liquid-gas heat exchanger HE.
  • the cooled liquid in the heat exchanger is also directed to cool the thermal sponges of the first stage densifiers.
  • the first-stage isothermalizer can also be inserted into this tank. Through the valve 8.2r, the amount of liquid in this circuit can be supplemented to cool the sponge
  • the gas discharging from the last densifier of the first stage begins each time the gas pressure equals the p r pressure in the storage tank.
  • the volume of compressed gas between the solid piston and the tank acts on the gas in the tank as a gas
  • Another strategy that can be applied for an isothermal compression is to use a single isothermal
  • the tank 8.2 may be the primary of a plate heat exchanger (Fig.27), the secondary of which is part of a cooling/heating
  • This inner section also includes the sprinklers 8.8a mounted on the top wall of the tank, and the sprinklers mounted on the vertical supports 8.8v, each spraying liquid or foam, in the corresponding horizontal plane between the solid sponge plates.
  • horizontal plates may have perforations to create longer routes for fluid leakage. To achieve this, the surface of the plates gradually decreases, from the lower plates to the upper ones.
  • the density and complexity of the thermal sponge system differ greatly, from one application to another, depending on the final compression ratio and the temperature T iz .
  • the tank may also be cooled with foam, if a foam generator is mounted on the pipe that introduces the gas into the tank and/or a connection is installed between the compressed gas pipeline at the inlet and the liquid layer mixed with surfactants at the bottom of the tank. On this pipe
  • blowers S necessary to create a local overpressure and to adjust the flow rate of the gas-piston.
  • advantages over the state-of-the- art compressors keeping a large area of heat exchange at all times, superior possibilities for distributing and regenerating the foam.
  • the second stage of the isothermalizer in Fig.39A is equipped with a thermal sponge made of
  • the envelope of the sponge is constituted by the very walls of the tank 8.2i, and this sponge is continuously cooled by injection of coolant by an adjustable sprinkler system 8.8a, which maintains the average gas temperature at a predetermined value T iz , the coolant being maintained at an average T amb temperature thanks to an HE heat exchanger. Since the
  • the sprinkler system can be replaced by a system of horizontal perforated pipes, mounted on the top cover, from which, under the action of gravity, water droplets emerge and they drip at low speed, on the vertical plates. In this way, the contact time between the heat transfer liquid and the metal plates, as well as between the liquid and the gas between the plates, is much higher than in the state-of-the-art systems.
  • the tank 8.2i in Fig.40 is also parallelepipedal, the gas 8.2a in this tank being cooled by a system composed of a deformable metal strip 8.2b, (similar to the strips used for transporting small materials), with a width almost equal to the width of the tank, which runs permanently on a system of rollers mounted inside or outside it.
  • the entire system is inserted into a parallelepiped tank 8.2 filled with coolant 8.21. Due to the mechanical energy received from the outside by the drive rollers, the metal strip travels a
  • the route can be distributed alternately in both tanks.
  • sealing gaskets or sealing rollers with a smaller diameter are provided. The longer this path and the closer the adjacent portions of the band segments are, the greater the thermal energy absorbed from the
  • the rollers can be arranged so that the metal strip has the cooled portions (outside the tank) as long as possible.
  • Fig.40a another configuration of the location of the drive rollers is represented: they are mounted side by side, in contact each other, so that the two sets of rollers form the two walls: upper and lower, and the metal strip passes as tightly as possible between two such rollers. In this configuration, seals are required only for the peripheral rollers.
  • the outer tank 8.2 is filled with fluid to a certain level, above the level at which the tank 8.2i and the cooling strip drive systems are located. At the top of the tank 8.2 remains a layer of gas which, through the pipe 8.2c, constantly communicates with the gas in the inner tank and into which the metal inserts 8gs are mounted. By achieving equal pressure between gas and liquid layers, the tank 8.2i can be made with much thinner walls, regardless of the total compression ratio. Secondly, most importantly,
  • 2565 seals between fixed and moving parts, or between moving parts, are not subject to high pressures, which allows the volume of fluid 8.21 entering the tank 8.2i and the equal volume of gas passing into the top layer of the tank 8.1 to be reduced to low values.
  • the fluid level in the tank 8.2 is kept constant with the pump 8.6m. This way, liquid infiltration in the second stage helps to improve the quality of cooling.
  • the liquid in the tank 8.2 is cooled by including it in a system that also contains the pump
  • the compressed gas can be exhausted by the valve 8.2s into the constant pressure tank 8.7, by inserting additional liquid into the tank, with the pressure equal to the final pressure of the gas, until the entire amount of gas has been transferred.
  • 2580 metal bands a system of horizontal perforated pipes or a horizontal metal plate, provided with perforations, can be mounted in the upper part of the 8.2i tank, which will insert heat transfer fluid for additional cooling of the metal strip and the gas inside.
  • gas piston densifier represented in Fig.41.
  • this densifier also consists of an
  • 2585 isothermal first stage 8.1 (here, a solid piston densifier with thermal sponge made of horizontal plates 8.1 a. mounted on harmonic-type supports 8.1 r), or any combination of compressors and densifiers that discharge into collector 8.3.
  • first stage 8.1 here, a solid piston densifier with thermal sponge made of horizontal plates 8.1 a. mounted on harmonic-type supports 8.1 r
  • Fig.41 is also installed a polytropic compressor 8.4 (which can be a screw compressor), followed by a heat exchanger, circuit that can deliver to the output his, an additional flow of
  • the liquid in this pipe which is constantly recirculated, also ensures the evacuation of the compressed gas from the densifier.
  • the gas from the other compressors discharging into line 8.3 can be cooled in the same way, or with the help of an additional HE heat exchanger.
  • the conditions that ensure a superior efficiency of the system are: a large and unobstructed section of the gas inlet and discharge paths, a sensitive, fast opening and with low pressure losses of the discharge path, the existence of a
  • the opening of the inlet and outlet valves 8.1s (Fig.41 ) is done at the command of the regulation system.
  • the discharge valve can open at a certain pressure of the gas in the manifold (fixed position of the T point of the piston 8.1 ), in which case the compressed gas is directed to the tank under constant
  • each tranche of gas will replace a tranche with the same volume of hydraulic liquid in the tank), or at a variable pressure, equal to that in the tank 8.2, in which case the compressed gas is directed to the tank 8.2, where it will be further compressed, and after reaching the desired pressure in this tank, it can be directed to the tank under constant pressure 8.8.
  • the first stage an isothermal compression is carried out, up to a pressure p f1 , seeking to achieve a difference AT as small as possible (when the main compressor of this stage is a densifier) .
  • a pressure p f1 seeking to achieve a difference AT as small as possible (when the main compressor of this stage is a densifier) .
  • thermal energy cogeneration systems, high power density energy storage systems obtained by thermal energy storage, etc.
  • the device in Fig.41 can also perform the task of storing thermal energy.
  • the first stage of the device contains an adiabatic compressor that discharges into the primary of a HE heat exchanger, at a pressure equal to that in the collector (pressure that increases as the pressure increases in the second stage.
  • the HE exchanger has the role of reducing the temperature of the gas at the compressor outlet at a temperature T/ z , as close as possible to T amb and the thermal energy given to
  • the agent in the heat exchanger secondary is stored in a thermal reservoir, while the mechanical energy required to compress the gas in the second stage (due to the gas piston formed) is stored, after reaching the pressure p fi in a constant pressure tank, where the temperature is close to T amb .
  • the recovery of the excess thermal energy in the exchanger justifies the use of simpler (and cheaper) compressors ), with higher outlet temperatures, but with higher flow rates.
  • compression is realized as a gas piston densifier, with a compression ratio on a short cycle, very low (the inlet pressure p f1 and the intermediate pressure p f2 being close in value), associated with an efficient system for removing excess thermal energy.
  • the gas pressure in the second stage may suffer small oscillations caused by the cooling system. Also, the total surface area of the solid elements of the sponge that take heat energy from the gas being compressed does not change.
  • the second stage of the gas piston densifier 8.2 in Fig.41 is the primary of a plate heat exchanger, whose gaskets must be sized for pressures higher than p f2 .
  • a coolant circulates continuously in closed circuit, at a speed depending on the
  • This circuit includes, externally, a fluid/ambient HE heat exchanger.
  • the input of the primary circuit is coupled to the collector line which in turn is coupled, via exhaust valves, with the compressors in the first stage, or to the HE heat exchanger, while the output of the primary circuit (used for discharging the compressed gas) is usually closed.
  • 2640 be achieved by introducing a refrigerant in a state of equilibrium liquid/vapor, in the secondary of the densifier, and implementing in this system the necessary pipes for this secondary to become a thermal tubes cooling system (gravitational or with capillarity).
  • a refrigerant in a near equilibrium liquid/vapor state in the secondary, it can become the evaporator of a Rankine-cycle, or ORC thermal engine.
  • the heat exchanger HE is replaced by a condenser, which
  • 2660 also offers the possibility that, for large compression ratios (with the consequence of increasing the heat to be exhaustefd) it will go into forced mode, in which only one of the first-stage densifiers is in operation, and its second stage is cooled by the heat exchanger formed by the tank of the second densifier and the heat exchanger concerned.
  • FIG.42A another type of gas piston densifier is shown, consisting of compressors system 8.1
  • Gas supply at temperature T, and variable pressure p f1 is made by a system of densifiers, compressors and heat exchangers, with a common collector 8.3, similar to that of Fig.41. Inside the tank
  • a thermal sponge according to the invention, designed according to the characteristics and requirements of the system, adapted to the gas supply system and to the cooling system.
  • the main part of the sponge is a system 8.8v of bars of various thicknesses and vertical plates of various widths, arranged at sufficiently small distances from each other (to achieve a good capture of the heat accumulated by the gas in the tank), but large enough to allow a slight leakage of the coolant.
  • the liquid required to cool the gas is distributed between the bottom of the tank 8.2, between the heat exchanger HE, the sprinkler system 8.8a, the pump body 8.6 and the piping system 8.3.
  • the system allows the installation of any type of sprinkler from the state of the art, their number and distribution, the flow rate and pressure difference, the dispersion angle, the size of the drops produced and other characteristics, being chosen according to the characteristics of the application. In many applications, a
  • the coolant contributes to the compression and cooling of the gas sucked from the first stage, through the bubble densifiers 8.11 , made right on the coolant transport pipes, by injecting them with gas from the first stage collector. From here, the gas reaches the sprinklers and is exhausted in the densifier, even in the areas with the highest temperatures of the gas. If the gas from the first stage is injected into a beam of pipes with a sufficiently small diameter 2690 it will not form bubbles, but will form successive layers of gas, due to the surface tension of the liquid, alternating with layers of liquid, which improves the conditions of heat transfer between the two media.
  • the chosen configuration also uses other cooling processes from the technical stage, namely, increasing the heat transfer surfaces by introducing, or creating aqueous foam.
  • increasing the heat transfer surfaces by introducing, or creating aqueous foam.
  • gas piston isothermalizers are their ability to store in their thermal sponge, energy in the form of heat, which greatly expands the range of applications in which they can be used
  • the method that allows the accumulation of all the energy brought as input by the piston of the isothermalizer C2 and the energy from other sources is to abandon the HE heat exchanger (which had the role of giving up the excess heat to the environment) and replace it with a thermal energy storage tank.
  • the mechanical work performed by the pistons in the first stage is converted into thermal energy, part of which remains in the thermal sponges, in the walls of the compressors and in
  • the invention also discloses a highly efficient active thermal insulation method. Note that, in the case of a perfectly thermally insulated system, including the energy consumed in gas-gas, gas-liquid, gas-solid, liquid-solid and solid-solid friction processes in compressors and connecting pipes, the resistive and magnetic losses of the motors mounted inside the system, which
  • Fig.43 proposes a possible configuration of a gas piston isothermalizer with heat storage.
  • the adiabatic compressor C3 is added, which takes from C2 the gas discharged by it, with a temperature T iz1 and a pressure p 2 and compresses it
  • the system In order not to destroy the available exergy, the system must be made in such a way that the instantaneous temperature of the gas at the outlet of the compressor C3 is equal to T iz2 .
  • the energy storage is done independently in each stage, and the role of the C3 compressor is to, together with the C2 densifier, achieve a ratio of compression that ensures equality between the gas pressures at the exit from the first stage and the one entering the second stage.
  • the system is very flexible, allowing the choice in each step of how the temperature of the thermal sponges evolves, and allows changing the temperature difference (therefore of the power consumed) in each cycle, in each
  • T iz1 can be permanently maintained at the same value, with the first stage cooling system releasing excess heat to the environment.
  • T iz2 changes as heat accumulates in the thermal sponge of tank 8.2.
  • p r increases, both the compression ratio of the C2
  • the densifier 02 is of the rotary type, with adjustable speed, commanded by the controller, the tank R and the compressor C3 are no longer needed. Due to the heat build-up, the temperature of the 02 densifier thermal sponge increases simultaneously with that of the second stage thermal sponge, and its compression ratio must be changed to supply the second stage with gas at the pressure and temperature existing in the reservoir at that time.
  • the compressor 01 must supply the densifier 02 with gas at the temperature Tiz, which is why its compression ratio must be changed accordingly.
  • the temperature differences between the two cooling systems can reach large values. For this reason, it is recommended that the 03 adiabatic compressor be a rotary one, or even a dynamic one (for high compression ratios, several pressure steps are introduced).
  • thermometer mounted in the liquid layer 8.2I of the second stage 8.2 will indicate, in real time, its temperature T r , approximately the same as the temperature of the thermal sponge.
  • the compression ratio of the compressor C3 will be equal to 1 , and the gas pumped into the tank 8.2 will have the temperature T iz1 and the constant pressure p 2 of the gas in the densifier, and will gradually replace the hydraulic liquid in the tank, until the liquid level of here it reaches an optimal value, then the liquid discharge valve from this tank is closed.
  • the compression will be continued in the tank 8.2i, by the gas piston supplied by the compressor C3, with
  • the thermal sponge of the second stage of the isothermalizer must be sized for these special conditions.
  • the second stage of the isothermalizer in Fig.43 is a closed and thermally insulated tank 8.2, cylindrical, spherical,
  • the sponge can be made from metal inserts and fabrics, from metal foam, metal balls, from containers with perforated walls filled with metal fragments of any form (including filings,
  • thermal sponge in Fig.43 is made of 8mf metal foam, with another thermal sponge on top, made of metal fabrics, to allow an easy distribution of the coolant. In the configuration in the figure, both thermal sponges completely occupy all the respective plane sections. The lower part of the tank, separated from the solid sponge by a filter that stops the movement of small particles dislodged
  • the isothermalizers work in the rarefier mode, gradually transforming part of the accumulated thermal energy (part of it can be transferred to thermal energy consumers, in the cogeneration mode) into mechanical energy that they supply to such consumers.
  • 2800 Another configuration, simpler to execute, of the thermal sponge can be achieved by mounting on the upper cover of the tank a large number of vertical plates, bars, or vertical wires, single or in bundles, deformable or not, metallic, or from other materials, and through the distribution system to introduce, with a minimum pressure difference, the heat transfer liquid, which will slowly drain on the vertical elements of the thermal sponge, towards the lower cover, from where it is taken over by the recirculation pump.
  • Fig.44 shows another possible configuration for this class of isothermalizers.
  • the system consists of the adiabatic compressor C1 that sucks in gas (for example, atmospheric air) at the ambient temperature and raises it to the temperature T iz1 , equal to the temperature of the gas in the second stage, and introduces it into the reservoir R1 , from where it is extracted by a group of low- pressure isothermalizers C2 which, after an isothermal compression, stores it in the tank R2.
  • gas for example, atmospheric air
  • the gas is taken by the high-pressure isothermalizer C3, compressed to the pressure p 1 and introduced into the second stage 8.2.
  • the discharge phase of this compressor is not done at a rigorously constant pressure, the discharge being, in fact, an adiabatic compression with a very low compression ratio, which also leads to a slight increase in the temperature of the gas in this stage.
  • the heat transfer fluid that circulates through both stages equalizes the temperatures of the thermal sponges of all the
  • the second stage of the isothermalizer in Fig.44 is a closed and thermally insulated tank 8.2, cylindrical, spherical, parallelepipedal, or any other convenient shape. Its main component is the thermal sponge 8.2b, composed of bundles of metallic wires, very thin and very dense, with the smallest possible distances between them. The wire bundles are fixed to a horizontal 8.2g metal plate with numerous
  • This sponge has a very high capacity to accumulate heat and allows for easy circulation of gas and coolant.
  • the lower part of the tank is reserved for temporary storage as well as coolant circulation, containing a 8.6m pump that sucks liquid from here, to send it through a series of pipes to the upper part of the tank, located between the tank cover and plate 8.2 g, from where it drains gravitationally and
  • cooling is the use of aqueous foams, surfactant liquids and foam regeneration gas.
  • gas piston isothermalizers are their ability to store in their thermal sponge, energy in the form of heat, which greatly expands the range of applications in which they can be used successfully.
  • the method that allows the accumulation of all the energy brought as an input by the C2 compressor piston and the energy from other sources is to abandon the HE heat
  • thermal energy including the energy consumed in gas-gas, gas-liquid, gas-solid, liquid-solid and solid-solid friction processes in compressors and connecting pipes), from which, a part remains in the thermal sponges, in the walls of the compressors and in the liquid in the system tanks, and another important part
  • the mechanical work done by the pistons in the first stage is converted into thermal energy (including the energy consumed in gas-gas, gas-liquid, gas-solid, liquid-solid and solid-solid friction processes in compressors and connecting pipes), from which, a part remains in the thermal sponges, in the walls of the compressors and in the liquid in the system tanks, and another important part is transported by the heat transfer agent to the external storage
  • the first stage of compression will consist of one or more series rotary compressors or, in the case of large tanks, of one or
  • the heat transfer agent is found in one or more full RL tanks and is introduced through the sprinkler system 8.8a and recirculated using the hydraulic device 8.6, which here plays the role of a liquid recirculation pump.
  • the power required to introduce the gas in the first stage increases, and the rotor of the compressor C1 rotates at a lower speed. As long as the reservoir pressure remains low, the
  • second stage gas compression can be made isothermal.
  • the pressure in the tank 8.2 increases, the power consumed by the compressors in the first stage increases, so does the temperature of the gas at the entrance to this tank (equal to the temperature corresponding, in an adiabatic transformation, to the pressure in the tank), which leads and to the gradual increase, in the
  • the pressure of the gas in the tank increases simultaneously with the pressure of the gas at the outlet of the adiabatic compressor, until the maximum pressure is reached, but the temperature T 1 of the gas stored in the tank will be lower than that at the inlet of the tank, and the temperature T fcf of the sponge thermal will be inferior to it.
  • Fig. 45A for the insulation of an isothermalizer, involves placing the thermal insulation material, intended to limit heat losses from the system components, on a series of structures (frames, nets, plates, etc.) mounted around the object to be insulated, arranged in this way so as to form successive layers of insulating plates, parallel to one of the surfaces of the insulated object,
  • the route can be composed of registers of tubular pipes with a rectangular section 15.5, also made of insulating materials, as in Fig. 45C, placed in parallel planes, the outlet of the last register of each plane being connected to the inlet 15.6 of the first register of the next plane.
  • 2925 registers is occupied by heat-insulating materials (preferably in the form of foam, cotton wool, or granules).
  • the end 15.4 of each route continues with a branch which connects to a common collection pipe.
  • the insulation is passive (like prior art insulations), consisting of a succession of solid and liquid insulating layers (in some areas, heat flow can only pass through solid
  • 2940 can be found at which, on each of the flow tubes, the temperatures on the two faces of the solid layers are nearly equal , and the outward heat flux to be almost zero on as many flux tubes as possible. On the whole assembly, one can reach close to the ideal situation in which, of the total thermal energy that would be transferred to the environment when the liquid is stationary, only the thermal energy from the outer solid layer, whose average temperature is the smaller the number of insulation layers is higher. The rest
  • thermodynamic transformation of the gas in the exchanger can be isobaric, at a P M pressure as high as possible, according to Fig. 15B, being part, in the case of a warm system 15.1 , of an isobaric-adiabatic-isothermal engine cycle, in which the adiabatic expansion occurs between the temperature of the isolated system and that of the environment, that is, between the pressure P M and a corresponding pressure P m , and the isothermal compression occurs at
  • the mechanical energy obtained can be supplied to an engine that sets in motion an organic Rankine cycle heat pump, having the condenser located inside the isolated system 15.1 , or a gas heat pump, having the densifier located inside the system 15.1 , keeping constant (after a small addition, of additional thermal energy that covers the inherent losses) its temperature.
  • the efficiency of the system increases if there is a source of waste heat, whose temperature T rez is too low to
  • the fluid used in the thermal insulation system as a recuperator passes, before entering the system, through a heat exchanger 15.3, where it absorbs thermal energy from this source, then enters the insulation system after a more consistent insulating layer.
  • Another option, especially for small temperature differences, is to store the fluid in a reservoir, at the temperature at which it leaves the system, or at a higher temperature after receiving additional
  • Channels with a rectangular section, for the circulation of the fluid through the insulating material can be executed at the installation site, through profiles dug into the insulating plates, or by fixing some spacers and directing the fluid, on the insulating plates 15.7, methods by which it is carried out several parallel flows (Fig.45D), or a single winding flow (Fig.45E).
  • Fig.45D several parallel flows
  • Fig.45E single winding flow
  • a cross-section (1 -1 ) through this subassembly is shown in Fig.45F.
  • this type of alveolar thermal sponges can be inserted into the channels made for the circulation of the heat carrier fluid of the systems in figures 45C, D, E, F, as well as in the fluid layers between the insulating plates of the systems in figures 45F, G.
  • the chosen materials for 2985 the realization of these thermal sponges, they are chosen according to the state of aggregation of the heat transfer agent.
  • the heat transfer fluid slowly travels a long path, between the suction device and the opposite end of the path, collecting on this path the thermal energy (positive or negative) that travels the same path in the opposite sense.
  • the flow velocity is low, therefore energy losses due to friction are low, and the heat generated is also absorbed by
  • FIG. 45H shows a system similar to the one in Fig. 45G, designed for vertical insulating plates, in which the heat transfer agent is of the combined type. At the outer end of the path is mounted a compressor that introduces the gaseous heat transfer agent and causes it to travel the entire path. At the upper level of the intermediate gas layers and the alveolar insulating plates 15.2, a series of pipes 15.8 are mounted,
  • both gaseous and liquid heat transfer agents arrive at the end of the path with a temperature equal to that of the isolated object, and additional thermal energy can be extracted and stored.
  • a third objective of the invention is to propose, for a number of applications, different plant systems, made by incorporating the types of densifiers and rarefiers described in the invention. This incorporation can be done without changing the sequence of operations in the state-of-the-art systems, replacing only some parts of the system with more efficient ones made according to the invention, either
  • Isothermalizers can be successfully used in the construction of gas/gas and gas/liquid heat exchangers. Since the heat exchange is more intense when the fluids in question have a higher density,
  • the gases are isothermally compressed in a densifier D, and at the exit they are expanded, also isothermally, in a rarefier R, up to the initial pressure. If these operations, at acceptable power densities, were performed with the prior art compressors and expanders, a large amount of exergy would be destroyed.
  • Fig. 46 the scheme of a gas-gas exchanger, without energy accumulation, made with isothermalizers according to the invention is
  • the low-pressure gas LPgasI located at the temperature T amb -AT, is isothermally compressed to an economic pressure PM1 (which also takes into account the thickness of the walls and the necessary safety measures at high pressures), and the other low-pressure gas LPgas2, located at temperature T m is compressed isothermally to an economic pressure PM2.
  • PM1 which also takes into account the thickness of the walls and the necessary safety measures at high pressures
  • PM2 which also takes into account the thickness of the walls and the necessary safety measures at high pressures
  • the two gases then exchange thermal energy in the HE heat exchanger at a temperature difference AT.
  • the new exchanger will have a
  • Fig.46A a similar process is shown, when one of the agents is the waste gas of a chemical
  • the system is similar to the previous one, but the R2 tank is replaced by the second stage of a gas piston heat storage isothermalizer.
  • the temperature of the waste gas LPgas2 is first brought to the variable temperature T iz of the thermal sponge in the tank 8.2 by quasi-adiabatic compression in the compressor C1 , then the gas is
  • the heat transfer gas LPgasI enters the circuit with the temperature T amb -AT and leaves the
  • the control system can increase their compression ratio and the temperature difference between the gas and the rarefier's thermal sponge and of the accumulator, the necessary excess energy being stored in its entirety, in the form of thermal energy. Under these conditions, the amount of gas stored in the RG gas tanks is smaller, therefore, in order to extract the stored thermal energy, it is necessary that the surplus
  • thermal energy be extracted by the isothermal compression, at the ambient temperature T am b, of a corresponding amount of heat transfer gas, or by putting into operation a thermal engine, made from the compressor E4, former expander, from the rarefier D2, former densifier, from the expander C, former compressor and from the densifier R1 , former rarefier, by the appropriate modification of the connecting pipes among them.
  • a principle scheme is represented for the recovery and storage of thermal energy from liquid agents (domestic or industrial waste, liquids discharged from cooling installations, geothermal waters, etc.).
  • the scheme is similar to the previous one, but the HE is a gas-liquid heat exchanger, the liquid agent can also fulfill the role of a heat transfer agent, cooling the thermal sponge of the densifier, and the densifier is preceded by an adiabatic expander, for a more accurate correlation of temperatures
  • the one that must be correlated with the temperature in the second stage is the gas temperature at the outlet of the exchanger.
  • Fig.47 shows a process for avoiding exergy losses in the case of interaction between two gases with different pressures. Concretely, the constructive and functional differences between two heating
  • system 47a the ambient air, after being compressed in a densifier (curve 2-3 from Fig.47c), is introduced into a expander E, where it expands adiabatically (curve 3-6 from Fig.47c) until the pressure P o of the environment, the pressure at which the gas temperature reaches the T m value.
  • the gas discharged by the expander is mixed in the mixer 15.24.1 and after exchanging heat with the liquid evacuated from the enclosure, it is taken over by the densifier.
  • the invention proposes a method by which all heat exchanges are made at a minimum temperature difference, chosen by the operator, which difference is the most convenient compromise between the energy efficiency of this exchange and the duration of time allocated to this operation.
  • the installation uses a gaseous intermediate transfer agent (for example, air), distributed between a heat engine (whose components are denoted in Fig.47B with the index 1 ) and a
  • the rarefier R1 is inserted into a reservoir of liquid at the temperature T amb , and the rarefier R2 into the reservoir of liquid at the temperature T M , while both densifiers are inserted on the outlet line (or into the reservoir) of the liquid at the temperature T m .
  • Each of these circuits of the working gas also contains a compressor C, respectively a expander E, as well as the secondary of the heat exchangers HE1 , in the primary of which the liquid enters with the
  • FIG.48 3115 stationary regime, is represented in Fig.48, in which heat exchangers are no longer needed.
  • the system consists of a tank with liquid 15.18 in which horizontal perforated plates 15.29 are mounted from place to place, with the role of facilitating the thermal stratification of the liquid inside.
  • the tank also contains a series of vertical pipes 15.28 on which, at different levels, a series of valves 15.28.1 are mounted, the closing and opening of which is done by the system's automatic regulation system.
  • the system are the smaller tanks 15.19, mounted inside the main tank, which are coupled, by means of pumps P, to the vertical recirculation pipes.
  • the rarefiers and densifiers of the system are mounted in these tanks, and between these devices are mounted the functional elements that ensure the jumps between isothermal temperatures.
  • these devices are compressors and adiabatic expanders, but, in the case of Stirling or Ericsson cycles, regenerators or heat exchangers can
  • the commands sent to these valves have the role of establishing the number and temperature of the liquid layers controlled by the system.
  • the tank In stationary mode, the tank is connected in its lower part with the tank with liquid at ambient temperature, and in its upper part with the tank with liquid at the storage temperature, ensuring at the exit liquid at the desired temperature.
  • the systems described are energy storage systems. For their operation, they consume mechanical energy, when it is available, to deliver it later, when there is a demand, in the form of mechanical energy and/or heat. They can also work without receiving energy from the outside, if it is provided by a thermal engine, with equivalent Carnot cycle, which has the residual fluid as a hot source and the ambient environment as a cold source.
  • the gases before being introduced into the exchanger, the gases (respectively, the gas) are compressed isothermally in the densifier D (the gas at temperature T amb exchanges heat with a liquid at temperature 7), up to an economic pressure P M (which also takes into account the thickness of the walls and the necessary safety measures at high pressures), and at the exit they are expanded, also isothermally, in the rarefier R, up to the initial pressure P o , operations which, at acceptable power
  • 3140 densities, if carried out with the compressors and prior art expanders, destroy a lot of exergy.
  • the new exchanger will have a much smaller volume, and the lower speed of fluid circulation causes a reduction in exergy losses.
  • one of the fluids is liquid, it can also act as a heat transfer agent, cooling the densifier sponge and giving up heat to the rarefier sponge.
  • Energy exchanges in isothermalizers take place in the presence of temperature differences (in Fig. 49, in both devices, this difference is 217), and
  • the temperature jumps necessary to carry out this exchange can be executed polytropically inside the devices by increasing the piston speed, with the help of compressors, respectively quasi-adiabatic expanders, or with isothermalizers made according to the invention.
  • the temperature differences of the transformations in the two isothermalizers, as well as the one in the exchanger must be as small as possible. Larger temperature
  • 3155 are represented in T-s coordinates, the transformations of the gas during an isothermal-isobaric- isothermal cycle 1 -2-3-4-5.
  • heat exchangers can also be used with very good results to create heat recovery units designed to extract residual heat from hot gases and liquids resulting from various domestic thermal processes (local heating and air conditioning installations, food preparation, etc.), or industrial, from the
  • Fig.50 shows the configuration of a recommended installation for the recovery of thermal energy contained in some hot liquids, existing in nature (geothermal waters), or resulting from various industrial processes.
  • the plant can be used for the intergal recovery (by bringing the liquid to ambient temperature) of the heat contained in the
  • 3165 coolant, at the exit from the condenser of a plant that produces electricity through a Rankine cycle, or similar, eliminating cooling towers , or other types of large heat exchangers from the state of the art.
  • the special performances of this system are generated by the fact that both isothermalizers and heat exchangers can be made with acceptable power densities, even when working with very small temperature differences between the elements participating in the heat transfer processes.
  • 3170 recovery is done in the liquid-atmospheric gas heat exchanger HE in which the liquid, circulated with the help of pump P, enters with the temperature T m from the condenser outlet and leaves with the ambient temperature T amb , and the gas, with the temperature Tamb, after its isothermal compression in the densifier D, followed by a polytropic expansion, enters with the temperature T amb -AT (the temperature jumps at the entrance and exit of the device are performed polytropically, in the densifier, by changing the
  • a gas piston densifier and energy storage composed of the adiabatic compressor C, the gas piston densifier DAac, and the reservoirs RL and RG.
  • the 1 -2-3-4-5-1 cycle performed by the working gas is composed of a 6-5-2 engine cycle and a 6-3-4 heat pump cycle.
  • T m and high piston speeds the heat extracted from the liquid and given to the environment can be greater than the useful mechanical work developed by the adiabatic expander, so the energy balance is negative,
  • 3195 densifier and accumulation of energy composed of gas piston densifier D, RL and RG tanks, E expander and hydraulic pumps/motors M/P (Fig.51 ).
  • a mechanical separation of the solid particles can also take place in the densifier, if it has a liquid piston, or if the liquid recirculated for cooling the sponge is suitable for this task and if a suitable filter is installed in the liquid recirculation path.
  • a fine separation of solid residues can also be done at the exit from the densifier, or at the entrance to the rarefier, the filters
  • Fig.51A is an example of the operation mode of this type of recuperator, without energy storage, mounted on a CT thermal plant intended for the production of the thermal agent
  • the installation is equipped with a densifier D with a solid or liquid piston, which also has the role of sucking in the gas required for combustion from the ambient environment.
  • the burning of fuel (solid, liquid, or gas) is done in the firebox 15.14, the central being equipped with devices for regulating the flow of fuel and gas. After the heated gas, together with the combustion gases, give up most of their thermal energy to the
  • 3210 heat transfer agent in the heat exchanger 15.15 to be transported to the component elements of the heating installation 15.17, it is taken with the temperature T m from the outlet of exchanger, by the densifier D, which compresses it isothermally, to an economically chosen pressure pm, then it is taken over by a expander (gas turbine) which transforms this thermal energy into mechanical energy, to set a electric generator in motion, or is stored in thermally insulated tanks.
  • the installation is also equipped with the
  • isothermalizers can be successfully adapted to improve the performance of state-of-the-art air conditioning and refrigeration installations.
  • the Rankine and inverted Rankine cycles used most often in the state of the art, can be replaced by Carnot, Stirling, or Ericson cycles, and the isothermal processes of evaporation
  • Fig.52 and Fig.53 show the schemes of heating and cooling installations of an ambient space, installations for which the source (hot, respectively cold) from/in which the required/excess thermal energy is absorbed/given up is chosen from a wide range of sources available, against which there must be a large enough temperature difference to ensure an acceptable power density.
  • Fig.52 shows a useful heat pump for heating systems. It mainly consists of two containers, 15.18 and 15.19, filled with the heat
  • 3235 transfer liquid e.g. water, antifreeze
  • a pump for recirculating the liquid in which a pump for recirculating the liquid is mounted, a Riz rarefier in the first container, respectively a Diz densifier in the second, provide each with a thermal transfer system from the liquid to the thermal sponge, respectively from the thermal sponge to the liquid, for example, a sprinkler system.
  • each of the two pumps discharges and sucks the heat transfer liquid to/from a liquid-air heat exchanger HE1 and HE2, respectively.
  • the heat 3240 exchanger HE1 is mounted in an environment that constitutes the cold source, and HE2 in the hot source. In the case of heating systems, HE1 is mounted outside the space to be heated, and HE2 inside it.
  • the HE2 exchanger can be a series of convective radiators, radiators with infrared radiation, fan coils, etc., and the HE1 exchanger is sized accordingly.
  • the heat transfer between the walls of the two heat exchangers (and any fins attached) is done by natural convection, or forced with
  • V fans the help of V fans, and in some cases, one or both exchangers are inserted in closed enclosures, through which circulates a heat transfer fluid.
  • This fluid can be, for example, geothermal water, waste water, water resulting from an industrial cooling process, antifreeze fluid from a liquid-soil heat exchanger, liquidgroundwater, liquid-seawater, liquid-water from surface lakes, or from flowing waters, etc.
  • the working gas which can be air, at a
  • thermodynamic processes occurring in this system are represented in Fig.52A.
  • the temperature of the heat transfer agent in container 15.18 is T m , the same as that of the agent in the HE1 exchanger and the thermal sponge of the rarefier, lower by 21 , than the temperature of
  • the temperature of the heat transfer agent in container 15.19 is T amb , the same as that of the agent in the HE2 exchanger and the thermal sponge of the densifier, higher by 21 3 than the temperature of the external environment and lower by A T 4 than that of the gas in the rarefier. Both the temperatures of the agent in the two containers and these temperature differences can be changed, depending on the heat requirement, by changing the
  • 3270 Fig.53 shows a heat pump usable in refrigeration and air conditioning installations, made with the help of two isothermalizers, which have the role of extracting heat from an overheated environment, or from an environment that we want to cool it and transfer it to another environment with a higher temperature.
  • the principle scheme of this pump is identical to
  • the heat is extracted from the enclosure of a refrigerator, a cold room, or any other type of closed enclosure (with temperature T amb +AT) and is given to an environment outside this enclosure, and in the case of air conditioning installations, the heat is extracted from the air intended for
  • thermodynamical since the temperature difference between the two sources is small, we have given up the compressor and the adiabatic expander, following that the respective temperature jumps are executed polytropically, in the isothermalizers, still controlling the speed of movement of the pistons of the isothermalizers, through the same regulator that ensures their isothermal trajectories.
  • the 3290 be mounted between the external wall of the room and a thinner screen mounted at some distance, on the entire surface (without the glazed surfaces) corresponding to the wall, as in Fig. 54.
  • the temperature of the gas in the container can be maintained at or near the temperature of the outside environment.
  • the screen can be mounted inside the room, at a small distance
  • 3300 the outer double wall, and through the container to circulate a gas flow (similar to some types of buildings with low energy consumption), which on its route exchange thermal energy directly, through intermediate layers, or through an air-soil heat exchanger, with layers of soil, with relatively constant temperatures, located at depths greater than 80 cm, or through an air-liquid heat exchanger with a ground water cloth.
  • a gas flow similar to some types of buildings with low energy consumption
  • both systems can also be made in the version with refrigerant and with phase change, if in Fig.52 and Fig.53 the isothermalizers are replaced with condensers and vaporizers with refrigerant, and instead of the expander E a heat exchanger and an isoenthalpic
  • 3310 expander All these devices are mounted in enclosures 15.18 and 15.19, filled with heat transfer liquid, the heat transfer being faster than in the case of a gaseous heat transfer agent, as are many of those in the state of the art.
  • the heat exchanges with the cold and hot sources are done through the HE1 and HE2 heat exchangers, by recirculating the heat transfer liquid. This offers a great advantage in the case of large systems, by replacing bulky tubing that recirculates air with small pipes that circulate liquid agent.
  • Fig.53B and Fig.53D are identical to the previous ones ( Fig.52 and Fig.53), but the energy source with which the ambient environment is correlated is, explicitly, the external
  • 3345 rarefier has the same entropy as that of the atmospheric gas (Fig.53C, curve 6-7), and the pressure recovery is done adiabatically, with the help of a compressor (adiabatic compressor C1 in Fig.53B).
  • adiabatic compressor C1 in Fig.53B the amounts of heat given off, respectively received in one cycle, by the two gas enclosures, from the two isothermalizers, were represented by dashed bold lines.
  • the quantitative ratio between the thermal energies transferred between the two enclosures can be modified together with the 3350 temperature differences between the liquid agent in the two enclosures and the sources with which it interacts.
  • Fig.53F and Fig.53G are represented the T-s diagrams of these two systems, in the version where the working gas is the air extracted from the outside environment. Similarly, configurations can be made in which the air is extracted both from the enclosure and from its exterior, is compressed, and the
  • the configuration in Fig.54 contains a series of additional features: to increase the transfer speed of thermal energy and to reduce the dimensions of the heat exchangers (both the inside and the outside), in containers 15.20 and/or 15.21 there are no install the HE1 and/or HE2 heat exchangers, but one R iz 1 , respectively R iz 2 rarefier, each coupled with a D iz 1 , respectively D iz 2 densifier.
  • the rarefier R iz 1 is
  • the complete system consists of a heat pump R iz 2- D iz 2 through which circulates air extracted from the enclosure 15.21 , pump that extracts heat from this air, to give it to the liquid with temperature T,, in the enclosure 15.19 (Fig.54A), from the heat pump R iz 1 -D iz 1 , through which circulates air extracted from the enclosure 15.20, pump that extracts heat from the liquid of the tank 15.18.2, with temperature T,, to give it to the environment in the enclosure 15.20 (Fig.54C) and from the heat pump R iz -D iz , through which
  • a working gas which is a regulating pump, the mechanical work consumed by it being added in the form of thermal energy to the energy of the liquid in the tank 15.19.1 and given to the liquid in the tank 15.19.2 (Fig.54B).
  • T the temperature of the liquid, this having different values in each of the 4 enclosures with liquid, the temperature differences between the enclosures can also be changed by changing the recirculated flow rate.
  • the D iz and R iz isothermalizers are only needed when the difference between the T ext and T amb
  • 3400 temperatures is large. They form a heat pump (curve 1 -2-3-4 in Fig.54B), which transfers heat between the other two heat pumps.
  • This system is advantageous through the multiple possibilities of changing the working temperatures, the power of the installation and the speed of response to the commands received. All three types of systems described can work both as heating installations and as cooling installations, by changing the direction of gas circulation and transforming densifiers into rarefiers and vice versa.
  • Fig.55 shows a heat pump intended for heating an enclosure, in a simplified configuration, composed of a single D iz densifier, mounted in an enclosure with liquid 15.25, metallic, equipped with cooling fins, which sucks air from encloses and compresses it isothermally (curve 1 -2-3 in Fig.55A) and an adiabatic expander E, mounted in a gas mixer 15.24, which adiabatically expands this gas (curve 3-6 in Fig.19A).
  • the expander is a rotating one
  • the mixer consists of a
  • the mixer can be made in stages, each stage
  • a calcinator can also be attached to this system to sterilize the air at the inlet or outlet of the densifier.
  • Fig.55C The system in Fig.55C is similar to it, but the direction of movement of the working gas is reversed, therefore, this system extracts heat from the enclosure, with the help of the R iz rarefier, and transmits it to the outside environment, with the help of the D iz densifier (Fig.55D).
  • the working pressure can reach higher values, therefore the adiabatic expander is followed by a R iz rarefier, which ensures a better COP.
  • Another change compared to the previous system, is the increase of the thermal transfer
  • This coil is a spiral tube around the enclosure in which the isothermalizer is mounted, through which part of the heat transfer agent circulates.
  • Isothermalizers can be used, in general, for heating fluids, or solid bodies immersed in a fluid, and in particular, for the production of domestic hot water. Any of the previously described heating
  • 3440 systems can be adapted for this purpose.
  • a system can be the one described in Fig. 52, where enclosure 15.20 can be the domestic cold water pipe, or any other infinite cold source in gaseous or liquid state, and enclosure 15.21 can be a hot water tank household, which will accumulate in the form of thermal energy, all the thermal energy extracted from the cold source and all the mechanical energy consumed to operate the heat pump.
  • the isothermal speed control system will adapt the piston
  • any facility for heating a room in a building can work, simultaneously or successively, with two densifiers: one in the room to be heated, the other in the hot water tank.
  • the rarefier in the current state of the art, the evaporator
  • the densifier in the current state of the art, the condenser
  • the isothermalizer l z is mounted in a liquid-air heat exchanger, the liquid circulating, thanks to the pump P, through a series of metal tubes 15.27, and the warmer air, from the enclosure, thanks to the fan V, among these tubes, giving off heat.
  • the regulation system is programmed to maintain a constant
  • the l z 1 isothermalizer and the l z isothermalizer form a heat engine, which takes thermal energy from the ambient environment, part of which is transformed into mechanical energy, and the other
  • the rarefier l z disconnects from the densifier l z 1 and couples with the densifier l z 2, forming a heat pump that takes the excess heat from medium and transfers it to the liquid in the isothermalizer l z 2, raising its temperature
  • a thermal engine consisting of the rarefier l z 2 and the densifier l z 1 becomes active, which operates until the temperature of the liquid in the isothermalizer l z 1 reaches the ambient temperature
  • the described system can work continuously, if the capacity of the storage tank is large enough to take the heat produced in the most severe external climatic conditions, or if there are additional
  • the invention proposes the establishment in urban agglomerations of intelligent hot water supply networks, consisting of a central station in which to produce hot water for the surrounding urban area, with the help of heat pumps, operated by engines that they work with renewable energy and extract primary heat from the sun, from residual sources, geothermal water, ground water, etc. These stations deliver hot water through a network of pipes to all consumers in
  • prosumers who in turn produce, in some conditions, surplus hot water.
  • the central station must be provided with sufficient storage capacities and with the possibility to transform the surplus water delivered into electricity, when there are consumption peaks in the electricity network, or when an
  • a household refrigerator that has a heat exchanger fitted with a rarefier inside, and a sealed tank with a densifier on the outside, transfers all the absorbed thermal energy to a hot water boiler. For this, the attached tank is filled, in the first phase, with cold water from the network. Each time the heat pump is switched on, the temperature of
  • the water in the tank rises, and when it reaches the set temperature, the cold water supply valve and the valve connecting to the domestic hot water tank are opened.
  • the hot water in the attached tank is replaced with cold water and the cooling process coupled with the production of hot water is resumed.
  • domestic water drainage can be divided, with the help of a thermostat, into two branches, the hot branch passing through a hot waste water/cold domestic water heat
  • the ambient space cooling installation in Fig.57 extracts heat from the environment by means of the R iz rarefier mounted in a tank 15.19 with heat transfer agent, agent that circulates through one or more fan coils 15.21 , or similar devices.
  • the air circulating through the installation is extracted from the
  • the 3520 usage temperature is stored in the boiler B. If the hot water production exceeds the required consumption, the surplus hot water is supplied directly to the regional domestic hot water network, to be distributed to other potential consumers, or to feed a regional thermal energy storage facility, made according to the invention.
  • the domestic hot water administration system involves the creation of an organizational system
  • the ambient space heating installation in Fig.58 is composed of a working gas circuit under pressure, which operates according to a Brayton cycle. The installation extracts heat from the outside
  • Fig.59 whose operating cycle is represented in Fig.59A, is similar, but the liquid heat transfer agent is replaced by a working gas under pressure.
  • Fig.59B shows a configuration of an installation for the combined production of cold and domestic
  • 3545 hot water obtained by modifying a heat pump from the state of the art, made with phase change refrigerants (circuit 1 -2-3-4-5 from Fig.59C).
  • the modification consists of the introduction of the evaporator, respectively the condenser in the tanks where a densifier is mounted, whose thermal sponge is maintained at the isothermal temperature by circulating the heat transfer liquid in the tank.
  • the evaporator gas circuit (circuit e-f-g-h in Fig.59C) is closed by a rarefier mounted in the secondary of an
  • the sterilization system proposed in this invention can be implemented in any state of the art air conditioning system, local or centralized. It is a thermodynamic system for decontamination, by heat treatment, of the air intended for respiration, which we will call calcinator. Thermodynamic sterilization destroys pathogens by incineration. These systems are easy to implement and can be extremely efficient
  • both the compressor and the expander are electrically driven positive 3570 displacement devices.
  • the new process can be used in many areas, is flexible, allows a wide range of powers and dimensions, an easy adjustment of working flows, pressures and temperatures. These systems can also be made to a small size, so they can be applied to personal protective equipment as well as portable systems. Another great advantage of this system is that most of the mechanical energy consumed by the compressor is returned to the system by the expander, minimizing the energy
  • 3585 allows the adoption of a nonstop calcination strategy: the gradual increase of the air temperature up to the value of T adm above the value of T m , followed by an immediate decrease (without pause), below this value, so that the time when the temperature value exceeds the value of T m is higher than td value.
  • the configuration of the computer is chosen according to the chosen strategy.
  • the calcinator can be realized with any type of compressor and expander that can meet the requirements of the chosen operating variant, therefore its choice is made according to the performance of volume, weight, cost, convenience, etc. of the whole ensemble.
  • Fig.60A In the system described in Fig.60A we chose a nonstop strategy.
  • the cylinder of a compressor/expander can be equipped (Fig.60A) with a single inlet-outlet orifice (to achieve the largest possible diameter of the access path in the cylinder).
  • the cylinders of both devices are connected to the body
  • the 4-way valve (electrically or mechanically controlled) is a spherical valve 15.11 in which ball 15.9 the gas passageways are made, by creating cavities that also serve for the storage of compressed
  • valve ball is actuated by means of a shaft and a camshaft, for the correct synchronization of the operating stages.
  • valve can also be used successfully in any application described in this invention, when it is desired to create wide paths for gas and liquids circulation, thus a reduction of exergy losses (consequently, an increase in energy consumption).
  • a very useful application is to make a new type of
  • valve used is a 3-way valve: one way for the isothermal compressor/expander, one for the inlet pipe and the other for the compressed gas discharge pipe.
  • the discharge pipe is connected directly to the compressed gas storage tank (constant pressure tank) at its bottom and it is permanently filled with liquid.
  • compressed gas storage tank constant pressure tank
  • non- deformable thermal sponges are mounted, for example from interwoven wire nets, which have a large
  • the isothermal cylinder is equipped with a thermally deformable sponge and cooling systems, whose inlet flow is always equal to the outlet flow, so that the amount of liquid in the cylinder is constant, equal to the amount needed to eliminate the dead volume, when the piston is in TDC.
  • a in Fig.60B position in which is opened, through one of the ball cavities, a path with a large passage section, for the admission of the gas in the cylinder.
  • the other cavity of the ball is filled with liquid, in direct connection with the tank. In this phase, a heat transfer takes place between the liquid and the thermal sponge of this cavity.
  • the cavity that in the previous positions was in the upper position arrives in front of the suction pipe, the liquid from the cavity being evacuated through this pipe.
  • the amount of liquid introduced through the sprinklers is separated and introduced into the cooling circuit, the rest of the liquid (in an amount equal to the amount of liquid that left the tank, is stored in another tank, to be used when the gas stored in the first tank it is directed, under the same pressure, to a user (which can be this densifier,
  • 3650 stages two isothermal stages, one in the cylinder, the other in the valve cavity and a polytropic stage, performed by the liquid piston in the exhaust pipe.
  • the inlet to the discharge pipe is made through a wire mesh 15.18, which reduces the diameter of the air bubbles formed by the penetration of the liquid, bubbles that are cooled more strongly in the exhaust pipe and in the tank.
  • Fig.61 shows another configuration that allows the heating or cooling of the air in an enclosure.
  • the main loop works in a Stirling cycle, more advantageous, at least for the small installations, due to the abandonment of the adiabatic compressor and expander, more difficult to operate and adjust in case small temperature differences between hot and cold source, replacing them with a single recuperator (devices that have recently reached high performance).
  • all valves are removed in this loop.
  • the secondary loop through which the
  • R iz 2 and D iz 2 are isothermalizers made according to the invention: they are equipped with thermal sponge 15.15 and sprinklers 15.16 to ensure optimal heat transfer and an actuation system,
  • the gas in the densifier cylinder At a certain moment of movement, when a predetermined volume is reached, the piston of the R iz 2 rarefier starts at the same speed. In the next phase, the piston of the D iz 2 densifier reaches the BDC, where it stops, closing the respective end of the regenerator 15.17. Through this operation, the gas in the densifier pass into the rarefier, keeping the same volume, after changing the thermal energy with the regenerator and reaching the temperature T m . In the next phase, the piston of the
  • the proposed process is similar to the Siemens process: after passing through a treatment unit 15.20, the gas is compressed isothermally (curve 1 -2 in Fig.62A), in a D iz 1 densifier Fig.62, up to a pressure P 2 (for high pressures several stages may be preferable, without the need for intermediate heat exchangers).
  • the pressure P 2 corresponds to an entropy s 2 , slightly higher
  • the gas is released adiabatically (curve 2-3 in Fig.53), in a turbine T, to a pressure below the vapor saturation curve, close to P a , (in this area, the pressure P a is boiling pressure ) and a temperature below the critical point.
  • P a is boiling pressure
  • the liquid concentration be only a few percent.
  • the gas is exhausted in a condenser in which, by extracting the latent heat (curve 3-4 from Fig.53), the gas is
  • FIG.62 A proposed configuration for such an installation is shown in Fig.62.
  • the condenser of the installation is the secondary 15.21 of a plate heat exchanger, through whose primary 15.22 a heat transfer agent circulates, which at the pressure P a , in the vicinity of the temperature T 1 is in liquid state.
  • the heat exchange between the two regions is all the more intense, the larger the surface of the partition
  • the primary liquid is conveyed by a 15.27 pump and introduced into another 15.25 tank in which the D iz 2 rarefier of a heat pump operating in Carnot mode (curve 2'-5'-4'-3 'in Fig.62A) is mounted.
  • this heat pump (wich also consists of an adiabatic C2 compressor and T2 turbine, as well as a D iz 2 densifier) transfer the extracted heat, as well as the mechanical work consumed, to another heat sink, at temperature T a , or at a different temperature.
  • the installation may be provided with an additional system for cooling the gas of condenser, consisting of the D iz 3 densifier (which uses as coolant even the liquefied product, or the heat transfer agent from the tank 15.25) and the expansion valve 15.26.
  • This system extracts the warmer gas, from the upper area of the condenser and after a slight isothermal compression expand it isentropically to the pressure P a (curve 6-3 in Fig.62A).
  • Such a system is composed of several segments, one in continuation of the other.
  • Each segment is composed of two networks of parallel pipes, arranged transversely, with a length equal to the width of the track, through which the heat transfer agent (most often a mixture of water and glycerol) circulates, the upper network is mounted buried in the surface layer of concrete , as close as possible to the surface, and the lower network is mounted buried, under the concrete foundation.
  • heat source 3735 can be obtained using as a heat source solar energy of lower intensity, sources of residual industrial or household energy, with temperatures lower than those of the state of the art, geothermal sources (if the atmospheric temperature is negative, the heat source can be the soil at a depth of several meters, or a sheet of groundwater), etc., and as cold sources, the energy of ambient air, soil, flowing water, water of lakes and seas, groundwater, etc. They can even be made heat engines, or heat pumps, that work with
  • Isothermalizers are also useful in other configurations of engines and refrigerating installations that have different types of gases as their working agent, especially atmospheric air, such as internal combustion engines, operating in an open circuit (with the removal of combustion gases together with appreciable quantities of thermal energy), or closed and other systems that operate according to an Otto,
  • Fig.63G the T-s diagrams of a state-of-the-art Brayton cycle (curve 1 -2-3-4-1 ) and a modified system, by introducing, in the initial phase, an isothermal compression (curve 1 -2'-3'-4'-1 ).
  • Fig.63D was represented how the T-s diagram of an Otto, Diesel, or Brayton engine is modified, but was introduced an additional modification and flattened the heat absorption curve 3-4
  • thermodynamic evolution of the working gas no longer takes place in a set of cylinders (the same evolution in each of them, but with phase shifts to reduce vibrations), but in a sequence of devices (Fig. 63E):
  • the engine can work with a wide range of fuels, liquid, gaseous, or powdery materials, being able to utilize fuels with lower calorific values 3790 Fig.63F shows one of the possible engine configurations.
  • the adiabatic compressor C1 and the combustion chamber are located in the same cylinder (not necessarily with a circular section), being separated from each other by a drawer 15.7.
  • drawer 15.7 opens.
  • the CC combustion chamber piston at top dead center begins to move, with increasing speed, and the compressor piston
  • Fuel supply can take place, through nozzles located in different positions, until the piston reaches the end of the stroke.
  • the fuel dosing can be done in such a way that the expansion of the gas in this chamber is isothermal, at the highest T jz temperature that CC allows, which determines the maximum (ideal) engine efficiency.
  • Phase-change heat engines (operating on a Rankine cycle, or ORC) can also improve their performance with these devices.
  • ORC Rankine cycle
  • 3850 proposes a series of new system configurations, in which the fraction of thermal energy released to the environment, out of the total energy available for storage, can be significantly reduced.
  • Fig.64 shows an A-CAES type energy storage system (by adiabatic compression), suitable for the apparatus described in this invention.
  • the energy available for storage is used for adiabatic compression of the working gas by means of the isentropic compressor C1 (Fig.64B),
  • the gas in the tank R1 is expanded to the R iz 1 rarefier (curve 3-1 on the T-s diagram in Fig.64A), the resulting mechanical energy being taken over by the useful task (usually a electricity generator).
  • the thermal energy, stored in the R2 tank can be extracted at any time and can receive various uses. Still this energy can be transformed into mechanical energy using the heat engine described in Fig.64B, through a process that can be used to
  • the R iz 2 rarefier must be modified after each cycle. These changes are made by a controller, which receives signals from the pressure transducers inside the rarefier and from the temperature transducers in the rarefier and the storage tank. If the R iz 1 rarefier and the D iz densifier are placed in the same R3 tank, filled with a heat transfer agent and the R iz 1 rarefier starts simultaneously with the heat engine, the expansion of the gas in the rarefier is made by absorbing the thermal energy ceded by the D iz densifier.
  • the working gas pressure in the two isothermalizers is chosen as high as possible (to obtain a high power density), and the expansion and compression ratios can be optimized. With this storage system can be obtained a stored energy utilization factor close to 100%.
  • the temperature T m is presented in Fig.65: the gas aspirated from the atmosphere is compressed adiabatically by the compressor C1 , to reach the temperature T iz (curve 1 -2 in Fig.65A), then is compressed isothermally, with constant T iz , by the D iz 1 densifier, to the point with s iz entropy (curve 2-3 in Fig .65A) and again adiabatically, by the compressor C2, for to reach the temperature T m (curve 3-4 from Fig.65A).
  • the compressed gas having a
  • the process described in Fig.66 is a new energy storage process, based just on the high energy efficiency of densifiers and rarefiers with thermal sponge, proposed in this invention.
  • the advantage of the new method compared to those of the prior art is that almost all the thermal energy generated by the action of the piston is stored, with each compression cycle, in the working gas temperature, in the thermal
  • thermodynamic process that takes place in one direction (for example, from cold to hot) and in which it is consumed/produced a certain amount of energy E, will produce/will consume, during the development in inversely direction of the process, an amount of energy the closer to E, the
  • the proposed system consists of three distinct operating subsystems, arranged in three stages:
  • Energy storage is performed by a hybrid system that works after an reversed Carnot cycle (consisting of a heat pump and a refrigerator) and consists of the Diz densifier, the T isentropic expander,
  • T n is the temperature of the thermal sponge, of the walls of the compressor and of the storage agent in the tank R d , temperature that increases with each cycle.
  • the isothermal rarefier Riz takes over the gas (expanded in the isentropic turbine T) at a temperature T ⁇ T ⁇ -zl ⁇ (T i2 is the temperature of the walls, of the thermal sponge of the isothermal rarefier Riz and of the storage agent in the tank Rd) and expand it isothermally.
  • T i2 is the temperature of the walls, of the thermal sponge of the isothermal rarefier Riz and of the storage agent in the tank Rd
  • compressor C ensures the transition of the gas from the variable temperature T iz2 to the variable temperature T iz1 , by the appropriate modifications of the compression ratio, coordinated by the controller.
  • the four apparatus structure a reversed Carnot cycle, which is the most efficient cycle for two heat sources with temperatures TH+ATJ and T i2 -AT 2 .
  • the heat pump will operate in this mode until T iz1 and/or T iz2 reach the predetermined limit values, at which point an additional cooling/heating system of the densifier and rarefier is switched to a steady state mode in which both T iz1 and T iz2 , as well as the other state quantities do not change.
  • storage systems can be conceived with configurations in which to process the gas in a Stirling, Ericson, Rankine cycle, or even other cycles, if these cycles can be completed, in the recovery phase, in the sense conversely, with minimal exergy losses.
  • an isentropic compressor and an isentropic expander are required to force the gas to perform the (positive, or negative)
  • this material can be, in the initial phase in a solid state (for example, a salt, a paraffin, even a metal, etc.), which melts after the T iz exceeds the melting temperature.
  • the material in the tank Rr may initially be in the liquid state and solidify during the gas expansion cycles in the rarefier. When they are in the liquid state, these materials can circulate inside the respective isothermal device, contributing to the efficiency
  • liquid heat transfer agents are limited to a temperature range between the melting point and the boiling point, which requires, for storage at high temperatures, the use of a set of such agents, which complicates the configuration of the installations.
  • uniformity of the temperature of solid thermal sponges can be done by installing fans and related piping inside the tanks. At high tank
  • a transport circuit is opened, through which a flow of storage agent, with temperature T is introduced into the tanks Rd and Rr, replacing a similar amount of agent with temperature T amb , respectively T (introducing some temperature difference between the apparatus and the environment in the tank), quantity which is directed to be stored in the additional tanks R2 and R4 respectively.
  • the thermal insulation of all system components is an active insulation (third stage of the system) of the type described in Fig.45.
  • the cooling fluid is a gas, which yields its recovered thermal energy to a liquid agent, to be stored in the tanks R5 and R6, respectively.
  • a great advantage of the system is its flexibility. Note, for example, that the configuration described contains all the components needed for atmospheric gas storage operations, similar to those in prior art CAES systems.
  • the D.iz densifier, together with the compressor C, the expander T and a system of constant pressure tanks make up such a system, with an energy efficiency superior to the classical
  • the expander T is removed from the circuit, the system supplies compressed gas at the desired T iz temperature, the thermal energy of the gas can be stored together with the compressed gas, or it can be extracted in a exchanger and used for various purposes, same as the system described in Fig.45B. Switching from one configuration to another can be done at any time during the storage process. Moreover, in the initial phase, the R.iz rarefier can also change its sense (and role) and participate in the
  • the amount of stored thermal energy can be increased (increasing the amount of mechanical energy addressed for consumers), at any time of the process, from any available thermal source (fossil fuels, or biofuels, solar energy, geothermal energy, waste energy).
  • the storage system in Fig.67 is similar, its main components being the tank Rd, in which the temperature of the gas and of thermal sponge ts 1 it contains are kept at the temperature T iz1 and the tank Rr, in which the temperature of the gas and thermal sponge ts2 which its contents are kept at a T iz2
  • the 4010 temperature as low as possible.
  • they are equipped with an active Rec insulation, which causes the gas that retains the thermal energy that could be lost through a passive insulation, to transfer this energy to the liquid agent in the tank R1 , respectively R2, or to a heat engine (respectively, a heat pump), which restores, with the help of a small additional energy input, the stationary temperatures in the two tanks.
  • the Diz p densifier installed in the Rd tank forms together with the adiabatic compressor C p , the Tp
  • the Diz 2 densifier mounted in the tank Rr, draws air from the atmosphere, compresses it isothermally at the temperature T iz2 and stores it, under constant pressure, in the tank R, at the atmospheric temperature T atm .
  • the transition from T atm temperature to T iz2 temperature and the reverse transition are performed by the isentropic T2 expander and the C2 compressor. Simultaneously with this compression operation, the
  • the 4020 heat pump also starts. It absorbs the heat delivered by the densifier to the tank Rr and transfers it to the tank Rd, together with an amount of heat equivalent to the mechanical work performed for this operation. As a result, the fluid in the Rd tank and its thermal sponge changes its temperature with each cycle of the pump, storing the mechanical energy received from a wind turbine, or from another source of mechanical energy.
  • the T iz1 temperature will rise to the permissible limit, while the T iz2 temperature will remain
  • the recovery phase of the two forms of stored energy is done by the rarefier with variable isothermal speed Riz1 , together with the adiabatic compressor C1 and the expander D1 , with variable compression ratio, controlled by a regulation system.
  • the temperature in the tank Rd will gradually decrease reaching, with the emptying of the gas in the tank R, a temperature close to T atm .
  • the solid thermal sponge made of materials resistant to high temperatures, will have in its composition a series of sealed tanks, in the form
  • the system of these tanks can also be used to store heat transfer agents, which at the starting temperature are in solid state and change to liquid state before the starting heat transfer agent reaches the boiling temperature, so that, at
  • the starting agent is collected and stored in a reservoir, being replaced by the new agent.
  • the thermal energy thus stored can then be used as such, or it can be transformed, almost entirely, into mechanical energy.
  • the outstanding energy efficiency of these devices and the possibility of obtaining them at a low cost, offer the possibility to make very simple, small energy storage systems,

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Abstract

A first objective of the invention is to propose a new process for carrying out isothermal compression and expansion (isothermalization) operations in which the increase (respectively, decrease) of the gas pressure in a work enclosure, from a starting value p„ at a desired value pf, achieved while keeping constant the average temperature of the gas in the enclosure, to take place with compliance of an optimal sequence of partial transformations, by imposing an optimal trajectory viz(t) of the piston (solid, liquid, gaseous, or combined) and the other moving parts, simultaneously achieving the best compromise between energy efficiency and power density of the installation. A second objective of the invention is that the isothermalizers proposed for the isothermal compression and expansion of gases and vapors, whose moving parts contribute to the achievement of the isothermal speed viz(t), benefit from a satisfactory power density for the application served, which compared to quasi-isothermal devices with the same power, results in a shorter duration of the compression/expansion cycle and allows the circulation of higher gas flows, with the constant preservation of the temperature difference AT. The isothermalizers disclosed in this invention propose new processes for making the thermal sponge and cooling it as well as the gas in the enclosure, as well as innovative methods by which the energy efficiency of the isothermalizers is increased. A third object of the invention is to propose new complex installations made by incorporating the types of densifiers and rarefiers disclosed in the invention. By using the new facilities, due to the increase in the performance of the compression/expansion processes, the performance of all technologies in which the compression and/or expansion of gases and vapors play an important role increases: the transport, the storage and liquefaction of gases, the extraction of residual heat from gaseous and liquid sources, the production of mechanical work by using classic sources and, above all, renewable and residual ones, improving the operation of internal combustion engines and hydrogen ones, storing the energy from these sources, treating and conditioning the air, etc.

Description

NEW PROCESS FOR ISOTHERMAL COMPRESSION AND EXPANSION OF GASES AND SOME DEVICES FOR ITS APPLICATION
Technical field
The invention refers to a compression process and a similar expansion process of gas and vapors, processes which leads to a progressive increase (respectively decrease) of the gas pressure in the working enclosure, from a p, starting value to a target pf value without significantly affecting its average Tm temperature. The methods revealed in the invention for the implementation of the proposed procedures use existing installations (or parts thereof) in the prior art, but also new installations, proposed by this invention. Through the rigorous application of this process, when they are used using only techniques and devices experienced in the prior art, a significant increase in the energy efficiency of these devices is obtained, therefore, a significant reduction in the energy consumed for the gas compression, respectively an increase in the energy supplied as a result of the expansion. The proposed invention does not stop at these results, but proposes a series of new devices, using which the exergetic performance of the compression and detention processes increases even more strongly, by increasing the performance of all technologies in which the compression and/or expansion of gas and vapors have an important share: transport, storage and liquefaction of gases, production of mechanical work using classical and, especially, renewable and waste heat sources, efficient storage in various types of tanks of thermal and mechanical energy from these heat sources, installations for heating living spaces and for providing domestic hot water, refrigerators and heat pumps, installations for treatment and conditioning of air, etc. The invention also contains the description of some installations in these technological fields, to which the application of the described procedure involves a series of structural changes, through which new technologies are obtained, with superior results to those in the prior art.
State of the art
The thermodynamic processes in compressors and expanders of the current state of the art are carried out with a polytropic index, between the isothermal and the isentropic index, and differ from one type of apparatus to another, but also vary during the process. There are few technologies whose optimal development requires such an intermediate polytropic index, but polytropic transformations are resorted to even when isothermal or isentropic ones are preferable, because of considerations of minimum cost and technical limitations of the respective configuration. As a rule, in practical applications there is the need to obtain transformations as close as possible to those with isentropic index (for the complete conversion, without energy changes with the environment, of thermal energy into mechanical energy and vice versa), or with isothermal index (in which the internal energy of of the gas not to change). Obtaining transformations close to isentropic ones is possible by using the highest possible speeds of the processes (to limit the duration of the thermal energy exchange between the working gas and its environment), in the state of the art there are various types of devices that approach this desired. Instead, a transformation close to the isothermal one can be obtained when almost all the energy taken by the working gas from the mobile organs, in the form of thermal energy, is given to its surrounding environment. For this, either the duration of time affected by this transfer must increase (which implies a decrease in the power density), or the global heat transfer coefficient between the gas and its surrounding medium (the polytropic index approaches the unit value). This change in the polytropic index is the objective pursued by the quasi-isothermal compression/expansion devices of the state of the art.
An ideal, perfectly isothermal compression, in which the gas and the surrounding medium (solid or liquid) have the same temperature (temperature of the surrounding medium Tamfc), throughout the process, is only possible in the theoretical case of a process of infinite duration. Through the compression process, all the amount of gas in the closed enclosure receives from the mobile organs (solid or liquid piston, diaphragm, liquid ring, movable or deformable walls, elements of a rotor, fluid jets introduced into the enclosure, etc.) amount of mechanical energy that it instantly converts into thermal energy which, if accumulated, would lead to an instantaneous increase in the average temperature of the gas in the enclosure. The evacuation of this thermal energy from the working gas (with an average temperature Tg, higher than the temperature Tamb) becomes possible, in a first phase, for the areas in the immediate vicinity of some heat-absorbing bodies (component elements of the device where the compression occurs and other elements located in the closed enclosure), in the following phases, the heat being taken up by conduction, or by convection, also by the other areas in the enclosure. Therefore, in an isothermal compression that takes place in an ambient environment with a constant Tamb temperature, the mean working gas temperature Tiz must also be constant and higher than Tamb. Therefore, the gas temperatures at some point and the average temperatures in different regions may undergo large variations, but the global average temperature must remain unchanged. In this process, the energy received by the gas from the piston and instantly discharged into its ambient environment is greater than in the ideal case, where Tiz=Tamb. The speed at which this energy is discharged (and therefore the power of the compression plant) depends on the size of the difference between the two temperatures, on the physicochemical characteristics of the gas and the materials from which the compressor is made (which contributes to the definition of an overall heat transfer coefficient CGT), on the size of the contact surfaces between the gas and its environment (which contributes to the definition of a total heat transfer area AGT) and on the distribution of temperatures within the gas and by its convective movements. The size of the contact surfaces and how they vary during the compression/expansion process, as well as how temperatures are progressively distributed inside the apparatus, are constructive features. Maintaining a constant average temperature of the gas during the compression process is perfectly possible and can be achieved by maintaining the equality between the mechanical energy ceded by the piston to the gas (dependent on its velocity) and the thermal energy ceded by the gas to heat-absorbing surfaces in its environment. This equality can be achieved by the appropriate modification, through rigorously directed actions, of the speed of heat transfer between the gas and its environment. Fundamental theoretical studies, confirmed by the experimental results obtained, have concluded that, for an energy efficiency (n=Tg/Tamb, where Tg is the average, for the entire duration of compression, of the instantaneous average temperatures) imposed, the most efficient strategy by which a mg quantity of gas with pressure p1 and temperature Tamb (the same in all the mass of the gas), located in an enclosure with temperature Tambi is brought to p2 pressure and Tamb temperature within a time frame tiz is a three step process (AIA cycle):
- an isentropic compression of the entire amount of gas, within a time interval At—>0, to the pressure p3, corresponding to a working temperature^
- an isothermal compression, within a time interval At=tiz, to a pressure p4, which corresponds to the entropy of the gas in the state (p2, Tamb , during which the average temperature of the gas remains TiZ:, for which continuous change of the piston speed (isothermal trajectory) is required. Any other compression between the state (p3, T,z) and the state (p4, Tiz) occurs within a time interval At>tiz, or requires higher energy consumption. In the process, the bigger tiz, the smaller is Tiz. At high Tiz values, high energy values are obtained
- an isentropic expansion of the entire volume of gas, from the average temperature Tiz to the average temperature Tamb, from pressure p4 to pressure p2, within a time interval At—>0.
If this sequence of operations is followed, the choice of working temperature Tiz is made according to the characteristics of each particular case and is a compromise between the amount of energy consumed in addition to the ideal compression and the duration of the compression cycle (therefore with the power of the compression plant).
For solid piston compressors in the state of the art, without circulation of cooling fluids, for a prescribed average Tiz gas temperature and a constant Tamb temperature for all surfaces of the compressor components / which are in direct contact with the working gas, if known (or approximated sufficiently precisely), at any time, the values of A ) and hfl) for each /, that is, the dimensional and physical characteristics of the material (both those of the components of the compressor and those of its ambient environment), the imposition of the condition to perform the isothermal compression leads us to a first-order differential equation (the equation of the isothermal velocity), in which the only unknown function is the viz(t) the time variation over a cycle of the piston speed, (according with: Tdrdk A., Petrescu S., Popescu G., Feidt M., Isothermal compressors and expanders, Revista Termotehnica nr. 2, 2012, http://www.aair.ro/ buletine/1684.pdf: Tdrdk A., Petrescu S., Popescu G., Feidt M., Quasi- izothermal compressors and expanders with liquid piston, Revista Termotehnica nr. 2, 2013, In the current state of the art, obtaining this equation and its solution are done with some degree of approximation, depending on the power of the device. After solving (analytically or numerically) the equation, can be built an actuation system, that imposes this motion relation viz(t) on the piston and a process can be obtained, in which the average gas temperature remains close to the Tjz value throughout the duration of tiz oi the compression phase, and the energy consumed for this isothermal compression is minimal, compared to any other motion relation v(t) where the duration of the polytropic compression phase is equal to tiz. There are proposals (RO2013/128401 - Motoare Stirling, RO2013/128402 - Motoare Ericsson de mare randament), for the creation of installations that apply this procedure, by inserting some cams in the kinematic drive chain, but no installations were created to apply them.
Another series of theoretical studies and experimental trials (Farzad A. Shirazi, Mohsen Saadat, Bo Yan, Perry Y. Li, Terry W. Simon, Iterative Optimal and Adaptive Control of a Near Isothermal Liquid Piston Air Compressor in a Compressed Air Energy Storage System, 2013 American Control Conference Washington, DC, USA, June 17-19, 2013; Caleb J. Sancken, Perry Y. Li, Optimal efficiency-power relationship for an air motor-compressor in an energy storage and regeneration system, Proceedings of the ASME 2009 Dynamic Systems and Control Conference, DSCC2009, October 12-14, 2009, Andrew T. Rice, Perry Y. Li and Caleb J. Sancken, Optimal Efficiency- Power Tradeoff for an Air Compressor/Expander, Journal of dynamic systems measurements and control, 2017; Mohsen Saadat, Perry Y. Li, Terry W. Simon, Optimal trajectories for a liquid piston compressor/expander in a Compressed Air Energy Storage system with consideration of heat transfer and friction, Conference Paper in Proceedings of the American Control Conference, June 2012), starting from the average values of the global thermal transfer coefficient hj(t)Aj(t), or from a sequence of values of the main parameters measured during the displacement of the piston, proposes "almost isothermal" transformations using the variable piston speed method, with focus on liquid piston devices (by varying the flow rate of liquid using variable section nozzles).
Other processes in the state of the art aim to increase, through different methods, the global heat transfer coefficient CGT= hAi'. by using materials with a heat transfer coefficient between gas and compressor components as high as possible, by increasing the heat transfer surfaces (as in US4446698A - Izotermalizer system and IT1105771 - Reciprocating compressor with finned piston and cylinder head), by introducing additional solid or liquid components into the structure of the device such as lubricants, suspended liquid droplets, or jet (among which, US20130327033 - Forming liquid sprays in compressed- gas energy storage systems for effective heat exchange, US20110296822 - High-efficiency liquid heat exchange in compressed-gas energy storage systems, US20120297776 - Heat exchange with compressed gas in energy-storage systems), by introducing aqueous foam (US20090301089 - System and method for rapid isothermal gas expansion and compression for energy storage), or metal inserts (US20100018196 - Open accumulator for compact liquid power energy storage), by recovering the discharged heat energy (for example, US8851043B1 - Energy recovery from compressed gas), etc. The common feature of all these processes is their uneven and limited effect, which does not take into account the variation of the energy supplied by the piston. In the state of the art, these processes are mainly oriented towards liquid piston devices, several types of solid piston devices, scroll and screw compressors and expanders.
The invention we propose starts from the fact that, for a predetermined temperature Tiz, any compressor with positive displacement in the state of the art can behave perfectly isothermally, if the trajectory of the moving component, the one that determines the variation of the gas volume in the enclosure, is properly established (respects the isothermal trajectory), and the methods of increasing the global heat transfer coefficient contribute to increasing the power density of the device. For large temperature differences AT=Tiz-Tamb, isothermal transformations are obtained at instantaneous piston speeds with high values and with high energy consumption, but for small temperature differences 21 T, the compression cycle time can be very high and the compressed gas flow rate (implicitly the compressor power) can be very small. The duration of the isothermal compression cycle can be considerably reduced, if the contact area between the gas and the ambient is large throughout the compression cycle, and especially when the piston approaches the top dead center TDC.
From a theoretical point of view, we started from the premise that an isothermal compression of the gas in a compressor, at the average temperature Tiz, can be obtained if, at any time t of the compression process, the instantaneous thermal energy taken by the gas from the instantaneous work of the moving assembly W(t) is equal to the instantaneous thermal energy Qi(t) taken from the gas by the compressor components (piston solid, liquid, or gaseous, cover, walls, thermal sponge, lubricant) which are in direct contact with this gas through the surfaces of Ap, Ac, Aw, AL respectively. These processes conforms to the first principle of thermodynamics and to Newton's law: (1 ) where Q, is the instantaneous heat transfer rate for component / (/=p, c, w, L), h,(heat transfer coefficient) is a conventional coefficient, characterizing the intensity of this transfer, and W(t) is the instantaneous piston work allocated to compression for moment t (in some cases, as with devices fitted with an elastic thermal sponge, part of the work can be stored temporarily in the elastic elements). Therefore, in an isothermal transformation, the instantaneous mechanical work and the instantaneous power are proportional to Tiz-Tamb, so by ensuring r/?, t t =constant, we ensure a constant temperature difference.
Compared to the devices in the state of the art, this type of devices presents major differences in terms of the objective pursued and the constructive characteristics of the device. To emphasize this, for the perfectly isothermal compressors and expanders proposed by this invention, we will continue to use the common name of "isothermalizer", respectively "densifier"/isocompressor for the perfectly isothermal compressor and "rarefierVisoexpander for the perfectly isothermal expander. Due to the fact that the process can transform any of the compression/expansion devices into an isothermal device, it can be applied to any type of mobile device in the state of the art that generates changes in the gas volume in the enclosure, as well as to devices with new, innovative configurations
Summary of the invention
A first objective of the invention is to propose isothermal compression and expansion apparatus with full control over the deviations from the ideal trajectory, respecting a sequence of operations through which the proposed objective can be achieved, simultaneously with achieving the best compromise between energy efficiency and power density of the installation. The process described in the present invention and the devices proposed for its implementation have as their objective the consistent use of the sequence of the 3 phases, previously listed, of compression and expansion with maximum energy efficiency, in order to obtain constant temperature differences AT between the gas and its ambient environment. The proposed devices are designed in such a way that the instantaneous heat transfer values from the gas to/from the constitutive elements of the isothermalizer and from them to the cold/hot source are always as high as possible, thus ensuring maximum efficiency for the technological installations that use them.
Although in many examples that we will describe to exemplify the application of this process we will very often refer to densifiers, the device in question also works perfectly as a rarefier. For this, it is necessary that the temperature difference T=Tiz-Tamb be negative and that the displacement of the piston be done in the opposite direction, in order to take over the expansion energy of the gas and transmit it to the device to be operated.
A second objective of the invention is that the devices proposed in the invention benefit from a sufficiently high power density, so that the high energy efficiency is not paid for by diminishing other performances of these devices. Since in the case of a perfectly isothermal transformation, the instantaneous thermal energy taken by the gas from the mechanical energy of the moving components is equal to the instantaneous heat flow between the gas and its environment, we can achieve this goal by methods of sharp reduction of the polytropic coefficient of the transformation. The instantaneous velocities for the entire isothermal trajectory viz(t) (for any compression/expansion ratio) must be as high as possible, which will allow the circulation, in the same time interval, of higher gas flows, keeping the same gap of temperature AT, so an increase in both power and energy efficiency of the device.
A third objective of the invention is to propose new complex installations, made by incorporating the types of densifiers and rarifiers described above. By using new installations, due to the increase in the performance of compression and expansion processes, the performance of all technologies in which the compression and/or expansion has an important weight increases.
Brief description of the
The description of the invention shall be made in relation to the following figures:
- fig. 1 : the principle scheme and the T-s diagram of the ideal 3-stage isothermal compressor
- fig. 2: the principle scheme of the linear motor-driven isothermal compressor driven in real time by a pressure regulator
- fig. 2a: process of replacing the automatic system with a preset kinematic system
- fig.3, 3a: process for achieving the AIA transformation (adiabatic-isothermal-adiabatic)
- fig.3b, 3c telescopic rod for hydraulically operated solid piston
- fig. 4: principle scheme of the isothermalizer with thermal sponge made of elastic foams
- fig. 5: izotermalizer with thermal sponge made of helical springs with rectangular section
- fig. 6: izotermalizer with thermal sponge made of elastic cords and horizontal metal plates - fig. 7: izotermalizer with thermal sponge from peripherally mounted coil springs and metal plates
- fig. 8: izotermalizer with thermal sponge made of helical springs and horizontal metal plates with vertical fins
- fig. 9: izotermalizer with thermal sponge made of horizontal metal plates between which vertical helical springs are mounted
- fig. 10: izotermalizer with thermal sponge made of horizontal metal plates between which telescopic fins are mounted
- fig. 11 : schematic diagram of the diaphragm densifier
- fig. 12 izotermalizer with thermal sponge made of rigid flat metal plates supported on springy elastic plates
- fig. 13: izotermalizer with thermal sponge made of horizontal metal plates between which the elastic blades are mounted
- fig. 14: izotermalizer with thermal sponge from flat metal plates and arched elastic plates with central storage space; cross section and plan section
- fig. 15: izotermalizer with thermal sponge made of flat horizontal metal plates mounted on sliding supports
- fig. 16: izotermalizer with thermal sponge made of flat horizontal metal plates mounted with bolts on bellows-supports
- fig. 17a: schematic diagram of the densifier without dead volume
- fig. 17b: mini-densifier for collecting the compressed gas, attached to a densifier
- fig. 18: system for collecting compressed gas by replacing it with liquid
- fig. 18A: system for extracting a discharge box at a helical spring densifier
- fig. 19: schematic diagram of the isothermalizer with two compartments, with a deformable thermal sponge and a non-deformable one
- fig. 20: izotermalizer with solid piston, in which the increasing adiabatic jump is made by the sudden expansion, under the influence of an elastic spring, of a bellows that is filled with cooling liquid, immediately after the end of the absorption phase
- fig. 21 : densifier with storied liquid piston
- fig. 21 A: horizontal section of storied liquid piston densifier
- fig. 22: densifier with storied liquid piston and sprinklers
- fig. 22A: detail of the installation of the sprinklers in the densifier with storied liquid piston
- fig. 23: configuration resulting from the combination of a solid piston densifier with deformable sponge and a liquid piston densifier
- fig. 23A, B: configuration resulting from the combination of a solid piston densifier with sponge made of horizontal metal plates mounted on elastic elements and a liquid piston densifier
- fig. 24: isothermalizer with liquid piston, composed of elementary isothermalizers that have the thermal sponge composed of vertical telescopic fins compressible by moving a plate - fig. 24A: schematic diagram of a liquid piston isothermalizer, composed of elementary thermal sponge isothermalizers composed of telescopic vertical fins whose compression is done by moving the solid piston of another elementary isothermalizer
- fig. 25: isothermalizer with liquid piston, with double effect, having the thermal sponge made of horizontal metal plates under which non-deformable sponges are mounted
- fig. 26: configurations resulting from the combination of a liquid piston densifier and a group of solid piston and thermally deformable sponge densifiers with elastic elements
- fig. 27: densifier with liquid piston introduced into the enclosure by spraying
- fig. 27A: : liquid piston densifier with uneven distribution of the elements that make up the thermal sponge
- fig. 28A rotary densifier with a blade in the rotor and cooling with sprinklers
- fig. 28B: pressure step densifier, with 3 overlaid densifiers with a blade in rotor
- fig. 29: pressure step densifier, with 2 side by side densifiers with a blade in the rotor
- fig. 30: processes for the improvement of the rotary densifier with a blade in the rotor
- fig. 31 : rotary densifier with rolling rotor and telescopic blade in rotor
- fig. 32: rotary densifier with a blade in the rotor and thermal sponge made of cylindrical tubes
- fig. 32A: thermal sponge drive system of the rotary densifier with a blade in the rotor
- fig. 32B: piston-rolling densifier
- fig. 33: rotary vane densifier and cooling with external sprinklers
- fig. 34: rotary vane densifier and cooling with internal and external sprinklers
- fig. 35: rotary vane densifier, with thermal sponge and cooling with external sprinklers
- fig. 36: double-effect solid piston densifier with profiled cam drive, with horizontal plate, thermal sponge and sprinkler cooling
- fig. 37A: gears densifier and sprinkler cooling
- fig. 37B: cams densifier with sprinkler cooling
- fig. 38A: scroll densifier with thermal sponge from spiral plates
- fig. 38B, C: mounting variants of the thermal sponge of the scroll densifier
- fig. 38D:: peristaltic densifier
- fig. 38E, F: mounting variants of the thermal sponge of the peristaltic densifier
- fig. 39: schematic diagram of a gas piston densifier, horizontal plate thermal sponge, sprinkler cooling and foam generators
- fig. 39A: gas piston isothermalizer, in which the inner tank contains a thermal sponge of closely spaced vertical plates, and the coolant, coming from the main (outer) tank, is sprayed through a sprinkler system, or moves exclusively under the action of its own weight
- fig. 40: isothermalizer with gas piston, where the inner tank contains a thermal sponge consisting of a metal conveyor belt, and the coolant, coming from the main (external) tank, is sprayed through a sprinkler system - fig. 40A: alternative mounting configuration of the roller belt to the densifier in Fig.40
- fig. 41 : gas piston densifier with gas/liquid and liquid/medium heat exchangers
- fig. 42: gas piston densifier made by coupling two identical densifiers
- fig. 42A: gas piston densifier with bubble precooling and with foam and spray cooling
- fig. 43: isothermalizer with gas piston, where the second stage is a tank with a thermal sponge made of metal foam and another of metal inserts, mounted on top, and the gas is cooled by a sprinkler system with variable flow
- fig. 44: principle diagram of a gas piston isothermalizer with heat storage in a tank equipped with a thermal sponge made of vertical wires
- fig. 45: active system for thermal insulation of isothermalizers, with recovery fluid
- fig. 46: schematic of a gas-gas heat exchanger, without energy storage, made with isothermalizers according to the invention
- fig. 46A: schematic of a gas-to-gas heat exchanger with energy storage made with isothermalizers according to the invention
- fig. 46B: schematic diagram of a thermodynamic system for energy recovery and thermal storage from residual liquid agents
- fig. 47: system for recovery of exergy losses in case of interaction between two gases with different pressures
- fig. 47A: T-s diagram of a system for recovery of exergy losses in case of interaction between two liquids with different temperatures
- fig. 47B: system for recovery of exergy losses in case of interaction between two liquids with different temperatures
- fig. 48: system for the recovery of exergy losses with the installation of isothermalizers in the tanks with the thermal stratification of the working fluid
- fig. 49: gas/fluid heat exchanger, with isothermal rise of gas pressure
- fig. 50: heat recovery from hot liquids
- fig. 51 : heat recovery from hot gases
- fig. 52: air-to-air heat pump for heating, with isothermals
- fig. 53: air-to-air heat pump for cooling, with isothermals
- fig. 54: sterilization and indoor air conditioning systems
- fig. 55: systems for sterilizing and conditioning indoor air, with heat recovery by passing through an active insulation
- fig. 56: domestic hot water production system
- fig. 57: installation for cooling the ambient space and production of domestic hot water
- fig. 58: ambient space heating installation consisting of a working gas circuit under pressure, operating according to a Brayton cycle
- fig. 59: ambient space heating installation consisting of a pressurized working gas circuit, operating after a Brayton cycle and in which the liquid heat transfer agent is replaced by a working gas under pressure
- fig. 60A: system for thermodynamic sterilization of air, composed of an adiabatic compressor and expander, between which a 4-way valve with large passage sections is mounted
- fig. 60B: isothermalizer whose discharge/intake is via a 3-way valve with large cross-sections into a tank under constant pressure
- fig. 61 . system for cooling, respectively heating the air in a room, similar to the one in Fig. 50, in which the working gas loop works according to a Carnot cycle
- fig. 62: gas liquefaction system, which works according to a Siemens cycle, the condenser being cooled with a heat pump
- fig.63: internal combustion engines with isothermalizers
- fig. 64A: T-s diagram of the processes in the ACAES energy storage system
- fig. 64B: the new ACAES system for energy storage
- fig. 65A: T-s diagram of the processes in the ICAES+ACAES energy storage system
- fig. 65B: ICAES+ACAES combined system for energy storage
- fig. 66: ICAES system with hybrid thermodynamic cycle, for thermal energy storage
- fig. 67: ICAES system with hybrid thermodynamic cycle, for thermal energy storage in reservoirs and in gas tanks under constant pressure
- fig. 68: combined CAES-gravity system for energy storage
- fig. 68A: underwater transporter with compressed gas storage tanks
- fig. 69: isothermal compression system in pressure steps
- fig. 70: hydrogen fuel engine, with an inert working gas
- fig. 70A: isobaric cycle with pressure increase
- fig. 70 B: isobaric cycle for compensating a Rankine cycle
- fig. 71 : complex off-shore system for energy storage
- fig. 72: wind system with sliding blades
- fig. 73: semi-submerged turbine for capturing the energy of fluid currents
- fig. 74: wave energy capture system with horizontal bellows
- fig. 75: wave energy capture system with vertical bellows
Description of the best embodiments of the invention
As it was shown, in the state of the art, solving the problem of obtaining an isothermal trajectory is based on theoretical models in which it is difficult to introduce all the interdependencies between the main physical quantities, all the non-linear dependencies and the particularities of the working gases as well as of all the materials used in the construction of the device, the variable environmental conditions, etc. and, consequently, the solutions found and implemented, lead to the realization of some differences considered acceptable compared to the proposed objective, only at low speeds, or at reduced efficiencies. Contrary to this process, the invention proposes a different solution: instead of determining from the design phase, with an inherent margin of error, how the instantaneous power of the moving device must vary, we introduced an automatic device to the system adjustment, which, based on information collected in real time by a series of transducers, determines, regardless of environmental variations and other disturbances that may occur, the meaning and amount by which the quantity must change at that moment of the force acting on those moving parts that influence the average temperature of the working gas. In addition, we have generalized the procedure, in order to apply it not only in the case of piston devices, where the main regulated quantity is the displacement speed of the piston, but in the case of all devices with positive displacement, by introducing automatic systems for regulating the angular speed of the rotor and/or the inlet and outlet flow rates of the cooling fluids (transferring even the piston function to these fluids). In addition, the invention also describes a gas piston isothermalizer, which isothermally compresses gases in a closed enclosure with constant volume, by controlling the gas flow rate introduced, simultaneously with the control of the input and output flow rates of a heat transfer fluid. As in the case of rotary isothermalizers, in order to achieve an isothermal transformation, it is necessary to adjust several parameters simultaneously.
These control systems become the main component of all the isothermalizers described in this invention. The shorter the response time and the smaller the deviation of the supplied signals, the closer the transformations controlled by them are to an isotherm. The use of real-time tuning devices does not exclude a theoretical determination of the optimal trajectory and the use of the results obtained in the design phase of the actuators to decrease the instantaneous deviation and increase the response speed of the controller.
The movement of the movable parts of the actuator at such a speed that a constant difference AT (between the average gas temperature and the temperature of the components bordering it) is maintained at all times, results in maximum exergetic efficiency for any known type of compressor/expander with positive displacement, especially of the alternative ones.The simple replacement of the most commonly used drive system, the connecting rod-crank type (which imposes a quasi-sinusoidal variation in the speed of the piston), with a system driven by a variable speed motor, driven by a regulation system made in such a way that to generate the isothermal velocity viz(t) (in the case of piston devices, this is described mathematically by a decreasing exponential curve) can lead, in any installation, to significant reductions in energy consumption. In addition, the movement of the piston at this speed (in the case of reciprocating devices), or changing the speed of the rotating organs (in the case of rotary devices), allows the constant maintenance of the average gas temperature between two close limits and offers the possibility of harnessing the energy contained by low-potential sources (whose absolute temperature exceeds the ambient temperature by only a few percent).
It should be noted that with this type of compressor configuration and this mode of displacement of the piston (keeping the gap AT constant), even if the compression ratio is very high, dividing the compression process into several stages becomes unnecessary. It remains useful, however, a compression in stages, without intermediate heat exchangers, for a more efficient distribution of the available power. We make it clear again that the previous statements, made in the case of gas densifiers, remain perfectly valid in the case of rarefiers, except that in their case the gas is colder than the thermal sponge and than its component elements, as a result heat transfer is from them to the gas.
In the case of rotary devices that contain, for gas compression/expansion, a single working enclosure (for example, compressors with a single vane in the rotor) maintaining the temperature between two close limits is done by varying the angular speed of the rotor during each rotation of it , and in the case of those with several enclosures with different volumes, each in different phases of compression (lobe compressors, gear compressors, screw compressors, scroll compressors, etc.), the speed can be kept constant, but in each enclosure, the coolant/heater flow rate and the piston fluid flow rate are varied.
As we have shown before, the transformation with maximum efficiency is achieved if the three stages of the isothermal evolution are strictly followed, stages which can be achieved in a single device, or by combining several distinct devices. For example, the device in Fig.1 is composed of three distinct components: isentropic expander 1 , isothermal densifier 2 and isentropic compressor 1. Depending on the direction of gas flow through the device, it performs isothermal compression or expansion. In Fig.lA, on a T-s diagram, the corresponding temperature variation (curve 1 -2-3-4) is shown, as well as the mechanical work consumed to compress the gas in one cycle (hatched area). In this case, the working gas is sucked into the isothermalizer at the isothermal temperature, and when the final pressure is reached, the compressed gas is discharged also at this temperature.
In practical applications, the isentropic compression step can be replaced, for a small AT difference, with a polytropic compression. This compression can take place inside the enclosure, through commands given by the controller to the actuator that changes the volume of the enclosure, imposing such a rapid change that the isothermal temperature is reached from the first moments of the process. In other applications, for a small temperature difference AT, the polytropic compression is indicated to be done outside the enclosure, in a compressor with a simple configuration (thus with a low cost price), for example in a blower, or other rotating device (which also involves the use of a small intermediate tank), which ensures the rapid attainment of the isothermal temperature, at a higher power density, simultaneously with a higher flow rate and a higher degree of use of the isothermalizer enclosure. In this way, the rapid displacements of the piston, those at high speed in the initial phase, are also avoided.
Similarly, the isentropic expansion step can be replaced, for a sufficiently large AT difference, with an isobaric cooling step, if the thermal energy thus recovered can be used efficiently (for example, for the production of mechanical energy). The corresponding temperature variation in these cases is shown on the T-s diagram in Fig. 1 B (curve 1 -2-3-4). Fig.l C shows, by comparison, the variation curve with the maximum efficiency for the same compression ratio, 1 -2o-3o-4. In order for the isothermal compression/expansion stage to take place in the most economical conditions, it is necessary that the determination of the isothermal viz(t) of the actuator is as accurate as possible (which implies, in the case of the theoretical approach, an exact determination of all the quantities involved in the differential equation that describes the phenomenon and of all the correlations between these quantities), and the actuation mechanisms intended to achieve this speed should be robust, have a response time as short as possible and a sufficiently large adjustment range. The appropriate modification of the piston speed can be done by making appropriate kinematic chains, actuated by motor assemblies with variable speed (preferably, direct current motors, linear motors, stepper motors, or with high-performance hydraulic devices for energy transmission). In the case of the liquid piston, and in that of the hydraulically actuated solid piston, the variation of the working liquid flow rate can be done with the help of nozzles with variable section (pressure-swirl nozzles).
If, for some types of installations these determinations can be made with acceptable accuracy (especially for large areas of the thermal sponge, for large diameters and small lengths of the cylinder of the device concerned, and for small temperature differences AT), other types require empirical research and a significant amount of test bench measurements. In carrying out these determinations, it should be borne in mind, that a small improvement in the operation of the prototype translates into significant total energy gains when switching to series production.
One method that can lead to the best results is the automatic real-time adjustment of the actuator power, which avoids performing calculations, sometimes laborious, for determining the viz. For this, it is necessary to instantly measure the main gas state parameters using small sensors, that are enough sensitive and whose response time is fast enough. The collected signals shall be transmitted to a computing device, which produces an adequate response, transmitted to the actuator. For example, if the gas temperature is measured at enough points to determine an instantaneous average temperature Tmj, the controller (the calculating and adjusting device) compares the measured value with prescribed value Tiz calculates the deviation and transmits an acceleration, or slowing (or even stationary) command to the actuator to achieve equality Tmi=Tiz. Also, temperature sensors mounted in certain regions of the apparatus can transmit signals that, after processing, cause to be sent commands to the flow regulators on the coolant agent pipes. In this way, an isothermal evolution can be achieved, regardless of the temperature of the external environment, the temperature of the components of the compressor, of the thermal sponge or of the cooling agents. Moreover, the information collected can be used to adjust the lubricant flow, the spray coolant flow, etc.
Automatic power adjustment of the actuator does not exclude in its entirely, the calculation of the isothermal speed curve viz(t), or the corresponding variation of the position Xiz(t). On the contrary, knowing as accurately as possible a presumptive value allows computing devices to respond more quickly, the signal from the measurement transducers being used to correct the trajectory provided by the theoretical knowledge of the speed (or position) of the piston.
At any given time, between the instantaneous values of volume Vh pressure p, and average temperature Tmi. there is a well-established relationship, depending on the molecular structure of the working gas, given by the law of gases. For example, if we can consider that the ideal gas model can be chosen for this gas, we will use the PiVi=nRTmi, relationship, where n is the number of gas moles in the cylinder, and R is the molar gas constant. Therefore, in order to achieve the equality of Tmi= ^constant, there must be a direct correlation between p, and V Also, for every volum /• there is a well-determined piston position: Consequently, a computational relationship can be inferred: Xi=f(Pi). It is preferable to choose the pressure as the control parameter, as its value is the same throughout the gas mass, and a single sensor is sufficient for its measurement. There are also simple, inexpensive, sufficiently precise pressure transducers that give an extremely fast response (for example, piezoelectric ones). Fig.2 shows the scheme of principle for this type of installation. In this case, the processor (controller) 12.4 (DC), compares the pressure measured by the pressure sensor 12.5, with the corresponding one in an isothermal transformation, the working temperature Tiz and the position Li of the piston at that time. Depending on the result, DC sends the appropriate control to the drive system 12.3 (here, based on a linear motor), moving the piston 12.2 which moves in cylinder 12.1. The excess heat of the gas is absorbed by the thermal sponge 12.6 and the other components of the densifier.
As an example, in Fig.2 is represented the principle diagram of such type of installation, made on the configuration of a solid piston compressor from the state of the art. In the figure, was represented with a continuous line the circuit of the heat transfer fluid, with a broken line the signals transmitted by the transducers, and with a dotted line (ACAD 10W100) the commands transmitted by the controller to the executive bodies. In this case, the system is managed by the processor 12.4, DC, which constantly compares the pressure measured by the pressure sensor 12.5, with that corresponding to the working temperature T,z and the volume / at each moment. The calculation of the instantaneous volume is based on the position at that moment of the piston, provided by the position transducer L, from which the volume occupied by the liquid (calculated on the basis of the signals provided by the two flowmeters) is subtracted. Depending on the result, the DC transmits the appropriate command to the drive system 12.3 (here, based on a linear motor), setting in motion the piston 12.2, which moves in the cylinder 12.1. The linear rotating field is created by a cylindrical set of electric coils and a set of permanent magnets (or a second set of coils) positioned on the piston rod, which moves along the axis of the cylinder. We have avoided the use of the classical system, from the state of the art, in which the return movement of the piston is ensured by a system of springs, a system whose optimal operation is conditioned by its resonance coefficient (the return movement of the piston will be ensured by reversing the magnetic field , with a speed commanded by the controller, to ensure the highest possible efficiency. The excess heat of the gas is absorbed by a thermal sponge 12.6, deformable under the action of the piston and the other component elements of the densifier, to be given to a heat transfer liquid flow, which includes the variable flow pump 7P, the heat exchanger HE, the sprinklers 8.6A mounted on the enclosure walls, the servovalves 7v and the flowmeters 7f (one for the input flow, the other for the output). The change in the heat flow rate taken by the liquid is made, simultaneously with the modification of the volume of the enclosure, through controller commands to the two 7v servovalves and possibly to the pump drive motor. Operation at maximum efficiency is ensured by the correlation of the two processes of isothermal temperature regulation. The realization of these optimal viz(t) trajectories requires the use of regulation systems that are as sensitive as possible and with the shortest possible reaction time, based, for example, on linear electric motors, on direct current electric motors with permanent magnets, or with a spooling rotor , on Ward-Leonard groups, on motors equipped with ESC (electronic speed control) circuits, synchronous motors powered by a frequency converter, DC servo motors for valve actuation, etc. If the controller used has a very small response time, the way the kinematic chain between the engine and the piston is made has less influence on the piston trajectory. If the total heat exchange surface between the gas and its surrounding medium remains approximately constant during the compression phase, an almost isothermal transformation can be obtained if, throughout this phase, the power transmitted to the piston is rigorously constant (in the case of an engine of direct current, this can be achieved by keeping the supply voltage and current constant), as long as the overall heat transfer coefficient is kept constant.
The cost of these types of drives, especially for high-flow and high-power-density isothermalizers, is quite high, and the maintenance costs are not negligible. For this reason, drives that use alternating current motors, single-phase or three-phase, asynchronous or synchronous, become preferable. The best compromise between the price and the performances of these drive systems is obtained by using in the design phase, successively, both solutions for driving the moving parts: the prototype of the isothermalizer designed for a specific application will be equipped with all the elements of the drive system tuning: the signal transducers, the actuators and the controller with the algorithm corresponding to the respective prototype and application. Then, the system is tested on the test bench, being able to experiment for a series of different values of Tiz, so for different average piston speeds and different values of energy efficiency. Depending on the results obtained, the optimal trajectory and the signal to be transmitted to the drive system are chosen. As a result of these experiments, the decision can be made to increase the absorbent surface of the thermal sponge, or to increase the heat transfer agent flow rate. Based on the curve describing, for one cycle, the position of the piston as a function of time L(t), i.e. the isothermal trajectory, the dimensions and shape of the motor cam and the characteristics of its guide spring, or the path of the profiled channel on one face of the disc drive are determined, or the configuration of another mechanism chosen to achieve this trajectory, devices which, together with a constant-speed drive motor, will replace, for isothermalizers manufactured in series, the direct current motor and its automatic speed regulation system. Fig.2a shows a solid piston isothermalizer, whose piston is actuated by means of a disk 12.7 through whose profiled channel the bearings 12.8 run. The disks are driven by means of the axis 12.9 driven by the alternating current motor 12.10. In Fig.2A is represented, as a function of time, the allure of a curve viz(t), and in Fig.2B, the way in which this curve is transformed into a closed curve (the profiled channel) whose trajectory depends on the angle of rotation of the engine. The mounting variant with a single disk, with the execution of the profiled channel on one side of it and the coupling system to the motor shaft executed on the opposite side, represented in Fig. 2C is the variant that occupies the smallest volume, compared to the variant with two disks from Fig. 2a, with the version with a profiled cam, or with the connecting rod-crank drive system. When it is desired to drive several devices by the same motor, a gear wheel is mounted on its axis, next to each device, which drives the rod of the respective device through the teeth on the periphery of each profiled disc. The system is very versatile, allowing these discs to be driven at different speeds, with different trajectories, or to couple different types of devices to the motor shaft. Other modes of motion transmission can also be used, for example through belts.
This technical process for making isothermalizers can be applied to any type of compressor/expander in the state of the art, including those which, being driven by a high-speed engine, through a connecting rod-crank system, have a sinusoidal variation of the piston speed and a trajectory close to the adiabatic one, and the average polytropic coefficient of the transformation is high. The prototype of this compressor will be equipped with a complete real-time speed control system, then the AIA (adiabatic-isothermal-adiabatic) curve of the transformation is raised, in which Tjz is chosen in such a way that the compression cycle time is the same. By replacing the automatic system with a kinematic device that determines the realization (with an acceptable approximation) of this optimal transformation, a compressor with a lower consumption of mechanical energy is obtained.
Using this process, laboratory devices can also be built to study how the average temperature varies in a closed enclosure with variable volume, and to obtain the curve of variation as a function of time, as well as as a function of the instantaneous position of the moving body, of the average temperature of the gas in the respective enclosure. For this, a series of pressure and translation transducers are mounted in this enclosure to determine the instantaneous position of the mobile organ that determines the change in the volume of the enclosure. Based on the pair of data obtained with the help of these transducers, the time variation, for different input signals, of the average temperatures of the gas in the enclosure is subsequently determined, with the help of gas laws.
Another possibility offered by the solution proposed in the invention is that of the global change, during operation, of the average speed of the drive motor (change valid for longer periods of time, of the order of multiples of the duration of a cycle), change that can be obtained by the variation of the supply voltage/current of the direct current motors, or of the frequency of the current with which the alternating current motor is supplied (single-phase, or three-phase, synchronous, or asynchronous) with the help of frequency converters. These speed changes are required in a wide range of applications:
- in situations where, due to external conditions, the evacuation of the thermal energy absorbed by the sponge becomes deficient, keeping the temperature difference AT (and the imposed energy efficiency) and the isothermal character of the transformation requires a corresponding reduction of the power (average speed) transmitted by mobile organ
- in situations where the sponge of the isothermalizer also has the role of storing energy, the heat transfer agent is only used to equalize the temperature of the thermal sponge and the walls of the device, keeping the temperature difference AT, and changing its temperature by the same value as the temperature of the thermal sponge increases . Keeping the temperature difference AT under the conditions of increasing the temperature of the thermal sponge entails the need to increase the power contribution of the piston - in the case of energy storage systems from renewable sources, the energy to be stored can show important variations over time. Similarly, during the supply of stored energy, load changes occur frequently. These variations can be compensated for by changing the speed of the piston drives
- if the hot source, the cold source, or both, involved in the operation of heat engines, refrigerators, or heat pumps are finite sources, or with random variation, the operation of the installation leads to changes in the temperature of these sources. The need to maintain the temperature difference AT implies a change in power, which can be achieved by changing the speed
All these changes in the average speed of the drive motors do not influence the shape of the isothermal velocity trajectory, but only the total duration of a complete cycle (and implicitly, the energy efficiency of the transformation). In Fig.2a, the frequency converter is represented by the rectangle 12.11 . In order to ensure a certain nominal difference ATn between the temperature of the gas and that of its environment (and therefore, a certain nominal energy efficiency qn of the thermodynamic transform) a certain nominal optimal trajectory vizn of the piston speed is required which, for a complete cycle, determines a nominal average speed vm.„, and for the drive motor a nominal uniform angular speed w,„ which is provided by a certain nominal frequency of the motor supply current. Reciprocally, maintaining the isothermal character of the transformation in the case of increasing the isothermal temperature Tiz, without changing Tamb, leads to the need to increase the average duration of the compression cycle. The constructive elements of the isothermalizer are chosen in such a way that the desired efficiency i n (and the desired speed wn), corresponding to the nominal isothermal temperature Tiz.n (therefore, the nominal energy efficiency) is obtained at the standard frequency of the network (50 or 60 Hz) but, every time the external conditions require it, the global speed can be changed: w= wn (where k is the ratio of the two speeds) with the help of the frequency converter, simultaneously with the change of the temperature difference AT=kATn=k(Tiz n-Tamb), or (in the case of gas piston isothermalizers) with the increase in nominal discharge pressure, which also leads to changes in energy efficiency, nominal power and power density. Moreover, in a number of applications (e.g., where the thermal sponge also becomes a thermal energy accumulator) the temperature of the gas in the enclosure Tiz is deliberately changed, continuously, depending on the temperature of the sponge, without changing and the temperature difference AT, thus keeping the shape of the isothermal trajectory unchanged and changing only the temperature and pressure level at which the transformation takes place, thus the nominal working power.
Another way to change the temperatures, pressures, power and energy efficiency during operation, for long periods of time, is to mount on the isothermalizer some mechanical devices, existing in the state of the art, by which the strokes of the main piston and of the pistons of additional adiabatic devices to can be changed during operation, thus changing the value of the temperature at which the isothermal transformation takes place
Since a perfectly adiabatic transformation is not possible (the acceleration required for the speed jump from v=0 to v=viz0 being limited), another improvement in the process of realizing the perfectly isothermal transformation refers to the way in which the two adiabatic jumps between Tamb and Tiz are approximated. To make the first jump, several solutions can be considered:
- changing the instantaneous speed of the piston, after the end of the intake phase, by changing the instantaneous angular speed of the drive motor (a very strong acceleration, followed by a strong braking to reach the isothermal starting speed viz.o (with the help of electronic regulators), method that requires the use of high-performance regulators (with as little overshoot as possible and with as high a response speed as possible) and the use of mobile mechanical components with low inertia
- by introducing a cam whose profile determines such a movement: the lateral surface of the cam must contain a portion of great length (approximately equal to the maximum radius of the disc) which makes as small an angle as possible with the plane passing through the center of rotation and through the tip of the cam. This configuration requires the guide roller located at the end of the piston rod to have as small a radius as possible. In both configurations, an important part of the device enclosure (designed for isothermal transformations), the greater the higher the 21 T, is used inefficiently, from this point of view
- the final adiabatic expansion can be done, after the end of the compression phase, inside the enclosure (by reversing the direction of travel of the piston and strongly accelerating it, followed by a strong deceleration to a stop, immediately followed by a new change of direction to carry out the evacuation of the gas from the enclosure. Throughout these stages, it is necessary to achieve very high accelerations and decelerations. For this reason, it is preferable that the return of the gas temperature to the Tamb value is done in an external expander. In the state of the art, most frequently, this cooling is done in a heat exchanger, at constant pressure, or directly in the storage tank, processes in which, if the thermal energy lost during cooling is not fully recovered, for its use in other useful applications, these processes occur with significant exergy destruction.
- by performing these two transformations outside the working enclosure, which requires the use of an adiabatic compressor and expander
A way of approximating these adiabatic jumps, by which large accelerations and decelerations are avoided, is described in Fig.3. To achieve the increasing adiabatic jump (with increasing temperature and pressure), the isothermalizer is provided with an additional enclosure 12.15, which communicates through a section as large as possible with the main enclosure. The additional enclosure 12.15 is constructed in such a way that, together with the surface S1 ; through which it communicates with the isothermalizer enclosure, it constitutes a compressor. It is equipped with the piston 12.19 which, at the top dead center (TDC), perfectly restores the surface cut from the wall 12.1 of the main compressor, allowing its piston 12.2 to move without obstacles. The axis of the isothermalizer is vertical, allowing a layer of liquid 12.22 to be deposited on the upper surface of the piston 12.1 , which has the role of occupying the entire dead volume, when this piston is at the top dead center. At the same time, this layer of liquid ensures the sealing of the enclosure, against possible liquid leaks past the gaskets of pistons 12.2 and 12.19. At the beginning of the suction phase, the piston 12.19 of the secondary compressor 12.15 moves from TDC to BDC, allowing the gas sucked in through the valve 12.12 to fill both enclosures. At the beginning of the compression phase, when the piston 12.2 moves accelerated, to reach the speed V/Z.o, the piston 12.19 suddenly moves, at high speed, to its TDC point and almost adiabatically compresses all the gas in the enclosure, bringing its temperature to the Tjz value and allowing piston 12.2 to rapidly reach isothermal velocity. When the piston 12.2 reaches the position T (Fig.3a), the gas pressure reaches the desired value and the discharge valve 12.13 opens, allowing the compressed gas to be transferred into the enclosure 12.16 of the adiabatic expander which, at this moment, is full of liquid. This expander is equipped with the piston 12.21 , which at this moment is in position Ti and the discharge valve 12.23, mounted at the highest elevation, through which the expander communicates with the storage tank under constant pressure. Therefore, transvasation has the effect of replacing the compressed air in the isothermalizer with an equal amount of liquid of the same pressure. The position Ti of the piston 12.21 is calculated so that in this position, the volume of the expander is equal to the volume of the compressed gas in the respective cycle. With piston 12.2 in this position, discharge valve 21.13 closes and intake valve 12.12 opens, allowing new intake and compression phases to begin in the main chamber. During these phases, the piston 12.21 of the adiabatic expander 12.14 moves suddenly, under the action of the spring 12.18, from the rest position T1 ; to its position BDC, and the gas in the enclosure expands adiabatically, its temperature reaching the Tamb value, after which the valve opens 12.23 and an appropriate amount of cold liquid replaces the gas in the expander. Then the piston moves again to position T1 ; returning to the storage tank the amount of liquid received in addition to the volume of compressed gas. Through these operations, in each cycle, a quantity of the heated liquid is replaced with an equal quantity of colder liquid. This does not exclude that, at high compression ratios, a constant flow of cooling liquid circulates through the isothermalizer enclosure (keeping in the enclosure the amount of liquid necessary to eliminate the dead volume). In the example in the figure, the pistons of the two adiabatic devices are actuated by elastic springs, which have the advantage of obtaining high accelerations, of the possibility of storing the mechanical energy consumed to compress the spring and its quick release. They allow the easy installation of simple mechanical devices, through which the BDC positions of the adiabatic devices can be easily changed during operation, changing the volumes of these devices and, consequently, the value of the temperature Tiz. Also, for systems in which the sponge has the role of storing thermal energy, automatic regulation systems can be designed in which the volume of the two adiabatic devices is adjusted according to the instantaneous temperature of the thermal sponge, or of the coolant.
Another improvement of the isothermal transformation process can be obtained by using, whenever possible, double-acting piston isothermalizers. These isothermalizers are composed of two independent enclosures, separated by a common piston: when the compression and discharge phases take place in one of the enclosures, the intake phase takes place in the other enclosure. As a result, there is some reduction in cycle time (with increased compressed air flow) and a considerable simplification of the profiled cam that guides the movement of the piston, by eliminating the steeply sloped portions, which makes it possible to permanently drive the piston with a power constant, eliminating shocks and vibrations. Another consequence of using the double-effect piston is the modification of the temperature variation curve, due to the longer duration of the intake phase, a phase in which the gas in the enclosure and the sponge interact thermally, which leads to changes in the optimal trajectory of the piston speed. Changes in the optimal trajectory are also caused by the presence of the piston rod in only one of the isothermalizer enclosures, a presence that changes both the gas volume and the total heat transfer surface in that enclosure. The magnitude of these changes must be determined by experiments with the actuation of the piston with automatic devices, controlled in real time, followed by the manufacture of appropriate cams.
We must also mention that in the case of thermal energy storage systems, the additional mechanical energy required to compress the gas to a higher isothermal temperature is stored entirely in the thermal sponge, in the form of heat.
Large compression ratios involve large differences between the instantaneous work done by the piston at the start and at the end of the compression process. For this reason, in these cases it is recommended to use a hydraulic drive, in which the force that causes the piston to move is the pressure of a fluid supplied by a hydraulic motor with variable speed (or variable flow). As an effect, the large difference in mechanical work between different stages of the process translates into large variations in fluid flow. In some state-of-the-art systems, reducing the flow gap between the different process moments is solved by the combined use of a solid and a liquid piston. In Fig.3b and Fig.3c, a telescopic solid piston, hydraulically actuated, is shown, with the piston at the BDC position and in an intermediate position, respectively. The master piston 13.2 slides inside the master cylinder 13.1. In the configuration shown in the figure, the first part of the piston stroke is divided, by the telescopic construction attached to the piston rod 13.3, into 4 segments of equal length and a segment of variable length, but both the number of segments and their length are at the discretion of the designer. The portion 13.3 of the rod is rigidly attached to piston 13.2 in its center, and an outer ring 13.4 is attached to the opposite end, larger than the diameter of the rod. Sections 13.3i (13.3a, 13.3b, 13.3c and 13.3d), are ring cylinders, which at the top have attached an inner ring 13.5i, and at the bottom have attached an outer ring 13.4i. These cylinders have an inner diameter equal to the outer diameter of the outer ring of the previous segment, and their outer diameter is equal to the inner diameter of the inner ring of the next segment. The outer rings of each segment slide on the inner surface of the next segment, and the inner rings slide on the outer surface of the previous segment, the seals 13.6 providing the sealing. In the configuration shown in the figure, the space between the bottom surface of the piston and the upper surfaces of the inner rings, as well as that between the outer surfaces of a segment and the inner surfaces of the next segment are vacuumed. Configurations may also be made, in which this space is occupied by a liquid or gaseous fluid at atmospheric pressure or a different one, if to this fluid is assigned an external reservoir and a series of flexible and fixed pipes for its proper circulation. A suitable range of piston speed trajectories can also be obtained from the combination of the motion of a solid piston driven by a mechanical device, with the additional movement of extending its rod, movement due to the hydraulic power of a liquid agent.
When the piston is in the BDC , the fluid supplied by a hydraulic motor penetrates through gate 13.7 and presses on the outer ring of segment 3.3, the surface of which is much smaller than the surface of the piston and, as the gas pressure in the densifier is reduced, the piston speed will be high. As the gas pressure in the densifier increases, the piston speed decreases. When this outer ring steps on the lower surface of the inner ring of the segment 13.3a, the displacement is also transmitted to this segment, which causes the working fluid to penetrate below the lower face of its inner ring. As the active surface increases, the piston speed tends to make an upward jump, but a suitable change in the liquid flow at this time keeps it constant, to decrease again as the piston moves. The upward jump of the active surface of the piston is repeated each time an outer ring of a segment steps on the lower surface of the inner ring of the next segment. When the segment 13.3d is driven in motion, the active surface becomes equal to that of the piston and its movement continues, without jumps, at a decreasing speed, until the piston power equals that required to compress the gas to the desired pressure.The engine power may be exceeded if the telescoping continues in the same way, with ring segments with the inner surface of the outer ring larger than the diameter of the piston (and the densifier cylinder), adding an additional cylinder with the corresponding diameter.
As it follows from relation (1 ), the power of an isothermal process depends, in addition to the initial pressure of the working gas and the compression ratio, on the isothermal temperature (established initially, depending on the desired energy efficiency) and on the speed with which it is removed /absorbed excess/necessary heat of the working gas, rate which depends on the device configuration. The second objective of the invention is to propose a series of design techniques for isothermalizer configurations, which will ensure them competitive power densities. The isothermalizers proposed in this invention use to a greater or lesser extent, according to the concrete applications served and the constructive variant chosen, one or more techniques for reducing the polytropic compression/expansion index, from the current state of the art, used as such , or with innovative modifications that increase the performances obtained.
One of these techniques, also used in quasi-isothermal devices in the state of the art, is to increase the surfaces of the compressor components that come into contact with the working gas: by changing the ratio between the diameter and the length of the device, by making the outer walls in a bellows-type configuration, or harmonic, which allows changing the length of the cylinder as a result of the displacement of the piston, without reducing the surface that is in contact with the working gas, by adding cooling fins on the inner faces of the piston and/or the cap, fins which, at displacement of the piston, they can interpenetrate, the introduction of deformable metal inserts inside solid piston devices, the introduction of non-deformable metal inserts inside liquid piston or gas piston devices (inserts that do not prevent the movement of this type of piston). The method to which the present invention appeals most often, due to its simplicity and effectiveness, is the use of a "thermal sponge", a type of deformable insert (12.6 in Fig.2), mounted between the inner face of the piston (solid or liquid) 12.2 and the cylinder cover 12.1. In isothermalizers with a high compression/expansion coefficient, the deformable thermal sponge will be supplemented with a non-deformable thermal sponge, with the largest possible absorption surface, mounted in an area outside the range of action of the solid piston, preferably in the area located at the highest elevation, an area where, due to convection, there is a tendency to accumulate gases with an above-average temperature. The deformable sponge is made up of one or more solid components (in many configurations a liquid component is also introduced, in a fixed amount) with variable volume and/or position. The solid components of the thermal sponge have the total surface Si, which is in direct contact with the working gas, approximately the same throughout the compression, and their degree of deformation is constantly controlled by the position of the piston, with each position of the piston corresponding to a different shape of the sponge, a property ensured by the elasticity of some of its component elements, or by kinematic devices controlled by (or in correlation with) the movement of the piston. The liquid components of a thermal sponge mounted in alternative devices can also play the role of a transport agent of excess thermal energy, if during the discharge and intake phases, they are replaced by cooled components, or they can take over the role of a liquid piston, if during thermodynamic transformation, the amount of liquid introduced is different from that discharged. In most rotary devices, thermal sponges made of solid elements are more difficult to make, but liquid elements of the sponge can be introduced and discharged, both from the inlet phase and during the thermodynamic transformation, in the form of a jet, in the form of drops, in powder form, in foam form, etc. When introduced/evacuated during thermodynamic transformation, they can combine the role of liquid piston with that of coolant/heater.
A compressor equipped with a solid elastic thermal sponge, in the form of a helical spring, having inside a channel through which a liquid flow circulates for the evacuation of excess thermal energy is disclosed in WO2014005229 - Temperature management in gas compression and expansion (US20140007569). This type of sponge meets some of the listed characteristics, but presents a number of disadvantages: the complexity of its technical realization, the overall low heat transfer coefficient and the low power density. Also, in some thermal engines, in some regenerators and in some heat exchangers from the state of the art, devices with constructive characteristics similar to the thermal sponges described previously are used, their temperature is variable in an oscillatory manner, they most often have the role of transfer heat from one area of the appliance served to another, and back.
For the densifiers and rarefiers described in this invention, more energy efficient solutions tailored to the application served are proposed. The easiest to implement is a sponge in the form of a multi- alveolar system, with cells that communicate with each other (foam obtained from elastic, metallic or non- metallic compounds, from natural rubber, synthetic rubber, from elastomers, from elastic polymers, from isoprene, etc. , deformable sheets, with large, regular or irregular surfaces). A more complex thermal sponge is made of materials with controllable deformation, with high thermal capacity, preferably metallic materials (these also have high thermal conductivity): various bands, or metal plates, elastic cords or bars, springs, spongy materials, membranes, fabrics from elastic and non-elastic materials, metallic and non-metallic, various products with open cells (which all communicate with each other), from other types of elastic or non-elastic metallic components, from inflatable bags with elastic or deformable walls, etc. The alveoli of such a system are constituted in the voids formed between the solid or liquid elements of the sponge. Ideally, the actual volume of the sponge VB, composed of the total volume of its solid components and the volume of the gas in the closed alveoli (disregarding the volume of gas in the open alveoli) should not change significantly, for any of the piston positions, not even after a large number of compression processes. Some small variations are acceptable, however, if they are oscillations around an average value.
In an isothermal compression, the mechanical energy transferred by the piston to the gas in the cylinder (whose average temperature is T,z) and transformed into heat is fully taken over by the component elements of the compressor (among which, those that are also in contact with the ambient environment, they transmit part of this energy to it) and by the thermal sponge (which is in contact with the component elements of the compressor on very small portions, the rest being in contact only with the gas in the cylinder). Therefore, if the active surface of the sponge is much larger than the active surfaces of the compressor, most of the excess thermal energy is taken up by the sponge, whose temperature gradually increases. If the thermal energy taken up by the sponge is not removed, the temperature T,z of the gas will not be able to be maintained at this value, unless the speed of the piston is gradually reduced accordingly. Due to the very small ratio between the specific gravity of the gas and that of the solid and liquid elements of the sponge, if the equation of motion viz(t) is constantly respected, the increase in their temperature (and, consequently, the temperature of the gas in the cylinder) is slow, an important number N of work cycles being necessary, for this change to be noticeable in the increase of the mechanical power absorbed by the piston actuation mechanism (N is the greater the mass of the thermal sponge is greater). If the densifier continues to operate without removing the heat accumulated by the sponge, the mechanical energy received from the outside by the piston is stored both as potential energy stored in the compressed gas reservoir and in the thermal sponge as internal energy. When the temperature of the sponge reaches high values, the energy stored in the sponge is equivalent to the potential energy stored in a reservoir of appreciable dimensions, containing compressed gas at an appreciable pressure.
These theoretical considerations justify two different strategies for using the densifier:
- no cooling system of the sponge, which allows the accumulation inside it of an appreciable amount of thermal energy and its subsequent use
- with sponge cooling system, which allows keeping it at a constant temperature Tamb, therefore keeping the gas temperature at a Tjz value.
It can be noted that devices equipped with a thermal sponge with a large absorption surface and efficient cooling systems can compress/expand, even at speeds comparable to those of polytropic devices in the state of the art, large volumes of gas without exceeding a gap temperature preset, acceptable in many applications. If the excess temperature is not discharged permanently, but at determined time intervals, the inclusion in the circuit of a compressor and an isentropic expander, with a controlled compression and expansion ratio, allow (apart from the role of executing the steps isentropics of the AIA cycle), by gradually changing this ratio according to the change in the temperature Tamb of the sponge, keeping the temperature difference AT=Tiz- Tamb and storing an appreciable amount of thermal energy in the relatively small volume of the sponge.
Regarding the construction of the thermal sponge, a very large variety of configurations can be realized (depending on the characteristics of the application in which the need for isothermal compression occurs). Figures 4 - 20 show some procedures for mounting a solid thermal sponge in solid piston isothermalizers (they can also use fixed amounts of liquid, used for lubrication, for cooling the sponge during compression and for the evacuation of the compressed gas remaining in the cylinder when the piston reaches TDC (having no role in gas compression).The procedures are identical for newly designed isothermalizers as well as for those obtained by modifying the configuration of prior art compressors.
The principle diagram of the isothermalizer equipped with a deformable thermal sponge is shown in Fig.3a. It consists of the housing 12.1 (composed of the cover with the check valves 12.12 and 12.13 and the side walls), the piston 12.2 and the thermal sponge 12.6. Check valves (one-way valves) are only one of the solutions for the intake and discharge of working gases. The complexity of the installation requires that, in many configurations, the circulation of gases, lubricants and cooling fluids be directed by a greater variety of taps, dampers, mechanically or electrically operated valves, etc., but in most drawings these will be represented only schematically, by the same symbol.
The operation of the densifier is the same as that of a sponge-free compressor: the gas suction is via the inlet valve, by moving the piston from top dead center TDC to the bottom dead center BDC, with the exhaust valve closed and compression to the desired pressure pf, by moving it from the bottom dead center BDC to the T point, with both valves closed, during which time the heat transfer from the gas to the sponge takes place. When the piston reaches this point, the exhaust valve opens so that the gas with pressure pf is exhausted to the desired destination, by moving the piston from T point to top dead center TDC. Most of the time, the gas is exhausted under constant pressure pf consuming only the displacement work Wd=pr- Vgf.
Fig.4 shows a very simple and very cheap type of isothermalizer with a thermal sponge, similar to bath sponges, made of elastic materials (various polymers, elastomers, various types of rubber, nets, or elastic metal foams), where, still from the manufacturing stage, measures are taken to increase the mechanical resilience and the heat transfer coefficient. The sponge is given the shape and dimensions of the isothermalizer enclosure right from the factory. The isothermalizer in the figure is cylindrical, and the thermal sponge will have the same geometric shape, but with a smaller diameter (calculated so that when its compression is maximum, its outer diameter reaches almost equal to the inner diameter of the cylinder of the isothermalizer, and with a height smaller (to achieve, in case of compression, a first polytropic phase and to allow an easier and faster movement of the piston at the beginning of the compression phase). In the simplest version, described in Fig.4B, when, after the compressed gas discharge phase, the piston reaches TDC, the sponge is compressed to the non-destructive limit allowed, and a small volume of gas remains in its alveoli (the dead volume), gas that expands, before the intake valve opens, when the piston moves to BDC. The cooling of the thermal sponge is done intermittently, when, after a fixed number of cycles, the measured temperature of the housing reaches a predetermined value, by replacing the gas in the enclosure with a cooling liquid (the ratio of the number of gas/air cycles can reach 1 :1000). A continuous cooling can be done, at the expense of the total volume of admitted gas, by introducing into the cylinder a quantity of liquid, the volume of which must be equal to the volume of the gas with the final pressure, liquid that needs to be recirculated and cooled, by any of the processes mentioned in the invention.
The one-piece elastic sponge can be replaced by a batch of elastic elements of much smaller sizes (the cells formed between these elements always communicate with each other) inserted in an elastic deformable bag, made of materials with high tear resistance, having a series of holes of reduced diameter, through which the gas inside communicates with the outside (the deformable bag can be replaced with a network with very small meshes).
Fig.4 and Fig.4A show a configuration of this isothermalizer, in which the cooling is permanent. Compared to the previous configuration, it additionally contains an additional thermal sponge 12.24, which may be non-deformable, made of a metal mesh (or foam), fixed to the cover of the enclosure, and at its lower part, a grill 12.25, or a small mesh net. The lower, deformable sponge (up to the limit where the volume of gas in the alveoli is very small) is provided with a series of vertical channels 12.27, with a small diameter, the gas in these channels being in communication with the neighboring alveoli of the sponge. In the upper part, these channels are closed by covers 12.26, which open and close simultaneously with the inlet valve 12.12. These channels can be helical elastic springs, hollow inside, made of a harder material, the sections of the coils being circular or ellipsoidal, embedded in the structure of the sponge since the manufacturing phase. In the rest phase, the height of the springs is equal to the height of the thermal sponge, and in the maximum compression phase (in which the coil sections are flattened ellipses) they have a height equal to that of the fully compressed sponge. The walls of the helical channel thus made are also provided with gas communication perforations. The axis of the cylinder is vertical, and a fixed quantity of liquid 12.22 is arranged on the upper face of the piston. As the piston moves toward TDC, the gas in the sponge alveoli and the vertical channels, as well as the gas outside it, are progressively compressed. As the liquid in the lower layer is absorbed by the sponge, the compressed gas in the alveoli, as well as that in the peripheral zone and that in the vertical channels (forced into the neighboring alveoli) is moved to the upper sponge, with a large absorption surface. When the piston reaches the T position, the gas reaches the desired pressure. At this moment, the alveoli of the lower sponge may contain a mixture of gas and liquid, or they may contain only liquid, and part of the liquid may pass through the grate into the upper sponge. When the discharge valve is opened, the enclosure is flooded, from the constant pressure tank (open system) with liquid of the same pressure, all the amount of gas in the cylinder, regardless of its location, being directed to the storage tank under constant pressure. In this configuration, the dead volume is zero. The piston continues its stroke until TDC (Fig. 4A), the position in which the lower sponge reaches the maximum degree of compression allowed and pushes the amount of liquid remaining in the sponge towards the discharge pipe. After closing the discharge valve, the inlet valve and the covers of the vertical channels are opened, all the liquid from the upper part drains to the lower one, and the working gas enters the enclosure.
In another constructive variant, the non-deformable thermal sponge is reduced to a ring with the outer diameter equal to the inner diameter of the cylinder and with the inner diameter equal to the outer diameter of the deformable sponge, the intermediate grill being no longer necessary but, to regulate the degree of compression of the elastic sponge, on the underside of the upper cover of the cylinder, between it and the upper surface of the deformable sponge, a plate with the same horizontal section as that of the sponge is mounted. When the piston moves towards the TDC position, which is at the lower limit of the ring represented by the non-deformable sponge, the liquid on its upper face is progressively absorbed by the deformable sponge, so that, in this position, all the amount of gas and liquid is in the non-deformable sponge. In this position, the opening of the discharge valve leads to the flooding of the annular sponge by the liquid from the pipe connected to the storage tank (under constant pressure), the compressed gas being pumped into the tank. After closing this valve and opening the suction one, the movement of the piston causes the entire amount of liquid to flow out on the upper face, where an amount equal to that coming from the tank is removed. One of the possibilities to eliminate this quantity is to create a small reservoir in the side wall, immediately above the upper face of the piston, with a volume equal to that of the gas being pumped, a quantity that remains here, until the side surface of the piston passes this cavity, after which it drains into the crankcase.
The liquid used for cooling must have a low adhesion to the material of the thermal sponge, for which purpose a series of additives can be added to the liquid and/or to the structure of the thermal sponge. In this configuration, very good results are obtained if the coolant has a high specific gravity (e.g. mercury), has the highest surface heat transfer coefficient, the highest thermal conductivity, and the contact angle between the liquid and the material used to make the sponge should be less than 90°.
The isothermalizer in Fig.5 illustrates a configuration in which: the thermal sponge 5.4 of this isothermalizer is a helical elastic metal spring with the rectangular turn section, one end of the spring being fixed to cover 5.1 , the other to piston 5.2. With the piston in the BDC , the spring 5.4 is in a detensioned state (or slightly pretensioned). In the illustration, the piston is in an intermediate position. We note that even in the situation where, with the piston in the TDC, the spring turns are in perfect contact (the tension is maximum), the dead volume of the isothermalizer (which includes a central cylindrical space with a diameter equal to the inside diameter of the spring) is quite large. It can be reduced by introducing additional springs with ever smaller diameters into this space (the inner diameter of an intermediate spring is approximately equal to the outer diameter of the next spring), increasing the heat transfer surface. Another method is that proposed in US20140007569, in which a cylindrical piece with a diameter almost equal to the inner diameter of the helical spring with the smallest diameter (Fig. 5a) is attached to piston 5.2 (or to the cylinder cover 5.1 ). The length of this part is chosen according to the densifier compression ratio. At the limit, when the piston is in the TDC, it can fully occupy the volume inside the fully compressed spring. In this device, on the compressed gas discharge pipe (which, most often, connects to the storage tank under constant pressure, a tank in which its lower part contains a hydraulic fluid) a quantity of liquid is permanently kept, sufficient to, upon opening the exhaust valve 5.5, to completely replace the amount of gas remaining in the cylinder. In this way, the dead volume of the device is reduced to zero, and a quantity of liquid enters the cylinder, sufficient to absorb the thermal energy accumulated by the sponge during one cycle. After closing the discharge valve, during the intake of fresh gas, the liquid washes the coils 5.4 of the thermal sponge. For more efficient washing, these coils are provided with a series of 5.4o holes. At the end of the suction phase, almost all the amount of liquid accumulates on the upper face of the piston, from where it is evacuated, with the help of a pump, through the pipes 5.2c. If the amount of liquid introduced through the discharge valve is not sufficient for effective cooling, additional liquid can be introduced through these pipes, during the compression phase, or a constant level of liquid on the piston can be maintained.
Another constructive variant is represented by the isothermalizer in Fig.6 and the one in Fig.7, where the main components of the thermal sponge are horizontal parallel plates (preferably metal) 5.1 1. These plates shall be attached to a vertical elastic rope system 5.7a, inserted into the elastic bellows 5.7b, respectively to a vertically mounted helical spring system 5.12, with an outer diameter smaller than that of the isothermalizer springs in Fig.5 (where they were the main component of the thermal sponge). The space inside the elastic helical springs 5.12 (circular, rectangular, etc. section) can be occupied by a sequence of springs 5.13, with appropriately decreasing diameters, and the space inside the spring with the smallest diameter can be partially occupied by a cylinder similar to cylinder 5.2a of Fig.5. In Fig. 7 a guide rod system 5.7c attached to the piston is proposed, which pierce the cover through holes fitted with sealing gaskets 5.8 (or vice versa). If tubes are installed instead of rods, through which cooling fluids circulate, the problem of removing the excess heat can be solved completely or partially. These tubes must be provided also with sprinklers to spray droplets of coolant, or aqueous foam. A process that achieves the same objective is the use of thermal tubes with vapor at the saturation limit. These rods also eliminate the possibility of side movements of the thermal sponge and the possibility of it coming into contact with the cylinder walls.
In order to keep the elastic properties of springs 5.12 unchanged, it is most often necessary to fix plates 5.11 by means of another deformable material, which can be rigidly fixed to horizontal plates 5.12a. This coating can be continuous, over the entire length of the turns, or it can be made of rings 5.12a, mounted one for each turnof the elastic spring. The fitting must be made in such a way that the horizontal plates are not subjected to mechanical stresses when, due to the compression of the spring, the outer diameter of the spring undergoes slight changes. The fixing of each plate 5.11 on each of its support springs 5.12 is done at a single point. To allow the free movement of the plates, at each spring, crossing windows are practiced in this plates. The windows that provide the largest absorbent surface of the board are circular, with a diameter equal to the maximum diameter of the support spring. In Fig. 7a, the fastening system is illustrated when the springs are located at the edge of the cylinder, and in Fig. 7b
1000 the fastening system when the springs are inside them. The number of plates mounted in these systems can be very large, thus ensuring a large heat transfer surface and, consequently, a high piston speed, or a very small difference between the temperature of the gas and that of the thermal sponge. At the limit, when the piston is at TDC, the horizontal plates can be very close together, even touching each other (Fig. 7c). Even when the horizontal plates come into contact with each other, many spaces remain filled
1005 with compressed working gas. For this reason, especially in the case of large compression ratios, it is appropriate to execute horizontal channels and vertical holes, so that, in the discharge phase, these spaces can be filled with a liquid agent, and in the suction phase, it can be evacuated, with minimal energy consumption.
In order to ensure a maximum density of the plates, the number of elastic cords 5.7a (Fig.6),
1010 respectively of the support springs 5.12 (Fig.7) can be multiplied, they form distinct sets, and the corresponding sets of horizontal plates 5.1 1 are mounted interspersed on these.
Figure 8 shows an isothermalizer configuration in which, starting from the configuration in Figure 7, the surface area through which heat absorption, by the thermal sponge, from the gas being compressed is considerably increased by the installation of fins, or other vertical elements 5.10. This
1015 leads to very high compression ratios without the excessive increase in the working gas temperature. The density, the arrangement on the support plate, the thickness of the plates, etc. may differ from one horizontal plate to another. There is a very high degree of freedom in the shape of these fins, their dimensions (a large thickness ensures a slower increase in the temperature of the sponge, a smaller distance between the fins ensures better cooling of the gas, a larger width of these reduces the required
1020 number of horizontal plates, a larger diameter of the holes practiced in both vertical and horizontal plates ensures efficient convective circulation and less friction losses in the liquid to be inserted at the end of the compression, for the exhaust of compressed gas, etc.). This variety also provides multiple options for choosing the isothermal trajectory viz. The smaller number of horizontal plates ensures that the horizontal stability of the thermal sponge is ensured by guide rollers, with little friction in the bearings. In Fig.8, the
1025 thermal sponge of such an isothermalizer is shown, with the piston in an intermediate position, and in Fig.8a is the same sponge with the piston in the TDC.
Configurations may be made in which vertical fins are walls separating laterally distinct areas of the cylinder (in a horizontal section, they are a sequence of concentric circles, or rectangles with increasingly smaller sides, or other geometric figures placed in each other, or side by side). In this case,
1030 the horizontal plates have the raised edges along the entire contour (5.1 1c, figure 8), like trays. This allows a liquid layer to accumulate permanently or periodically on the horizontal plates (5.11 d, figure 8b). The height of the tray edges will determine the size of the liquid fraction of the sponge 5.1 1 a. The vertical movement of the piston causes the vertical fins to move, penetrate the liquid layers on the horizontal plates and discharge liquid into the liquid layer formed on the top surface of the piston. In this way, the
1035 liquid is circulated among the plates, favoring the transfer of heat.
In another configuration of the isothermalizer in Figure 8 the thermal sponge is made without the elastic springs between the horizontal plates. In this way, the action of the piston is transmitted from one horizontal plate to another successively, not simultaneously. In this configuration, mechanical or electromechanical locking-unlocking mechanisms of horizontal plate movement can be installed in the
1040 side walls of the densifier, allowing a diversification for the shape of the curves of the isothermal velocities. If the liquid fraction of the thermal sponge remains, during a cycle, at the same value, holes for gas circulation can be applied in the vertical plates. If a permanent flow of the liquid fraction is ensured and the size of this fraction can be changed and controlled by changing the flow rate of fluid introduced and discharged from the cylinder, holes can be also applied in the horizontal plates,. This results in the
1045 change of both the speed variation curve and the time and space distribution of the heat energy accumulated by the thermal sponge.
The isothermalizers in Fig.9 are isothermalizers provided with horizontal plates 5.11 , fixed on the helical springs 5.12, inside which the springs 5.13 are mounted, with progressively decreasing diameters, and the space inside the spring with the smallest diameter may be partially occupied by a full cylinder. A
1050 system of guide rods 5.7c, fixed to the piston, are mounted in the enclosure and pierce the cover through holes secured with gaskets 5.8 to prevent uncontrolled lateral movements. In order to ensure a large absorption surface as well as a large thermal energy storage capacity, the number of horizontal plates is less and the distance between them is greater, but the space between the plates is interlaced with other helical springs 12.26, of lower height. A fixed amount of liquid 12.22 is arranged on the upper face of the
1055 piston, which has the purpose of replacing the compressed gas, reached the desired pressure, and sending it to the storage tank. When the piston is at TDC, all the springs are fully compressed and the spaces between them and those inside them are completely filled with fluid. To achieve high compression ratios, to provide easy paths for liquid movement and to create convective gas currents, horizontal channels and vertical holes are executed in horizontal plates. That the gas in the upper part of the
1060 enclosure is warmer, in this area it is recommended to mount an additional sponge 12.24, made of metal foam, or of dense metal networks, sponge with superior heat absorption capacity. The isothermalizer in Fig. 9A is similar, but the fixing of the horizontal plates is done on deformable supports of the bellows type (harmonica) 5.22, 5.24.
Fig.10 shows a configuration of an isothermalizer whose thermal sponge is made with telescopic
1065 vertical elements 5.10. The lateral displacement tendencies of the horizontal plates are eliminated by fixing elements 5.1 Od of each plate of one or more lower wings of the telescopic sections 5.10, or by guide rollers, which slide on the side walls of the enclosure, mounted on the periphery of each of plates. In this way, the action of the piston is transmitted from one horizontal plate to another successively, and when the piston is withdrawn, the sections return to their maximum length, due to their weight. In this 1070 way, the density of the vertical plates is higher in the lower part of the enclosure. For this reason, the distance between the horizontal plates of the sponge, the number of telescoping sections and their height must be reduced towards the top. In addition, a high-density thermal sponge 12.24 (metal fabric, metal foam, etc.) is mounted between the last horizontal plate and the cylinder cover, which takes the compressed gas from the sponge with plates, as the gas space in this sponge is greatly reduced , and on
1075 the upper face of the piston a layer of liquid 12.22 is placed, consisting of an appropriate amount of lubricant, or heat transfer liquid, whose volume must equal the dead volume when the piston is at TDC. Instead of using this amount of liquid, a smaller amount can be introduced (remaining in the chamber throughout the compression phase), but in the walls of the chamber, or in the piston, one or more holes are made, sealed by a valve, which opens when the lower sponge is fully compressed, allowing a quantity
1080 of liquid (acting as a piston) to enter the cylinder, which will continue to compress to the desired pressure, then expel the compressed gas. The amount of residual liquid remains, in both configurations, during compression with the solid piston, at the same value, even if there is a flow of coolant through the device: during this time interval, the inlet and outlet flow are equal. Between the vertical and horizontal plates, there are holes for gas and liquid circulation.
1085 In Fig. 10a, an embodiment of the vertical telescopic fins is exemplified. In Fig. A is represented, in an intermediate position of the compression phase, a vertical section (1 -1 ) through one of these fins for the case where the horizontal section through the "cylinder" of the densifier is rectangular. In Fig.C is represented a vertical section through an intermediate section of the fin, and in Fig.B, a horizontal section through the fin. It is noted that each section 5.10 of the wing (minus the lower one, which is a simple plate
1090 with a stop) is a rectangular parallelepiped with a hollow interior. The thickness of the walls is chosen according to the purpose of the isothermalizer, being thicker in the case of isothermalizers with the accumulation of thermal energy. Each section (except the first, where it is not necessary) has in the upper part some lateral talons 5.10b, which slide through a series of corresponding channels 5.10a, practiced in the inner walls of the upper section, having a length equal to the height of the section, less the stopper
1095 thickness. When the piston is at TDC, the plate density is very high.
State-of-the-art diaphragm compressors, in any of the constructive variants, can become isothermal by easy-to-make modifications, by inserting a thermal sponge according to the invention into the working chamber and by adding an cooling/heating system based on a heat transfer fluid. Fig.11 shows a diaphragm densifier consisting of the upper housing 5.1 , the lower housing 5.1 b and the elastic
1 100 diaphragm 5.33. In the configuration shown, the densifier is operated directly by the piston 5.2, but can also be operated by means of a volume of hydraulic oil, in which case the housing 5.1 b has perforations for the oil circulation. The shape of the two housings is modified, the enclosure between them having a shape close to that of a rectangular parallelepiped, with "softening" the edges, which allows aspiration of a larger volume of gas and offers more choice for the type of thermal sponge. In Fig.1 1 we chose a
1 105 sponge composed mainly of flat metal plates 5.1 1 , supported on supports mounted on the carry-supports of harmonic type, composed of flashplates 5.22, 5.23 and 5.24. The thermal sponge can also have a permanent liquid component, with the role of avoiding the formation of a dead volume, and an itinerant liquid component, that with the help of 6.9b sprinklers, cools the gas, subject to compression. This component can also be used as a liquid piston, with flow rate adjusted in such a way as to obtain an110 isothermal speed for the compression.
In order to better control the volume changes of the enclosure, the diaphragm 5.33 is mounted between two metal plates, 5.30 and 5.32, and is rigidly fixed to the two plates, by means of the plates 5.34, along a median axis, the outer edges of the the diaphragm being rigidly fixed between the two halves of the housing. The free part of the elastic diaphragm is extensible under the action of the piston,115 and the part between the two plates can slide on some rollers 5.31 .
The isothermalizers in Fig. 12, Fig.13 and Fig.14 also have in their composition thermal sponges made of elastic and inelastic metallic components which, when the piston is in TDC, occupy almost entirely the internal volume of the isothermalizer. The one in Fig. 12 is constructed from horizontal plates 5.1 1 , which rest on a system of lamellas 5.14a and 5.14b or elastic half-plates 5.14. When fully tensioned,120 two paired half-plates 5.14 almost entirely cover the surface of a horizontal plate, as does the group of plates 5.14a and 5.14b installed between two horizontal plates. From Fig.12A it is noted that part of the blades are fixed (for example, by welding 5.14c) to the horizontal plate below it at the periphery of the plate, while the other blades, having reverse curves, are arranged alternately with the first blades and are fixed near the midline of the plate. In the case of half-plates, they can be welded in any of the two listed125 positions. When in this type of isothermalizer, the piston is at TDC, the space not occupied by the thermal sponge of the isothermalizer enclosure, in the case of precision machining of plates, lamellae and halfplates can be made very small. This space is occupied by the dead volume of the compressed gas. or by the coolant 12.22 arranged on the upper face of the piston.
The one in Fig. 13 is made by alternating flat plates 5.1 1 , which slide on an equal number of130 arched plates 5.14, so dimensioned that when the piston is at top dead center, all these plates overlap one another and occupy as large a part as possible from the cylinder of the device. The whole assembly is stabilized by a rod 5.7c, which has one end fixed to the cylinder cover, and the other end pierces the piston through a hole made in the piston and sealed with the gasket 5.8. The rod 5.7c, in all configurations in which it is used (eg Fig.7, Fig.9) has a non-circular section and is located in the axis of135 the cylinder, preventing the rotation of the stabilized plates. Since the plates rub against each other during the movement of the piston, a larger amount of lubricant, judiciously distributed, is necessary to lubricate and cool the piston. In order to achieve this objective, the cover of the device is provided with a series of holes distributed on the contour of the cover, holes to which the supply pipes with lubricant 5.2c are connected, and the sponge plates are provided with a series of holes 5.1 1 o and 5.14o, distributed in such140 a way that the oil is distributed to all regions where friction occurs. The lubricating liquid drains on the plates and accumulates on the upper face of the piston, from where it is evacuated with the help of a pump, through the pipes 5.2c and is sent to the cooling bath. A quantity of lubricant is introduced or kept permanently on the piston, which will supplement the cooling power and which will reduce the dead volume of the device to zero. 145 The isothermalizer shown in Figure 14 is similar, but the solid sponge plates are made by the welded connection 5.15a between the rigid planar plates 5.1 1 and a series of arched elastic bands 5.15, among which they can be mounted (on a single level, if in the TDC the arched plates become flat, or on several levels if in the TDC the arched plates maintains adequate curves) horizontal plates with corresponding widths, with two slip points (bolts mounted on the elastic bands, sliding into grooves150 executed in the thickness of the horizontal plates). The compressed gas is collected in the space between the horizontal plates, as well as in an inner parallelepiped space 5.2b. The dimensions of the collection space are set by the width of the arched strips 5.15, and the height of the inner parallelepiped collection space is adjusted by the dimensions of the piece 5.2a, fixed on the moving piston. In section 1 -1 , an inner, top view of the system is shown. 155 The isothermalizer in Fig.14a is similar to that of Fig.13, but arched plates with a different number of curves with different radii of curvature, are mounted between the horizontal plates. As in previous configurations, the highest absorption power is obtained when all the plates have the same surface, close to the section surface through the cylinder, and in the TDC they perfectly overlap.
This type of thermal sponge can be used to reduce the energy consumption of state-of-the-art160 compressors having a superunitary polytropic index, compressors for which the main objective is not to achieve an isothermal compression, but to obtain a large volume of compressed gas in as short time as possible. This objective can be achieved in a more economical way than at the state of the art (where the desired compression ratio is obtained by staged compression, intercaling some heat exchangers between these stages), by inserting a thermal sponge with a maximum absorption surface into the compressor,165 obtained with heat-accumulator elements having a minimum volume, associated with a continuous flow lubrication system, which also takes over the sponge cooling function and reduces the dead volume as much as possible when the piston is in the TDC. In addition, the introduction of a piston actuator system which (at a compression cycle time equal to that of a conventional compressor) introduces a variable piston speed, higher in the exhaust, in the suction and in the first part of the active piston stroke, and170 smaller toward the end of the compression process, it further reduces energy consumption and also makes the cooling system more efficient.
To any compressor with a superunit polytropic index in the prior art, the energy consumption necessary to obtain a given compression ratio may be reduced, in a given time, if a properly sized thermal sponge is inserted inside its cylinder. Most of the time, this involves some constructive changes to175 the original compressor (for example, shortening the stroke of the piston, or lengthening the useful part of the cylinder, with a Gb value, equal to the thickness of the sponge in a fully compressed state, as well as the adequacy of the lubrication system to the new requirements). In the conditions of the evolution of energy prices and the objectives of reducing thermal pollution and noxious pollution, the expenses necessary for these adaptations will be rewarded. 1 180 For isothermalizers can also be made configurations without elastic components. The isothermalizer in Figure 15 (horizontal section through a vertical cylinder with a rectangular section) consists of a thermal sponge made of metal plates 5.11 , made with a thickness as small as possible (if a high power of the isothermalizer is desired), but large enough for the plates not to be subjected (due to their own weight, or too sudden movements) to some residual deformations. To ensure that the plates
1 185 have a stable horizontal position and a variable value spacing (dictated by the position of the piston), a sufficient number of movable carry-supports 5.19 are fitted inside the cylinder, located in close proximity to the cylinder walls, in such a way that the movements of either of them would not impede the movements of the others, nor the movement of the plane plates. In Figure 1 1 , section 1 -1 is a vertical section through the cylinder, executed in the area where the carry-supports are mounted. Each carry¬
1 190 support is made in the form of blades, or narrow rods, on whose inner side (facing the inside of the cylinder) are mounted, (by welding, riveting, embossing, etc.) supports 5.20 of the plates 5.1 1 , made of sheet metal, wire, pieces processed by machining, etc. On each port-support is mounted, at different levels (usually equally spaced), a number of supports equal to the total number of plates, or equal to the number of plates in a set, if the interlaced plate technique is used. One end of the carry-support shall be
1 195 secured, by means of a movable joint, to a fishplate 5.18 rigidly attached to the piston. On the other end of the blade, a short swivel arm is attached, also via a movable joint, which has a guide roller 5.16 attached, which can run on a rail, or in a channel 5.17 of the cylinder cover.
The horizontal plates 5.11 are rectangular, occupying almost the entire horizontal section area, but they have practiced in the corners a series of cuts to avoid collision with the carry-supports and
1200 supports on the neighboring levels, as well as to create the consoles 5.21 that are laying on the supports on that level.
When the piston is at the BDC , the carry-supports make the minimum angle (almost 0°) with the vertical axis, and the distance between the plates is maximum. As the piston moves, the angle made by the longitudinal axis of carry-supports whith the vertical axis increases, and the distance between the
1205 plates decreases. When the piston is at the TDC, the carry-supports make the maximum angle (almost 90°) with the vertical axis, and the distance between the plates is minimal. When carefully processing components, the plates can perfectly overlap without intermediate spaces, ensuring a small dead volume and easy circulation for the fluid intended to replace this gas.
Fig.16 shows a horizontal section through the cylinder of an isothermalizer which also consists of
1210 a thermal sponge made of very thin metal plates 5.11 layed on a carry-supports system. Here, the carrysupports are made of a sequence of pairs of fishplates 5.23 and 5.24, placed superimposed in the same vertical plane. Both fishplates of these pair have a centrally located hole through which a pin passes, around which both fishplates can rotate. At the same time, this pin, having a corresponding length, can be the support for one of the horizontal plates 5.11. In another configuration, the length of the pin is
1215 approximately equal to the thickness of the two fishplates and is an empty cylinder, which constitutes a bearing for the support attached to the plate 5.1 1 by a rigid joint. The ends of the fishplates are coupled by moving joints, with two other pairs of fishplates (one lower and one upper). The extreme fishplate pairs are shorter and are coupled by moving support joints 5.22, one fixed to the piston, the other to the cylinder cover. In the “magnifying glass” of Fig.16A a front view of the carry-supports system is shown in
1220 the position corresponding to the piston at the TDC, and in section 1 -1 a front view of the entire sponge corresponding to the piston in an intermediate position. In this configuration, a high density of horizontal plates 5.11 can be ensured by increasing the number of carry-supports, associated with a process of interspersed mounting of horizontal plates.
When the piston of an isothermalizer is at the TDC, the compressed gas is unevenly distributed in
1225 the volume of the cylinder. Significant volumes of compressed gas (so-called “dead volume”) may remain inside the sponge and in the space between the sponge and the cylinder walls, which will drop to the pamb presure once the piston is moved in reverse, before the inlet valve is opened. This fact, like the situation with the compressors in the state of the art, it leads to a decrease in the final flow of compressed gas, a decrease as much important as the compression ratio of the compressor is greater. Careful selection of
1230 the shape and dimensions of the thermal sponge components, with taking into account their modifications during the processes in the densifier, must lead to the achieving of the most regular shapes of the volumes in which the compressed gas is found, when the piston is in position T and a minimum dead volume when the piston is in the TDC position. The gas is now exhausted by moving the piston from position T to the TDC position, where the dead volume reaches the lowest value.
1235 A total elimination of the dead volume can be achieved, as in the principle diagram in Fig.17a, by introducing into cylinder 5.1 a, as early as the initial phase, a liquid phase 5.3b of the 5.3a thermal sponge, consisting of an appropriate amount of lubricant, or a heat transfer liquid, the volume of which equals the dead volume. At this densifier, when the 5.2a piston reaches the T position, the 5.5a output valve opens (due to the valve adjustment, or due to a command received from the control system), located at the
1240 highest enclosure elevation. From this point on, the piston movement causes the compressed gas to transfer into the discharge line, which is completed when the piston reaches the TDC position, where only the thermal sponge (solid and liquid phase) remains inside the cylinder.
Another possible configuration for the exhaust of compressed gas is shown in Figure 17b, in which 5.1 a is a small densifier, whose inlet window 5.6a is at the same time the discharge window for a
1245 larger densifier 5.1 , with which it has a common wall. This mini-densifier is equipped with the 5.2a piston and a thermal sponge made of flat plates 5.11 a. Moving the piston 5.2 from the TDC to the BDC leads to the inlet of the working gas at p, pressure in both cylinders. The first phase of compression is achieved by moving piston 5.2 from the BDC to the point T, interval during which the volume of gas in the densifier 5.1 a does not change, but the gas in this cylinder is compressed in the same relation to the gas in
1250 cylinder 5.1 , and its sponge contributes to the accumulation of excess thermal energy. The compression conditions in the two densifiers being different, will be different throughout the compression period, also the temperatures of the gas and of thermal sponges they contain. The liquid fraction of the sponge in the densifier 5.1 can be chosen so that when piston 5.2 reaches the TDC , it completely occupies the volume of the cylinder not occupied by the solid fraction, without entering in cylinder 5.1 a at all. At this point, all
1255 the initial volume of gas in the two cylinders is transferred to cylinder 5.1 a, and its pressure reaches the final pf value. In cylinder 5.1 a, the gas can be subjected to a new isothermal compression stage, or it can be exhausted into the storage tank by moving the 5.2a piston from the BDC to the TDC. In applications where severe purity conditions are imposed on the compressed gas, the thermal sponge of the densifier 5.1 a shall be carried out only with a solid fraction, so that the dead volume is as low as possible. In
1260 cylinder 5.1 a, the gas can be subjected to a new isothermal compression stage, or it can be exhausted into the storage tank by moving the 5.2a piston from the BDC to the TDC . In applications where severe purity conditions are imposed on the compressed gas, the thermal sponge of the densifier 5.1 a shall be carried out only with a solid fraction, so that the dead volume is as low as possible.
In Fig.18, the compressed gas exhaust operation is carried out, with the piston in position T,
1265 through the window 5.6r and the pipe 5.6c, which connects directly to the storage tank located at a higher elevation (or with another useful destination) and which contains, at its base, liquid from associated hydraulic circuit, with pressure pf . If liquid is found in the pipe, the 5.6a window serves only for gas suction. In this case, when the piston, in his movement to the TDC reaches the point T, opening of the window 5.6r allows liquid from the 5.6c pipe to enter the cylinder and replace the compressed gas at the
1270 pressure pf This, due to archimedic forces, reaches the top of the storage tank, being replaced by an equal volume of liquid. This volume of fluid is then exhausted into the associated hydraulic circuit, for example through a valve fitted in the piston, or is exhausted back into the pipe, by moving the piston from point T to the TDC . Acest volum de lichid este apoi evacuat in circuitul hidraulic asociat, de exemplu printr-o supapa montata in piston, sau este refulat inapoi in conducts, prin deplasarea pistonului din
1275 punctul T in TDC. When the piston reaches the TDC and the exhaust window closes at the control system command, the exact amount of liquid required to remove the dead volume remains in the cylinder. This amount of liquid remains permanently in the cylinder as a liquid fraction of the thermal sponge. This process can also be applied to solid piston compressors of the state of the art. Both valves 5.5 in Figure 4 have been replaced by the wide windows 5.6a and 5.6r created in the side walls of the cylinder (in the
1280 case of a cylinder with a rectangular section, the window width may be equal to the thickness of the sponge when the piston is in the TDC, and its length can be equal to the width of the wall), which allows for rapid circulation with reduced losses of exergy. The figure shown the piston in TDC , with the suction valve 5.6a open. The valve remains open until the piston reaches the BDC .
When, for the purpose for which the isothermalizer is used, it is useful to recover a larger fraction
1285 of the energy supplied to the system by means of the piston, may be implemented compressed gas exhaust devices, with the recovery of the thermal energy accumulated by the thermal sponge. In Fig.18A is shown such a process applied to the isothermalizer in Fig .5. The thermal sponge of this isothermalizer is composed of a solid fraction (a rectangular section helical spring) and a liquid fraction that completely eliminates the dead volume of the cylinder when the piston is in the TDC. Gas suction and exhaust are
1290 made through valves 5.5 in the cover of the isothermalizer (Fig.5), or through the 5.6c windows in the side walls. As we mentioned above, in the case of a densifier that is not equipped with a cooling circuit, the temperature of the gas (and implicitly, the mechanical work required to compress it during a cycle) and of the thermal sponge increase progressively with the number N of compression cycles. The heat the gas receives is accumulated, along with the mechanical energy received, in the compressed gas storage
1295 tank. After a sufficiently large number N of compression cycles, when the temperature of the sponge reaches a convenient value, the densifier sponge can be extracted entirely from the densifier, stored in an isolated enclosure and replaced with an identical sponge having the temperature of Tamb . This is possible, for example, if the cylinder has a rectangular section, the 5.6c side windows and the caps closing them have the width equal to the compressed sponge and the length equal to the side wall and if, at the
1300 moment immediately before extraction, the side caps 5.6c and the edge plates of the sponge 5.4a shall be mechanically coupled together so that they can be translated, sliding on the surface of the piston then on the outer rails (for example, pushed by a 5.2d piston, or by towing).
Another way to extract the “overheated” sponge is to extract the piston fully, coupled with the sponge, through the casing of the device. After extraction, the piston-sponge assembly is cooled in a fast
1305 cycle, or is stored and replaced with another identical one, having the temperature Tamb .
These processes for the reuse of excess thermal energy can be used in temporary storage systems for the renewable energy. In other types of uses, thermal sponge cooling is one of the most important problems of building high-performance densifiers. The thermal energy absorbed by the thermal sponge can be eliminated, depending on the characteristics of the densifier, by any of the prior art
1310 procedures listed in the previous paragraphs.
For compression/expansion systems in simple installations, when the compression/expansion ratio is low, a lubrication system combined with a lubricant cooling system leads, after a transitional regime, as in polytropic compressors, to a stabilization of the sponge temperature at a Tb value, higher or lower depending on the coolant flow and the coolant temperature. The presence of solid thermal sponge
1315 due to the additional heat absorption surface and the transfer to the lubricant (in this case the total heat transfer coefficient between the solid sponge and the liquid sponge is much higher than that between gas and lubricant) results in a lower value of Tb temperature and, implicitly, in a lower instantaneous work. In the case of large compression ratios, the temperature value of Tfccan be maintained at a low value if, e.g., the system operation is supervised by an controller which, at predetermined intervals (or dictated by a
1320 feed-back adjustment system) stops the piston movement and leaves running, for a short time, only the cooling system and the lubrication system. During this time, the thermal sponge exchanges heat with the liquid in motion, and the densifier turns into a heat exchanger. To avoid too frequent stops, the on-off mode can be replaced by a multi-speed mode.
Similar to the thermal energy absorption system used in screw compressors, a volatile liquid can
1325 be used for lubrication, which at Tiz temperature to generate a high concentration of in suspension particles. These particles have a great ability to limit the temperature rise above the working temperature of Tiz. When certain purity performance of the compressed gas is required, or for reasons of simplification, the lubrication function is taken over by a coolant circulating through densifier with constant
1330 flow or, it is inserted intermittently into the cylinder (once for a number of N cycles). During the cooling operation, the piston can improve the efficiency of this operation by means of short back and forth movements. Also, in densifiers where the gas in the apparatus is exhausted by replacing it with the hydraulic agent in the compressed gas storage tanks, this agent can be included in a cooling circuit, continuously or intermittently and can also take over the cooling function of the thermal sponge. For
1335 example, for the densifier in Fig.17a, the elimination of excess heat is done by replacing the compressed gas with colder liquid during the exhaust operation (the warmer liquid being subject to upward forces). The stopping time of the liquid agent in the compressor can be prolonged (periodically or every cycle) by the commands sent to the piston by the controller. Another possibility of replacing this fraction, increasing the flow rate of the gas, is to remove the remaining fluid in the cylinder, during the fresh gas suction
1340 operation, by opening a valve located in the piston (with the liquid spilling into the densifier housing), or by absorbing it through a component pipe of a cooling circuit equipped with a suitable heat exchanger. If the temperature of the liquid is lower than the temperature of the sponge, it will take up and discharge some of the thermal energy accumulated by the sponge, the higher the longer it stays in the cylinder.
With the imposition of higher performance criteria for the isothermalizer, the performance required
1345 for the thermal sponge cooling system increases. This is achieved by implementing more efficient cooling systems. The state of the art proposes a wide range of such procedures: by continuously introducing a coolant aqueous foam, with the elimination of excess fluid in each cycle, by continuous or intermittent spraying (toward the end of the compression cycle) of a coolant, or by any other method which advantageously combines the action of the solid piston with that of a liquid piston. The liquid piston
1350 appears whenever the instantaneous flow of coolant inserted is greater than that exhausted. If the flow rate of coolant circulating through this circuit is correspondingly correlated with the instantaneous gas pressure in the cylinder and the instantaneous piston speed, the compression carried out is almost isothermal.
As an example, in the densifier in Fig.5, in the empty space in the center of the helical spring 5.4
1355 a tubular helical spring can be mounted, coupled by flexible tubes to an external cooling circuit with heat exchanger, through which, under the effect of a hydraulic pump, a coolant circulates. The hollow spring is fitted with horizontal spray nozzles. It introduces suspended particles between the metal plates, which considerably increases the cooling speed of the gas and of the thermal sponge. Excess fluid accumulated above the piston is eliminated in continuous flow. This system can also be applied to densifiers in Fig.6,
1360 Fig.7, Fig.8, Fig.9 and Fig.10, if the horizontal plates 5.11 provide, by design, overlapping circular holes in such a way as to create the space necessary to mount the tubular spring with spray nozzles. Pentru izotermalizatoare cu diametrul mai mare, se pot monta mai multe arcuri elastice elicoidale cu aspersoare, concentrice, cu diametre diferite, iar in spatiile inelare dintre ele se monteaza sisteme de placi orizontale (Fig.19). 1365 If the horizontal plates 5.1 1 and vertical plates 5.10 of the isothermalizer in Fig.8 are fitted with judiciously placed perforations, this type of isothermalizer can be cooled by inserting the coolant through the top of the apparatus, at a pressure equal to that of the apparatus. This type of densifier is very suitable for cooling with aqueous foam. Foam regeneration can be done by introducing at a pressure slightly higher than instantaneous gas pressure in trays 5.11 c.
1370 At the isothermalizers shown in Fig.15 and Fig.16, the additional coolant introduction system may be mounted on the skeleton supporting the horizontal plate system, inside it, or on similar independent structures.
Fig.19 shows an isothermalizer whose cylinder consists of two segments with different diameters. In the main cylinder 5.1 , whose inner diameter is equal to the outer diameter of the piston, a thermally
1375 deformable sponge is mounted, whose volume in a fully compressed state is only a fraction smaller than the volume of the cylinder corresponding to this position of the piston. The total volume, the configuration of this thermal sponge and its total surface are chosen according to the speed of the piston, the desired efficiency of the device and the compression ratio. In the configuration in the figure, a thermal sponge composed of horizontal metal plates 5.1 1 is represented, which rests on the large helical springs 5.12 and
1380 small 5.13, with a circular section and in which, in a fully compressed state, the helical springs are compressed to the lower limit, and the horizontal plates overlap one another. For high piston speeds, this type of sponge may be missing, in which case the piston moves up to the top cover and pushes the entire amount of gas into the second 5.1s segment of the cylinder. The thermal sponge mounted in this second segment is a dense, metallic foam type, or reticulated network, with a large absorption surface, having the
1385 main role in heat extraction. In the simplest constructive variant, the two sections and their thermal sponges are sized so that when the piston reaches the upper position and the compressed gas is in the upper compartment, the desired compression ratio is reached, in which case the discharge valve opens 5.5e and the cylinder is flooded with liquid on the discharge line. This liquid has the role of absorbing the heat accumulated by the thermal sponges and is discharged, during the absorption phase, through the
1390 holes and related pipes 5.2c. Starting from this configuration, devices can also be made in which no liquid is found on the discharge pipe, but a fixed amount of liquid is permanently found on the upper surface of the cylinder, in such a quantity that when the piston is found in the upper position, this liquid should occupy the volume left free by the total expulsion of the gas from the cylinder. In this situation, the liquid in the cylinder is recirculated, with a flow rate calculated according to the desired efficiency for the
1395 compression operation, while keeping the liquid level in the cylinder constant, variant in which the gas flow in each cycle is lower. This impediment can be removed by completely evacuating the liquid from the cylinder and introducing the appropriate amount of liquid, with a pressure at least equal to the instantaneous gas pressure, during the compression phase. It is also possible to resort to the previously described method: in the enclosure, in an annular cylindrical space made by cuts made in the horizontal
1400 plates 5.1 1 , helical elastic springs 5.131, provided with sprinklers 8.6A, are mounted. The cooling system with sprinklers that distribute jets of liquid in a horizontal plane can be applied (Fig. 20) together with the technical procedures described in Fig. 3, proposed to increase the energy efficiency of isothermalizers. The procedure applies especially to isothermalizers with thermal sponges that are composed of horizontal plates 5.1 1 (with elastic springs, with ropes, with bellows-type support
1405 systems, etc., with or without vertical fins). The isothermalizer contains one or more vertical bellows
12.28, made of elastic, resistant materials, such as rubber, for their installation, vertical channels are made in the thermal sponge, from the piston to the upper cover. These channels have a horizontal section of the same shape as the horizontal section through the bellows which has the largest area, but with slightly larger dimensions to allow compression and expansion of the elastic bellows. In the wall of
1410 the bellows, at the level of each space bordered by two horizontal plates, one or more spraying systems
12.29, of the sprinkler type, are mounted. One of the ends of these bellows is fixed to the piston and the other to the upper cover, one of these ends being provided with a valve 12.12a through which the bellows communicate with the coolant circuit. In each bellows there is a spring 12.30, whose length, in an untensioned state, is equal to the height of the enclosure when the piston is in the BDC position, and in a
1415 tensioned state, equal to the height of the cylinder when the piston is in the TDC position (Fig. 20a).
During the aspiration phase, the inner spring keeps the liquid bellows in a tensioned state, by means of an unlocking mechanism. When the piston reaches the BDC position, the mechanism unlocks, the height of the bellows suddenly reaches its maximum length, and as a result, the gas in the enclosure is suddenly compressed, almost adiabatically. Valve 12.12a being open, the heat transfer liquid fills the
1420 bellows. In the gas compression phase, the pressure of the liquid is maintained above the pressure in the enclosure, so that it also enters the isothermalizer enclosure, being distributed by the sprinklers in the spaces between the horizontal plates. The flow rate of the agent is regulated by the main regulator, so that, together with the displacement of the piston, it ensures the isothermal trajectory. If the liquid acts only as a heat transfer agent, the height of the liquid layer accumulated on the upper face of the piston is
1425 maintained at a constant value by equalizing the inlet and outlet flow rates. If these two flow rates are not equal, the introduced liquid also acts as a piston. When the piston reaches TDC, the locking mechanism is actuated again, and the gas remaining in the enclosure is discharged, through the discharge valve 12.13, either due to the liquid still sprayed or due to the liquid existing in the discharge line.
The technical procedure proposed for increasing the energy efficiency of isothermalizers,
1430 described in Fig.1 -3, can be applied to all isothermalizers in the state of the art, with the particularities imposed by the configuration of the respective device and its destination (it can be used to impose a wide range of transformations thermodynamic, not only for isothermal ones). As in the case of isothermal transformations, in the memory of the regulation system mounted on the prototype of the device we want to optimize, the proposed equation V=f(p) of the transformation is entered and, based on the
1435 measurement signals transmitted by the position transducers and pressure, appropriate acceleration/deceleration commands are sent to the drive system. Thus, for the entire range of identical devices, the optimal trajectory for the respective application is obtained. For example, in the case of double-acting reciprocating piston isothermal isotherms, the inlet of the working gas takes place in a compartment different from that in which the compression takes place, but in which a series of previous
1440 compression operations have taken place, so that the walls and the thermal sponges of the isothermalizer have unevenly accumulated an unknown amount of heat, part of which is transferred to the sucked gas, as it enters the device. If in the case of single-acting pistons, the achievement of the optimal trajectory requires a duration of the suction phase as short as possible, in the case of double-acting pistons, this duration is equal to the durations of the compression and discharge phases in the other compartment.
1445 The measurement of the gas pressure at the end of the intake phase, indicates exactly (regardless of the temperature distribution in the working gas and in its ambient environment), if its average temperature is equal to that prescribed by the optimal trajectory and, consequently, if it needs to be cooled or additionally heated, as well as how much the piston must be accelerated to reach the optimal isothermal speed.
Also, when a higher power density is desired (higher gas flows and implicitly higher piston
1450 speeds), in addition to the speed regulation system, the isothermalizers are also equipped with a cooling/heating system, which introduces in their enclosure, during compression, through devices provided for this purpose (for example, through sprinklers), a flow of heat transfer liquid, which is evacuated through another pipe and transported, with the help of a pump, to a heat exchanger. If the flow rate of liquid discharged is equal to that supplied, the liquid agent is only the heat transfer liquid, but in
1455 case of inequality, the liquid agent is also the piston. The inlet and outlet flow rates of the liquid agent are determined by a regulator, through commands sent to the servomotors that actuate the hydraulic valves on the inlet and outlet pipes and commands that are correlated with the commands sent by the piston actuator regulator. To avoid an oscillating operating regime, one of the regulators must be prioritized.
If it is desired, in this case of multiple regulation, to replace all the direct current motors of the
1460 installation, it can be done successively. In the first stage, with the circulation of the heat transfer liquid interrupted, the optimal trajectory (close to the IAI regime, adiabatic-isothermal-adiabatic) of the solid piston for the chosen temperature Tiz is determined, with the help of the regulation system, the mechanical converter of the rotation is built (from the sinusoidal speed , at the isothermal speed), for example the profiled guide channel of the piston rod and the solid piston drive system is replaced. Then,
1465 the flow of the liquid agent is released, a suitable starting flow rate is chosen and the regulator that directs the opening of the flow control valves is unlocked. In this way, the flow variation curves can be obtained for different starting flow rates and for different outlet flow rates, fixed or variable. After building the mechanical speed converter of the control servo motors, replace the automatic system, based on direct current motors, with a simpler system based on alternating current motors, with the frequency of the
1470 supply current equal to that of the mains, or with a variable frequency, with the help of frequency converters.
In liquid piston and ionic compressors from the state of the art, the piston compressor is a liquid agent supplied by a hydraulic motor, the liquid having remarkable properties: lubrication and sealing, minimizarea of dead volume and a two-way transmission agent of heat and mechanical energy. The 1475 disadvantage of these systems is that, if thermal sponges made up of solid elements only are inserted into the cylinder, at some point they will be covered by the liquid whose level increases, so the heatabsorbing surface of the sponge shrinks (just when its temperature rises faster).
One way to overcome this inconvenience is to make a solid piston compressor, in which a thermal sponge is mounted with N fixed horizontal plates (Fig.21-25), with the surface slightly smaller
1480 than the horizontal section through the cylinder, which multiplies approximately N times, the absorption surface of the device. The plates are arranged in such a way that after the introduction of liquid into the cylinder, all the plates are covered with a layer of liquid, this becoming a liquid component of the thermal sponge, as it takes over from the upper surface of the solid plate, the role of absorbing the energy excess heat. Moreover, on the periphery of the horizontal plates, low-height stoppers can be mounted to
1485 permanently retain a thin layer of liquid, thus sacrificing part of the useful volume of the isothermalizer, in favor of a more efficient thermal sponge. The further penetration of the liquid agent leads to an increase in the thickness of the liquid layers, without significantly affecting the absorption surfaces, but causing a further increase in the pressure in the cylinder (the liquid thermal sponge continuing its role as a piston).
An example is the liquid piston isothermalizer whose longitudinal section is represented in Fig.21 ,
1490 and the transverse section in Fig.21 A. The isothermalizer is constructed like a prior art liquid compressor, with a cylindrical casing 7.1 and a liquid piston 7I actuated by the solid piston 7.5, actuated, in turn, by a speed control system, the imperative of which is to provide the isothermal speed viz The solid piston is used in combination with an electric, variable speed drive assembly. In the case of a hydraulic drive with variable liquid flow, cylinder 7.2 is connected directly to the liquid pipe.
1495 The thermal sponge has a solid and a liquid component. The solid component consists of the horizontal plates 7.3a provided, on most of the circular sector, with the peripheral skirts 7.3b, intended for the division of the liquid piston. The other side of the circular sector, separated from the first by the vertical walls 7.6b (Fig.21 ), has the skirts facing towards the upper part. The liquid component consists of the liquid layers that form on each of these solid horizontal plates. The liquid component of the thermal
1500 sponge has mainly the role of a piston and acts simultaneously in the N elementary compressors formed by splitting the main compressor. In cylindrical-shaped compressors, the gas contained in each of these elementary compressors yields heat, mainly to two circular surfaces 7.3a, which have a diameter almost equal to the diameter of the master cylinder. If the horizontal plates 7.3a were missing, cylinder 7.1 , cap 7.2 and the liquid piston would constitute a liquid piston compressor, with an initial volume approximately
1505 equal to the sum of the volumes of the N elementary compressors, but with a (variable) heat transfer surface only slightly larger than that of an elementary compressor. Due to the way these plates are arranged, the liquid piston acts simultaneously in each of the N elementary compressors, leading to the formation of N elementary pistons, the piston speed in each of them being N times less than the speed of the unique piston, and the heat energy corresponding to this power is distributed over a contact area of N
1510 times greater. Therefore, compared to the non-thermal sponge compressor, we can achieve an isothermal compression of the same amount of gas, with a hydraulic motor speed (which supplies the liquid agent) of about N times higher (the same power, distributed over a time interval of N times shorter).
Starting from this constructive scheme, a multitude of similar configurations can be made. Fig .21 shows a cross section of a cylindrical densifier (in many applications a rectangular section is more
1515 advantageous), and Fig.21A shows a horizontal flat section of it, at the level of an elementary compressor. The horizontal plates 7.3 and 7.4 separate the compressor from the constant pressure tank 7g and the liquid piston 7I, respectively. In this type of densifier, the liquid piston consists of a fixed volume of liquid agent (the same type of liquid as the one in the tank 7g), equal to the free volume of the compressor (the volume of gas in the cylinder, immediately after the suction phase). In the first phase, the
1520 densifier absorbs a volume of gas through the valve 7a (located in the upper elementary compressor), when the piston 7.5 moves from the TDC to the BDC. At the same time, an identical volume of liquid, located in the densifier, is transferred into the tank 7L. The compression phase follows, in which, after closing the suction valve, the liquid from the tank 7L enters the densifier cylinder and, according to the law of the communicating vessels, is distributed into the N elementary compressors. The instantaneous
1525 gas pressure in these mini compressors is almost the same (the difference is given by the height of the liquid column between the compared compressors. The compressed gas is exhausted at the level of each elementary compressor, through the windows 7.6a, practiced in the partition wall 7.6. This partition wall, together with the side wall 7.1 and the two vertical intermediate walls 7.6b (Fig.21 ), borders a ring sector 7s, which communicates freely with the tank 7g, being constantly flooded by the liquid agent with the
1530 pressure pf in the tank. Opening windows 7.6a is done by moving a movable cap (piston 7.7) that is running tight (by means of seals) on wall 7.6 and is controlled by a differential pressure switch 7p, when the piston 7.5 is in position T and the pressure pf of the liquid in the densifier is equal to the pressure in the tank 7g. At this time, the gas pressure in each elementary compressor is equal to the pressure pf, to which is added the pressure given by the liquid column between the measuring point and the elevation of
1535 the respective elementary compressor. The displacement of piston 7.7 causes the entire amount of compressed gas in the densifier to be replaced by liquid agent in the tank 7g (in this way, the entire volume bordered by horizontal plates 7.3 and 7.4 is occupied by the liquid) and causes the level of the liquid in this tank to lower. If piston 7.5 continues to move to the TDC, the amount of fluid between the level T and the TDC level (equal to the total volume of compressed gas during a stroke of the piston) is
1540 exhausted through the pipes 7r to another device with pressure pf (for example, a tank, or a hydraulic generator).
In another configuration, the tank 7g and the piston 7.7 may be missing, the wall 7.3 becomes the outer wall and the windows 7.6a are replaced, each of them with a check valve. The tank 7s is replaced by a simple pipe in the wall of which valves are mounted to each elementary compressor, the lower end
1545 being opened through a slot to the lower compressor, and the upper end is opened to the upper gas layer, located in the upper elementary compressor. In this case, the check valves open successively, the respective mini compressor being immediately flooded by the liquid in the main column, the thickness of the air layer increasing accordingly.
The liquid and, indirectly, the plates in the densifier can be cooled by keeping the liquid in the
1550 tank 7g at temperature Tamb or/and by the recirculation, continuous or intermittent, of the liquid agent from the tank 7L. An increase in heat transfer surfaces can be achieved if the gas inlet pipe in the densifier is supplied by a foam generator. It can also introduce, at the right time, foam or compressed gas, directly into the liquid of each elementary compressor, using thin pipes. An increase in heat transfer surfaces can be achieved if the gas inlet pipe in the densifier is supplied by a foam generator. It can also introduce, at
1555 the right time, foam or compressed gas, directly into the liquid of each elementary compressor, using thin pipes. In some configurations, this compressed gas can even come from the tank 7g. Additional cooling is obtained if different types of metal inserts, strips, metal nets, etc. are inserted into each elementary densifier at its top, or if various types of vertical metal elements are mounted on its ceiling. Due to the natural convective currents, as the gas is compressed it tends to thermally stratify, so that in areas where
1560 the instantaneous temperature can reach higher values, the absorption surface of the thermal sponge will be increased.
The densifier in Fig.22 is built on the same principle of overlapping a large number of elementary liquid piston compressors, made by interspersing their upper and lower walls, 7.3s and 7.3I, respectively. Compared to the previous configuration, two types of liquid piston mini-densifiers appear in this type of
1565 densifier: a mini-densifier 7c between the upper and lower walls, with higher height, cooled by liquid spray and a mini-densifier 7d between the lower and upper walls, with a lower height, without spraying. The liquid piston is inserted into the elementary compressors 7d directly, through the windows 7.3f. A vertical skirt 7.3g is inserted to separate a layer of gas into each densifier. The introduction of the liquid piston into the elementary compressors 7c is done by a distributor of agent 7.1 1 , from which the liquid agent is
1570 sprayed into the compressor, intensifying its role as a gas coolant. Similar to the previous configuration, a horizontal wall 7.3 separates the densifier area from the tank 7g, which is in direct communication with a tank 7s located, this time, in the center of the densifier, having a cylindrical shape and being separated from the densifier by the cylindrical wall 7.6, in which the windows 7.6a are executed, on each basic compressor. To each of these windows corresponds a similar window, located at the same level, in the
1575 cylindrical wall 7.7, located inside cylinder 7.6, so that these windows overlap in the "open" position and allow gas and liquid to pass from one compartment to the other. The closing of these windows in the "closed" position is done by turning with an appropriate angle, or by moving vertically the cylindrical wall 7.7, in such a way that the seals mounted on the outer surface of cylinder 7.6 around the windows 7.6a block all gas passageways. The same transfer system is also applied in the last phase of the cycle, in
1580 case of fluid exhaust from the cylinder and gas suction with the pressure p,. A sector 7.8 in the side wall, with the height equal to that of the densifier, has a series of windows 7.8a practiced at each mini- densifier. During the compression operation, these windows are closed by a piston 7.8b of the appropriate size and shape, fitted with suitable seals and a horizontal displacement system. Through these windows the liquid agent is removed from the elementary compressors (after the compressed gas
1585 is removed and the valves 7.6a are closed) and the working gas is introduced at the initial pressure.
In some configurations, when the densifier is carried out by means of parallel plates very close to each other, or by means of small alveoli inserts, or very small mesh woven nets, in the case of liquids whose viscosity exceeds a certain limit, it is effective to implement some devices to accelerate the discharge of liquid from the densifier after the compressed gas exhaust phase. In Fig.22A, we
1590 represented on a large scale details of a possible configuration for such a device, as well as for a variant of gas inlet valve and a variant of diffuser. The accelerator is made by making the upper plates 7.3s of each elementary compressor out of three distinct components: an outer peripheral ring 7.3sa, with the outer diameter equal to the inner diameter of cylinder 7.8 to which it is rigidly attached, an inner peripheral ring 7.3sb, with internal diameter equal to the outer diameter of cylinder 7.6 on which it is rigidly fixed, and
1595 a movable flat ring 7.3s, with outer diameter greater than the inner diameter of the peripheral ring 7.3sa and inner diameter smaller than the outer diameter of the inner peripheral ring 7.3sb. The movable ringsections shall all be fixed on one or more rods 7.9 in a position below the corresponding plate 7.3s in such a way that, through the seals mounted on the edges of the upper surface, air and liquid are not allowed to flow to the lower compressor when the rods 7.9 are in the “closed” position. Moving the entire
1600 rod-plate system to a lower position causes wide access paths to open, in which the friction between the liquid and the plates is greatly diminished. An increase in the efficiency of fluid circulation is achieved when these movements are made at high speeds, with sudden starts and stops, so that the forces generated by the surface tension of the liquid are overcome.
Here is the sequence of the phases in the densifier: the liquid agent is inserted through a gate 7a
1605 into the tank 7L, from which it is distributed naturally between the lower plates 7.3I of an elementary compressor 7c and the upper plates 7.3s of the following elementary compressor. The mini-densifiers 7d are thus formed, which also become liquid dispensers for mini-densifiers 7c. The gas layer between these plates is compressed and pushed through the holes 7.10 of the bottom plates inside the compressor 7c located above them. In the configuration in Figure 20, a simple plug 7.1 Od with the appropriate seal
1610 (whose collar steps on springs 7.1 Or) completely obscures the entrance path to the respective minicompressor 7c. The springs 7.1 Or are pretensioned, so that the valves can be opened at a set pressure. Various devices for regulating the flow of gas passing through these holes can also be implemented. If both air inlet and gas inlet valves are remotely controlled, priority may be given to either of the two types of densifiers.
1615 During operation, regardless of the flow of liquid agent introduced (or the speed imposed to the solid piston), the fluid pressure at the periphery of the master cylinder is equal to that of the gas and liquid in the densifiers 7d, and that of the gas and liquid in the densifiers 7d is slightly lower, this causing the valves 7.10 and 7.1 1 to open, depending on their adjustment. If the flow of liquid entering the sprinklers is lower than that introduced through the windows 7a, the gas pressure increases, which also causes the 1620 gas inlet from the densifiers 7d to open into the densifiers 7c, causing the gas pressure in these densifiers to increase. When passing through the liquid layer, this gas undergoes additional cooling.
In the configuration shown in Fig.21 A and in the additional details of Fig.22A, the inlet valve is a sealing plug directed by a pre-loaded spring, and the sprinkler can be a simple disc-shaped cap, with a horizontal spray holes 7.11 on the side. This valve opens when the difference between the dispenser fluid
1625 pressure and the gas pressure in the corresponding elementary compressor is greater than a preset value. This difference remains almost constant throughout the compression period, but may have slight differences from one elementary minicompressor to another.
After a few such cycles of alternating the compression process between the two types of mini- densifiers, all the mini-densifiers of type 7d transfer the compressed gas to the mini-densifiers of type 7c ,
1630 and when the pressure reaches the final one, the compressed gas is evacuated to the tank 7g.
Using the same idea (for their use in cases where large single-step compression ratios are desired), many configurations of liquid piston densifiers can be made, with one or more solid-piston densifiers as the main subassembly. Any of the solid piston densifiers described above, or made on the same constructive principles, may be used. In these configurations, the functionality of the system and its
1635 adaptation to different particular applications depend on the initial number and volume of densifiers, as well as the amount of mechanical energy available. For example, for any of the solid piston densifiers in Fig.4-16, the exhaust of compressed gas by replacing it with a liquid that has the gas pressure is an efficient process for removing compressed gas bags and for extracting and discharging excess heat from the solid thermal sponge. The process can be implemented by fitting valves in the piston that allow
1640 controlled fluid intake and proper extension of the master cylinder. In this way, the solid piston densifier becomes a liquid piston densifier. For example, Fig.23 shows the densifier in Fig.16, in which vertical perforations 5. 11 o are made in the components of the solid thermal sponge (horizontal metal plates 5.1 1 ), in such a way that when the plates overlap under compression, these holes also overlap and form continuous channels. In order to increase the initial volume of gas admitted in the cylinder and to increase
1645 the absorption surfaces of the thermal sponge, as well as to decrease the total mass of the densifier, a series of horizontal channels 5.11c are executed between these vertical holes. As the liquid piston only enters into operation at the final stage of compression, when a slower speed is required, the friction forces between the liquid and the solid sponge are lower. An identical solution is shown in Fig.23A, for the solid piston densifier in Fig.7, by making the holes 5.11o and the channels 5.1 1c. In both cases, in
1650 piston 5.2 the valves for fluid intake 5.2s are fitted and the possibility of liquid agent passing through the windows 7a is realized. In Fig.23B, the liquid circulation channels are highlighted, after positioning the solid piston in its upper position.
A considerable increase in the flow of gas circulated under conditions of an isothermal compression is obtained if, after a corresponding increase in the height of the densifier, an additional
1655 thermal sponge 5gs is inserted in the upper part of the densifier, made of metal foam, metallic fabrics, other metal inserts with a large absorption surface (in Fig.23 it is made of woven blankets made of metallic wire, superimposed, without separation intervals, mounted on a horizontal support system, of bars, rods, perforated plates, etc.), which absorbs a large amount of heat energy in all phases of compression. The liquid piston penetrates into the holes and grooves executed in the horizontal plates, as
1660 well as in the alveoli of the additional sponge, only in the final phase of compression, when the forward speed of the liquid piston is quite low and the liquid-sponge friction forces are reduced.
Fig.24 shows a constructive variant of the liquid piston isothermalizer with vertical telescopic fins. It is composed of a series of identical elementary isothermalizers, separated from each other by horizontal walls 7.17, and from the enclosure for introducing the active liquid, by a common vertical wall
1665 7.18, provided for each elementary isothermalizer with a slot, located at its base. In each elementary isothermalizer there is a horizontal plate 5.1 1 , with a slightly smaller surface area than the lower wall, which rests on a plate 7.21 , at some distance from the base of the mini-isothermalizer (because in the first phase of compression, the thermal energy to be absorbed is less). All these plates are fixed on a vertical rod 7.20 which penetrates all the base plates and its lower end is fixed to the displacement
1670 mechanism. If the distances between the horizontal plates are the same, at a full stroke of the piston, all the plates (along with the horizontal plates resting on them) move the same length, compressing the vertical fins to their minimum height. The configuration of the telescopic fins differs from those described in Fig.10, by the addition of horizontal plates 5.10c between the vertical fins and inside them, a change by which the total surface of the thermal sponge increases considerably. In Fig.24A is represented such a
1675 fin in the fully compressed position, and in Fig.24B, the same fin in the maximum length position.
In this configuration, the rod 7.20 on which the plates 7.21 are fixed passes through the lower wall of the enclosure, and its end is connected to an actuation mechanism, with the role of simultaneously moving the inner plates and the horizontal plates with which they are in contact and with this to perform the compression of all telescopic sections. This passage through the housing of the isothermalizer, like all
1680 passages through the walls of its compartments, is sealed, for example with rubber gaskets. The movement of the vertical rod must be correlated with the opening of the regulating valve on the pipe through which the liquid piston moves, in such a way that the isothermal trajectory is achieved, and the horizontal plates remain permanently on the surface of the liquid in the respective compartment. The movement of the drive rod must stop when the telescopic sections are compressed to the maximum, even
1685 if the liquid piston continues to advance, until the gas is completely removed between the sponge plates and introduced into the vertical pipes 7.23 (possibly filled with metal fabric), which makes the connection between each mini-isothermalizer and the compressed gas tank under constant pressure 7r (in which, along with the gas 7g, the liquid 7I is found), installed at a higher elevation. (In the drawing in the figure, only pipe 7.23 is represented, located in the plane of the vertical section, corresponding to the elementary
1690 isothermalizer) Each of these pipes is closed (separately or collectively) by a valve 7.22 that opens only when the discharge phase is triggered. In this phase, the compressed gas remaining in the premises and in the connecting pipes is replaced by liquid from this tank. After closing the valves 7.22, the gas inlet valve 12.13 is opened and the way to empty the liquid from the premises is opened. The dimensional design and the choice of materials used to make the thermal sponges, as well as the choice of the liquid
1695 used, will take into account ensuring the fastest possible removal and with minimum energy consumption of the liquid from the isothermalizer enclosure.
The method of driving the horizontal plates from the outside can be replaced by a method by which the horizontal plates are moved by the liquid piston itself. For this purpose, instead of the plates and the rod used in the previously described configuration, to move the fins of the telescopic sections,
1700 inflatable, hermetic, gas-filled bags are mounted, in such a way that their volume is large enough to displace a volume large amount of the liquid entering the respective compartment, and the Archimedean forces should be large enough to support the full weight of the fins of the telescopic sections and move them up to the maximum compression limit. To achieve this goal, the materials used to make the vertical sections are chosen from those with the lowest specific mass (aluminum, plastics, etc.). A solution that
1705 can give very good results for the realization of other configurations of thermal sponges, when they do not have an accumulation role, is the use of light materials, but with high thermal conductivity, after they have been plated with thin layers of metal.
To regain the space lost by using these inflatable bags, they can be replaced by light metal containers, in which a light, non-deformable thermal sponge is inserted. All metal containers in the
1710 enclosures communicate with each other through flexible pipes, and one of them is connected to an external isothermalizer, which introduces compressed gas into these containers and also compresses the gas inside the containers, together with which it forms an gas piston isothermalizer, independent from the one with a liquid piston, having different inlet and discharge pressures. Also, the pumping of the compressed gas is done independently, by replacing it with liquid from its own storage tank. The gas
1715 piston acts in tandem with the liquid piston, in such a way that the intake and discharge of the gas from the containers is done almost simultaneously with that of the gas in the compartments of the liquid piston isothermalizer. If the penetration of liquid into the gas piston isothermal vessels for discharge occurs before the time when the pressure in the liquid piston isothermal reaches the final pressure, the liquid displaced by sinking, as an effect of the additional weight, of the gas piston isothermal vessels , acts like
1720 a piston, compressing the gas to the discharge pressure, or to a pressure lower than this.
A similar isothermalizer is shown in Fig.24C, in which the compression of the fins is done by the solid piston (which, at the same time, is its upper wall), of an elementary compressor mounted on the lower wall of the main elementary compressor. The liquid piston, entering the cylinder 7.2, is distributed through the slots 7.18f in each elementary isothermalizer 7.18i and compresses the gas inside. When the
1725 gas pressure reaches a sufficient value, the movement of the piston 7.18p begins, which is, at the same time, its upper wall and the lower wall of the elementary isothermals 7.18s. At this stage, the isothermalizer 7.18s is a solid piston isothermalizer. The gas here is compressed, simultaneously with the displacement of the piston 7.18p and the compression of the telescopic fins 5.10. Compression in this isothermalizer ceases when the piston reaches its maximum height, corresponding to full compression of
1730 the fins, at a gas pressure lower than that in the lower isothermalizer. The liquid piston continues to enter the isothermalizer until the gas pressure reaches the desired value, that of the gas in the compressed gas tank 7r, under constant pressure, in which, along with the gas 7g, the liquid 7I is found. At this moment, the valves 7.22.2 on the vertical pipes 7.23 (possibly filled with metal fabric) are opened successively (due to the differences between the hydrostatic pressures of the isothermalizers, located at different
1735 levels), which make the connection between each mini-isothermalizer and the tank 7 r. In the drawing in the figure, only pipes 7.23 are represented, located in the plane of the vertical section, corresponding to a single pair of elementary isothermalizers. In this phase, the compressed gas remaining in the enclosures and in the connecting pipes is replaced by liquid from this tank. After closing the valves 7.22.2, the pressure in all elementary isothermalizers 7.18i becomes equal to that in the tank, sufficient to keep the
1740 piston in the maximum position. At this point, the valves 7.18v are opened and the liquid from the main column enters each of the isothermals 7.18s, isothermals that become liquid piston. The compression also continues in these isothermalizers, until the pressure in the tank is reached, when the valves 7.22.1 are opened, through which the gas is discharged and the entire device is filled with liquid. At this moment, the gas inlet valve 12.13 is opened and the way to empty the liquid from the enclosure is opened.
1745 An increase in the total surface area of the thermal sponge is achieved by mounting flat vertical fins 7.18i on the upper surface of the piston 7.18p. They are dimensioned and mounted in such a way that when the piston moves forward it enters the telescopic fins, in the free space between the sections of the vertical fins.
The isothermalizer in Fig.25 is a double-effect liquid piston isothermalizer resulting from joining
1750 two isothermalizers similar to those described in Fig.21 , but the configuration can be adapted to any other type of liquid piston isothermalizer, including those with telescopic fins. The isothermalizer in the figure is composed of two identical liquid piston isothermalizers, with a parallelepiped body 7.1 , the fixed vertical walls 7.6 provided with gas outlet holes for each elementary isothermalizer 7c, the movable vertical walls 7.7, attached to them but which can slide laterally, provided with similar exhaust holes, located in such a
1755 way as to allow gas circulation only in a certain position, a solid thermal sponge, formed by the horizontal plates 7.3a provided to the outside with the peripheral skirts 7.3a, intended to divide the liquid piston, and towards the inside, in contact to the vertical wall 6.6, with skirts oriented towards the upper part. The liquid piston 7I is actuated by the hydraulic motor 7p, with variable flow due to the actuation of a regulating flap, commanded by a regulation system, whose imperative is to provide the liquid piston with
1760 the isothermal speed viz. Through a 7v valve system, the liquid agent is directed alternately to the two main isothermalizers, where they are divided into elementary pistons.
Composite systems may also be made by simultaneously or successively accumulating the effects of a liquid piston and a set of solid pistons. Such configurations are described in Fig.26, Fig.26a and Fig.26b and can have as a model any of the solid piston densifiers described above, if their volume
1765 variations can be fully controlled. The densifiers 7g in Fig.26 are equipped with a thermal sponge composed of elastic plates 7.3, mounted on a support 7.3s. They are placed in hermetic bags 7.14, made of elastic materials, or other tear-resistant materials, with high heat transfer coefficient, but slightly deformable, even at low pressures.These bags are fastened to a metal plate 7.2s fitted with holes 7.2o, that make the connection between the gas in these densifiers to the gas layer 7gs, located above the
1770 plate 7.2s, layer which contains another thermal sponge, for example, a dense cross-linked metal network. The volume of each bag 7g is reduced, with the advance of the liquid piston, according to the reduction in the volume of the thermal sponge it contains and may reach a minimum value when the deformation of the sponge is naturally blocked (when all the inner plates overlap), or by externally controlled devices. The so-created enclosures communicate with the top gas layer 7gs in the compressor
1775 cylinder, but can be separated from it by the valves 7.2ps. The metal plate 7.2s separate, in this way, all the gaseous regions 7s containing thermal sponges with large absorbency area, from the rest of the enclosure 7i. In some configurations, various pipes that create additional communication paths can be inserted into the enclosure, such as pipe 7.16, which makes the equalization of pressures between layers 7i and 7s by means of an intermediate liquid piston. The device is equipped with a series of valves for the
1780 circulation of gas and liquid flows: valves 7.2a for the initial gas intake, valves 7.2e for the discharge of compressed gas, valves 7a for introducing the liquid piston into the enclosure, for the circulation of the cooling liquid (same as the piston liquid) and for liquid evacuation, simultaneously with the entry of gas into the lower chamber 7i, valves 7.2i which control the entry and exit of the liquid piston into densifiers 7g, valves 7.2ps which can block communication between densifiers 7g and the layer 7gs, valves 7.2 pm
1785 and 7.16pi, which control the communication between the lower and upper enclosure of the cylinder. The play of these valves allows us to choose between a multitude of variants of unfolding the functional stages
In a first configuration, we chose the option in which the liquid piston penetrates successively, in all the enclosures of the densifier. In the first phase, with valves 7.2ps and 7a open, with all other valves
1790 closed, with an initial pressure p, in the lower enclosure 7i and with an initial pressure p2 in the upper enclosure 7gs and in the enclosures 7g, the liquid piston enters the enclosure 7i. In order to achieve high plant performance, it is recommended that the initial gas pressures be as high as possible. At an external pressure p1 and an internal pressure p2, the volume of the inflatable bags has a maximum value (a value as close as possible to the total volume of the lower enclosure 7i is recommended), and the elastic
1795 elements of the sponges 7g are tensioned, because they take part of the mechanical energy necessary to introduce the gas. In this way, the hydraulic pressure (and power) required to penetrate the liquid piston is high, and the isothermal start speed is much lower than in the case of gases at atmospheric pressure. The liquid compressing agent, with an initial pressure p/; will gradually replace part of the gas here by compressing it. At the same time, the pressure of the liquid agent is exerted on the walls of the 7g
1800 compressors, reducing their volume and increasing the pressure of the gas inside.
When the gas pressure in the enclosure 7i equals that in 7s, the inflatable bags return to their rest form, the one with the elastic elements not tensioned, the mechanical energy previously accumulated in the elastic elements of the thermal sponge diminishing the mechanical energy required by the liquid piston to achieve this compression. Part of the thermal energy produced by compressing the gas in the 1805 chamber 7i, is given to the walls 7.14 (inflatable bags) and the walls of the chamber 7i, with heat transfer surfaces that decrease with the advance of the liquid piston and with the increase of the gas pressure in all compartments, and a part of the thermal energy resulting from the compression of the gas in the enclosures 7s and 7gs is transferred to the thermal sponges 7g and 7gs, respectively. The gas pressure in the enclosures 7s can increase, at a rate dependent on the Young's modulus of the plates 7.3, until the
1810 plates that make up the sponge overlap and occupy a minimum volume, the mechanical energy accumulated by the elastic plates being ever greater. In this position of the piston, if liquid elements of the thermal sponges are found inside each inflatable bag to occupy all the voids inside the respective bellows (the excess liquid also spilling into the 7gs layer), the closing of the 7.2ps valves causes the temporary separation of the lower enclosure 7i (in which the gas pressure has reached the value p3) than regions 7g
1815 and 7gs (in which the gas pressure has reached the value p4>p3). In the next phase, the equalization of the two pressures p4 and p3 is carried out, an operation that can be carried out in different ways, including:
- the opening through the windows 7a of a communication path with a hydraulic motor and the opening of the valves 7.2pi, causes the disappearance of the forces that kept the elastic elements of the densifiers
1820 7g tense, and their return to the rest volume, with the discharge of the elastic energy stored in the mechanical work produced by engine, simultaneously with the expansion of the gas in the enclosure 7i, with the absorption of thermal energy from the enclosure walls, up to the pressure p3 and the filling of the inflatable bags with the liquid in the enclosure; after equalizing the pressures and opening the valve 7.2I, the liquid piston continues to compress the entire volume of gas, from the pressure p3, to the pressure pF,
1825 then, after opening the valve 7.2e, the complete liquid filling of the entire enclosure and the discharge of the compressed gas.
- the opening of the valve 7.2pm and the sudden actuation of the liquid piston (for example, by suddenly introducing a piston, solid or liquid, into the pipe that brings the liquid to the inlets 7a) leads to the formation of a gas piston that adiabatically compresses the gas in the enclosure 7gs up to pressure p4,
1830 keeping the volume of the airbags unchanged. The mechanical energy used to move this volume of gas is transformed into thermal energy fully absorbed by the gas in the room. The temperature and pressure differences between the mixing gases cause non-negligible exergy losses. Further movement of the liquid piston has the effect of increasing the gas pressure in the enclosure from the pressure p3 to the pressure pF, then, after opening the valve 7.2e, completely filling the entire enclosure with liquid and expelling the
1835 compressed gas. The mechanical energy stored in the elastic elements can be recovered during the pumping phase of the liquid from the enclosure 7i, by directing it to a hydraulic motor.
- if the pipe 7.16 is full of liquid, when the valves 7.2pm, 7.161 and 7.2pi are opened, it is pushed towards the layer 7gs, with the lower pressure (by the pressure of the gas in the enclosure 71 and by the pressure exerted by the displacement of the walls of the inflatable bags 7g due to the elastic mechanical energy
1840 accumulated in the elastic elements of the thermal sponges) and compresses it to an intermediate pressure between p3 and p4, the same for all the gas in the enclosure; the 7g bags fill with liquid and return to their resting volume, and the elastic mechanical energy accumulated in the bag elements is fully recovered as compressive mechanical energy
The thermal sponge of the densifier in Fig.26a is made of plates, or elastic metal strips 5.14,
1845 which have a series of corrugations with different radii of curvature. They can be arranged in organized structures, or they can be arranged, as in Fig.23a, in a more or less random way and inserted in deformable hermetic bags 7.14, made of elastic materials resistant to breaking. Hermetic bags can have the form of mattresses, with the width I equal to one side of the device enclosure and the length equal to a multiple of the other side L. These mattresses are placed in overlapping layers, along the entire height
1850 of the device, by bending 180 degrees after each length L, without restricting the free movement of the gas. If necessary, the mattresses communicate with each other through rigid tubes. In another configuration, instead of a single very long bag, an appropriate number of mattresses are used, with a suitable thickness (depending on the type of thermal sponge) and with an area equal to the horizontal section of the cylinder, the mattresses communicating with each other, through more rigid tubes. In
1855 another configuration, instead of each mattress, tubes of length I or L are used (both sizes arranged in alternating layers can be used). All these tubes (and mattresses) communicate with each other, through rigid, metallic or deformable tubes, forming a single enclosure. In the configuration in Fig.23b, the isothermalizer contains layers of cylindrical bags or rectangular mattresses, but the thermal sponge is made of elastic metal sheets rolled into more or less helical rolls. In all cases, although the bags can be
1860 mounted without support, from the point of view of the speed of heat transfer from the premises, it is preferable to arrange them on some supports.
In the initial state, the bags are inflated, at an initial pressure which may be different from the atmospheric one. The liquid is introduced and discharged into/out of the enclosures, through gates 7a, with hydraulic pumps, the ratio between the inlet and outlet flow being variable, but always unitary (the
1865 liquid also plays the role of cooling), or superunitary (the liquid plays the role of a piston ). The operation of this compressor is similar to that of the compressor in Fig.23, with similar components having the same notation.
At the compressor in Fig.27, the introduction of liquid into the compressor is done by spraying. The upper thermal sponge is a reticulated mesh 7gs, under which the lower thermal sponge is mounted,
1870 in the form of a tubular pipe 7.16, with double walls, one end of the pipe being located in the area of the metal mesh, and the other near the base of the cylinder. Both ends of the pipe pass through the walls of the cylinder and can be closed by valves, both on the liquid inlet/outlet path 7a and on the gas inlet/outlet path 7.2pi and 7.2ps. This pipe can be very long and, if laid in parallel layers (for example, cylinders of length L/l, placed next to each other, can occupy a large fraction of the total volume of the cylinder. It is
1875 provided with a large number of nozzles 7.10, or/and single holes, with a higher density at the top.The liquid piston (which is also the coolant and is part of a cooling circuit, equipped with a hydraulic pump and a heat exchanger), is introduced through the valve 7a from one end of the pipe, in the space between the walls of the pipe. In the walls of the pipe, spray nozzles are mounted, both to the inside of the pipe and to its outside, which spray liquid in all areas occupied by gas, compressing it. The constant temperature of
1880 the gas is maintained by adjusting the ratio between the flow of liquid in and out. At some point, the entire lower part of the cylinder is occupied by liquid, the gas accumulating in layer 7gs. Compression continues with a low liquid flow until the final pressure pF is reached, when valve 7.12e opens, the liquid flow increases rapidly and the compressed gas is fully discharged to the tank/consumer. After compression, the liquid is discharged by suddenly withdrawing the lower cylinder head 7.4 and draining the liquid into a
1885 reservoir.
The compressor in Fig.27a is a liquid piston densifier, without solid moving parts, with a cylinder 7.1 of circular section, in which the mass of the thermal sponge is distributed in such a way that the thermal energy from the transformation of the mechanical energy of the piston is absorbed into - a way as uniform as possible. Since, according to the invention, the piston moves with the isothermal speed viz, the
1890 temperature difference AT is maintained throughout the compression. In this device, until the desired compression ratio is reached, the liquid agent travels a longer route than in the case of liquid piston compressors from the state of the art. The cover 7.2 of the cylinder is equipped with an exhaust valve 7.12r (the gas is discharged by replacing it with a liquid agent), and the lower wall 7.4 with an inlet valve 7.12a and with valves 7.4e for exhausting the liquid from the cylinder, at the end of compression. The
1895 thermal sponge is made of vertical cylinders (their cross-section may not necessarily be circular) concentric 7.3v, arranged at greater distances in the central part of the densifier, but increasingly closer towards its periphery. Moreover, the peripheral cylinders are provided with elements to enhance heat absorption (in the figure, horizontal fins 7.3f). Another component of the thermal sponge is the 7gs reticulated metal network, located in the upper part of the cylinder. The pressure in the cylinder is kept
1900 constant by drilling 7.3o holes in the upper part of the vertical plates. Such holes can also be made at lower altitudes, to control the paths taken by the liquid agent, to increase the absorption power of the excess thermal energy and to accentuate the upward convective currents.
Another advantage of this configuration is the creation of the possibility that, after the evacuation of the compressed gas from the densifier, the cylinder 7.1 and its cover 7.2 can be raised for a short
1905 period of time, during which the liquid agent is removed and replaced by the working gas.
Isothermal transformations can also be achieved with rotating devices, starting from the rotary compressors of the state of the art, applying the procedures described in this invention. These transformations can also be carried out with rotary devices which, at the technical stage, are most often used as liquid pumps or internal combustion engines, if appropriate measures are taken to ensure the
1910 tightness between the enclosures with different pressures.
As mentioned above, in the case of single-enclosure rotary devices (as with the blade compressor in Fig.28), the isothermal speed can be obtained by continuously changing the angular speed of the rotor, in such a way as to maintain at all times the equality between the instantaneous work delivered to the gas by the piston (in this case, the sliding blade in the rotor) and the instantaneous
1915 thermal energy transferred by the gas to its environment. In the case of rotary devices with several closed enclosures, the change in the angular speed of the rotor has different effects in each of them, so that it is preferable to maintain a constant rotor speed and to modify separately for each enclosure other characteristics that allow this equality to be achieved (e.g. coolant inlet and outlet flow rate). The permanent change of angular speed can be abandoned also in applications where the power of the
1920 device is a decisive factor, applications where efficient thermal sponges should be used, with large absorption surfaces and cooling installations of high performace.
The isothermalizer described in Fig.28 is a variant of the blade compressor, compressor described in detail in the patent application RO128041 (A2). It is characterized by the fact that it uses only one blade in the rotor. It consists of a stator (empty cylinder 6.2), inside which it rotates around its center
1925 shaft, the rotor 6.1. In this configuration, the rotor cylinder is empty and its diameter is larger than the radius of the stator. In the rotor is mounted a pocket, usually parallelepiped 6.4, obtained by installing side walls along the entire length parallel to the plane formed by the rotor diameters, on one side and on the other side of it, equally spaced from it. In this pocket (housing) in which the parallelepiped blade 6.3 is inserted, the length of which is equal to the inner length of the housing in which it is inserted (so of the
1930 stator), whose height is equal to the depth of the pocket and the thickness is equal to the inside thickness of the pocket (the four side surfaces of the blade slip tightly onto the inside surfaces of the housing). The length of this housing is equal to the inner length of the stator (the surfaces of the base of the blade also slip tightly onto the inner surfaces of the stator bases). As a rule, the rotor is tangent to the inner surface of the stator wall. In the configuration in Fig.28, the radius of curvature of the stator wall is modified, over
1935 the entire length of the stator, on a sector 6.5, being equal to the radius of the rotor. In this way, the contact portion between the rotor and the stator is no longer limited to a straight segment, but extends to a curved surface with the desired width. The blade can slide along the entire height of the notch, and when its tip touches the stator wall, it divides it into two chambers, sealed between them. This extreme position of the blade is ensured by the centrifugal force generated by the rotation of the blade, as well as
1940 by the pressure of the fluid 6.11 closed between the blade and the bottom of the notch (a lubricant, which is also coolant of the gas and which is inserted using a pump, through a flexible pipe 6.41 (Fig.31 )) and circulates between the housing in the rotor, a heat exchanger and a tank) or/and, as in the current state of the art, by elastic springs. The liquid in this housing can penetrate, due to high pressure, through a groove made in the blade (channel 6.3a in Fig. 29), to the contact surface between the blade and the
1945 stator, mitigating the effects of friction and providing a superior seal. The sealing between the two compartments with different gas pressures can also be improved by fitting elastic seals (6.3b in Fig. 29) whenever possible.
The height of the rotor can be equal to the inside height of the stator, in which case the surfaces of the two bases of the rotor slide over the surfaces of the two stator bases. In the configuration in Fig.31 ,
1950 section 1 -1 , this height is higher and the sliding movement between the stator bases and the rotor walls is provided by bearings 6.91 and segments or seals, etc. The rotor of the machine is mechanically coupled with an engine (electric or mechanical), and in the case of a expander, with a generator or other mechanical load. On both sides of the tangent surface, there are two rectangular slots (6.6d and 6.7d in Fig.28), connected to pipes 6.6 and 6.7 respectively, for the siction and for the exhaust of the working
1955 fluid. If the machine acts as a densifier, the inlet 6.6d can be free, and on the discharge line 6.7 a valve 6.7a is mounted, automatic or operated by a coil 6.7b (Fig.28). If the machine acts as a rarifier, the suction is through a valve or drawer and the exhaust is usually free. In the configurations in which the axis of the stator is vertical, the suction and the exhaust of the working gas is made by cut-outs executed in the two circular plates that constitute the bases of the stator: a cut for intake in the 6.6v area of the lower
1960 base and a cut for exhaust in the 6.7v area of the upper base of the stator.
With these constructive components, the described device is a rotating polytropic compressor that can achieve good performance in certain specific applications. Like any polytropic compressor, it can perform isothermal compression operations when its angular velocity is equal to the isothermal angular velocity u)(t) over the entire duration of a rotation, but even for large temperature differences AT this
1965 speed is very low. Achieving higher speeds and, consequently, higher compressed air flow rates is possible by reducing the polytropic coefficient of compression. In the densifier shown in Fig.28, this objective is achieved by injecting an abundant liquid thermal agent, which serves as a liquid thermal sponge when the liquid inlet flow rate is equal to the outlet flow rate, and also as a liquid piston when the input flow rate is higher. The injection process can start outside the machine in a humidification
1970 antechamber (AC from Fig.30B), where the volume is constant and there are no moving parts, it is easier to control. In this antechamber, the working gas introduced via a rotary compressor is cooled, its temperature being brought to the working value Tiz,. In the antechamber is abundantly sprayed the coolant, process which continues in the stator enclosure, through the nozzles 6.9b, mounted at the end of the sprinklers 6.9a, supplied from the pipe 6.9. The flow rate of each sprinkler can be changed with the
1975 adjustment tools 6.9v mounted at the entrance to the main line 6.9, or on each sprinkler. The rotor cylinder can also be filled with coolant, directed to sprinklers mounted in the rotor wall. Also, part of the fluid 6.11 is driven by the rotor blade and discharged together with the compressed gas onto the exhaust pipe and after separating it into the pipe 6.7c is collected in the tank 6.10, which is part of a cooling circuit together with the heat exchanger HE and the pump 6.7M. All actuators of flow-regulating devices, as well
1980 as those that determine the angular speed of the rotor, are controlled by a central device DC that receives signals from piezoelectric pressure transducers 6.8 mounted in the work room. The central device shall be programed in such a way as to ensure the equality between the mechanical power given to the gas and the thermal power given by the gas to its environment.
Due to the fact that at high gas pressures and low speed of rotation of the blade, the separation
1985 between the two compartments is more difficult to achieve, the compression ratio of the gas obtained with a single densifier being limited. In order to obtain large compression ratios, it is necessary to series several densifiers, which perform a compression in pressure stages and a judicious distribution, within a full rotation, of the total engine torque. Fig.29 shows how multiple apparatus can be arranged in series, and Fig.28B, how densifiers can be superimposed. In this latter case, the liquid collection tanks are 1990 mounted between densifiers, and the rotors drive of all densifiers is done by a single motor, on the axis of which the gear shift GS are mounted, which makes the transition from the angular speed of the motor to the angular speed of that device.
The isothermalizer in Fig.30 cumulates a series of changes that can be made to the isotermalizer with a blade, changes that can be applied separately, or cumulating several of them, depending on the
1995 objective pursued. These changes are:
- changing the shape of stator 6.2 so that a section parallel to the bases is no longer circular, but the new section allows a continuous and watertight slide of the blade, and leads to a favorable change in the isothermal angular speed cuff The stator shape and the curve direction of the blade, determine different isothermal speeds. From this point of view, the construction of the densifier may differ from that of the
2000 rarifier.
- entering a radius of curvature for the blade 6.3 (here, consisting of two sections, 6.31 and 6.32) and implicitly, for the corresponding notch 6.4 in the rotor, change which also results in a change in the isothermal angular speed cuff
- reduction of the rotor radius 6.1 , combined with the realization of a telescopic configuration of the blade
2005 6.3, to increase the useful interior space of the stator. As a rule, this change must be made by making an integer ratio between the length of the inner circumference of the stator and the length of the outer circumference of the rotor.
- making in the rotor of some internal cavities. In the configuration shown in the figure, these cavities are designed to make, through valve 6.65, a passage between the uncompressed gas tank and the low
2010 pressure chamber of the compressor, and through valve 6.66, a path between the high pressure chamber and the compressed gas tank
- inserting the stator into a tank 6.10 filled with coolant 6.101 , a fluid that, with the help of a pump, circulates through the cooling system SR. Sprinklers 6.9b are supplied via pipes 6.9a directly from the tank 6.10.
2015 - changing the kinematics of the apparatus: the inner cylinder 6.1 is held fixed and the outer cylinder 6.2, together with the tank 6.10 and the cooling system SR, mounted on one of the caps, rotate around it.
In Fig.31 the operation of a rotary isothermalizer with a rotor blade is exemplified, when the outer diameter of the rotor 6.1 , with circular section, is equal to the inner radius of the stator 6.2, also with circular section. In the configuration shown in the figure, the friction movement between the rotor and the
2020 stator, existing on the machine in Fig.28, is replaced by a rolling motion of the rotor on the inside walls of the stator. During this free rolling motion, the rotor moves along a circular path on the flywheel 6.81 (Fig.31 , section 1 -1 ) around a shaft perpendicular to the flywheel. In turn, thanks to an adjustable speed motor system, the flywheel rotates around the axis of the stator, at a distance equal to the length difference between the two rays (at the apparatus shown in the figure, where the radius of the stator is
2025 twice the radius of the rotor, this distance is equal to the radius of the rotor). By turning the flywheel, the rotational axis of the rotor moves on a circle with the center on the axis of the stator cylinder, the walls of the two cylinders being permanently in contact on a generator. In this way, the rotor is driven to rotate around its axis. At a complete rotation of the flywheel, the rotor performs exactly two rolls on the inner wall of the stator, passing through two main points on the circumference of the stator (the position in which the
2030 blade is entirely inside the rotor), the position of which is the same at each rotation. On either side of these points are mounted in the rotor wall, inlet holes 6.61 and 6.63, and exhaust holes 6.62 and 6.64, respectively. The opening and closing of these holes is done by means of rotating drawers 6.83, which are driven by the engine 6.8 via the axes 6.82, with a rotation speed equal to the flywheel rotation speed. The position of the drawers in 4 different positions (l-IV) of the rotor is indicated in Fig.31 B, by horizontal
2035 bars for the “closed” position and by vertical bars for the “open” position. Cooling of the gas during compression is done by the sprinklers 6.9b mounted in the wall of the stator, or by itscovers 6.22. As with the compressor in Fig.28, coolant circulation can also be made through the inside of the rotor 6.1 , if it is not used for other purposes. At this ratio of 2:1 between the two diameters, the rotor blade cannot be executed in one piece, and a telescopic blade consisting of two sections 6.31 and 6.32 respectively is
2040 required. Also, in the configuration shown, where the rotor height is higher than that of the stator, to allow the rotor to move, the stator covers 6.2 (the two bases) must be movable in relation to the walls 6.21 : they rotate through bearings 6.91 mounted on the stator walls and through bearings 6.92 mounted on the rotor walls.
Fig. 32 shows how a solid sponge can be implemented, composed of almost parallel plates in
2045 such an isothermalizer. The plates are cylindrical metal sheets 6.12, each with a notch along a generator, with an opening slightly larger than the width of the blade, with unequal diameters, with values between stator and rotor diameter, mounted between these two cylinders, so that, compared to the assembly in Fig.28, the central axis of the rotor is moved towards the central axis of the stator, in the plane containing them, with a distance equal to the total thickness of all these plates, without leaving gaps for gas leaks.
2050 By simply rotating the rotor about its axis, the cylindrical plates of the thermal sponge are engaged in a rotational motion in which the peripheral points of contact with the rotor and those of contact between successive plates move at the same speed, which would lead to different angular velocities of the plates and at pressures exerted on the rotor blade. If these plates are light enough and elastic enough, they can be driven by the rotor blade in a rotational motion synchronized with that of the rotor. Another way to
2055 ensure that all cylindrical plates have the same angular velocity is their successive reciprocal drive. To achieve this mechanism, gear teeth 6.1 m are mounted from place to place, on the outer surface of the rotor cylinder, for example in the shape of triangular prisms (see also Fig.32A). Each such tooth corresponds, in the same plane perpendicular to the central axis, on each plate of the thermal sponge, a hollow 6.12m (obtained, for example, by punching), or a hole, slightly longer than the rotor tooth. These
2060 holes are made in such a way that the holes near the dead center overlap over the corresponding rotor tooth (so the distance between the holes increases as the diameter of the cylindrical plate increases). In this way, as it rotates, the rotor engages with the first plate, this with the next, and so on, equalizing their angular velocities. In the configuration in the figure, the rotor is empty and serves to convey the gas with the inlet
2065 pressure pa, through the valve 6.6a and of the one at the exhaust pressure pr through the valve 6.7a. The coolant 6.4I is conveyed through the rotor blade 6.3, through the pipes 6.9a and through the sprinklers 6.9b, from where it is injected between the sponge plates, the liquid in the stator being discharged, by means of a pump, through a hole made in the lower cover of the stator. .
A very similar construction has the rolling piston isothermalizer in Fig.32B. It is made with the
2070 same design as the state-of-the-art rolling piston compressors, to which has been added the thermal sponge, the sponge cooling system and the measuring-regulation-control devices that ensure the isothermal angular velocity. In this configuration, the vane 6.1 12 that separates the different pressure zones is operated from the outside, using the spring 6.113 and performs back and forth movements in the cylinder 6.1 11 , along a fixed axis. Rotor 6.1 performs a rolling motion inside the stator 6.2. Under these
2075 conditions, the 6.12 plates of the thermal sponge are not engaged in the rotational movement, no additional mechanisms are needed for synchronization.
The isothermalizer in Fig. 33 has a design similar to vane pumps. It consists of a stator 6.2 inside which, tangent to one of the generators of its inner surface, the rotor 6.1 rotates. In the body of the rotor are made several notches 6.4, equally spaced, in which slide the parallelepiped blades 6.3, which, some
2080 springs mounted at the bottom of the notch, keep them in constant contact, along a generator, with the inner surface of the stator. In this way, the internal volume of the stator is divided into several regions whose volume undergoes successive increases and decreases, depending on the rotation angle of the rotor. For judicious use of the entire available volume, in the case of densifiers, the extended regions communicate with each other through the stator wall in a 6.6a portion open to the environment (if
2085 padmisie=Patm') , or to the suction line 6.6 (if Padmisie^Patm) ■ The exhaust valve 6.7a, actuated by a solenoid valve 6.7b, is located in the region where the volume of gas between two successive blades reaches the minimum value. In the case of rarefiers, the direction of rotation of the rotor and the role of the valves are reversed.
As with the single vane isothermalizer, the change in working gas pressure is due to both the
2090 change in the internal volume of these regions due to the movement of the vane and the change in this volume due to the spray of coolant using 6.9b sprinklers fed through pipes 6.9a. The solenoid valve 6.9b is controlled separately for each volume of gas contained between two successive blades. As with the single-blade isothermal, the working gas can also be cooled by replacing the sprinklers with foam generators.
2095 A more efficient use of the internal volume of the stator 6.2 is achieved by a simplified construction of the rotor 6.1 , keeping only its central axis, which can be full (Fig.35), or empty (Fig.34), on which they are mounted pockets 6.4, in which the blades 6.3 slide. Between the surface at the base of these pockets and blades are inserted elastic springs and/or lubricating fluid 6.4I. On the isothermalizer rotor in Fig.34, a cylindrical tank 6.4r is also mounted with a smaller radius than in the case of the device
2100 in Fig.33, and on the outer surface of this cylinder, the flat plates (preferably metal) 6.1 p are mounted radially, which forms the solid thermal sponge, and the radial pipes 6.11, at the end of which the sprinklers 6.9b are mounted. On the isothermal rotor, can be mounted solid thermal sponges which can have also other configurations. In Fig.35, the thermal sponge is made of metal wires 6.3b, which can occupy all the space that in Fig.33 the rotor of the device occupies.
2105 In both cases, the suction of the working gas is made as in the case of the isothermalizer in Fig.33, through a wide opening 6.6a in the stator wall, but for exhaust a 6.7 wide opening is used, with width, measured on the circumference of the circle, slightly larger than the distance between two successive blades.The exhaust opening continues with a pipe mounted at the highest elevation of the stator, if its axis is horizontal. At the stator end of this pipe, a layer of liquid is permanently maintained,
2110 which ensures a complete emptying of the compressed gas, with the final pressure pf, located in the interpaletary space next to the exhaust port (interpaletary space with minimum volume) together with the 6.2I liquid injected by the 6.9b sprinklers. In addition, if there is a pressure difference between the gas in the pipe and the gas in the apparatus when the position of the blades forms a communication path, this layer of liquid acts as a liquid piston, avoiding the loss of exergy that would occur in the case of its
2115 absence. Moreover, with this method, a second, polytropic compression step can be introduced, followed by a cooling of the gas as it passes through the liquid layer. Further rotation of the rotor leads to the emptying of the liquid from the inter-blades space into the tank 6.10 and the intake of the working gas. If the stator axis is vertical, the inlet pipe is mounted on its lower base, through a 6.6v opening, and the discharge pipe is mounted on its upper base, through a 6.7v opening, of the shape and positioning
2120 indicated in Fig. .34 and Fig.35, respectively.
Quasi-isothermal compressions can be obtained just as easily, starting from state-of-the-art liquid ring compressors, with obtaining, at the same discharge temperature, higher compressed gas flow rates, if between the rotor blades of this type of compressor efficient thermal sponges are introduced, similar to those described in Fig.34 and Fig.35.
2125 The construction of the isothermalizer in Fig.36 brings together the characteristics of several types of isothermal described above. It is a solid double-acting piston device consisting of a cylinder 5.1 (not necessarily circular in section) and two covers: one upper 5.1 s and one lower 5.11. Together, they delimit a closed enclosure, divided into two compartments by the piston 5.2, the sealing between them being made with elastic gaskets, segments, etc. The piston is moved between a bottom dead center BDC
2130 and a top dead center TDCn, where n=1 for the lower compartment and n=2 for the upper compartment. Obviously, E>DC1 =TDC2 and TDC1 =E>DC2. The position in figure 36 corresponds to the situation TDC1 =E>DC2. This movement is due to a drive motor, mounted inside one of the compartments, or outside, on the cylinder wall. The motor shaft operates one or more profiled cams 6.14, the profile being executed in such a way that the displacement of the piston is done with the isothermal speed v1 iz.
2135 Through kinematic connections, these cams lead to the telescoping of the rods that move the piston. In the configuration in the figure, the cams move a horizontal bar 6.15 over a short distance, which in turn, pressing the bolts 6.13, lead to the rotation of all the splints 5.23 and 5.24, splints that form the telescopic rods of the piston. When the point furthest from the camshaft reaches its highest position, the piston reaches TDC1. During its ascent, the piston compresses the springs 6.16, mounted between the cover
2140 5.1 s and the horizontal bar 6.15a, which rests on the bolts 6.13a, mounted on the upper telescopic rods. In this way, the downward displacement of the piston is also determined by the profile of the profiled cam 6.14, with the speed v2,z, which may be different from v1iz, due to the constructive differences between the two compartments (these differences can also cause differences in volume, or pressure, which may require the use of different tanks for the storage of compressed gas).
2145 The telescopic rods in the two compartments also serve as carry-supports for the horizontal plates 5.1 1 of the two thermal sponges and are made according to the model of the carry-supports in the isothermalizers in Fig.16. These rods also serve as supports for the 6.9b sprinklers and the 6.9 pipes that feed them (these pipes can be placed right inside the splints that make up the telescopic rods. In this way a continuous cooling of the thermal sponges can be achieved. The cooling circuit is composed of the
2150 6.7M pump which feeds the 6.9 ducts passing through the lower cover, then through the movable piston (strips 5.22), from the lower to the upper circuit, from the liquid layers 8.2s and 8.2I, which collect the liquid from sprinklers, from the 5.22c telescopic rod that collects the liquid from these layers, keeping their volume constant, and from the heat exchanger HE.
In addition, in the upper part of the upper compartment, the free volume that forms in the area of
2155 the elastic springs when the piston is in TDC1 , an additional 8gs sponge is mounted, made of a metal wire. Suction and exhaust of gas are made through valves 6.6I, 6.6s and 6.7I, 6.7s. The compressed gas is exhausted, for both compartments, through pipes containing a layer of liquid, with the role of completely filling the gas bags and with the role of liquid piston for equalizing, without loss of exergy, the pressures.
Fig.37A and Fig.37B show some of the modifications by which other state-of-the-art devices, the
2160 and the cam pump, respectively, can be transformed into rotary isothermal densifiers. In these configurations, the 6.9b sprinklers are mounted in the housing (they can also be mounted in the rotor body 6.14, 6.15, respectively 6.16, 6.17) which inject coolant into the space between the gear teeth. This liquid is the liquid piston of the compressors, the flow rate through each sprinkler being controlled by means of 6.9c valves (for devices with larger volumes, these can be adjustable valves with servomotors),
2165 which receive commands from a central unit, depending on the pressure in closed enclosure corresponding to the respective sprinkler, pressures indicated by piezoelectric transducers 6.8.
As in the case of the rotary apparatus described above, a layer of liquid 6.2I is maintained on the discharge line 6.7, to equalize, without loss of exergy, the pressure in the last inter-blades chamber with that in the storage tank. Moreover, this process can be applied to all types of isothermalizers described
2170 (reciprocating, rotary, solid-piston, liquid-piston) to introduce an additional compression stage, if the compression operation ends at a pressure lower than that of the gas in the storage tank (or of another upstream device). In this case, when the discharge valve is opened, the liquid from the exhaust pipe enters the respective dansifier and acts as a liquid piston, suddenly compressing the remaining gas in the device. As a result, the upstream gas (usually in a large volume) undergoes a slight quasi-adiabatic 2175 expansion, accompanied by a slight cooling, and the gas in the apparatus, a quasi-adiabatic compression, accompanied by a slight heating, followed by heat exchange with the layer of liquid through which it is forced to pass. Also, upon contact between the two quantities of gas, a uniformization of temperatures takes place. The polytropic index of the two operations, compression and expansion, depends, each, on the size of the surfaces with which they exchange heat at that moment.
2180 Fig.38 contains some proposals for the implementation of thermal sponges in the configuration of some types of scroll compressors and some peristaltic compressors of the prior art, in order to bring as close as possible the polytropic coefficient of the transformations that take place in these devices to the unit value. As with the other isothermalizers described above, the proposed objective can be achieved if, in addition to this process, are applied procedures for the angular velocity modification and for the
2185 controlling the cooling processes of the gas in the compression phase by introducing a suitable liquid, in the spray state, by the introduction or generation of foam, or by the introduction of substances in suspension. In addition, in order to obtain an increase in the compressed gas flow rate, the procedures described above can be implemented for the complete discharge of the compressed gas and for the elimination of the dead volume. Since, for the implementation of these processes for scroll and peristaltic
2190 compressors, no procedures other than those described above are proposed, in the representations in Figure 38, they are not described.
Fig.38A shows a cross section through a compressor, with the two spiral volutes 6.18 and 6.19 (here, Archimedean spirals) interspersed. Usually, one of the volutes is fixed, the other performing an eccentric orbital motion, without rotating, but there are also compressors in which, to ensure a safer seal
2195 between the compartments with different pressures, the two volutes rotate simultaneously, in the same direction, but with different centers of rotation. The sealing between the compartments with different pressures is achieved by using 13.6 spiral-shaped gaskets, mounted on grooves made on the ridges of the two spirals.
The thermal sponge is composed of thin elastic metal plates, having the same spiral shape as the
2200 main spirals, about the same length, the same height and the same step, but with a smaller thickness g. They are located between the two main spirals, the distance between them being a multiple of a whole fraction of the distance between the loops of a single volute. For example, if this distance is b, the distance between two successive spirals of the sponge is b/N, where N is the number of plates that make up the sponge. In this way, between every two loops of the spiral considered mobile, are found a fixed
2205 spiral and a number of 2N spirals of the thermal sponge. To prevent the movement of the thermal sponge plates along the spirals, displacement that may be the result of frictional forces that occur at points where the distance between the two main spirals is zero (contact points), the thermal sponge spirals will be longer, exceeding both extremities the respective end of the main spiral, as in Fig.38A, and will be perforated, through the resulting holes being inserted a rod 6.22, fixed to the respective main wing, which
2210 allows the perpendicular sliding of these spirals under the action of elastic forces, but prevents other types of travel. Configurations in which both main spirals rotate in the same direction, in addition to the known advantages of compressors equipped with this type of thermal sponge, offer another advantage: the sliding friction of the contact points is replaced by a rolling friction, which leads to reduce the stresses exerted on the spirals.
2215 If n is the number of plates between the two spirals, plates that have the thickness g, on both sides of one of the main spirals (the fixed one, if only one is movable) a thickness reduction with depth Ng is performed along the entire length of this spiral, over the entire height t of this volute. At the points where the distance between the two main spirals is minimal, this distance is equal to Ng and is filled, in its entirety, by the spirals of the thermal sponge. If the two covers 6.18 and 6.19 are arranged in a
2220 horizontal position, the sponge coils are supported with their lower edge on the lower cover of the compressor, their upper edge being in contact with the upper cover. All these contacts must be tight, at least when the distance between the two main spirals is minimal. The thermal sponge is mounted by temporarily fixing the N thin spirals on each of the faces of the fixed spiral, its thickness thus becoming equal to the thickness of the movable spiral, then the introduction of the movable spiral, followed by the
2225 release of the sponge spirals, spirals that will distance due to their elasticity. Due to the tendency to return to the original shape, the spirals of the thermal sponge will be arranged, more or less evenly, inside the compressor. This configuration has two major shortcomings: the difficulty of eliminating gas leaks between areas with different pressures and the relatively uneven distribution of sponge spirals, the distance between the spirals decreasing successively, as the distance between the main coils decreases.
2230 The compressor shown in Fig.38A manages to remove these deficiencies. First of all, although the spirals of the thermal sponge are also made of elastic metal plates, on each spiral 6.21 is fixed a series of identical plates 6.20, with the width equal to the width of the spirals, but with a much smaller length. When not tensioned, these plates have a shape close to the letter S, as shown in Fig.38C. In this figure is represented unfolded, a portion of the set of main spirals, thin spirals and elastic plates, when the
2235 distance between the main spirals is maximum. The representation of the same assembly, when the distance between the main spirals is minimal is shown in Fig.38B.
The thickness reduction of the fixed spiral is performed only on a portion of its side faces, which leads to the formation of two channels 6.23 (visible in cross section 1 -1 ), one on each side, along the entire length of the spiral. The depth of this channel has the value 2-N-g, if the thickness of the elastic
2240 plates is equal to that of the spirals of the thermal sponge. The height of the thermal sponge spirals is equal to the width of this channel, and the spirals are arranged in such a way that when the distance between the two main spirals is minimal, the sponge spirals, together with the spacer plates penetrate these channels and occupy the entire volume, and in the parts in which the main spirals are spaced at a distance L, the sponge spirals are spaced from each other, with a fraction L/N, the same for all
2245 interspaces. As the number of contact points is greater than 3, the sponge plates always have at least 3 support points in that channel, so that they will never come into direct contact with either the lower or the upper cover. With these modifications, the only locations where pressure losses may occur are located at the contact points, if the respective section of the channel in the fixed spiral is not completely occupied by the
2250 thickness of the secondary spirals and spacer plates. The sealing method proposed in Fig.38A, section 1 - 1 , is the mounting, on the bottom and on the edges of the respective channels, of some membranes made of elastic materials, slightly deformable. The volume formed between this membrane and the channel walls is filled with a fluid and is tightly divided, by deformable walls, into regions with a width not greater than necessary to cover the contact surface between the main spirals. When a point of contact
2255 approaches such a region, the spirals entering the channel press the membrane at the bottom of the channel and push the fluid between the membrane and the bottom of the channel into the regions between the side walls and the membrane, obstructing possible gas leaks, as can be seen in “magnifying glass” 2, which shows an enlarged image of the region in question, when the movable spiral 6.18 steps on the fixed spiral 6.19.
2260 Fig.38D shows one of the possibilities to transform a peristaltic compressor into an isothermalizer, by implementing a thermal sponge. A thermal sponge (Fig.38E shows a cross section through its unfolded shape) consisting of metal plates 5.14, elastic, corrugated, similar to those described in Fig.10a and metal plate 5.11 , flat, rigid, occupying a central position, between two sets of corrugated boards. All these plates have the same dimensions when fully tensioned. In order to prevent uncontrolled
2265 movement of these plates, they are fixed by an elastic cord 5.7, or by an elastic spring, located between the lower and the upper plate. The thermal sponge is inserted into a deformable peristaltic tube, the shell of which is elastic. When the tube is pressed between two jaws, the corrugated plates become flat, and the sponge acquires a shape similar to that shown in Fig.38B, having in addition, the peristaltic tube, which tightly tightens these plates (position 3 in Fig.38E). The same result is obtained, in unfolded form
2270 and when the sponge is made similar to that of the screw compressor (Fig.38C).
One of the disadvantages of compressors with peristaltic tubes is the need to periodically replace, due to wear, their coating. One way to remove this inconvenience is to replace these tubes with metal channels (troughs), for which can be used the configuration used to insert the thermal sponge into the channels made in the fixed spiral of the scroll compressor. A section through such a channel is shown in
2275 Fig.38F. The shaft of trough 6.2, with a rectangular cross section, is a section of an arc of a circle, having its center having the center in the center of rotation of the arm 6.27, the arm on which is mounted, by means of the bearing 6.32, the pressing roller 6.26. As in the case of the peristaltic tube, the ends of the trough are curved and have a path, towards the downstream and upstream device, outside the range of the pivoting arm. Inside the trough is mounted a thermal sponge, similar to those in Fig.38E, or Fig.38F,
2280 and its walls are lined with an elastic sealing membrane, similar to membrane 6.23a section 1 -1 in Fig.38A, Above the thermal sponge, in its completely unstressed state, along its entire length, a lamella is mounted with a width equal to the inner width of the trough. Continuous channels are drilled on the side edges of the lamella, in which a sealing gasket is mounted, transforming the lamella into a real piston. The material from which the lamella is made (metal, or plastics similar to those from which peristaltic 2285 tubes are made) must be sufficiently malleable, so that, under the action of the force exerted by the roller, combined with that exerted by the elastic elements of the sponge, to follow a smooth route, with acceptable radii of curvature, but be hard enough to withstand the stresses to which it is subjected for a long time. The end of the lamella in the inlet area of the pressure roller must be shaped in such a way as to allow gradual entry of the roller and must be fixed in relation to the trough so as to prevent movement
2290 along it. Likewise, the other end of the blade must be shaped in such a way as to allow easy and complete evacuation of the compressed gas. If the material of which it is made does not have sufficient longitudinal elasticity, this end may be allowed to slide freely on the bottom of the trough.
Although there are also linear peristaltic compressors, the most used are those with circular or spiral tubes. In all these variants, thermal sponges can be mounted inside the tubes. In the configuration
2295 of Fig.38D, two peristaltic tubes 6.25 are used which have the shape of segments from a circular ring and are fixed on a metal bed 6.24, so as to allow the penetration into the circular space between them, a pressing roller 6.28. This roller is mounted on a movable arm 6.27, which is continuously rotated by a drive motor 6.26. A higher power density can be obtained if several pressing rollers are mounted on the same arm, each such roller performing the action of compression on two circular peristaltic tubes, with
2300 different radii of curvature, equipped with an internal thermal sponge. Each of the peristaltic tubes is provided, at the end of the support on which it is mounted, with a check valve (in some applications, the inlet valves can be dispensed with). A single pressing roller acts on each peristaltic tube. When this roller reaches the peristaltic tube, both valves are closed and the tube is filled with gas at the initial pressure. As soon as the roller passes the first end of the tube, the inlet valve opens so that another portion of gas
2305 enters the rear portion of the tube at the initial pressure p, while the gas in the tube is progressively compressed to the final pressure pf At this point, the exhaust valve opens and the roller acquires the role of evacuating the compressed gas to the user. The two valves close simultaneously when the press roller reaches the second end of the peristaltic tube. Configurations can also be made in which, on the metal bed, there are N circular tubes of equal length, shorter, each provided with its own valves. For each such
2310 tube a pressing roller is mounted, at the same distance from the center of the device, on separate arms, which make equal angles between them.
The use of peristaltic tubes can be extended to other types of compressors. For example, on the inner wall of the stator of vane compressors, peristaltic tubes with thermal sponge can be mounted, the tube being pressed and the gas being compressed by the rotational movement of the vanes. The final
2315 pressure of the gas in the peristaltic tube is usually different from that of the gas inside the compressor and has a different destination, but by careful sizing, the two compressors can become two stages of compression of the same process.
Another type of compressor in which peristaltic tubes equipped with thermal sponges can be used are scroll compressors, which by this method simplifies the problem of friction and tightness of the
2320 compartments. For example, in the scroll compressor in Fig.38A, is made the thickness reduction of fixed blade, on both sides, up to the level 2 Ng+2g1, (where g1 is the wall thickness of the peristaltic tube in which a thermal sponge is inserted whose thickness in fully compressed state is 2 Ng), then the two tubes in fully compressed state are fixed on both sides of this spiral, then the movable spiral is inserted, after which the two thermal sponges are released. Under the action of the forces exerted by the elastic
2325 elements of the thermal sponges, the space inside the compressor will be occupied, almost entirely, by the peristaltic tubes. Note that in this configuration, the distance between the two covers can be increased, so that the tips of the two spirals no longer touch the lid of the opposite spiral. The orbital movement of the movable spiral, or the rotation in the same direction of both spirals, has the effect of a peristaltic pressure and leads to the compression of the gas in these tubes.
2330 Rotary compressors are especially useful for high gas flow rates, for low compression ratios. To obtain higher pressures can be used the pressure step method. They are ideal for supplying precompressed working gas, at temperature Tiz, for densifiers with high compression ratios.
The proposed procedure for increasing the energy efficiency of isothermalizers, described in Fig.2, can also be applied to rotary isothermalizers. Since these isothermalizers usually contain several
2335 enclosures in which the compression/expansion phase occurs at different times, it is advisable to use, from the very beginning, an alternating current motor to drive the rotor of the device, and the isothermal trajectory to be obtained by regulating the flow of the heat transfer agent, both at the entrance and at the exit of the enclosures, with the help of flow regulation valves, operated by automatically controlled servomotors, according to the gas pressure in each enclosure. The DC motors that control these flow
2340 rates can be replaced with AC motors, as in the case of reciprocating piston isothermalizers.
Another type of compressor that has multiple gas compression/expansion chambers is the rotary screw . An isothermal compression for this type of compressor can be achieved by using rotors with variable pitch, decreasing towards the exit (in this way, the ratio between the surface of the heat absorbing elements and the volume of the respective enclosure increases, as it does approach the
2345 end of compression operation of the gas in enclosure), and the installation of systems for carrying out heat exchange by abundant spraying with liquid, the inlet and outlet flows in/from each enclosure being determined by the automatic control of the valves on the liquid circulation pipes. The sprinklers for dispersing the heat transfer liquid are mounted in the housing of the device and/or in the metal body of the rotors, fed from a channel that passes through their axis.
2350 Other methods for carrying out heat exchange use substances in the form of foam or oil droplets (or other liquid with high volatility at working temperature) in suspension as a heat transfer agent.
A new type of isothermalizer, capable of performing energetically efficient isothermal transformations, is the gas piston isothermalizer, whose principle diagram is represented in Fig.39. In principle, this isothermalizer has a first stage 8.1 , which includes one or more compressors and solid or
2355 liquid piston isothermalizers, connected in series or in parallel, which discharge into a tank 8.2i, which constitutes the second stage and whose volume is significantly larger than that of the devices in stage 8.1 , both stages being equipped with heat transfer control systems between the gas and its ambient environment, the pressure and temperature of the gas in each stage being controlled by changing of the speeds of the actuation devices and of the flow rates of the heat transfer agent, by an automatic
2360 regulation system.
For an isothermal compression, it is necessary that the average temperature of the gas in each compression stage be constant, equal to the preset value Tiz. In Fig.39A and Fig.39B, two more variants are proposed for achieving the optimal AIA trajectory of isothermal compression, in which, an amount of thermal energy equal to that accumulated/given up by the gas as an effect of the compression/expansion
2365 in each stage, is given/absorbed instantly into/from its environment. The enclosure of both stages is continuously cooled/heated to keep the gas temperature constant, and the amount of energy input to the two enclosures is varied by the controller's control of the isothermal speed.
In a long compression cycle (the time required to obtain the final pressure in the second stage), the first stage compressors must ensure, at its output, in each short cycle (we will call a short cycle, a
2370 cycle of the first stage devices), a quantity of gas with a temperature equal to Tiz, but with a variable pressure, always equal to the gas pressure of the second stage, a requirement that can be satisfied by various combinations of compressors, whose valves open automatically, or are operated by commands issued by a controller, depending on various continuously measured parameters. For this, in the first stage 8.1 , to perform the adiabatic jump, the procedure described in Fig.3 can be applied or an alternative
2375 compressor C1 can be used, whose polytropic coefficient is close to the adiabatic one. At low values of Tiz, a quasi-adiabatic compressor is preferable (eg a rotary blower, which can provide a sufficiently high flow rate at the Tiz temperature, with reduced dimensional characteristics than a reciprocating adiabatic compressor). This compressor supplies compressed gas with a pressure p1 and a temperature Tiz=Tamb+AT. The isothermal compressor (isocompressor) C2 raises the gas pressure from the value p1 to
2380 a variable value pr, equal to the gas pressure of the second stage, keeping the gas temperature constant. Since, in the adiabatic compressor C1 , the working gas is compressed in a short time, and the isothermal compression that follows requires a much longer period, a high-speed adiabatic compressor with a much smaller internal volume than the isothermal compressor is indicated, which flows into an intermediate tank R, in which the pressure p1 is kept constant, by equalizing, through the play of the valves, or by changing
2385 the revolutions, the input flow with the output one. The pressure p1 is the inlet pressure in the isocompressor (densifier) C2, and the discharge pressure (and implicitly, the compression ratio) is variable, being equal to the pressure pr of the second stage, a pressure whose value changes very little during a cycle short of the C2 isocompressor. The pressure pr, entering the second stage, varies during a total cycle (long cycle), between the pressure p, and the final pressure pf. If the final pressure pf has a
2390 high value, the C2 isocompressor can be made by connecting in series several densifiers. For the supply of the second stage with as short interruptions as possible, or without interruptions, it is recommended that the C2 densifier work with a double-effect piston, or with several densifiers in parallel.
If, when starting the system, gas is found in the tank 8.2i at ambient pressure, the compressor C1 will directly pump into this tank (or, the piston of the compressor C2 will only carry out the transvasation of
2395 the gas, without compressing it), until in this tank (which constitutes second stage) pressure p1 is reached. Due to the large amount of heat absorbed by the sponge, part of which is given to the cooling system, the temperature of the gas in this tank increases very little (the change in temperature difference AT being insignificant). Then, the discharge of the compressor C1 switches to the tank R, at this time the densifier C2 also starts, initially with a compression ratio equal to 1 (C2 discharges all the gas drawn into
2400 the tank 8.2i). Gradually, this compression ratio increases until the pressure pf is reached in the 8.2i tank. In this phase, the speed of the piston of the C2 densifier, during compression, follows the optimal trajectory, corresponding to the coolant flow and the heat exchange surfaces between its sponge and the coolant (if the total of these surfaces does not change, nor the curve according to which it varies the flow rate of the heat transfer liquid, the power developed by its drive motor in the compression phases, must
2405 remain constant). In the discharge phase, the piston actually performs a slight compression of the gas in the 8.2i tank, it having to evolve on a new isothermal trajectory (the new isothermal compression process will be exerted on the compressed gas pumped out by C2, to which is added all the gas existing in tank 8.2, also taking into account the heat given up in this tank, to the heat transfer liquid).
In an open system, similar to those used in ICAES type storage systems, in the first stage the
2410 tank 8.2i is filled with a hydraulic agent, the densifier 02 works with a minimum discharge pressure p2, corresponding to an economic regime . It supplies in the second stage gas with the constant pressure p2, gradually replacing the hydraulic liquid here (the mechanical energy consumed for pumping being recovered in a hydraulic motor/pump 8.6M, Fig.39A) , until the liquid in the tank is exhausted, when close the discharge valve. Then the compression continues with the tank 8.2i closed, with the gradual and slow
2415 increase of the compression ratio of the densifier 02, the discharge pressure having the value pr, intermediate between p2 and pf. Another recommended solution is that this first phase, in which the densifier works with a constant compression ratio, pumping the compressed gas into the tank with constant pressure p2, has a limited duration, after which the pumping operation of each tranche of gas compressed, from the first stage, to be accompanied by the removal from the reservoir 8.2 of a smaller
2420 volume of liquid than in the previously described configuration, which has the effect of a slower increase in the compression ratio of the gas in the reservoir 8.2 and provides the possibility of to control, more effectively, the increase in compression power. The timing of these operations can be made so that the pressure pf is reached with the discharge of the last tranche of liquid. At this moment, the discharge valve 8.2s and the valve 8.2r are opened to admit the additional amount of liquid (after closing the cooling
2425 circuit), which will transfer the gas from the tank 8.2 to the tank 8.7 under constant pressure. The replaced fluid is collected in a reservoir after passing through the 8.6M hydraulic motor. Then, these phases (long cycles) are repeated until the tank 8.7 is filled under constant pressure. Throughout the compression, the temperature of the gas in both stages is maintained, thanks to the cooling system, at the Tjz value.
2430 For systems that do not require storage, tank 8.7 is no longer needed, but an additional tank 8.2 is added to the system. In this way, after filling one of the tanks 8.2 with compressed gas, the discharge of the first stage switches to the second tank 8.2 and begins the compression of a new tranche of gas, while the gas from the first tank is delivered to the upstream device, through the transvasation of the liquid in the RL tank, using the 8.6M1 pump.
2435 In Fig.40, in the first stage 8.1 , two densifiers C2 are used, in parallel, to double the flow of compressed gas and, through their alternative operation, to reduce the periods when, in the second stage, no gas is introduced. When the pr pressure is high enough, through an automatically controlled valve system, the two densifiers can switch to series operation.
In the configuration of Fig.39 and Fig.39A, the tank 8.2i is inserted into another parallelepipedal
2440 tank 8.2, partially filled with a cooling liquid 8.2I. The outer tank 8.2 is filled with liquid up to a certain level, above the upper level of the tank 8.2i. A layer of gas is left in the upper part of the tank 8.2, which, through pipe 8.2c, constantly communicates with the gas in the inner tank. Therefore, this gas layer is also compressed, at the same pressure, by the action of the densifier piston 8.1. To cool this gas, a thermal sponge is installed consisting of a wire metal inserts 8gs, which in turn can be cooled
2445 continuously, or periodically. The fluid level in the tank 8.2 is kept constant by the 8.6M pump. In turn, the liquid in the tank 8.2 is cooled by its inclusion in a system that also contains the liquid-gas heat exchanger HE. The cooled liquid in the heat exchanger is also directed to cool the thermal sponges of the first stage densifiers. Moreover, for rapid cooling, the first-stage isothermalizer can also be inserted into this tank. Through the valve 8.2r, the amount of liquid in this circuit can be supplemented to cool the sponge
2450 component 8gs, to discharge the gas when it reaches the required pressure, or to transform this agent into a liquid piston.
For an isothermal compression, it is necessary that the average temperature of the gas in each compression stage be constant, equal to the preset value Tiz. The most effective solution for the gas in the 8.2i tank to be maintained at this value is to insert a non-deformable thermal sponge into the tank,
2455 whose envelope is made up of the very walls of the 8.2i tank, a sponge that is permanently cooled by a adjustable system to maintain the average gas temperature at the preset value Tiz.
In the second phase of compression, the gas discharging from the last densifier of the first stage begins each time the gas pressure equals the pr pressure in the storage tank. During the exhaust, the volume of compressed gas between the solid piston and the tank acts on the gas in the tank as a gas
2460 piston, increasing with each cycle the compression ratio. In the tank 8.2i, part of the thermal energy contained in the compressed gas from that given by the piston during the exhaust (exhaust which does not occur at constant pressure, but is in fact a compression with a small compression ratio) it is taken over by the thermal sponge and is given over to the cooling system, including the time for the next compression phase from the enclosure of the solid piston compressor, when its exhaust valve is closed
2465 and the tank 8.2i becomes a simple heat exchanger. Due to the imposition of an isothermal speed viz1 for the densifier 8.1 for the entire compression duration tiz1 and an isothermal speed viz2, for the same piston for the entire compression duration tjz2 o the gas exhaust (equivalent to an isothermal compression of the gas in the tank), the temperature of the gas in the tank can be maintained at a valuer^, close to the ambient temperature. During the following phases executed by the piston, the suction and the 2470 compression (with a slightly higher compression ratio), the exhaust valve is closed and the gas in the tank continues to cool, suffering also some pressure reduction.
In another configuration, when the densifier 8.1 terminates its exhaust operation, another similar isothermal compressor can reach the exhaust phase, the gas in this compressor having also the temperature of Tjz and a pressure equal to the new pressure in the tank. The exhaust phase for this
2475 compressor requires a higher mechanical energy consumption, therefore the thermal energy that must be absorbed by the tank sponge increases, and the time required for the operation is a little longer. At the end of this phase, another isothermal densifier, after his compression is completed, is ready for a new exhaust operation, at a slightly higher pressure and with a slightly longer duration. This densifier is followed by as many densifiers as necessary for the first densifier to complete the isothermal compression
2480 cycle, with a compression ratio higher than in the first cycle, and be ready for the next exhaust. The sequence of these phases is repeated until the desired pressure is reached in the tank 8.2, then the gas in the tank is transferred through the valve 8.2s into a storage tank 8.7, by replacing it with pressurized fluid.
Another strategy that can be applied for an isothermal compression is to use a single isothermal
2485 compressor in the first stage, but to accelerate the speed of the piston, increasing the temperature difference between the gas and the sponge, and thus obtain in the tank 8.2i a temperature Tiz+ higher than Tiz, so that during the time that the compressor piston 8.1 performs the intake and isothermal compression phases at the temperature Tiz+, the gas in the tank 8.2i cools below the temperature Tiz. permanent non-deformable thermal sponge can be placed in the densifier 8.2i, as its piston is a
2490 gas piston. Any solution used at the prior art to reduce the polytropic coefficient can be chosen for its realization, but superior efficiencies are obtained when installing thermal sponges with a heat absorption surface as large as possible, where there is the possibility of easy circulation of gas and liquid through the channels and holes of the sponge, reducing to the maximum mechanical energy losses by friction, losses which would be converted into thermal energy which should be eliminated to the environment and where
2495 there is an efficient coolant well distributed in the enclosure, the flow rate of which is correlated with the thermal capacity of the sponge and can be adjusted as the pressure in the tank increases (i.e. the heat to be eliminated). In the case of this type of densifier, in the absence of a moving piston, the surface of the thermal sponge in direct contact with the gas and can be greatly increased. For example, the tank 8.2 may be the primary of a plate heat exchanger (Fig.27), the secondary of which is part of a cooling/heating
2500 circuit equipped with circulator pumps and another heat exchanger, which gives the absorbed heat to another medium, thus benefiting from a large heat exchange area. Of the same large surface area, but with a higher heat transfer coefficient, has a part the secondary of the exchanger in which a fluid at the saturation limit is inserted, which by evaporation, followed by a condensation in an external condenser eliminates excess heat.
2505 For the tanks 8.2i in Fig .39 we chose parallelepipedal configurations, the gas 8.2a in these tanks being cooled by a fixed sponge, composed of tubular cylindrical supports 8.8v of some sprinklers inside which coolant for sprinklers circulates, and of horizontal metal plates 8.8o fixed to these supports. The spray of the sprinkler and that accumulated at the base of the tank in a layer whose level is kept constant by the pump 8.6m forms the itinerant liquid sponge which, together with the liquid in the tank 8.2, is the
2510 section inside the tank of a cooling circuit. This inner section also includes the sprinklers 8.8a mounted on the top wall of the tank, and the sprinklers mounted on the vertical supports 8.8v, each spraying liquid or foam, in the corresponding horizontal plane between the solid sponge plates. In turn, horizontal plates may have perforations to create longer routes for fluid leakage. To achieve this, the surface of the plates gradually decreases, from the lower plates to the upper ones.
2515 The density and complexity of the thermal sponge system differ greatly, from one application to another, depending on the final compression ratio and the temperature Tiz. For example, for many applications it is sufficient, as with quasi-isothermal compressors of prior art, only by a itinerant liquid sponge distributed by ceiling sprinklers (but, in the case of gas piston isothermalizers, the density of the fluid droplets can be maintained at very high values throughout the isothermal transformation). In
2520 contrast, the input and output flow of the liquid agent are equal, the role of the solid or liquid piston that would lead to the decrease of the heat exchange surfaces being taken over by the gas piston, which keeps these surfaces unchanged. The tank may also be cooled with foam, if a foam generator is mounted on the pipe that introduces the gas into the tank and/or a connection is installed between the compressed gas pipeline at the inlet and the liquid layer mixed with surfactants at the bottom of the tank. On this pipe
2525 and on those supplying the sprinklers are fitted the blowers S, necessary to create a local overpressure and to adjust the flow rate of the gas-piston. In this case too, there are advantages over the state-of-the- art compressors: keeping a large area of heat exchange at all times, superior possibilities for distributing and regenerating the foam.
The second stage of the isothermalizer in Fig.39A is equipped with a thermal sponge made of
2530 vertical plates 8.8v, very close to each other, having the same shape as the respective cross-section through the tank, but with slightly smaller dimensions. The envelope of the sponge is constituted by the very walls of the tank 8.2i, and this sponge is continuously cooled by injection of coolant by an adjustable sprinkler system 8.8a, which maintains the average gas temperature at a predetermined value Tiz, the coolant being maintained at an average Tamb temperature thanks to an HE heat exchanger. Since the
2535 pressure of the liquid in which the 8.2i tank is immersed is equal to that of the gas in the tank, the sprinkler system can be replaced by a system of horizontal perforated pipes, mounted on the top cover, from which, under the action of gravity, water droplets emerge and they drip at low speed, on the vertical plates. In this way, the contact time between the heat transfer liquid and the metal plates, as well as between the liquid and the gas between the plates, is much higher than in the state-of-the-art systems.
2540 The tank 8.2i in Fig.40 is also parallelepipedal, the gas 8.2a in this tank being cooled by a system composed of a deformable metal strip 8.2b, (similar to the strips used for transporting small materials), with a width almost equal to the width of the tank, which runs permanently on a system of rollers mounted inside or outside it. In turn, the entire system is inserted into a parallelepiped tank 8.2 filled with coolant 8.21. Due to the mechanical energy received from the outside by the drive rollers, the metal strip travels a
2545 winding path, most of it inside the tank 8.2i, and the other part in the tank 8.2, being able to have (Fig.40a), only two crossing positions from -one area to the other. In another constructive variant, the route can be distributed alternately in both tanks. When the tape passes through the metal walls, sealing gaskets or sealing rollers with a smaller diameter are provided. The longer this path and the closer the adjacent portions of the band segments are, the greater the thermal energy absorbed from the
2550 compressed gas at a given speed. For bigger strip lengths, the rollers can be arranged so that the metal strip has the cooled portions (outside the tank) as long as possible. In Fig.40a, another configuration of the location of the drive rollers is represented: they are mounted side by side, in contact each other, so that the two sets of rollers form the two walls: upper and lower, and the metal strip passes as tightly as possible between two such rollers. In this configuration, seals are required only for the peripheral rollers.
2555 In this situation, after passing between two rollers, some segments of the metal strip can travel an additional route through the coolant. Also shown is a configuration where all but two of the rollers are located inside the tank. The system is not difficult to achieve, as the passages between the two reservoirs, 8.2 and 8.2i do not have to be particularly tight, the liquid spills on the surface of the metal strip having a favorable effect on the elimination of excess heat
2560 The outer tank 8.2 is filled with fluid to a certain level, above the level at which the tank 8.2i and the cooling strip drive systems are located. At the top of the tank 8.2 remains a layer of gas which, through the pipe 8.2c, constantly communicates with the gas in the inner tank and into which the metal inserts 8gs are mounted. By achieving equal pressure between gas and liquid layers, the tank 8.2i can be made with much thinner walls, regardless of the total compression ratio. Secondly, most importantly,
2565 seals between fixed and moving parts, or between moving parts, are not subject to high pressures, which allows the volume of fluid 8.21 entering the tank 8.2i and the equal volume of gas passing into the top layer of the tank 8.1 to be reduced to low values. Also here, the fluid level in the tank 8.2 is kept constant with the pump 8.6m. This way, liquid infiltration in the second stage helps to improve the quality of cooling. In turn, the liquid in the tank 8.2 is cooled by including it in a system that also contains the pump
2570 8.6M and the HE liquid-gas heat exchanger. When the pressure in the second stage of the compressor reaches the desired value, the compressed gas can be exhausted by the valve 8.2s into the constant pressure tank 8.7, by inserting additional liquid into the tank, with the pressure equal to the final pressure of the gas, until the entire amount of gas has been transferred.
In the configuration in Fig.40A, a thermal sponge made of metal strip 8.2b is inserted into the
2575 tank 8.2i, but due to the fact that the connection between the tank 8.2i and the upper gas layer (in which the sponge 8gs is mounted) is made through a channel with a larger section, the route of the cooling band no longer passes through the walls of the 8.2i tank, but through this channel, to be directed to the coolant volume.
Similar to the cooling system in Fig.39A, and in the case of metal sponges made with movable
2580 metal bands, a system of horizontal perforated pipes or a horizontal metal plate, provided with perforations, can be mounted in the upper part of the 8.2i tank, which will insert heat transfer fluid for additional cooling of the metal strip and the gas inside.
Another type of compressor, capable of performing isothermal compressions with high energy efficiency is the gas piston densifier represented in Fig.41. In principle, this densifier also consists of an
2585 isothermal first stage 8.1 (here, a solid piston densifier with thermal sponge made of horizontal plates 8.1 a. mounted on harmonic-type supports 8.1 r), or any combination of compressors and densifiers that discharge into collector 8.3. For an effective control of the flow rate and temperature of the gas absorbed by the second stage, in Fig.41 is also installed a polytropic compressor 8.4 (which can be a screw compressor), followed by a heat exchanger, circuit that can deliver to the output his, an additional flow of
2590 gas with temperature Tiz and pressure pr. During a short cycle, the pressure in the densifier 8.1 starts from an initial value p, (usually atmospheric) and increases to the instantaneous reservoir pressure pr, and the reservoir pressure increases little by little, in the intervals where, after opening discharge valve, the compressor piston discharges the compressed gas. In the configuration in Fig.41 , we chose the solution of an additional cooling of the compressed gas, by introducing it in the form of bubbles in the discharge
2595 pipe 8.3. Also, the liquid in this pipe, which is constantly recirculated, also ensures the evacuation of the compressed gas from the densifier. The gas from the other compressors discharging into line 8.3 can be cooled in the same way, or with the help of an additional HE heat exchanger. The conditions that ensure a superior efficiency of the system are: a large and unobstructed section of the gas inlet and discharge paths, a sensitive, fast opening and with low pressure losses of the discharge path, the existence of a
2600 liquid fraction of the thermal sponge of whose volume should be adjusted (before, or during operation) so that every time the piston reaches TDC, the dead volume of the compressor cylinder 8.1 is equal to zero.
The opening of the inlet and outlet valves 8.1s (Fig.41 ) is done at the command of the regulation system. The discharge valve can open at a certain pressure of the gas in the manifold (fixed position of the T point of the piston 8.1 ), in which case the compressed gas is directed to the tank under constant
2605 pressure 8.7 (each tranche of gas will replace a tranche with the same volume of hydraulic liquid in the tank), or at a variable pressure, equal to that in the tank 8.2, in which case the compressed gas is directed to the tank 8.2, where it will be further compressed, and after reaching the desired pressure in this tank, it can be directed to the tank under constant pressure 8.8.
The best efficiency of using the mechanical energy received by the system is obtained when in
2610 the first stage an isothermal compression is carried out, up to a pressure pf1, seeking to achieve a difference AT as small as possible (when the main compressor of this stage is a densifier) . There are many applications where there is also a demand for thermal energy (cogeneration systems, high power density energy storage systems obtained by thermal energy storage, etc.). In such situations, as well as when a high energy conversion speed is required, the use of adiabatic compressors in the first stage of
2615 the system is justified. The device in Fig.41 can also perform the task of storing thermal energy. For this, the first stage of the device contains an adiabatic compressor that discharges into the primary of a HE heat exchanger, at a pressure equal to that in the collector (pressure that increases as the pressure increases in the second stage. The HE exchanger has the role of reducing the temperature of the gas at the compressor outlet at a temperature T/z, as close as possible to Tamb and the thermal energy given to
2620 the agent in the heat exchanger secondary is stored in a thermal reservoir, while the mechanical energy required to compress the gas in the second stage (due to the gas piston formed) is stored, after reaching the pressure pfi in a constant pressure tank, where the temperature is close to Tamb. Most of the time, the recovery of the excess thermal energy in the exchanger justifies the use of simpler (and cheaper) compressors ), with higher outlet temperatures, but with higher flow rates. The second stage of
2625 compression is realized as a gas piston densifier, with a compression ratio on a short cycle, very low (the inlet pressure pf1 and the intermediate pressure pf2 being close in value), associated with an efficient system for removing excess thermal energy. Until the intake valve opens, the gas pressure in the second stage may suffer small oscillations caused by the cooling system. Also, the total surface area of the solid elements of the sponge that take heat energy from the gas being compressed does not change. After
2630 reaching the pressure pF (after a long cycle), the gas is transferred to the tanks 8.8. by means of a liquid piston.
The second stage of the gas piston densifier 8.2 in Fig.41 is the primary of a plate heat exchanger, whose gaskets must be sized for pressures higher than pf2. Through the secondary 8.2a of the exchanger 8.2, a coolant circulates continuously in closed circuit, at a speed depending on the
2635 instantaneous compression ratio. This circuit includes, externally, a fluid/ambient HE heat exchanger. The input of the primary circuit is coupled to the collector line which in turn is coupled, via exhaust valves, with the compressors in the first stage, or to the HE heat exchanger, while the output of the primary circuit (used for discharging the compressed gas) is usually closed.
. A considerable increase in the efficiency of the entire system, by reducing the difference 21 , can
2640 be achieved by introducing a refrigerant in a state of equilibrium liquid/vapor, in the secondary of the densifier, and implementing in this system the necessary pipes for this secondary to become a thermal tubes cooling system (gravitational or with capillarity). In another variant, by putting a refrigerant in a near equilibrium liquid/vapor state in the secondary, it can become the evaporator of a Rankine-cycle, or ORC thermal engine. To realize the thermal engine, the heat exchanger HE is replaced by a condenser, which
2645 receives the agent through the turbine 8.4 and exhausts the condensate with the pump 8.6. In this configuration, the speed of all pistons in the compression phase can be increased, because the additional mechanical energy consumed, due to the increase in temperature difference 21 , is fully recovered in the thermal engine
The existence of gas tanks under constant pressure 8.7 creates the possibility of a new
2650 compression stage. After storing the gas in constant pressure tanks 8.7 and 8.8, the compression may continue with the extraction of compressed gas from these tanks and its compression to a higher pressure value, the additional compressed gas may be stored in the tanks from which it was extracted (via a intermediate tank), or in tank 8.2. A variant of this densifier is shown in Fig.42. The first stage of this densifier consists of two
2655 identical densifiers 8.1 , similar to that of Fig.41 , and the second stage, from the tank 8.2a, into which a tank 8.2g is inserted, made in the form of a comb. The combination of the two tanks is a heat exchanger 8.2. Each of the two tanks is a gas piston densifier, the exhaust of which can be directed during the working regime, to a cooling system, containing an ambient-gas heat exchanger and a blower C for driving the gas or, after reaching the desired pressure, toward constant pressure tanks 8.7. The system
2660 also offers the possibility that, for large compression ratios (with the consequence of increasing the heat to be exhaustefd) it will go into forced mode, in which only one of the first-stage densifiers is in operation, and its second stage is cooled by the heat exchanger formed by the tank of the second densifier and the heat exchanger concerned.
In Fig.42A another type of gas piston densifier is shown, consisting of compressors system 8.1
2665 and the tank 8.2. An advantageous solution for cooling the walls of the tank is to install this tank inside a larger tank with liquid, in which the pressure is maintained at all times equal to that of the tank 8.2. This allows the walls of this tank to be thinner, allowing faster heat escape.
Gas supply at temperature T, and variable pressure pf1 is made by a system of densifiers, compressors and heat exchangers, with a common collector 8.3, similar to that of Fig.41. Inside the tank
2670 is installed a thermal sponge according to the invention, designed according to the characteristics and requirements of the system, adapted to the gas supply system and to the cooling system. In Fig.42A, the main part of the sponge is a system 8.8v of bars of various thicknesses and vertical plates of various widths, arranged at sufficiently small distances from each other (to achieve a good capture of the heat accumulated by the gas in the tank), but large enough to allow a slight leakage of the coolant.
2675 The liquid required to cool the gas is distributed between the bottom of the tank 8.2, between the heat exchanger HE, the sprinkler system 8.8a, the pump body 8.6 and the piping system 8.3. The system allows the installation of any type of sprinkler from the state of the art, their number and distribution, the flow rate and pressure difference, the dispersion angle, the size of the drops produced and other characteristics, being chosen according to the characteristics of the application. In many applications, a
2680 sprinkler system is enough to generate a dense and permanent fog of very small drops, the support of sprinklers taking on the role of solid sponge. It is also advisable to create areas with different temperatures that generate ascending gas currents. In the configuration in Fig.42A, a sprinkler system mounted on the top of the densifier was chosen, which spreads the liquid droplets horizontally. A significant amount of this liquid remains attached to the solid sponge elements, helping to limit the
2685 temperature increase. In addition, in this configuration, the coolant contributes to the compression and cooling of the gas sucked from the first stage, through the bubble densifiers 8.11 , made right on the coolant transport pipes, by injecting them with gas from the first stage collector. From here, the gas reaches the sprinklers and is exhausted in the densifier, even in the areas with the highest temperatures of the gas. If the gas from the first stage is injected into a beam of pipes with a sufficiently small diameter 2690 it will not form bubbles, but will form successive layers of gas, due to the surface tension of the liquid, alternating with layers of liquid, which improves the conditions of heat transfer between the two media.
The chosen configuration also uses other cooling processes from the technical stage, namely, increasing the heat transfer surfaces by introducing, or creating aqueous foam. To achieve this desideratum, in the liquid phase 8.11 of the thermal sponge at the base of the densifier, and in the liquid
2695 trays 8.9 mounted inside the thermal sponge, a number of surfactants are added to reduce the surface tension of the liquid, this favors the formation of foam when compressed gas is introduced into the liquid from the collector of the first stage
The great advantage of gas piston isothermalizers is their ability to store in their thermal sponge, energy in the form of heat, which greatly expands the range of applications in which they can be used
2700 successfully. In storage mode, the method that allows the accumulation of all the energy brought as input by the piston of the isothermalizer C2 and the energy from other sources is to abandon the HE heat exchanger (which had the role of giving up the excess heat to the environment) and replace it with a thermal energy storage tank. The mechanical work performed by the pistons in the first stage is converted into thermal energy, part of which remains in the thermal sponges, in the walls of the compressors and in
2705 the liquid in the system tanks, and another important part is transported by the heat transfer agent to the external storage tanks. In these conditions, in which there are no necessary exchanges of heat with the external environment, for the proper functioning of the system, the best possible thermal insulation of the entire system is necessary. A thermal insulation with superior characteristics brings important advantages (increased energy efficiency) in the case of all the isothermalizers described, but in the case of
2710 applications with heat accumulation, especially if the storage time is longer, the insulation of the system plays a very important role. The invention also discloses a highly efficient active thermal insulation method. Note that, in the case of a perfectly thermally insulated system, including the energy consumed in gas-gas, gas-liquid, gas-solid, liquid-solid and solid-solid friction processes in compressors and connecting pipes, the resistive and magnetic losses of the motors mounted inside the system, which
2715 allows round-trip storage efficiencies very close to 100%.
For applications where thermal energy storage is also carried out, Fig.43 proposes a possible configuration of a gas piston isothermalizer with heat storage. In its first stage 8.1 , to the previously described system, consisting of compressors C1 and C2, the adiabatic compressor C3 is added, which takes from C2 the gas discharged by it, with a temperature Tiz1 and a pressure p2 and compresses it
2720 adiabatically up to the pressure pr of the gas in the tank 8.2. In this tank, the instantaneous temperature of the gas is Tiz2. In order not to destroy the available exergy, the system must be made in such a way that the instantaneous temperature of the gas at the outlet of the compressor C3 is equal to Tiz2.
If the cooling system of the sponges (which ensures the uniformity of temperatures in each enclosure, both for gas Tizi, and for thermal sponges: Tamb=Tizi-ATi,, (where i=1 ,2) is common, as in the
2725 figure, equal temperature differences zlT, and 21 T2 can be obtained, in which case compressor C3 is no longer needed. Also, for equal Tamb temperatures, it is possible to work in the two stages with different temperature differences, in which case the C3 compressor ensures in the second stage an increase in the temperature difference, in order to increase the amount of thermal energy stored. If both stages have their own cooling system, the temperatures of the gas and the thermal sponges in the two stages are
2730 different, the energy storage is done independently in each stage, and the role of the C3 compressor is to, together with the C2 densifier, achieve a ratio of compression that ensures equality between the gas pressures at the exit from the first stage and the one entering the second stage. The system is very flexible, allowing the choice in each step of how the temperature of the thermal sponges evolves, and allows changing the temperature difference (therefore of the power consumed) in each cycle, in each
2735 step, without generating exergy losses. The only precaution, easily accomplished, is that the inlet pressure in each compressor does not differ from the discharge pressure of the downstream apparatus.
For simpler applications, Tiz1 can be permanently maintained at the same value, with the first stage cooling system releasing excess heat to the environment. In this case, Tiz2 changes as heat accumulates in the thermal sponge of tank 8.2. As pr increases, both the compression ratio of the C2
2740 densifier and the power consumed by C3 must increase to achieve equalization of both gas temperatures and pressures in the two stages. Due to the thermal sponge in the tank 8.2, and the variable flow of heat transfer liquid sucked from the RL tank (due to the commands given by the controller to the execution bodies of the flow flaps) and distributed to this thermal sponge, the gas temperature in these tanks and the liquid temperature in the RL tank are maintained at close values. In the next cycle of compressor C3,
2745 its discharge valve opens at the new pressure in tank 8.2.
When the gas pressure in the tank reaches the desired value pf, an additional amount of liquid from the tank RL is introduced to push the compressed gas into the expander E4, which ensures the storage of the gas in the tank RG under constant pressure at the temperature Tamb.
When the temperature equalization system is common, it is calibrated to ensure equal
2750 temperatures for the two thermal sponges. In this case, if the densifier 02 is of the rotary type, with adjustable speed, commanded by the controller, the tank R and the compressor C3 are no longer needed. Due to the heat build-up, the temperature of the 02 densifier thermal sponge increases simultaneously with that of the second stage thermal sponge, and its compression ratio must be changed to supply the second stage with gas at the pressure and temperature existing in the reservoir at that time.
2755 Also, the compressor 01 must supply the densifier 02 with gas at the temperature Tiz, which is why its compression ratio must be changed accordingly.
In some applications, the temperature differences between the two cooling systems can reach large values. For this reason, it is recommended that the 03 adiabatic compressor be a rotary one, or even a dynamic one (for high compression ratios, several pressure steps are introduced).
2760 To ensure maximum efficiency, in all these configurations, all changes in speeds and revolutions, as well as the opening of valves and regulation of cooling flows, are made by an automatic control system, which ensures the isothermal nature of the processes in C2 and in tank 8.2. For this, a thermometer mounted in the liquid layer 8.2I of the second stage 8.2 will indicate, in real time, its temperature Tr, approximately the same as the temperature of the thermal sponge. The operating system
2765 of the regulator immediately determines the average temperature Tiz= Tr+A T of the gas in the tank, and based on it, the optimal compression ratio for the compressors C1 and C3, to ensure the gas at its exit, this temperature. On this basis, the compression ratio of the densifier C2 is also calculated, which ensures the equality between the pressure pr in enclosure 8.2 and the discharge pressure of the adiabatic compressor C3.
2770 At the beginning of the compression operation, the compression ratio of the compressor C3 will be equal to 1 , and the gas pumped into the tank 8.2 will have the temperature Tiz1 and the constant pressure p2 of the gas in the densifier, and will gradually replace the hydraulic liquid in the tank, until the liquid level of here it reaches an optimal value, then the liquid discharge valve from this tank is closed. The compression will be continued in the tank 8.2i, by the gas piston supplied by the compressor C3, with
2775 the gradual and slow increase of the temperature, as well as the slow increase of the compression ratio of the compressor C3, its discharge pressure having the value pr , intermediate between p2 and pf .
If the isothermalizer is used in applications with heat accumulation, the thermal sponge of the second stage of the isothermalizer must be sized for these special conditions. The second stage of the isothermalizer in Fig.43 is a closed and thermally insulated tank 8.2, cylindrical, spherical,
2780 parallelepipedal, or any other convenient shape. Its main component is the 8mf thermal sponge. This sponge must have a very high capacity to accumulate heat, contain gas cells in connection with each other and allow a loose circulation of gas and coolant (which must be chosen according to the coefficient of surface tension) . Therefore, the sponge can be made from metal inserts and fabrics, from metal foam, metal balls, from containers with perforated walls filled with metal fragments of any form (including filings,
2785 or other metal waste), with fragments of refractory bricks, refractory sand , stones and boulders, bricks and ceramic pipes, etc. The thermal sponge in Fig.43 is made of 8mf metal foam, with another thermal sponge on top, made of metal fabrics, to allow an easy distribution of the coolant. In the configuration in the figure, both thermal sponges completely occupy all the respective plane sections. The lower part of the tank, separated from the solid sponge by a filter that stops the movement of small particles dislodged
2790 from the sponge, is reserved for temporary storage, as well as for the circulation of the cooling liquid, containing an 8.6m pump that sucks liquid from here, to send it through a series of pipes to the sprinklers 8.8a. When the pressure in the second stage of the compressor reaches the desired value, the compressed gas is pumped (by introducing additional liquid sucked from the RL tank), through the valve 8.2s, into an adiabatic expander E4 (to bring it to the ambient temperature), then in the tank under
2795 constant pressure 8.7. The replaced liquid from the tank 8.7 returns to the RL tank, or is taken up by a hydraulic motor, for the recovery of the mechanical energy of movement.
In the supply mode, the isothermalizers work in the rarefier mode, gradually transforming part of the accumulated thermal energy (part of it can be transferred to thermal energy consumers, in the cogeneration mode) into mechanical energy that they supply to such consumers. 2800 Another configuration, simpler to execute, of the thermal sponge can be achieved by mounting on the upper cover of the tank a large number of vertical plates, bars, or vertical wires, single or in bundles, deformable or not, metallic, or from other materials, and through the distribution system to introduce, with a minimum pressure difference, the heat transfer liquid, which will slowly drain on the vertical elements of the thermal sponge, towards the lower cover, from where it is taken over by the recirculation pump.
2805 Fig.44 shows another possible configuration for this class of isothermalizers. In its first stage 8.1 , the system consists of the adiabatic compressor C1 that sucks in gas (for example, atmospheric air) at the ambient temperature and raises it to the temperature Tiz1, equal to the temperature of the gas in the second stage, and introduces it into the reservoir R1 , from where it is extracted by a group of low- pressure isothermalizers C2 which, after an isothermal compression, stores it in the tank R2. From here,
2810 the gas is taken by the high-pressure isothermalizer C3, compressed to the pressure p1 and introduced into the second stage 8.2. The discharge phase of this compressor is not done at a rigorously constant pressure, the discharge being, in fact, an adiabatic compression with a very low compression ratio, which also leads to a slight increase in the temperature of the gas in this stage. The heat transfer fluid that circulates through both stages equalizes the temperatures of the thermal sponges of all the
2815 isothermalizers in the system and of the walls of the two tanks. In this way, a Tiz1 temperature is achieved throughout the system, a temperature that gradually increases until the gas pressure in the second stage reaches the preset value pr.
The sizes of compressors C2 and C3, their compression ratios and their energy efficiency (dictated by the temperature difference 17) are chosen in such a way (and changed during operation)
2820 that their cycles have approximately the same duration. Also, the compression ratio of the compressor C1 changes during the compression process, so that the temperature of the gas at its exit is always equal to the temperature Tiz1 of the second-stage gas. To ensure maximum efficiency, in all these configurations, all changes in speeds and revolutions, as well as the opening of valves and regulation of cooling flows, are made by an automatic control system, which ensures the isothermal nature of the processes in C2
2825 and in tank 8.2 .
The second stage of the isothermalizer in Fig.44 is a closed and thermally insulated tank 8.2, cylindrical, spherical, parallelepipedal, or any other convenient shape. Its main component is the thermal sponge 8.2b, composed of bundles of metallic wires, very thin and very dense, with the smallest possible distances between them. The wire bundles are fixed to a horizontal 8.2g metal plate with numerous
2830 additional 8.2o perforations. This sponge has a very high capacity to accumulate heat and allows for easy circulation of gas and coolant.
The lower part of the tank is reserved for temporary storage as well as coolant circulation, containing a 8.6m pump that sucks liquid from here, to send it through a series of pipes to the upper part of the tank, located between the tank cover and plate 8.2 g, from where it drains gravitationally and
2835 washes all the thermal sponge. When the pressure in the second stage of the compressor reaches the desired value, the compressed gas is pumped (by introducing additional liquid drawn from the RL reservoir), into an adiabatic expanlder E4 (to bring it to ambient temperature), then into the constant pressure reservoir 8.7. The replaced liquid from the tank 8.7 returns to the RL tank, or is taken up by a hydraulic motor, for the recovery of the mechanical energy of movement. Another solution for effective
2840 cooling is the use of aqueous foams, surfactant liquids and foam regeneration gas.
The great advantage of gas piston isothermalizers is their ability to store in their thermal sponge, energy in the form of heat, which greatly expands the range of applications in which they can be used successfully. In storage mode, the method that allows the accumulation of all the energy brought as an input by the C2 compressor piston and the energy from other sources is to abandon the HE heat
2845 exchanger (which had the role of giving up the excess heat to the environment) and replace it with a thermal energy storage tank. The work done by the pistons in the first stage is converted into thermal energy (including the energy consumed in gas-gas, gas-liquid, gas-solid, liquid-solid and solid-solid friction processes in compressors and connecting pipes), from which, a part remains in the thermal sponges, in the walls of the compressors and in the liquid in the system tanks, and another important part
2850 is transported by the heat transfer agent to the external storage tanks. The mechanical work done by the pistons in the first stage is converted into thermal energy (including the energy consumed in gas-gas, gas-liquid, gas-solid, liquid-solid and solid-solid friction processes in compressors and connecting pipes), from which, a part remains in the thermal sponges, in the walls of the compressors and in the liquid in the system tanks, and another important part is transported by the heat transfer agent to the external storage
2855 tanks.
In all applications where it is desired to increase the power density of the isothermalizer, especially in applications with thermal energy accumulation, it is possible to give up keeping a constant value of the temperature difference AT (at the cost of a small exergy consumption), and the first stage of compression will consist of one or more series rotary compressors or, in the case of large tanks, of one or
2860 more C1 compressors, adiabatic, of dynamic, axial or radial type, discharging into a closed system of tanks 8.2. as represented in Fig.44. In the simplest configuration, this compressor(s) is mounted, together with its drive system, inside the main tank, fixed to one of the walls. At the beginning of compression, the pressure in tank 8.2 being reduced (most often atmospheric pressure), the compressor works as a gas piston pump with a compression ratio close to 1 , pumping gas with a temperature close to that of the
2865 suction gas, in closed tank 8.2, without significantly changing the temperature of the gas here. In these conditions, the consumed power being very small, if we want to keep the power absorbed at the compressor shaft constant, its speed must be increased as much as possible (default gas flow rate), thanks to the corresponding change in the supply frequency. When the pressure in the reservoir begins to rise, the mechanical work consumed for compression is converted into heat, which is absorbed by the
2870 thermal sponge and the heat transfer agent in the reservoir 8.2r. The heat transfer agent is found in one or more full RL tanks and is introduced through the sprinkler system 8.8a and recirculated using the hydraulic device 8.6, which here plays the role of a liquid recirculation pump. As the pressure in the second stage increases, the power required to introduce the gas in the first stage increases, and the rotor of the compressor C1 rotates at a lower speed. As long as the reservoir pressure remains low, the
2875 compression ratio remains low and so does the amount of heat absorbed by the sponge. By controlled variation of the coolant flow rate, second stage gas compression can be made isothermal. As the pressure in the tank 8.2 increases, the power consumed by the compressors in the first stage increases, so does the temperature of the gas at the entrance to this tank (equal to the temperature corresponding, in an adiabatic transformation, to the pressure in the tank), which leads and to the gradual increase, in the
2880 tank, of the temperature difference between the gas and the thermal sponge. Also, due to the thermal sponge, the temperature difference between the gas in the tank and that of the gas discharged by the C1 compressor gradually increases, which also leads to a heat exchange between the existing gas and the newly introduced gas and to an increase in the temperature difference AT2 from the reservoir. The amount of heat absorbed by the sponge increases more and more, due to the additional thermal energy
2885 brought by the hot gas. Therefore, the pressure of the gas in the tank increases simultaneously with the pressure of the gas at the outlet of the adiabatic compressor, until the maximum pressure is reached, but the temperature T1 of the gas stored in the tank will be lower than that at the inlet of the tank, and the temperature Tfcf of the sponge thermal will be inferior to it.
When this pressure is reached, the valve between the two stages closes (the first stage can
2890 switch to another similar tank 8.2), and the mechanical energy stored in the gas by changing its pressure is recovered, with the help of the expander E, by gradually replacing the gas in the tank 8.2 with heat transfer fluid from a RL tank. After the full amount of gas has been discharged, the first stage is restarted. This will supply gas with temperature Tbt, at the pressure corresponding to the polytropic coefficient of the first stage, also taking into account the temperature of the thermal sponge in this stage. The first stage
2895 becomes a gas compression and storage facility in the tank 8.2, until the complete evacuation, in the RL tank, of the liquid from this stage, after which, since the introduction of the gas is not accompanied by the evacuation of an identical volume of liquid, the gas pressure in the tank 8.2 increases until the maximum value is reached again (accompanied by a further increase in the temperature of the thermal sponge). The previously described operations are resumed and repeated until the thermal sponge reaches its
2900 maximum temperature. At this moment, only a small amount of energy is stored in increasing the pressure of the sucked gas (energy stored in a small volume), the rest being stored in increasing the internal energy of solid and liquid thermal sponges (especially in RL tanks).
As we have shown in the case of isothermalizers with heat accumulation, a condition for the good operation of these types of devices and the systems in which they are used is the realization of active
2905 thermal insulation for the main elements of the system. The extension of this type of insulation to all devices in which heat is circulated, contributes not only to increasing the energy efficiency of the installations in which they are used, but also to an important reduction of global thermal pollution. In addition, through the active insulation system, elements of other systems, of any type, in which there are temperature differences between their interior and their ambient environment, such as various types of
2910 tanks, transport pipes, buildings, can be thermally insulated civil, residential and industrial, etc. The method proposed in Fig. 45A for the insulation of an isothermalizer, involves placing the thermal insulation material, intended to limit heat losses from the system components, on a series of structures (frames, nets, plates, etc.) mounted around the object to be insulated, arranged in this way so as to form successive layers of insulating plates, parallel to one of the surfaces of the insulated object,
2915 layers between which flows of heat-insulating fluids (preferably liquids) can circulate. The support structures of the plates, together with the plates and the spacers between the plates (single buffers, or rulers parallel to the direction of fluid flow) will form a compact insulating block around the object to be insulated. In Fig. 45A, the solid plates 15.2 are arranged around the object to be insulated 15.1 in such a way that between the plates are generated mains for moving the fluid on a snail-shaped route, having the
2920 object 15.1 in the center. For the isolation of large areas, a larger number of adjacent fluid mains is made. In addition to this structure (or for the isolation of smaller devices), or for the isolation of flat surfaces, the route can be composed of registers of tubular pipes with a rectangular section 15.5, also made of insulating materials, as in Fig. 45C, placed in parallel planes, the outlet of the last register of each plane being connected to the inlet 15.6 of the first register of the next plane. The space not occupied by the
2925 registers is occupied by heat-insulating materials (preferably in the form of foam, cotton wool, or granules). The end 15.4 of each route continues with a branch which connects to a common collection pipe.
When the liquid is stationary, the insulation is passive (like prior art insulations), consisting of a succession of solid and liquid insulating layers (in some areas, heat flow can only pass through solid
2930 layers), resulting in each tube of flux a total thermal resistance. The temperature differences between the layers being small, we can neglect the convective currents and consider that through each liquid microregion there is a heat flow that generates a temperature difference between the neighboring solid walls. When the liquid circulates along the winding paths imposed by the insulation structure, from the outside to the inside, starting with a temperature equal to that of the external environment, it passes
2935 through solid walls whose temperature increases progressively, which makes the temperature of the liquid, with an average value between the neighboring wall temperature values to also increase as the fluid approaches the end of the path. In this way, part of the heat flow that was going to the outside is retained by the heat transfer liquid and transported to the collection pipe. If the thickness and arrangement of the solid and liquid layers is carefully calculated (and verified experimentally), a velocity
2940 can be found at which, on each of the flow tubes, the temperatures on the two faces of the solid layers are nearly equal , and the outward heat flux to be almost zero on as many flux tubes as possible. On the whole assembly, one can reach close to the ideal situation in which, of the total thermal energy that would be transferred to the environment when the liquid is stationary, only the thermal energy from the outer solid layer, whose average temperature is the smaller the number of insulation layers is higher. The rest
2945 of the heat is taken up by the heat transfer fluid, whose temperature increases from the value of the ambient temperature, until close to the temperature of the isolated object (the difference being the greater the faster the liquid's movement speed than its ideal speed). After extracting it from the system, this liquid can be stored in a tank and, after extracting the thermal energy, reintroduced into the circuit. Extraction of thermal energy from the recuperative liquid can be done immediately after its exit from the system, in a
2950 heat transfer fluid-gas heat exchanger (Fig. 45B). The thermodynamic transformation of the gas in the exchanger can be isobaric, at a PM pressure as high as possible, according to Fig. 15B, being part, in the case of a warm system 15.1 , of an isobaric-adiabatic-isothermal engine cycle, in which the adiabatic expansion occurs between the temperature of the isolated system and that of the environment, that is, between the pressure PM and a corresponding pressure Pm, and the isothermal compression occurs at
2955 the temperature of the environment. The mechanical energy obtained can be supplied to an engine that sets in motion an organic Rankine cycle heat pump, having the condenser located inside the isolated system 15.1 , or a gas heat pump, having the densifier located inside the system 15.1 , keeping constant (after a small addition, of additional thermal energy that covers the inherent losses) its temperature. The efficiency of the system increases if there is a source of waste heat, whose temperature Trez is too low to
2960 be efficiently exploited for other purposes. In this case, the fluid used in the thermal insulation system as a recuperator passes, before entering the system, through a heat exchanger 15.3, where it absorbs thermal energy from this source, then enters the insulation system after a more consistent insulating layer. Another option, especially for small temperature differences, is to store the fluid in a reservoir, at the temperature at which it leaves the system, or at a higher temperature after receiving additional
2965 thermal energy from a heat pump , or to store this energy, at a higher temperature (with a higher power density) in the tank of an isothermalizer with energy storage (renouncing to the first stage adiabatic compressor)
If the surfaces that require thermal insulation are large, they can be broken down into independent sectors, each with its own fluid path, but with a common introduction and collection pipe.
2970 Channels with a rectangular section, for the circulation of the fluid through the insulating material, can be executed at the installation site, through profiles dug into the insulating plates, or by fixing some spacers and directing the fluid, on the insulating plates 15.7, methods by which it is carried out several parallel flows (Fig.45D), or a single winding flow (Fig.45E). A cross-section (1 -1 ) through this subassembly is shown in Fig.45F.
2975 Another way to achieving passive insulation is the use of plates parallel to each other and parallel to the insulated surface 15.2, placed at a short distance from each other, on peripheral spacer elements (on all four sides), in the form of sticks 15.7, which to create a sealed volume from the outside, this interspace being occupied by the heat transfer fluid (Fig. 45G). Insulating plates are made of compact heat-insulating materials, with a large number of small diameter perforations (which create
2980 communication paths between successive layers of fluid, but also additional fluid-solid contact surfaces), or of spongy materials with open cells, or from very flat containers filled with mineral or basalt wool, wool, textile waste, granular materials, etc. Also, this type of alveolar thermal sponges can be inserted into the channels made for the circulation of the heat carrier fluid of the systems in figures 45C, D, E, F, as well as in the fluid layers between the insulating plates of the systems in figures 45F, G. The chosen materials for 2985 the realization of these thermal sponges, they are chosen according to the state of aggregation of the heat transfer agent. By mounting a pump or a compressor at one end of the path, the heat transfer fluid slowly travels a long path, between the suction device and the opposite end of the path, collecting on this path the thermal energy (positive or negative) that travels the same path in the opposite sense. The flow velocity is low, therefore energy losses due to friction are low, and the heat generated is also absorbed by
2990 the heat transfer agent. This type of insulation is also suitable when the heat transfer fluid is air. Fig. 45H shows a system similar to the one in Fig. 45G, designed for vertical insulating plates, in which the heat transfer agent is of the combined type. At the outer end of the path is mounted a compressor that introduces the gaseous heat transfer agent and causes it to travel the entire path. At the upper level of the intermediate gas layers and the alveolar insulating plates 15.2, a series of pipes 15.8 are mounted,
2995 through which the liquid heat transfer agent circulates, and at the lower level, a series of collecting troughs 15.9 are mounted. The upper pipes are provided in the lower part with holes through which the liquid flows freely 15.1 1 and moves to the collecting troughs, cooling the gas from the gas layers and the alveolar materials that make up the insulating layers. The liquid collected in the lower troughs, after absorbing part of the thermal energy of the solid and gaseous components, is sucked by a pump 15.12
3000 and sent through the pipes 15.10 to the pipe in the upper part of the next layer. If the flow rate of both agents is ideal, both gaseous and liquid heat transfer agents arrive at the end of the path with a temperature equal to that of the isolated object, and additional thermal energy can be extracted and stored.
3005 Description of the possibilities of industrial application of the invention
A third objective of the invention is to propose, for a number of applications, different plant systems, made by incorporating the types of densifiers and rarefiers described in the invention. This incorporation can be done without changing the sequence of operations in the state-of-the-art systems, replacing only some parts of the system with more efficient ones made according to the invention, either
3010 by changing the sequence of operations, or by introducing new principles of operation, to obtain for the respectively applications, superior performance. The use of these devices, by increasing the energy efficiency and power density for the gas compression and expansion processes and/or by efficiently accumulating the thermal energy produced, leads to an increase in the performance of all the applications in which they are used.
3015 All the proposed new installations aim to make the most of the exergy available in the available energy sources, as well as in those existing in the environment. In these installations, we will permanently avoid direct contact between gas masses at different pressures or temperatures, as well as contact between any other elements participating in the process, if there are large temperature differences between them. Higher efficiencies are achieved when all thermal energy exchanges are made at the
3020 smallest possible temperature differences. The inability of prior art devices to achieve such temperature differences at acceptable power densities has meant that some of these system configurations have not even been studied.
In compression installations for supplying pneumatic devices, in compression installations used for storing and , the replacement of state-of-the-art compressors with densifiers
3025 according to the invention brings important advantages, by reducing the energy consumed, by increasing the flow rates, by eliminating heat exchangers between intermediate pressure stages, by reducing the size of the tanks, by significantly reducing chemical and thermal pollution, etc.
Isothermalizers can be successfully used in the construction of gas/gas and gas/liquid heat exchangers. Since the heat exchange is more intense when the fluids in question have a higher density,
3030 before being introduced into the exchanger, the gases (respectively, the gas) are isothermally compressed in a densifier D, and at the exit they are expanded, also isothermally, in a rarefier R, up to the initial pressure. If these operations, at acceptable power densities, were performed with the prior art compressors and expanders, a large amount of exergy would be destroyed. In Fig. 46, the scheme of a gas-gas exchanger, without energy accumulation, made with isothermalizers according to the invention is
3035 represented. The low-pressure gas LPgasI, located at the temperature Tamb-AT, is isothermally compressed to an economic pressure PM1 (which also takes into account the thickness of the walls and the necessary safety measures at high pressures), and the other low-pressure gas LPgas2, located at temperature Tm is compressed isothermally to an economic pressure PM2. The two gases then exchange thermal energy in the HE heat exchanger at a temperature difference AT. The new exchanger will have a
3040 much smaller volume, and the lower speed of fluid circulation causes a reduction in exergy losses. At the exit of the exchanger, the two gases are expanded to their initial pressure. The thermal energy given off by the two densifiers is recovered almost entirely by the two rarefiers with which they each share the same tank and the same cooling/heating circuit of the thermal sponges.
In Fig.46A, a similar process is shown, when one of the agents is the waste gas of a chemical
3045 process (for example, the gases resulting from a combustion), and the other is a heat transfer gas (for example, atmospheric air), intended recovery and storage of residual thermal energy. The system is similar to the previous one, but the R2 tank is replaced by the second stage of a gas piston heat storage isothermalizer. The temperature of the waste gas LPgas2 is first brought to the variable temperature Tiz of the thermal sponge in the tank 8.2 by quasi-adiabatic compression in the compressor C1 , then the gas is
3050 isothermally compressed and introduced into the heat exchanger HE with a temperature Tiz+AT, close to of this, and exits with a temperature Tamb+AT, slightly higher than the inlet temperature of the heat transfer gas. The residual gas is removed to the atmosphere, after expansion in the rarefier R in the reservoir R1 , with the temperature Tamb .
The heat transfer gas LPgasI, enters the circuit with the temperature Tamb-AT and leaves the
3055 heat exchanger HE with the temperature Tiz and is introduced into the storage tank 8.2, whose thermal sponge and the heat transfer agent in the tank RL accumulate its thermal energy, and the mechanical energy is stored in the RG gas tank after expansion to atmospheric pressure. The system operation control system must correlate, after each compression cycle, the discharge pressure from the gas piston densifier with the expansion ratio of the expander E4 in such a way that the gas entering the RG tanks
3060 always has the same pressure and same temperature (Tamb).
In principle, for the full recovery of the energy (mechanical and thermal) stored, the same phases of the storage process must be executed, but in the opposite direction. The presented system is so versatile that it can fully recover the stored exergy, even if the characteristics of some of the process phases are changed in the recovery process. For example, if the mechanical energy to be stored (the one
3065 that drives compressor C and expander D in the waste gas circuit) has large fluctuations, the control system can increase their compression ratio and the temperature difference between the gas and the rarefier's thermal sponge and of the accumulator, the necessary excess energy being stored in its entirety, in the form of thermal energy. Under these conditions, the amount of gas stored in the RG gas tanks is smaller, therefore, in order to extract the stored thermal energy, it is necessary that the surplus
3070 thermal energy be extracted by the isothermal compression, at the ambient temperature Tamb, of a corresponding amount of heat transfer gas, or by putting into operation a thermal engine, made from the compressor E4, former expander, from the rarefier D2, former densifier, from the expander C, former compressor and from the densifier R1 , former rarefier, by the appropriate modification of the connecting pipes among them.
3075 In Fig. 46B, a principle scheme is represented for the recovery and storage of thermal energy from liquid agents (domestic or industrial waste, liquids discharged from cooling installations, geothermal waters, etc.). The scheme is similar to the previous one, but the HE is a gas-liquid heat exchanger, the liquid agent can also fulfill the role of a heat transfer agent, cooling the thermal sponge of the densifier, and the densifier is preceded by an adiabatic expander, for a more accurate correlation of temperatures
3080 from the entrance to the exchanger. Since the temperature of the liquid in the exchanger does not change in the first stage, the one that must be correlated with the temperature in the second stage is the gas temperature at the outlet of the exchanger.
Fig.47 shows a process for avoiding exergy losses in the case of interaction between two gases with different pressures. Concretely, the constructive and functional differences between two heating
3085 systems of an enclosure are represented. In system 47a, the ambient air, after being compressed in a densifier (curve 2-3 from Fig.47c), is introduced into a expander E, where it expands adiabatically (curve 3-6 from Fig.47c) until the pressure Po of the environment, the pressure at which the gas temperature reaches the Tm value. The gas discharged by the expander is mixed in the mixer 15.24.1 and after exchanging heat with the liquid evacuated from the enclosure, it is taken over by the densifier.
3090 In system 47b, the adiabatic expansion is done only up to the Text temperature of the external environment (curve 3-4 in Fig. 47c), then it is taken over by a rarefier, which expands it isothermally up to the pressure Po (curve 4-5 in Fig. 47c). Through this process, the mechanical energy consumption required by the heat pump is reduced (the area between curves 4-5-6 in Fig.47c). The diagram in Fig.47A describes the operation mode of the liquid mixer in Fig.47B. Such a
3095 system is extremely useful when there is a need to provide thermal energy, in the form of a certain amount of liquid with a certain temperature Tm, when there is a tank in which the respective liquid is stored at a temperature higher than TM. In the state of the art, the problem is solved by mixing a quantity of liquid at ambient temperature Tarrto (or other available temperature) with the corresponding quantity of liquid at temperature TM. An important amount of exergy is lost by this method, the greater the difference
3100 between Tm and TM. The invention proposes a method by which all heat exchanges are made at a minimum temperature difference, chosen by the operator, which difference is the most convenient compromise between the energy efficiency of this exchange and the duration of time allocated to this operation. For this, the installation uses a gaseous intermediate transfer agent (for example, air), distributed between a heat engine (whose components are denoted in Fig.47B with the index 1 ) and a
3105 heat pump (whose components are denoted with the index 2). The rarefier R1 is inserted into a reservoir of liquid at the temperature Tamb, and the rarefier R2 into the reservoir of liquid at the temperature TM, while both densifiers are inserted on the outlet line (or into the reservoir) of the liquid at the temperature Tm. Each of these circuits of the working gas also contains a compressor C, respectively a expander E, as well as the secondary of the heat exchangers HE1 , in the primary of which the liquid enters with the
3110 temperature Tamb respectively the secondary of the heat exchanger HE2, in the primary of which the liquid enters with the temperature TM. The primary of both heat exchangers supplies liquid at the temperature Tm. In the initial phase, the two circuits operate with the liquid circulation pumps stopped, until the temperature of the liquid in the primary of the two exchangers reaches the working temperature.
A configuration in which exergy losses can be reduced, especially in the phase preceding the
3115 stationary regime, is represented in Fig.48, in which heat exchangers are no longer needed. The system consists of a tank with liquid 15.18 in which horizontal perforated plates 15.29 are mounted from place to place, with the role of facilitating the thermal stratification of the liquid inside. The tank also contains a series of vertical pipes 15.28 on which, at different levels, a series of valves 15.28.1 are mounted, the closing and opening of which is done by the system's automatic regulation system. The main elements of
3120 the system are the smaller tanks 15.19, mounted inside the main tank, which are coupled, by means of pumps P, to the vertical recirculation pipes. The rarefiers and densifiers of the system are mounted in these tanks, and between these devices are mounted the functional elements that ensure the jumps between isothermal temperatures. In the configuration in the figure, these devices are compressors and adiabatic expanders, but, in the case of Stirling or Ericsson cycles, regenerators or heat exchangers can
3125 be used.
In the first phase of operation, the commands sent to these valves have the role of establishing the number and temperature of the liquid layers controlled by the system. In stationary mode, the tank is connected in its lower part with the tank with liquid at ambient temperature, and in its upper part with the tank with liquid at the storage temperature, ensuring at the exit liquid at the desired temperature. 3130 The systems described are energy storage systems. For their operation, they consume mechanical energy, when it is available, to deliver it later, when there is a demand, in the form of mechanical energy and/or heat. They can also work without receiving energy from the outside, if it is provided by a thermal engine, with equivalent Carnot cycle, which has the residual fluid as a hot source and the ambient environment as a cold source.
3135 In Fig.49, before being introduced into the exchanger, the gases (respectively, the gas) are compressed isothermally in the densifier D (the gas at temperature Tamb exchanges heat with a liquid at temperature 7), up to an economic pressure PM (which also takes into account the thickness of the walls and the necessary safety measures at high pressures), and at the exit they are expanded, also isothermally, in the rarefier R, up to the initial pressure Po, operations which, at acceptable power
3140 densities, if carried out with the compressors and prior art expanders, destroy a lot of exergy. The new exchanger will have a much smaller volume, and the lower speed of fluid circulation causes a reduction in exergy losses. When one of the fluids is liquid, it can also act as a heat transfer agent, cooling the densifier sponge and giving up heat to the rarefier sponge. Energy exchanges in isothermalizers take place in the presence of temperature differences (in Fig. 49, in both devices, this difference is 217), and
3145 the temperature jumps necessary to carry out this exchange can be executed polytropically inside the devices by increasing the piston speed, with the help of compressors, respectively quasi-adiabatic expanders, or with isothermalizers made according to the invention. For the energy efficiency of this type of heat exchanger to be acceptable, the temperature differences of the transformations in the two isothermalizers, as well as the one in the exchanger, must be as small as possible. Larger temperature
3150 differences (which lead to the reduction of the dimensions of the devices), can occur with the increase of the piston speeds, but they reduce the energy efficiency, even if (at a temperature of the liquid at the exit of the exchanger equal to T b) they extract from the liquid the same amount of thermal energy. For this reason, at low liquid temperatures, this type of heat exchanger is economical only with isothermalizers with thermal sponges with a large heat exchange surface, made according to the invention. In Fig. 49A
3155 are represented in T-s coordinates, the transformations of the gas during an isothermal-isobaric- isothermal cycle 1 -2-3-4-5.
These heat exchangers can also be used with very good results to create heat recovery units designed to extract residual heat from hot gases and liquids resulting from various domestic thermal processes (local heating and air conditioning installations, food preparation, etc.), or industrial, from the
3160 chemical industry, metallurgy, thermal power plants, cement factories, the transport industry (combustion gases, furnace gases, exhaust gases, cooling gases, etc.). Fig.50 shows the configuration of a recommended installation for the recovery of thermal energy contained in some hot liquids, existing in nature (geothermal waters), or resulting from various industrial processes. For example, the plant can be used for the intergal recovery (by bringing the liquid to ambient temperature) of the heat contained in the
3165 coolant, at the exit from the condenser of a plant that produces electricity through a Rankine cycle, or similar, eliminating cooling towers , or other types of large heat exchangers from the state of the art. The special performances of this system are generated by the fact that both isothermalizers and heat exchangers can be made with acceptable power densities, even when working with very small temperature differences between the elements participating in the heat transfer processes. The heat
3170 recovery is done in the liquid-atmospheric gas heat exchanger HE in which the liquid, circulated with the help of pump P, enters with the temperature Tm from the condenser outlet and leaves with the ambient temperature Tamb, and the gas, with the temperature Tamb, after its isothermal compression in the densifier D, followed by a polytropic expansion, enters with the temperature Tamb-AT (the temperature jumps at the entrance and exit of the device are performed polytropically, in the densifier, by changing the
3175 speed of the piston) and exits with the temperature Tamb-AT, going through circuit 1 -2-3-4-5 from Fig.50A. Ensuring the isothermal trajectory is done by changing the speed of the piston and by changing the controlled flow of the cooling liquid, extracted from the liquid pipe from the exchanger outlet. The heat extracted by the gas can be converted into mechanical energy right at the exit of the exchanger, by its adiabatic expansion in the expander E, or it can be stored in tanks intended for this purpose, after its
3180 compression in a gas piston densifier and energy storage , composed of the adiabatic compressor C, the gas piston densifier DAac, and the reservoirs RL and RG. The 1 -2-3-4-5-1 cycle performed by the working gas is composed of a 6-5-2 engine cycle and a 6-3-4 heat pump cycle. At low temperatures Tm and high piston speeds, the heat extracted from the liquid and given to the environment can be greater than the useful mechanical work developed by the adiabatic expander, so the energy balance is negative,
3185 but this can be avoided by using thermal sponge isothermalizers with large heat exchange surface.
Both in heat exchangers in which the gas pressure is raised isothermally, and in the case of heat recuperators, it is indicated that the working pressure is as high as possible, and when the temperature differences are small, to use rotary isothermalizers, and then when temperature differences are large, liquid piston isothermalizers should be used.
3190 When the heat we want to extract is found in waste gases, they first pass through a pre-filtertreatment plant (marked with F in Fig.51), for the coarse removal of particles whose deposition could disturb the proper operation of the installation and to counteract the possible corrosive properties of the gas. The gases are then isothermally compressed at filter exit temperature in a Diz densifier, then adiabatically in a C compressor to bring the gas to the next stage pressure, and then in a gas piston
3195 densifier and accumulation of energy, composed of gas piston densifier D, RL and RG tanks, E expander and hydraulic pumps/motors M/P (Fig.51 ). A mechanical separation of the solid particles can also take place in the densifier, if it has a liquid piston, or if the liquid recirculated for cooling the sponge is suitable for this task and if a suitable filter is installed in the liquid recirculation path. A fine separation of solid residues can also be done at the exit from the densifier, or at the entrance to the rarefier, the filters
3200 mounted at this point having dimensions much reduced compared to those of the inlet filter, and where the high gas pressure allows a crossing of the filter lighter, with reduced speed and with lower total exergy losses. The installation in Fig.51A is an example of the operation mode of this type of recuperator, without energy storage, mounted on a CT thermal plant intended for the production of the thermal agent
3205 for the heating installations of residential, commercial, public utility, industrial buildings, etc. . The installation is equipped with a densifier D with a solid or liquid piston, which also has the role of sucking in the gas required for combustion from the ambient environment. The burning of fuel (solid, liquid, or gas) is done in the firebox 15.14, the central being equipped with devices for regulating the flow of fuel and gas. After the heated gas, together with the combustion gases, give up most of their thermal energy to the
3210 heat transfer agent in the heat exchanger 15.15, to be transported to the component elements of the heating installation 15.17, it is taken with the temperature Tm from the outlet of exchanger, by the densifier D, which compresses it isothermally, to an economically chosen pressure pm, then it is taken over by a expander (gas turbine) which transforms this thermal energy into mechanical energy, to set a electric generator in motion, or is stored in thermally insulated tanks. The installation is also equipped with the
3215 hydraulic pump P which, if the piston of the densifier is liquid, transmits its mechanical and thermal energy to this piston and, thanks to an adjustment flap, ensures the achievement of the optimal trajectory, and if the piston of the densifier is solid, circulates the agent heat carrier for cooling the thermal sponge. In both cases, the recovery of the thermal energy taken by the liquid in the densifier is carried out in the heat exchanger 15.16.
3220 Because they can handle higher gas flows than prior art compressors and quasiisothermal expanders, and because of the higher rates at which heat is transferred, isothermalizers can be successfully adapted to improve the performance of state-of-the-art air conditioning and refrigeration installations. In this sense, the Rankine and inverted Rankine cycles used, most often in the state of the art, can be replaced by Carnot, Stirling, or Ericson cycles, and the isothermal processes of evaporation
3225 and condensation of refrigerants (expensive and polluting substances) from the state of the art are replaced by processes of isothermal compression and expansion of gases in isothermalizers. Moreover, new, simpler and more economical configurations can be made with these isothermalizers, which can successfully compete with these state-of-the-art refrigerating installations, in which phase changes take place.
3230 Fig.52 and Fig.53 show the schemes of heating and cooling installations of an ambient space, installations for which the source (hot, respectively cold) from/in which the required/excess thermal energy is absorbed/given up is chosen from a wide range of sources available, against which there must be a large enough temperature difference to ensure an acceptable power density. Fig.52 shows a useful heat pump for heating systems. It mainly consists of two containers, 15.18 and 15.19, filled with the heat
3235 transfer liquid (e.g. water, antifreeze), in which a pump for recirculating the liquid is mounted, a Riz rarefier in the first container, respectively a Diz densifier in the second, provide each with a thermal transfer system from the liquid to the thermal sponge, respectively from the thermal sponge to the liquid, for example, a sprinkler system. Through a piping system, each of the two pumps discharges and sucks the heat transfer liquid to/from a liquid-air heat exchanger HE1 and HE2, respectively. The heat 3240 exchanger HE1 is mounted in an environment that constitutes the cold source, and HE2 in the hot source. In the case of heating systems, HE1 is mounted outside the space to be heated, and HE2 inside it. In the case of heating larger spaces, the HE2 exchanger can be a series of convective radiators, radiators with infrared radiation, fan coils, etc., and the HE1 exchanger is sized accordingly. The heat transfer between the walls of the two heat exchangers (and any fins attached) is done by natural convection, or forced with
3245 the help of V fans, and in some cases, one or both exchangers are inserted in closed enclosures, through which circulates a heat transfer fluid. This fluid can be, for example, geothermal water, waste water, water resulting from an industrial cooling process, antifreeze fluid from a liquid-soil heat exchanger, liquidgroundwater, liquid-seawater, liquid-water from surface lakes, or from flowing waters, etc. Inside the two isothermalizers and in the connecting pipes between them is the working gas, which can be air, at a
3250 pressure high enough to ensure the required power. In the configuration represented in Fig.52, recommended for large temperature differences between the two sources, inside the heat exchanger, the compressor C and the expander E are mounted that ensure the adiabatic jumps of the Carnot cycle, the heat exchangers required for an Ericsson cycle, or the recuperators required in the case of a Stirling cycle. The installation of these devices inside the hot container ensures the recovery of the heat given by
3255 them to the walls of the device, which means that two lower heat pumps are added to the main cycle (in Fig. 52A, a heat pump with an inverted Carnot cycle, represented with the help of broken lines).
The T-s diagrams of the thermodynamic processes occurring in this system are represented in Fig.52A. The temperature of the heat transfer agent in container 15.18 is Tm, the same as that of the agent in the HE1 exchanger and the thermal sponge of the rarefier, lower by 21 , than the temperature of
3260 the external environment and higher by AT2 than that of the gas in the rarefier. The temperature of the heat transfer agent in container 15.19 is Tamb, the same as that of the agent in the HE2 exchanger and the thermal sponge of the densifier, higher by 21 3than the temperature of the external environment and lower by A T4 than that of the gas in the rarefier. Both the temperatures of the agent in the two containers and these temperature differences can be changed, depending on the heat requirement, by changing the
3265 average speed of the isothermalizers (and implicitly, the power of the device), by changing the liquid flows transferred through the isothermalizers and by external heat exchangers, as well as the temperatures and fluid flows with which the external walls of the exchangers exchange thermal energy. All these changes ensure that maximum energy efficiency is maintained when the isothermal trajectories of the gas in the isothermalizers are maintained.
3270 Fig.53 shows a heat pump usable in refrigeration and air conditioning installations, made with the help of two isothermalizers, which have the role of extracting heat from an overheated environment, or from an environment that we want to cool it and transfer it to another environment with a higher temperature. As in the case of the systems in the state of the art, based on the processes of condensation and isothermal evaporation of a refrigerant, the principle scheme of this pump is identical to
3275 that of the heat pumps used in the previously described case of heating installations, but the direction of gas circulation in the heat pump is reversed compared to that in the heating installations (from Fig.52). In the case of refrigerators, the heat is extracted from the enclosure of a refrigerator, a cold room, or any other type of closed enclosure (with temperature Tamb+AT) and is given to an environment outside this enclosure, and in the case of air conditioning installations, the heat is extracted from the air intended for
3280 breathing from the enclosure of a circulated space. In the configuration in Fig.53, since the temperature difference between the two sources is small, we have given up the compressor and the adiabatic expander, following that the respective temperature jumps are executed polytropically, in the isothermalizers, still controlling the speed of movement of the pistons of the isothermalizers, through the same regulator that ensures their isothermal trajectories. The T-s diagrams of the thermodynamic
3285 processes occurring in this system are represented in Fig.53A. As in the case of heating installations, the transfer of thermal energy between the two enclosures is made through the liquid-air heat exchangers HE1 and HE2, located in the containers with perforated walls 15.20 and 15.21 , the exchange of energy between the walls of the liquid and air pipe making -by natural or forced convection. In the case of monosplit installations, intended for the air conditioning of a single room, the external container 15.20 can
3290 be mounted between the external wall of the room and a thinner screen mounted at some distance, on the entire surface (without the glazed surfaces) corresponding to the wall, as in Fig. 54. By adjusting the flow rate of liquid passing through the heat exchanger and the flow rate of air entering the container, the temperature of the gas in the container can be maintained at or near the temperature of the outside environment. In another configuration, the screen can be mounted inside the room, at a small distance
3295 from the outer wall, on its entire surface, the temperature in the container being maintained at a temperature equal to, or almost equal to, that of the room. This configuration, associated with active insulation (with heat recovery crossing the outer wall), ensures low energy consumption, both in the case of heating and cooling installations. Another possible configuration, with high efficiency, both in the cold and in the warm season, is to make the container 15.20 between the two layers of similar thicknesses of
3300 the outer double wall, and through the container to circulate a gas flow (similar to some types of buildings with low energy consumption), which on its route exchange thermal energy directly, through intermediate layers, or through an air-soil heat exchanger, with layers of soil, with relatively constant temperatures, located at depths greater than 80 cm, or through an air-liquid heat exchanger with a ground water cloth.
Both of these air conditioning systems work on the same principle as the prior art air conditioning
3305 systems that use a refrigerant, the condensers and vaporizers being replaced by rarefiers, respectively densifiers, and the latent heat being replaced by thermal energy from the mechanical work produced /consumed by the piston. Moreover, both systems can also be made in the version with refrigerant and with phase change, if in Fig.52 and Fig.53 the isothermalizers are replaced with condensers and vaporizers with refrigerant, and instead of the expander E a heat exchanger and an isoenthalpic
3310 expander. All these devices are mounted in enclosures 15.18 and 15.19, filled with heat transfer liquid, the heat transfer being faster than in the case of a gaseous heat transfer agent, as are many of those in the state of the art. The heat exchanges with the cold and hot sources are done through the HE1 and HE2 heat exchangers, by recirculating the heat transfer liquid. This offers a great advantage in the case of large systems, by replacing bulky tubing that recirculates air with small pipes that circulate liquid agent.
3315 In the case of some isothermalizers equipped with thermal sponges with a large absorption surface and with efficient cooling systems, heat transfer coefficients comparable to those between the refrigerant in the evaporation/condensation phase and the walls of the device can be obtained.
The advantages of the new systems are:
- the ability to quickly change any of the temperature gaps of the system, which provides a high speed of
3320 reaction of the system to external disturbances and prompt fulfillment of received orders
- the lack of the isenthalpic expanderr, the lack of isobaric cooling between the isentropic compression and the condensation phase, the need to cool the condensed liquid before the isenthalpic expansion, the possibility to compress/expand the perfectly isothermal gas, the possibility to give up expensive or dangerous refrigerants, etc. , provides, from the point of view of efficiency, a net advantage of isothermal
3325 installations.
- greater freedom in choosing the configuration and dimensions of the heat exchangers. From this point of view, the realization of flat heat exchangers, which can be mounted behind a wall or a screen, brings advantages of comfort and more judicious use of space. It also allows the use of more efficient exchangers (eg plate ones).
3330 High efficiencies, at acceptable power densities and sizes, can be obtained with rotary isothermalizers, where working gas is introduced at high pressures, and the compression/expansion ratio chosen is in the 2-5 range.
The configurations in Fig.53B and Fig.53D are identical to the previous ones ( Fig .52 and Fig.53), but the energy source with which the ambient environment is correlated is, explicitly, the external
3335 atmosphere, and the working gas from the two isothermalizers is no longer just any gas, but the very air sucked from the atmosphere of the enclosures and discharged into the same atmosphere, after processing. This allows the incorporation in the system of a calcinator conforming to CBI R02022/000007, consisting of the compressor C and the expander E, intended for sterilizing the air in the enclosures. The adiabatic jumps between the outlet of the densifier and the inlet of the rarefier are made,
3340 in both cases (both in the heating and in the cooling installation) with the help of the expander E. The T-s diagrams of the two systems are represented in Fig.53C and Fig.53E. In both diagrams, it is noted the introduction of an isobaric heat transfer between the air in the enclosure and the air discharged by the Riz rarefier (curve 7-1 from Fig. 53C, respectively curve 6-1 from Fig. 53E), when the discharge was made at atmospheric pressure. When a higher efficiency is desired, the pressurization is done until the gas in the
3345 rarefier has the same entropy as that of the atmospheric gas (Fig.53C, curve 6-7), and the pressure recovery is done adiabatically, with the help of a compressor (adiabatic compressor C1 in Fig.53B). In the two diagrams, the amounts of heat given off, respectively received in one cycle, by the two gas enclosures, from the two isothermalizers, were represented by dashed bold lines. The quantitative ratio between the thermal energies transferred between the two enclosures can be modified together with the 3350 temperature differences between the liquid agent in the two enclosures and the sources with which it interacts.
In Fig.53F and Fig.53G are represented the T-s diagrams of these two systems, in the version where the working gas is the air extracted from the outside environment. Similarly, configurations can be made in which the air is extracted both from the enclosure and from its exterior, is compressed, and the
3355 two gas tranches exchange heat between them, after which they enter an equivalent Carnot cycle.
The configuration in Fig.54 contains a series of additional features: to increase the transfer speed of thermal energy and to reduce the dimensions of the heat exchangers (both the inside and the outside), in containers 15.20 and/or 15.21 there are no install the HE1 and/or HE2 heat exchangers, but one Riz1 , respectively Riz2 rarefier, each coupled with a Diz1 , respectively Diz2 densifier. The rarefier Riz1 is
3360 mounted in enclosure 15.18 (together with the densifier Diz, the rarefier Riz2 is mounted in enclosure 15.21 , the densifier Diz1 in enclosure 15.18.1 , and the densifier Diz2 in enclosure 15.19. The liquid agent in containers 15.18 and 15.19, with temperature Tl, is circulated through these containers, through the isothermals mounted inside them and through enclosure 15.18.2, which contains the rarefier Riz1. In order to have a high power density, the gas entering the Riz-Diz heat pump has a high pressure. It is noted that
3365 the complete system consists of a heat pump Riz2- Diz2 through which circulates air extracted from the enclosure 15.21 , pump that extracts heat from this air, to give it to the liquid with temperature T,, in the enclosure 15.19 (Fig.54A), from the heat pump Riz1 -Diz1 , through which circulates air extracted from the enclosure 15.20, pump that extracts heat from the liquid of the tank 15.18.2, with temperature T,, to give it to the environment in the enclosure 15.20 (Fig.54C) and from the heat pump Riz-Diz, through which
3370 circulates a working gas, which is a regulating pump, the mechanical work consumed by it being added in the form of thermal energy to the energy of the liquid in the tank 15.19.1 and given to the liquid in the tank 15.19.2 (Fig.54B). Is noted by T,, the temperature of the liquid, this having different values in each of the 4 enclosures with liquid, the temperature differences between the enclosures can also be changed by changing the recirculated flow rate.
3375 A portion of air from the enclosure 15.21 , with temperature Tamb, is isothermally compressed in the densifier Diz2 (curve 2-3 in Fig.54A), mounted in the closed container 15.19.2, filled with liquid at a temperature T, (this temperature is equal to that of the thermal sponge of the rarefier) and gives up heat to the liquid in the enclosure. To produce compression, after suction, the temperature of the gas increases polytropically, with a difference AT (curve 1 -2 in Fig.54A). Upon entering the Riz2 rarefier, the
3380 air is expanded polytropically to a temperature Tamb-AT2, then isothermally, to the pressure Po of the room (curve 3-4-5 in Fig.54A), cooling the sponge of the rarefier. Arriving in enclosure 15.21 , the cooler air is isobarically heated, cooling the air in the enclosure (curve 5-1 in Fig.54A). The sponge of this rarefier can reach low temperature values, due to the lack of coolant, so that, during stop intervals (when the system has a sequential operation), by opening the valve 15.22, air with the Tamb temperature is
3385 circulated, to bring it back to the work temperature. The gas in the container 15.20.1 , with the temperature Text„ lower than the temperature 7} of the liquid in the container 15.18.2, is absorbed by the densifier Diz1 and after a polytropic compression, reaches the temperature Text+AT1 (curve 1 -2 in Fig.54C ) and is isothermally compressed (curve 2-3 in Fig.54C), then it is expanded adiabatically with the help of expander E (curve 3-4 in Fig.54C), then it is
3390 expanded adiabatically with the help of expander E (curve 3-4 in Fig.54C). Since the Diz1 densifier is not equipped with a liquid cooling system, the temperature of its sponge increases gradually (under the control of the central regulator of the system) which also leads to the increase of the difference AT1, up to a value at which, the heat input due to the displacement of the piston in a cycle is given, through the walls, to the external environment in the same cycle. If the operation of the entire system is designed to
3395 work intermittently, with the Tamb temperature maintained between two limits set by the action of some thermostats, during periods of densifier shutdown, by simultaneously opening both of its valves (inlet and outlet), by starting of expander E and opening valve 15.22, a forced convection cooling system is created, which removes some of the heat from the sponge.
The Diz and Riz isothermalizers are only needed when the difference between the Text and Tamb
3400 temperatures is large. They form a heat pump (curve 1 -2-3-4 in Fig.54B), which transfers heat between the other two heat pumps. This system is advantageous through the multiple possibilities of changing the working temperatures, the power of the installation and the speed of response to the commands received. All three types of systems described can work both as heating installations and as cooling installations, by changing the direction of gas circulation and transforming densifiers into rarefiers and vice versa.
3405 Fig.55 shows a heat pump intended for heating an enclosure, in a simplified configuration, composed of a single Diz densifier, mounted in an enclosure with liquid 15.25, metallic, equipped with cooling fins, which sucks air from encloses and compresses it isothermally (curve 1 -2-3 in Fig.55A) and an adiabatic expander E, mounted in a gas mixer 15.24, which adiabatically expands this gas (curve 3-6 in Fig.19A). In the configuration in the figure, the expander is a rotating one, and the mixer consists of a
3410 metal tube in which the expander is mounted, provided at one end with a fan that introduces a flow of air from the ambient environment, air which, after washing the walls of the rotating expander, it mixes with the gas pushed by the expander (both gas flows having the same direction of movement), the resulting mixture having a temperature close to that of the environment (depending on the diameter and length of the tube 15.24). To achieve higher heat transfer speeds, the mixer can be made in stages, each stage
3415 being equipped with an additional fan and a tube 15.24.1 with a larger diameter than that of the previous stage. After leaving the mixer, this gas exchanges heat with the ambient gas (curve 6-5 in Fig.55A) and returns to a temperature close to its own. If the insulation of the wall that separates these two environments of active type, and consists of successive layers of air, separated by insulating plates, the air from the outside environment is introduced into the enclosure with the help of a fan, passing
3420 successively through these layers, it gradually heats up, taking over most of the heat loss through this wall and reaches the interior at the temperature Tamb-AT2 (curve 5-1 in Fig.55A). The described system is recommended for circulated enclosures, only if the compression ratio of the densifier is low (the pressure inside must be lower than that stipulated in the safety rules for pressure equipment). For protected environments, the compression ratio can be increased, but to ensure a satisfactory COP performance
3425 coefficient, it is recommended to introduce an additional rarefier into the circuit (curve 4-5 in Fig.55A). In both cases, part of the mechanical energy required for the operation of the heat pump is exerted by the atmospheric gas.
A calcinator can also be attached to this system to sterilize the air at the inlet or outlet of the densifier.
3430 The system in Fig.55C is similar to it, but the direction of movement of the working gas is reversed, therefore, this system extracts heat from the enclosure, with the help of the Riz rarefier, and transmits it to the outside environment, with the help of the Diz densifier (Fig.55D). The working pressure can reach higher values, therefore the adiabatic expander is followed by a Riz rarefier, which ensures a better COP. Another change compared to the previous system, is the increase of the thermal transfer
3435 surface between the heat transfer agent and the ambient environment, by introducing the cooling coil 15.26. This coil is a spiral tube around the enclosure in which the isothermalizer is mounted, through which part of the heat transfer agent circulates.
Isothermalizers can be used, in general, for heating fluids, or solid bodies immersed in a fluid, and in particular, for the production of domestic hot water. Any of the previously described heating
3440 systems can be adapted for this purpose. For example, such a system can be the one described in Fig. 52, where enclosure 15.20 can be the domestic cold water pipe, or any other infinite cold source in gaseous or liquid state, and enclosure 15.21 can be a hot water tank household, which will accumulate in the form of thermal energy, all the thermal energy extracted from the cold source and all the mechanical energy consumed to operate the heat pump. The isothermal speed control system will adapt the piston
3445 speed to the actual temperature of the water in the storage tank every time. Because the temperature difference between ambient Tamb and liquid TL is not very large, we can dispense with the adiabatic compressor and expander. The T-s diagram of such a system is represented in Fig.55B. The heat obtained in this way has, in most cases, a lower cost price than that obtained by burning a fuel, the difference being the greater, the higher the coefficient of performance (COP) of the heat pump. More than
3450 that, any facility for heating a room in a building can work, simultaneously or successively, with two densifiers: one in the room to be heated, the other in the hot water tank.
Also, if in refrigeration and air conditioning installations in buildings, the rarefier (in the current state of the art, the evaporator) is mounted in a liquid-ambient heat exchanger in the enclosure, and the densifier (in the current state of the art, the condenser) is mounted inside a tank with liquid, all the thermal
3455 energy which, in the state of the art, is absorbed from the enclosure and dissipated in the ambient environment, thermally polluting it, is accumulated in the tank and can be consumed when needed, including for the production of mechanical energy, if a favorable cold source appears (after the outside temperature drops). The operation of this type of heater is based on the thermal stratification of the liquid in the tank. 3460 Fig.56 shows a configuration of a hot water production facility with increased efficiency. The installation consists of three isothermalizers, each with its tank and recirculation pump, and a hot water tank R. The isothermalizer lz is mounted in a liquid-air heat exchanger, the liquid circulating, thanks to the pump P, through a series of metal tubes 15.27, and the warmer air, from the enclosure, thanks to the fan V, among these tubes, giving off heat. The regulation system is programmed to maintain a constant
3465 temperature of the liquid in this isothermalizer and, consequently, also the temperature of the air in the enclosure. In the three isothermalizers, the working gas is found at an acceptable pressure from the point of view of safety and the power density of the installation. All the isothermalizers are connected, by means of valves controlled by the regulation system, to two connecting pipes, so that each of them can be coupled to either of the other two, and each of them can work both as a rarefier and as a densifier.
3470 The system works sequentially, at the time of start-up, the liquid in the isothermalizer lz1 having the temperature of the network, and that in the isothermalizers lz and lz2, having the temperature of the ambient environment.
- in the first stage, the lz1 isothermalizer and the lz isothermalizer form a heat engine, which takes thermal energy from the ambient environment, part of which is transformed into mechanical energy, and the other
3475 is transferred to the liquid in the lz1 isothermalizer, raising its temperature
- when the temperature difference between the liquids in the two isothermalizers is too small, through commands sent to the valves, the rarefier lz disconnects from the densifier lz1 and couples with the densifier lz2, forming a heat pump that takes the excess heat from medium and transfers it to the liquid in the isothermalizer lz2, raising its temperature
3480 - in the next phase, a thermal engine consisting of the rarefier lz2 and the densifier lz1 becomes active, which operates until the temperature of the liquid in the isothermalizer lz1 reaches the ambient temperature
- there follows, again, a phase in which the isothermalizers lz and lz2 form a heat pump, which operates until the temperature of the liquid in the isothermalizer lz2 reaches the storage temperature.
3485 - at this moment, the valves on the liquid circuit are opened, a tranche of cold water replaces the one from the isothermalizer lz1 , which it transfers to the isothermalizer lz2, and the hot water from this tank is transferred to the storage tank.
The described system can work continuously, if the capacity of the storage tank is large enough to take the heat produced in the most severe external climatic conditions, or if there are additional
3490 capacities to take the surplus hot water. In this sense, the invention proposes the establishment in urban agglomerations of intelligent hot water supply networks, consisting of a central station in which to produce hot water for the surrounding urban area, with the help of heat pumps, operated by engines that they work with renewable energy and extract primary heat from the sun, from residual sources, geothermal water, ground water, etc. These stations deliver hot water through a network of pipes to all consumers in
3495 that area, but these consumers become prosumers, who in turn produce, in some conditions, surplus hot water. In these situations, through an appropriate metering system, the prosumers connected to the network through which they are fed, deliver to this network, each one's surplus. The central station must be provided with sufficient storage capacities and with the possibility to transform the surplus water delivered into electricity, when there are consumption peaks in the electricity network, or when an
3500 advantageous cold source appears.
This idea can also be applied to refrigeration plants. For example, a household refrigerator that has a heat exchanger fitted with a rarefier inside, and a sealed tank with a densifier on the outside, transfers all the absorbed thermal energy to a hot water boiler. For this, the attached tank is filled, in the first phase, with cold water from the network. Each time the heat pump is switched on, the temperature of
3505 the water in the tank rises, and when it reaches the set temperature, the cold water supply valve and the valve connecting to the domestic hot water tank are opened. The hot water in the attached tank is replaced with cold water and the cooling process coupled with the production of hot water is resumed.
In residential buildings, domestic water drainage can be divided, with the help of a thermostat, into two branches, the hot branch passing through a hot waste water/cold domestic water heat
3510 exchanger, then returning to the waste water system, and the domestic water, after which takes its thermal energy, enters a storage system.
The ambient space cooling installation in Fig.57 extracts heat from the environment by means of the Riz rarefier mounted in a tank 15.19 with heat transfer agent, agent that circulates through one or more fan coils 15.21 , or similar devices. The air circulating through the installation is extracted from the
3515 ambient environment and after an isothermal compression in the densifier D (curve 1 -1 ' in Fig.57 A), it is sterilized with the help of the calcinator C-E mounted in the high-pressure tank 15.20.3 and cooled to the temperature Tamb (curve 3-4 in Fig.57A) in the secondary of the HE heat exchanger. Cold domestic water circulates through the primary of this exchanger, which after taking the heat generated by the sponge of the densifier D, takes the heat of the gas from the secondary of the exchanger, reaches the hot water
3520 usage temperature and is stored in the boiler B. If the hot water production exceeds the required consumption, the surplus hot water is supplied directly to the regional domestic hot water network, to be distributed to other potential consumers, or to feed a regional thermal energy storage facility, made according to the invention.
The domestic hot water administration system involves the creation of an organizational system
3525 in which individual consumers can become domestic hot water prosumers, which implies the creation of a double metering system that calculates the difference (positive or negative) between the hot water consumed in interest own and the one provided to the network. Such a system can solve the problem of domestic hot water consumption for extended areas around large consumers of electricity for the production of cold (refrigeration spaces, storage spaces, networks of outlets for frozen products, public
3530 spaces with air conditioning (buildings offices, gyms, performance halls, etc.) In this way, the electricity consumption for air conditioning is redirected to the production of domestic hot water, reducing the consumption of fossil fuels, or to storage and transformation into electricity when it occurs a load peak. The ambient space heating installation in Fig.58 is composed of a working gas circuit under pressure, which operates according to a Brayton cycle. The installation extracts heat from the outside
3535 environment by means of the heat transfer agent from the air-liquid heat exchanger HE1 mounted in a tank 15.20, agent entering the primary of the liquid-gas heat exchanger HE in the tank 15.18, transferring this heat to the gas from the secondary (curve 1 -2 from Fig.58A). The gas in the secondary is the working gas of the heat pump (Brayton circuit 1 -2-3-4 in Fig.58A) formed by the two exchangers and the compressor C and the expander E. The excess heat is taken over by the heat transfer agent in the heat
3540 exchanger liquid-gas heat HE mounted in the tank 15.18 and distributed to the air-liquid heat exchangers HE2 mounted in the tanks 15.20, or to the ambient environment.
The heating installation in Fig.59, whose operating cycle is represented in Fig.59A, is similar, but the liquid heat transfer agent is replaced by a working gas under pressure.
Fig.59B shows a configuration of an installation for the combined production of cold and domestic
3545 hot water, obtained by modifying a heat pump from the state of the art, made with phase change refrigerants (circuit 1 -2-3-4-5 from Fig.59C). The modification consists of the introduction of the evaporator, respectively the condenser in the tanks where a densifier is mounted, whose thermal sponge is maintained at the isothermal temperature by circulating the heat transfer liquid in the tank. The evaporator gas circuit (circuit e-f-g-h in Fig.59C) is closed by a rarefier mounted in the secondary of an
3550 ambient gas/heat transfer liquid heat exchanger mounted in the ambient environment, and the condenser gas circuit (circuit a-b-c-d in Fig.59C) , through a rarefier mounted in the secondary of a working fluid/heat transfer fluid heat exchanger mounted in a tank (in the present case, a boiler with domestic hot water). In the case of small temperature differences, the jumps between the isothermal temperatures of these two circuits do not require the use of adiabatic compressors and expanders, the control exercised by the
3555 piston speed regulation system being sufficient.
The sterilization system proposed in this invention can be implemented in any state of the art air conditioning system, local or centralized. It is a thermodynamic system for decontamination, by heat treatment, of the air intended for respiration, which we will call calcinator. Thermodynamic sterilization destroys pathogens by incineration. These systems are easy to implement and can be extremely efficient
3560 (100% efficiency, for a sufficiently high incineration temperature, applied for a sufficiently long period of time), therefore, they should not be missing from any air conditioning system. By this process of thermodynamic sterilization, the temperature of the air intended for respiration (air in a certain location, at ambient temperature and atmospheric pressure) is raised to the calcination temperature (at which the target pathogens are rendered harmless). The increase in air temperature is achieved by compressing it
3565 quasi-adiabatically to the pressure corresponding to this temperature, followed by a period of maintaining the air at this temperature in a regenerator, and by a quasi-adiabatic expansion "in the mirror" until the ambient temperature. A cooling, or heating (electrically or thermodynamically) of the air in the regenerator, allows its delivery to a predetermined temperature (which may be different from that of the environment). In this invention, both the compressor and the expander are electrically driven positive 3570 displacement devices. The new process can be used in many areas, is flexible, allows a wide range of powers and dimensions, an easy adjustment of working flows, pressures and temperatures. These systems can also be made to a small size, so they can be applied to personal protective equipment as well as portable systems. Another great advantage of this system is that most of the mechanical energy consumed by the compressor is returned to the system by the expander, minimizing the energy
3575 consumption required for sterilization. This is a crucial advantage over any other sterilization system, and the flexibility of the system makes it easy to implement in any ventilation, heating/cooling and air conditioning system for breathing air. It is compatible and can be coupled with most other state-of-the-art sterilization systems.
The design of the calcinator is made taking into account its precise destination: the class (or
3580 classes) of microorganisms to be combated. This determines the minimum temperature Tm up to which the air must be heated to obtain the calcination effect and the minimum duration tm to maintain it, necessary for the complete destruction of this/these classes of pathogens. This minimum temperature Tm can be exceeded, and a permissible value Tadm can be reached. There is a time duration td <tm, in which the air temperature is higher than the minimum temperature Tm, sufficient for a complete calcination. This
3585 allows the adoption of a nonstop calcination strategy: the gradual increase of the air temperature up to the value of Tadm above the value of Tm, followed by an immediate decrease (without pause), below this value, so that the time when the temperature value exceeds the value of Tm is higher than td value. The configuration of the computer is chosen according to the chosen strategy.
Obtaining the temperature required for calcination is done by quasi-adiabatic compression of the
3590 sucked gas from the environment made with a reciprocating compressor, and the temperature necessary for the air to be breathed, through a process of its expansion (combined, if necessary, with other additional thermal processes). The calcinator can be realized with any type of compressor and expander that can meet the requirements of the chosen operating variant, therefore its choice is made according to the performance of volume, weight, cost, convenience, etc. of the whole ensemble. In air conditioning
3595 systems, the most suitable are positive displacement compressors and expanders.
In the system described in Fig.60A we chose a nonstop strategy. In order to reduce the losses generated by the valves with a reduced passage section, the cylinder of a compressor/expander can be equipped (Fig.60A) with a single inlet-outlet orifice (to achieve the largest possible diameter of the access path in the cylinder). In the chosen configuration, the cylinders of both devices are connected to the body
3600 of a 4-way valve 15.11 , which has three main positions:
- the open position, noted a in Fig.60A, characterized by the creation of wide paths, both for the admission of atmospheric air and for the evacuation of sterilized air, carried out by the proper movement of the pistons
- the compression position, denoted b in Fig.60A, in which the piston of the compressor 15.8 performs a
3605 compression movement and that of the expander 15.10 is stationary in the TDC, - the expansion position, denoted c in Fig.60A, in which the piston of the expander 15.10 performs the expansion operation, and that of the compressor is stationary in the TDC.
The 4-way valve (electrically or mechanically controlled) is a spherical valve 15.11 in which ball 15.9 the gas passageways are made, by creating cavities that also serve for the storage of compressed
3610 air 15.12. The valve ball is actuated by means of a shaft and a camshaft, for the correct synchronization of the operating stages.
This type of valve can also be used successfully in any application described in this invention, when it is desired to create wide paths for gas and liquids circulation, thus a reduction of exergy losses (consequently, an increase in energy consumption). A very useful application is to make a new type of
3615 isothermalizer. In Fig.60B, the valve used is a 3-way valve: one way for the isothermal compressor/expander, one for the inlet pipe and the other for the compressed gas discharge pipe. The discharge pipe is connected directly to the compressed gas storage tank (constant pressure tank) at its bottom and it is permanently filled with liquid. In the cavities created in the ball of the valve, non- deformable thermal sponges are mounted, for example from interwoven wire nets, which have a large
3620 heat absorption surface. Sprinklers can also be installed in the outer walls of the valve to inject coolant as these cavities pass in front of them. In turn, the isothermal cylinder is equipped with a thermally deformable sponge and cooling systems, whose inlet flow is always equal to the outlet flow, so that the amount of liquid in the cylinder is constant, equal to the amount needed to eliminate the dead volume, when the piston is in TDC.
3625 In its continuous or sequential motion, the valve goes through three main positions:
- the open position, denoted a in Fig.60B, position in which is opened, through one of the ball cavities, a path with a large passage section, for the admission of the gas in the cylinder. The other cavity of the ball is filled with liquid, in direct connection with the tank. In this phase, a heat transfer takes place between the liquid and the thermal sponge of this cavity.
3630 - the compression position, denoted b in Fig.60B, position in which, to the lower cavity of the valve, the communication with the suction pipe is suppressed, the cylinder piston moves towards TDC, the cooling installation is started and the compressed gas is directed, entirely , to this valve cavity. In the upper cavity, a quantity of liquid with a volume equal to that of the cavity (and of the compressed gas from the other cavity) is displaced to the inlet pipe. In this position, immediately after the connection between the
3635 lower cavity and the suction pipe has closed, the liquid sprinklers that introduce the heat transfer agent into the cavity come into operation.
- the exhaust position, denoted c in Fig.60B, position in which the cylinder piston is in the TDC, the cavity which in the previous positions was in the lower position, arrives in front of the exhaust pipe and the compressed air is replaced with liquid from the pipes and it moves towards the compressed air tank, and
3640 the cavity that in the previous positions was in the upper position, arrives in front of the suction pipe, the liquid from the cavity being evacuated through this pipe. After evacuation, the amount of liquid introduced through the sprinklers is separated and introduced into the cooling circuit, the rest of the liquid (in an amount equal to the amount of liquid that left the tank, is stored in another tank, to be used when the gas stored in the first tank it is directed, under the same pressure, to a user (which can be this densifier,
3645 transformed into a rarefier).
Unlike other installations that store compressed gas in constant pressure tanks, the isothermalazer piston described above no longer consumes mechanical energy to transfer the compressed gas into the tank, therefore a hydraulic motor is not required in the installation configuration to recover displacement energy. Another advantage of this system is the possibility of gas compression in
3650 stages: two isothermal stages, one in the cylinder, the other in the valve cavity and a polytropic stage, performed by the liquid piston in the exhaust pipe. For the final cooling of the gas, the inlet to the discharge pipe is made through a wire mesh 15.18, which reduces the diameter of the air bubbles formed by the penetration of the liquid, bubbles that are cooled more strongly in the exhaust pipe and in the tank.
Fig.61 shows another configuration that allows the heating or cooling of the air in an enclosure.
3655 Compared to the systems shown in Fig.53B and Fig.53D, the main loop works in a Stirling cycle, more advantageous, at least for the small installations, due to the abandonment of the adiabatic compressor and expander, more difficult to operate and adjust in case small temperature differences between hot and cold source, replacing them with a single recuperator (devices that have recently reached high performance). In addition, all valves are removed in this loop. In the secondary loop, through which the
3660 atmospheric air circulates, because the adiabatic compressor and the expander also have the role of sterilizing the air, the configuration described above is kept, (cold and warm source links, which can be chosen from a wide range of options, are not represented).
In Fig.61 , Riz2 and Diz2 are isothermalizers made according to the invention: they are equipped with thermal sponge 15.15 and sprinklers 15.16 to ensure optimal heat transfer and an actuation system,
3665 which in addition to imposing isothermal speed during compression and expansion, ensures a correct correlation between the movements of the two pistons. In a first phase, in the Diz2 densifier there is gas at temperature Ta, equal to that of the liquid in which the Diz2 densifier is immersed, the Diz2 piston is in BDC, and the Riz2 piston in TDC, blocking the gas entry in this cylinder. Displacement of the Diz2 piston with the isothermal velocity corresponding to the temperature Ta, leads to the isothermal compression of
3670 the gas in the densifier cylinder. At a certain moment of movement, when a predetermined volume is reached, the piston of the Riz2 rarefier starts at the same speed. In the next phase, the piston of the Diz2 densifier reaches the BDC, where it stops, closing the respective end of the regenerator 15.17. Through this operation, the gas in the densifier pass into the rarefier, keeping the same volume, after changing the thermal energy with the regenerator and reaching the temperature Tm. In the next phase, the piston of the
3675 Riz2 rarefier continues its motion to BDC, with the isothermal velocity corresponding to the temperature Tm. A new transfer phase follows, in which both pistons move from one end to the other end of the respective cylinders and in which, when passing through the recuperator, the gas returns to Ta temperature. For larger installations, the collecting of the gas that is introduced into the densifier, as well as the
3680 distribution of the final product are done through piping networks, similar to the networks in the state-of- the-art installations. In these installations it is possible to work, in both loops, with higher pressures, the result being the processing of higher flows.
Another area in which the use of isothermalizers can bring an increase in the performance of the installations used is that of gas liquefaction. The T-s diagram in Fig.62A explains the new principle of
3685 operation, applicable to most gases and gas mixtures (air, natural gas, etc.), regardless of the pressure Pa and the temperature Ta at start-up. The proposed process is similar to the Siemens process: after passing through a treatment unit 15.20, the gas is compressed isothermally (curve 1 -2 in Fig.62A), in a Diz1 densifier Fig.62, up to a pressure P2 (for high pressures several stages may be preferable, without the need for intermediate heat exchangers). The pressure P2 corresponds to an entropy s2, slightly higher
3690 than the entropy of the critical point. Then, the gas is released adiabatically (curve 2-3 in Fig.53), in a turbine T, to a pressure below the vapor saturation curve, close to Pa, (in this area, the pressure Pa is boiling pressure ) and a temperature below the critical point. For reasons of anti cavitation protection, it is recommended that, when leaving the turbine, the liquid concentration be only a few percent. The gas is exhausted in a condenser in which, by extracting the latent heat (curve 3-4 from Fig.53), the gas is
3695 completely liquefied.
A proposed configuration for such an installation is shown in Fig.62. The condenser of the installation is the secondary 15.21 of a plate heat exchanger, through whose primary 15.22 a heat transfer agent circulates, which at the pressure Pa, in the vicinity of the temperature T1 is in liquid state. The heat exchange between the two regions is all the more intense, the larger the surface of the partition
3700 walls and the smaller the distance between the plates. The primary liquid is conveyed by a 15.27 pump and introduced into another 15.25 tank in which the Diz2 rarefier of a heat pump operating in Carnot mode (curve 2'-5'-4'-3 'in Fig.62A) is mounted. In turn, this heat pump (wich also consists of an adiabatic C2 compressor and T2 turbine, as well as a Diz2 densifier) transfer the extracted heat, as well as the mechanical work consumed, to another heat sink, at temperature Ta, or at a different temperature. To
3705 control the liquefaction flow, the installation may be provided with an additional system for cooling the gas of condenser, consisting of the Diz3 densifier (which uses as coolant even the liquefied product, or the heat transfer agent from the tank 15.25) and the expansion valve 15.26. This system extracts the warmer gas, from the upper area of the condenser and after a slight isothermal compression expand it isentropically to the pressure Pa (curve 6-3 in Fig.62A).
3710 The great advantage of this liquefaction process is that, in the case of gas regeneration, it can be applied in the opposite direction, by going through the same steps, recovering, in case of small temperature differences between the two paths, most of the energy consumed during liquefaction.
Another application where isothermals can lead to a remarkable increase in energy performance is the temperature control system of concrete airstrips, car and motorcycle racing tracks and even
3715 highway running surfaces. Such a system is composed of several segments, one in continuation of the other. Each segment is composed of two networks of parallel pipes, arranged transversely, with a length equal to the width of the track, through which the heat transfer agent (most often a mixture of water and glycerol) circulates, the upper network is mounted buried in the surface layer of concrete , as close as possible to the surface, and the lower network is mounted buried, under the concrete foundation. The two
3720 networks are linked together, either pipe by pipe, creating a spiral route, or the pipes in each network are connected to two horizontal pipes, located on the edges of the track, forming two separate ladder-type networks, which are then connected by several pipes vertical. Through a system of pumps, the heat transfer fluid circulates successively through the pipes of the two networks, carrying out a heat transfer between the concrete layer on the surface of the track and the soil layer located at the depth of the
3725 network. At small temperature differences between the soil and atmospheric air, heat exchange occurs naturally, but at larger differences, one (between the two layers), or two (for each layer) heat pumps/heat engines must be turned on which to facilitate this heat transfer.
Regarding the transformation of thermal energy into useful work, with the help of heat engines without phase change (gas-only), the use of densifiers and rarefiers described in this invention brings
3730 great advantages over the systems used in the state of the art. When they are used, their large heat exchange surface offers the possibility of obtaining high rates of transfer of thermal energy between different components of the system, which allows obtaining high power densities in relation to the volume or weight of the installations and offers the possibility of operation between closer temperature limits between hot and cold sources, greatly expanding the range of usable energy sources. Satisfactory yields
3735 can be obtained using as a heat source solar energy of lower intensity, sources of residual industrial or household energy, with temperatures lower than those of the state of the art, geothermal sources (if the atmospheric temperature is negative, the heat source can be the soil at a depth of several meters, or a sheet of groundwater), etc., and as cold sources, the energy of ambient air, soil, flowing water, water of lakes and seas, groundwater, etc. They can even be made heat engines, or heat pumps, that work with
3740 ideal (Carnot) efficiencies, based on atmospheric temperature differences between day and night, between atmospheric air and water of lakes and seas, or even between two close locations , differently sunny. At a given power, the volume of these motors and pumps increases as the temperature difference between the hot and cold sources decreases.
Due to the large exchange surfaces of the thermal sponges, engines and refrigerating
3745 installations operating according to Carnot, Stirling and Ericson cycles, direct or inverted, made with these types of isothermalizers allow obtaining high rates of thermal energy transfer between different components of the system, such as and between them and the external environment, which allows high power densities to be obtained. A great advantage of engines and heat pumps made with these types of isothermalizers is the possibility to choose and modify, during operation, the two regime temperatures Tiz1
3750 and Tiz2 inside the densifier, respectively the rarefier, by changing the compression ratio of the isentropic devices and the simultaneous change of the speed of the drive motors of the isothermalizers, keeping the optimal temperature difference AT, or adapting it to the change in the temperature of the external environment, or other operating conditions. This facility offers, for example, the possibility to exhaust, almost entirely, the thermal energy from a finite thermal reservoir, as happens with the energy storage
3755 system exemplified in Fig.65.
Isothermalizers are also useful in other configurations of engines and refrigerating installations that have different types of gases as their working agent, especially atmospheric air, such as internal combustion engines, operating in an open circuit (with the removal of combustion gases together with appreciable quantities of thermal energy), or closed and other systems that operate according to an Otto,
3760 Diesel, Atkinson, dual, Brayton, Humphrey, Lenoir, etc. cycle. In Fig.63 we represented the T-s diagrams of some of these cycles: Brayton (Fig.63A), Otto (Fig.63B), Lenoir (Fig.63C). A common feature of these cycles (and the others, previously mentioned) is that the entropy of the working gas varies between sm and sa values. The introduction in these thermodynamic cycles of an isothermal compression at the minimum temperature Ta (usually equal to the atmospheric temperature), to replace those phases of
3765 these systems in which the entropy of the working gas varies between sm and sa values, without changing the others, offers the opportunity to increase both the power density (by moving all the gas evolutions to the area of high pressures) and the efficiency of the engines, reducing, at the same time, to almost zero, their thermal pollution (gases removed in open cycles have, at atmospheric pressure, atmospheric temperature), without making other changes in the construction of the device, with the exception of
3770 changing the expansion ratio in the turbine, or the adiabatic expander which gives value on the absorbed thermal energy. In Fig.63G, the T-s diagrams of a state-of-the-art Brayton cycle (curve 1 -2-3-4-1 ) and a modified system, by introducing, in the initial phase, an isothermal compression (curve 1 -2'-3'-4'-1 ).
In Fig.63D was represented how the T-s diagram of an Otto, Diesel, or Brayton engine is modified, but was introduced an additional modification and flattened the heat absorption curve 3-4
3775 (adding to the isochore, or isobaric, a component of relaxation). For this purpose, the thermodynamic evolution of the working gas no longer takes place in a set of cylinders (the same evolution in each of them, but with phase shifts to reduce vibrations), but in a sequence of devices (Fig. 63E):
- the Diz densifier, with variable compression ratio, which moves the entire thermodynamic evolution towards the area of high pressures and results in the considerable reduction of thermal pollution, with the
3780 resulting increase in efficiency
- the C1 adiabatic compressor, whose compression ratio can be modified, at the operator's command, to obtain power increases
- the CC combustion chamber, which requires the use of materials with high resistance to high temperatures, equipped with a piston, a fuel supply system 15.5 and an ignition system 15.6, with
3785 variable fuel flow depending on the required power
- the isentropic turbine T which, together with the piston of the combustion chamber, harnesses the thermal energy resulting from fuel combustion
The engine can work with a wide range of fuels, liquid, gaseous, or powdery materials, being able to utilize fuels with lower calorific values 3790 Fig.63F shows one of the possible engine configurations. In this configuration, the adiabatic compressor C1 and the combustion chamber are located in the same cylinder (not necessarily with a circular section), being separated from each other by a drawer 15.7. When the gas in compressor C1 reaches the prescribed pressure and temperature, drawer 15.7 opens. At this point, the CC combustion chamber piston at top dead center begins to move, with increasing speed, and the compressor piston
3795 slows down to stop at TDC at the end of the first portion of the cylinder. Due to these movements with variable speeds, the volume of gas between the two pistons has some fluctuations, but it reaches the predetermined value at the moment when the compressor piston reaches the end of its stroke and the drawer 15.7 closes. Fuel spraying can start as soon as some gas has entered the CC, but its ignition only occurs after the drawer is closed. At the moment of ignition, the CC piston already has a sufficiently high
3800 speed so that the combustion of the fuel does not take place at a constant volume. Fuel supply can take place, through nozzles located in different positions, until the piston reaches the end of the stroke. In the ideal case, the fuel dosing can be done in such a way that the expansion of the gas in this chamber is isothermal, at the highest Tjz temperature that CC allows, which determines the maximum (ideal) engine efficiency.
3805 Phase-change heat engines (operating on a Rankine cycle, or ORC) can also improve their performance with these devices. Was already mentioned how the Carnot efficiency of these installations can be increased by introducing a heat pump with isothermalizers, functioning between the temperature in the condenser and that of the ambient environment (curve 1 -6-5-7-1 in Fig.63J), or from another cold source, as well as the advantages of replacing, in certain installations, condensers and evaporators with
3810 isothermalizers.
In Fig.63H (vapor superheated system), Fig.631 (mixing gas behavior diagram) and Fig.63J (system without mixing gas, with isothermal expansion and heat pump to extract additional energy from heat engine condenser) the operation of Rankine engines where the working fluid is a gas near the saturation limit on either side of it (e.g. water vapour) and another gas far from this limit are illustrated by
3815 the corresponding T-s diagrams. The advantage of this combination is obvious in the case of vapors that condense at low or very low pressures (for example, water vapor) and which, for this reason, require the use of bulky condensers. Another advantage is gained by introducing in the boiler, in the liquid close to the boiling temperature, a rarefier in which the isothermal expansion temperature of the gas-vapor mixture is ensured by washing the thermal sponge with this liquid. This process allows the partial
3820 pressure of the vapors in the boiler to be raised to the pressure that ensures boiling at the maximum temperature Th2 in the superheater, increasing the Carnot efficiency of the engine (or of the heat pump, if the direction of flow of the working fluid is changed). Another important modification is the replacement of the condenser with a heat exchanger (high efficiency can be obtained with an exchanger with a large number of plates, with small separation distances between them and with many acicular protrusions on
3825 the inner surfaces that border the mixing fluid ( or with a fine network of wires placed in these regions). Softened water circulates in the opposite direction through the adjacent interstices. Therefore, the mixture leaving the expander (the highest power density is provided by a turbine) at a temperature 7}, corresponding to the partial pressure of the vapor from which condensation begins, will gradually cool and with it the pressure will also decrease partial water vapour, up to the pressure corresponding to the
3830 coolant inlet temperature. In Fig.631, the T-s diagram describes the transformations undergone by the mixture gas during one cycle.
In compression installations for supplying pneumatic devices, in compression installations used for storing and transporting gases, the replacement of state-of-the-art compressors with densifiers according to the invention brings important advantages, by reducing the energy consumed, by increasing
3835 the flow rates, by eliminating heat exchangers between the intermediate pressure stages, by reducing the size of the tanks, by significantly reducing thermal pollution, etc.
The field in which the use of isothermalizers can make significant progress compared to the prior art is the storage of mechanical and thermal energy from renewable energy sources and waste heat from many industrial processes that still dissipate this energy into the environment, thermally polluting it. In any
3840 of the storage systems used in the prior art (D-CAES, A-CAES, l-CAES systems), in the storage phase, a significant fraction of the thermal energy of the working gas, resulting from the transformation of the mechanical energy of the piston during the compression phase is discharged into the environment at its temperature. In the energy recovery phase, the temperature at which the expansion takes place in the expander is lower than the ambient temperature. Therefore, this storage-recovery cycle occurs with
3845 significant loss of exergy, or at very low speeds. In some configurations, these losses are recovered by a significant supply of energy from other sources, usually from fossil fuels. In all these systems, especially those based on quasi-isothermal compression to high gas pressures, the replacement of compressors and expanders, currently used, with isothermal densifiers and rarefiers, leads to significant increases in energy efficiency and to improving all the parameters of these installations. Moreover, the invention
3850 proposes a series of new system configurations, in which the fraction of thermal energy released to the environment, out of the total energy available for storage, can be significantly reduced.
Fig.64 shows an A-CAES type energy storage system (by adiabatic compression), suitable for the apparatus described in this invention. As in the prior art systems, the energy available for storage is used for adiabatic compression of the working gas by means of the isentropic compressor C1 (Fig.64B),
3855 from the pressure Pa and the temperature Ta, to the pressure Pm and the temperature Tm (curve 1 -2 on the T-s diagram, Fig.64A). Unlike prior art systems, the resulting gas is cooled (curve 2-3 on the T-s diagram, Fig.64A), with additional mechanical energy consumption, in the HE gas/liquid heat exchanger. If Tm is high, the gas can be cooled in stages, with the heat transfer fluid change in each stage. The resulting liquid, at a temperature close to Tm, is stored in the tank R2, and the gas at the temperature Ta is
3860 stored in the tank R1. When a load peak occurs on the consumption network, the gas in the tank R1 is expanded to the Riz1 rarefier (curve 3-1 on the T-s diagram in Fig.64A), the resulting mechanical energy being taken over by the useful task (usually a electricity generator). The thermal energy, stored in the R2 tank, can be extracted at any time and can receive various uses. Still this energy can be transformed into mechanical energy using the heat engine described in Fig.64B, through a process that can be used to
3865 extract heat from any finite thermal energy tank. The stored thermal energy is consumed, almost entirely, for the production of mechanical energy, with the help of a heat engine running in a Carnot cycle, consisting of the Riz2 rarefier, the T turbine (or an adiabatic piston expander), the Diz densifier and the C2 adiabatic compressor. For the complete extraction of this energy, the compression/expansion ratios of the compressor and the turbine, respectively, as well as the duration of the isothermal expansion process in
3870 the Riz2 rarefier must be modified after each cycle. These changes are made by a controller, which receives signals from the pressure transducers inside the rarefier and from the temperature transducers in the rarefier and the storage tank. If the Riz1 rarefier and the Diz densifier are placed in the same R3 tank, filled with a heat transfer agent and the Riz1 rarefier starts simultaneously with the heat engine, the expansion of the gas in the rarefier is made by absorbing the thermal energy ceded by the Diz densifier.
3875 The working gas pressure in the two isothermalizers is chosen as high as possible (to obtain a high power density), and the expansion and compression ratios can be optimized. With this storage system can be obtained a stored energy utilization factor close to 100%.
As with prior art storage systems, one of the factors limiting the increase in the amount of stored energy is the working gas temperature at the outlet of the isentropic compressor, the high temperatures
3880 requiring expensive materials. The avoidance of high temperatures is made, in the prior art, by a staged compression: in each stage, the working gas is compressed from the temperature Ta to the temperature Tm, then it is cooled in a heat exchanger to the value of Ta. The resulting gas with temperature Ta is stored in a high pressure tank, and the liquid with temperature Tm will be stored in a large enough tank.
An efficient solution by which high storage pressures can be reached, without the gas exceeding
3885 the temperature Tm, is presented in Fig.65: the gas aspirated from the atmosphere is compressed adiabatically by the compressor C1 , to reach the temperature Tiz (curve 1 -2 in Fig.65A), then is compressed isothermally, with constant Tiz, by the Diz1 densifier, to the point with siz entropy (curve 2-3 in Fig .65A) and again adiabatically, by the compressor C2, for to reach the temperature Tm (curve 3-4 from Fig.65A). After cooling in the HE heat exchanger (curve 4-5 in Fig.65A), the compressed gas, having a
3890 temperature close to Ta, is stored in the tank R1 , and the liquid coolant, having a temperature close to Tm, is stored in the tank R2. Due to the heat absorbed by the thermal sponge and the walls of the densifier, their temperature 7}, gradually changes. Therefore, the temperature T^TH+AT differs from one compression cycle to another, the difference being equal to this increase in the ambient temperature of the gas. AT\s established still from the design phase, and is a compromise between efficiency and power
3895 density and can be modified during the storage phase, depending on external conditions. After a long standstill of the installation, at start-up, in the first cycle, Tn = Tamb, and if until then the energy source is not exhausted, the process lasts until, in the last cycle, T,, = Tm-AT, where Tm is the maximum temperature admitted. At this time, the temperature of the liquid in the tank R1 is Tm-zl that of the tank R2. The thermal energy stored in these tanks can be extracted and transformed into mechanical energy
3900 by means of two heat engines, equipped with isothermalizers. In the configuration of Fig.65B, the liquid from the tank R2 is moved to the tank R1 , the functionality of the Diz1 densifier is reversed becoming the Riz2 rarefier, and together with the turbine T, the adiabatic compressor C2 and the Diz densifier mounted in the tank R3, form a heat engine that harnesses the heat stored in the R2 tank. The mechanical energy stored in the gas pressure in the tank R4 is harnesses with the help of the Riz1 rarefier and with the help
3905 of the thermal energy released by the Diz densifier.
The process described in Fig.66 is a new energy storage process, based just on the high energy efficiency of densifiers and rarefiers with thermal sponge, proposed in this invention. The advantage of the new method compared to those of the prior art is that almost all the thermal energy generated by the action of the piston is stored, with each compression cycle, in the working gas temperature, in the thermal
3910 sponge with its solid and/or liquid components and in thermal tanks, from where it is taken over, almost entirely, in the expansion phase.
We know that a thermodynamic process that takes place in one direction (for example, from cold to hot) and in which it is consumed/produced a certain amount of energy E, will produce/will consume, during the development in inversely direction of the process, an amount of energy the closer to E, the
3915 closer the conditions of the process are to those during the direct process. This is the principle on which the operation of the system described in Fig.66 is based, a system which, in the energy storage phase, functions as a refrigeration system, storing thermal energy, both positive and negative, and in the energy recovery process, is a heat engine that works between the two heat sources created in the first phase. The result is an extremely efficient and flexible system, useful in a wide range of storage applications, for
3920 a wide range of energy supply processes, with variable energy and dimensional parameters in a very wide range.
The proposed system consists of three distinct operating subsystems, arranged in three stages:
Energy storage is performed by a hybrid system that works after an reversed Carnot cycle (consisting of a heat pump and a refrigerator) and consists of the Diz densifier, the T isentropic expander,
3925 the Riz isothermal rarefier and the C isentropic compressor. The mechanical energy to be stored is taken over by the drive system of the Diz densifier, in order to isothermally compress its working gas, which is initially at an optimum pressure pM (in other configurations the storage process can start with taking the gas out of the atmosphere, its compression at to pressure pM, and storage in a tank R with constant pressure). During the compression process, the gas pressure increases to a predetermined pressure pf,
3930 at a constant temperature during the cycle: 7}z/=7}/+21T/, where Tn is the temperature of the thermal sponge, of the walls of the compressor and of the storage agent in the tank Rd, temperature that increases with each cycle. The isothermal rarefier Riz takes over the gas (expanded in the isentropic turbine T) at a temperature T^T^-zl ^ (Ti2 is the temperature of the walls, of the thermal sponge of the isothermal rarefier Riz and of the storage agent in the tank Rd) and expand it isothermally. The adiabatic
3935 compressor C ensures the transition of the gas from the variable temperature Tiz2 to the variable temperature Tiz1, by the appropriate modifications of the compression ratio, coordinated by the controller. The four apparatus structure a reversed Carnot cycle, which is the most efficient cycle for two heat sources with temperatures TH+ATJ and Ti2-AT2.
If we take into account the case that the temperature of the thermal sponge and the walls of the
3940 densifier is Tn, in the first cycle of the process the mechanical energy input of the piston materializes in increasing the pressure in the densifier and in a thermal energy addition, which is taken over by the sponge and the walls of the apparatus. In turn, part of the heat taken up by the walls is transferred to the storage agent in the Rd tank. A part of the thermal energy taken over by the thermal sponge has the same destination, if a cooling circuit is installed between the inside and the outside of the densifier.
3945 Therefore, after each compression cycle, most of the gas state quantities and the temperature of the other components involved in the process change, most often with very small values. If we are required to permanently maintain the isothermal nature of the compression and the same AT1 (considered the most economical), each cycle will increase the Tiz temperature, inlet and outlet pressures, as well as the temperatures of the thermal sponge, of walls and of the agent in the tank. Similar phenomena occur in the
3950 rarefier, if we keep permanently the isothermal character of the expansion and the same A T2 (the most economical temperature difference).
The heat pump will operate in this mode until Tiz1 and/or Tiz2 reach the predetermined limit values, at which point an additional cooling/heating system of the densifier and rarefier is switched to a steady state mode in which both Tiz1 and Tiz2, as well as the other state quantities do not change.
3955 Starting from the same idea, storage systems can be conceived with configurations in which to process the gas in a Stirling, Ericson, Rankine cycle, or even other cycles, if these cycles can be completed, in the recovery phase, in the sense conversely, with minimal exergy losses. For example, in the case of the Stirling and Ericson cycles, for acceptable exergetic efficiency, an isentropic compressor and an isentropic expander are required to force the gas to perform the (positive, or negative)
3960 temperature jumps AT3 and A T4, equal to the temperature difference between the primary and secondary of heat exchangers that recover thermal energy in the isochoric/isobaric processes.
The second stage of the system consists in the two isolated tanks Rd and Rr, in which the Diz densifier and the Riz rarefier are immersed, respectively. Between the walls of the tank and the device in the tank is the material (liquid or solid) that has the role of accumulating thermal energy, positive or
3965 negative, released by the device to the environment. To increase the storage capacity, this material can be, in the initial phase in a solid state (for example, a salt, a paraffin, even a metal, etc.), which melts after the Tiz exceeds the melting temperature. On the other hand, the material in the tank Rr may initially be in the liquid state and solidify during the gas expansion cycles in the rarefier. When they are in the liquid state, these materials can circulate inside the respective isothermal device, contributing to the efficiency
3970 of heat exchanges.
The use of liquid heat transfer agents is limited to a temperature range between the melting point and the boiling point, which requires, for storage at high temperatures, the use of a set of such agents, which complicates the configuration of the installations. In these cases, the uniformity of the temperature of solid thermal sponges can be done by installing fans and related piping inside the tanks. At high tank
3975 pressures, the heat exchange between the working gas and the thermal sponge is efficient enough, and the mechanical energy required to operate the fans is transformed, through gas-gas and gas-solid friction, into thermal energy and is stored in the thermal sponges.
In addition to the thermal storage agent, initially in the tanks Rd and Rr, quantities of this material are stored in the additional tanks R1 and R3, respectively. When the temperatures Tiz1 and/or Tiz2 have
3980 reached the prescribed values, a transport circuit is opened, through which a flow of storage agent, with temperature T is introduced into the tanks Rd and Rr, replacing a similar amount of agent with temperature Tamb, respectively T (introducing some temperature difference between the apparatus and the environment in the tank), quantity which is directed to be stored in the additional tanks R2 and R4 respectively.
3985 To limit thermal energy losses (which can be significant in the case of very high or very low storage temperatures), the thermal insulation of all system components is an active insulation (third stage of the system) of the type described in Fig.45. In the configuration in the figure, the cooling fluid is a gas, which yields its recovered thermal energy to a liquid agent, to be stored in the tanks R5 and R6, respectively.
3990 The recovery of stored energy is done by reversing the cycles performed during storage.
A great advantage of the system is its flexibility. Note, for example, that the configuration described contains all the components needed for atmospheric gas storage operations, similar to those in prior art CAES systems. The D.iz densifier, together with the compressor C, the expander T and a system of constant pressure tanks make up such a system, with an energy efficiency superior to the classical
3995 systems. If the expander T is removed from the circuit, the system supplies compressed gas at the desired Tiz temperature, the thermal energy of the gas can be stored together with the compressed gas, or it can be extracted in a exchanger and used for various purposes, same as the system described in Fig.45B. Switching from one configuration to another can be done at any time during the storage process. Moreover, in the initial phase, the R.iz rarefier can also change its sense (and role) and participate in the
4000 process of storing the compressed gas, at the temperature of Tamb, or Tiz. Also, the amount of stored thermal energy can be increased (increasing the amount of mechanical energy addressed for consumers), at any time of the process, from any available thermal source (fossil fuels, or biofuels, solar energy, geothermal energy, waste energy).
Can also be created configurations in which some of the energy is stored in compressed gas
4005 tanks at high pressures, at Tiz temperature, and another part in warehouses where thermal sponges extracted from the densifier or rarefier are stored.
The storage system in Fig.67 is similar, its main components being the tank Rd, in which the temperature of the gas and of thermal sponge ts 1 it contains are kept at the temperature Tiz1 and the tank Rr, in which the temperature of the gas and thermal sponge ts2 which its contents are kept at a Tiz2
4010 temperature as low as possible. For this, they are equipped with an active Rec insulation, which causes the gas that retains the thermal energy that could be lost through a passive insulation, to transfer this energy to the liquid agent in the tank R1 , respectively R2, or to a heat engine (respectively, a heat pump), which restores, with the help of a small additional energy input, the stationary temperatures in the two tanks. The Dizp densifier installed in the Rd tank, forms together with the adiabatic compressor Cp, the Tp
4015 turbine and the Rizp expander installed in the Rr tank, a heat pump. When starting the system, the Diz2 densifier, mounted in the tank Rr, draws air from the atmosphere, compresses it isothermally at the temperature Tiz2 and stores it, under constant pressure, in the tank R, at the atmospheric temperature Tatm. The transition from Tatm temperature to Tiz2 temperature and the reverse transition are performed by the isentropic T2 expander and the C2 compressor. Simultaneously with this compression operation, the
4020 heat pump also starts. It absorbs the heat delivered by the densifier to the tank Rr and transfers it to the tank Rd, together with an amount of heat equivalent to the mechanical work performed for this operation. As a result, the fluid in the Rd tank and its thermal sponge changes its temperature with each cycle of the pump, storing the mechanical energy received from a wind turbine, or from another source of mechanical energy. The Tiz1 temperature will rise to the permissible limit, while the Tiz2 temperature will remain
4025 unchanged. The recovery phase of the two forms of stored energy is done by the rarefier with variable isothermal speed Riz1 , together with the adiabatic compressor C1 and the expander D1 , with variable compression ratio, controlled by a regulation system. The temperature in the tank Rd will gradually decrease reaching, with the emptying of the gas in the tank R, a temperature close to Tatm.
A great advantage of isothermalizers starts from the possibility of these devices to store the
4030 absorbed mechanical energy also in the form of thermal energy, an advantage offered both by a very good insulation (in the active system described above) and by inserting the isothermalizer in the liquid tank, to which it yields its surplus heat, thus limiting the storage temperature. More than that, the system allows easy energy storage in the form of latent energy. For this, the solid thermal sponge, made of materials resistant to high temperatures, will have in its composition a series of sealed tanks, in the form
4035 of bars, or cylinders, inside which materials with a lower melting point are inserted (for example paraffin , salts, etc.), materials that, at a certain temperature, go into a liquid state, then (if the walls of the tank are sized to withstand the pressures occurring inside) even into a gaseous state. The system of these tanks can also be used to store heat transfer agents, which at the starting temperature are in solid state and change to liquid state before the starting heat transfer agent reaches the boiling temperature, so that, at
4040 the right moment, the starting agent is collected and stored in a reservoir, being replaced by the new agent.
The thermal energy thus stored can then be used as such, or it can be transformed, almost entirely, into mechanical energy. The outstanding energy efficiency of these devices and the possibility of obtaining them at a low cost, offer the possibility to make very simple, small energy storage systems,
4045 useful for small applications. Able to reach high power densities, comparable to those of electric batteries, thermal energy reservoirs can be used to propel vehicles of various dimensions. If they are executed in standardized shapes and sizes, they can be replaced very easily after unloading. Ill
For example, can be made systems consisting of such a densifier (preferably in two pressure stages), operated by a wind turbine, or by the electricity from the public grid, outside the peak load, which
4050 sucks in atmospheric air, and after compression, exhausts it back, at constant pressure, into a high- pressure metal tank (to avoid accidents, the tank is buried, or placed in a safety enclosure). Most often, the need to extract stored energy occurs during periods of peak load and higher atmospheric temperatures than in the storage phase, which is favorable to increased efficiency. The system must also contain hydraulic fluid reservoirs, a pump-hydraulic motor to recover the energy consumed during
4055 exhaust, a system for controlling the temperature of the gas in the storage tanks and, possibly, an electric generator to introduce the surplus energy into the electric grid. The stored energy is released by reversing the operating cycle, whenever needed. For example, for a compression ratio of 1 :350, a storage capacity of 1 m3 can provide the daily energy requirement for several families and can return power to the grid. For larger consumers (residential buildings, office buildings, etc.) one of the systems described above can be
4060 used, systems that can also meet the need for hot water.
The advantage of this system is that, due to the fact that the working agent works in a closed circuit, the energy storage is given up by storing the gas at a high pressure, all the mechanical energy received from the outside (less that stored in the working gas) is transformed into energy thermal, sensitive and latent, and stored in well-insulated tanks. In this way, part of the energy used to compress
4065 the gas is recovered from this phase, the energy required for thermal storage being lower, therefore the speed of gas circulation inside the cycle (thus the absorbed power) increases accordingly. In addition, by controlling the temperature difference between the thermal sponges of the isothermalizers and the working gas, this speed can be changed whenever there are variations in the power provided by the devices whose energy is stored.
4070 The same thing happens when the heat pump uses the outside environment as one of its energy reservoirs (as a cold or hot source respectively) and stores only heat, or only cold. And in this case, all the mechanical energy is transformed into heat, but the efficiency of the system depends on the size of the temperature difference between the gas and the thermal sponge of the isothermalizer that exchanges thermal energy with the ambient environment.
4075 Fig.68 shows one of the metal tanks 15.28 of a combined storage system, for storing mechanical energy in the form of compressed gas at high pressure (equal to the mechanical energy that can be obtained by expanding this gas to atmospheric pressure), tank which can store, in the oceanic environment (near a field of wind turbines), on the surface of a storage lake (near a hydropower plant, or a pumped hydroelectric energy storage PHES), etc, gravitational energy, equal to the weight of the
4080 volume of water that can be stored in the reservoir, multiplied by the depth of the respective accumulation. The shape of the tank should be, in the vertical direction, aerodynamic, and its volume should be slightly larger than the volume required for the empty tank to float on the surface of the accumulation (when the weight of the volume of water displaced is equal to the weight of the tank on land). If the infrastructure of the system (skeleton 15.29, foundation 15.33 buried in the ground 15.35, 4085 which can be shared with that of the infrastructure and in which a positioning niche 15.34 is made, with the shape and dimensions of the lower part of the tank) allows, the lifting installations 15.30 are mounted at a higher elevation, allowing the tanks to be maneuvered up to some height above the water level, ensuring the storage of an amount of additional energy equal to the weight of the tank multiplied by this height. In addition, such a tank has the property of floating near the surface of the water, and when it is
4090 flooded with liquid through valve 15.32, the air is vented through valve 15.27), to sink to the recess arranged in the foundation (in the case of shallow accumulations, this recess can be executed at much greater depths than the bottom of the accumulation), driving during the descent a mechanism that develops a mechanical energy equal to the product of the weight of the volume of water entered into the tank and the depth to which the tank reaches. The energy developed can be used during this dive (the
4095 instantaneous power being easily timed by mechanical means), or it can be stored (for example, by driving an isothermal densifier). For a new reuse, the tank must be brought to the surface, when there is available energy, by reversing the direction of the descent mechanism, or by draining the liquid from the tank. For this, working gas is pumped into the tank, at a pressure equal to the water pressure at the respective depth, or a pump 15.31 is used for evacuation, these processes requiring a properly sized pipe
4100 15.37 (for gas under pressure, or for liquid). A method by which the need for this pipe can be avoided is to allocate an additional compartment in the tank, filled with gas at a pressure high enough to, after expansion to the appropriate hydrostatic pressure, evacuate the entire amount of liquid from the tank. In the case of pumped hydroelectric energy storage, the enclosure is emptied by opening an outlet, when the liquid level in the storage pool reaches the level at which the tank is located, and after closing this
4105 hole, the tank returns to the surface it is ensured by the increase of the level in the basin, during pumping. Whichever method is used, it requires an energy consumption equal to the energy gained during the diving phase.
Additionally, in addition to gravitational energy storage, the reservoir volume can be used, both at the surface and at depth, to store gas at high pressure, compressed with state-of-the-art systems
4110 technique, or with the help of densifiers. Gas storage at high temperatures (thermal energy storage) is possible when efficient insulation is used, using lightweight materials.
In the case of off-shore installations, during the periods when the tank is on the surface of the sea or the ocean, the tank can be used, with appropriate constructive arrangements, for the production of energy by using the energy of waves and tides, through one of the methods used in the state of the art,
4115 and the facility's infrastructure can support underwater turbines for harnessing the energy produced by deep currents.
The idea behind this use can also be used to create underwater transport facilities with low energy consumption (Fig.68A), for example for the transport of wind turbine components from a warehouse on the coast, to offshore assembly site. If the transported parts contain internal voids (for
4120 example, wind turbine rotor blades), the volume of the gas tanks, necessary to ensure the buoyancy of the container with parts 15.38, is correspondingly smaller. Also, this type of transport can be used for the transport of goods between two locations, located on one side and the other of a river, a canal, a strait, etc., and even between two locations on land, located at equal elevations, if an underground tunnel is built between them, later filled with water. It can also be combined with rail transport, without the need to
4125 transship the goods to continue the route over obstacles of this type. The conveyor consists of the container 15.38, which can be filled with goods of any kind, to which is attached an appropriate number of tanks 15.44 with atmospheric gas (or with gas under pressure, in the case of combined systems), in such a way that the system floats near the accumulation surface. The conveyor is provided with support systems similar to those of cable transport installations 15.39 and/or with running systems 15.40 on rails
4130 15.41 , supported by pillars 15.42. The conveyor is also provided with a maneuvering tank 15.44 from Fig.68B, compartmentalized by a wall 15.47. In one of the compartments there is air under pressure, and the other compartment allows the intake of water through the valve 15.47. The volume of this compartment is calculated in such a way that at a partial amount of liquid (for example, half of the total volume), the conveyor floats near the surface). In this way, the admission of an additional amount of liquid
4135 leads to the sinking of the conveyor, the descent speed being dependent on the additional amount of liquid (frictional forces braking the descent in proportion to the forward speed) and on the angle of inclination of the running system, and the reintroduction of the same quantities leads to the stop of descent. If at this point the suspension system has a minimum point, the admission of additional liquid leads to the creation of an upward force, dependent on the additional amount of liquid, which tends to
4140 bring the carrier back to the surface of the accumulation. The liquid level in this compartment (hence, the lift) is regulated with the help of the compressor/expander C and the air circulated through the valve 15.46. In the case of small depths and short distances, the running system can be positioned on the bottom of the basin, a single inflection point being sufficient. In order to reduce the energy consumed by the compressor, it is advisable that the entire assembly be provided with an aquadynamic profile (for
4145 example, by providing a common outer shell 15.38a, of low weight), and that the outer surfaces be treated to reduce frictional forces. It should also be emphasized that, thanks to the Archimedean forces, all the mechanical stresses to which the system infrastructure is subjected are considerably reduced.
In the case of off-shore storage systems, this type of transporter simultaneously transports the gas tanks (with atmospheric gas, or under pressure, the weight of the gas being negligible), from the off¬
4150 shore systems, to the coastal locations. In this way, it is no longer necessary to transport electricity between the two locations, the wind turbines supplying the densifiers with mechanical energy, and the rarefiers on the shore converting it into electrical energy. In this way, the storage and supply periods become independent and can be carried out simultaneously. Also, the existence of such type of transport offers the possibility for systems installed in favorable locations, with permanent winds, to provide them
4155 with replacement tanks after filling the ones in operation.
Fig.69 shows the principle diagram of an isothermalizer in which, in order to reach a high final pressure, the compression/expansion is done in pressure steps. Unlike the state-of-the-art devices, maintaining a constant gas temperature during the transformation, excludes the use of intermediate heat exchangers. Instead, because the isothermal speed of the piston implies, in the case of a high
4160 compression ratio, a significant reduction of it during the transformation, three different isothermal speeds are used, one for each stage. In the configuration shown in the figure, the cooling of each thermal sponge is done with a heat transfer agent from three different tanks, the temperature in each tank being kept constant with the help of thermal engines that eliminate the surplus heat due to the action of the piston in the surrounding environment. The different temperatures in the three reservoirs allow the isothermal
4165 velocity of each piston to increase from the initial velocity (set for a predetermined AT and energy efficiency) as the pressure increases. In the configuration in the figure, the growth ratio from one stage to another is 3. When the described system is used in an energy storage system, the heat transferred to the environment, in the compression phase, is accumulated, in each of the stages, in a properly sized reservoir (even if this implies, in each step, to maintain the preset temperature differences, an increase in
4170 the mechanical work required for compression) and is taken over entirely, in the expansion phase, increasing the efficiency of the entire system.
The schematic diagram in Fig.70 shows the composition of a thermal engine, with internal combustion, which uses hydrogen as fuel. It is stored in liquid form at an appropriate temperature and pressure in the insulated tank 15.51. The necessary fuel is pure oxygen, also stored in liquid form in the
4175 isolated tank 15.52, a solution that completely eliminates the noxes. The working gas can be any inert gas (which does not react with either oxygen or hydrogen) which at any temperature in the range Tc-Th2 (temperature of the cold source - temperature of the hot source) is found in a gaseous state, and is stored in engine components. The engine cycle is an isothermalized Brayton cycle, as previously described. The inert gas, at the temperature of the cold source, is compressed adiabatically by the compressor C (curve
4180 2-3 in Fig.70A) and introduced into the combustion chamber CC under constant pressure, through pipe 15.53 where hydrogen is also introduced, through a pipe tubular, mounted inside the working gas pipe, and the oxygen, through a direct pipe. By burning hydrogen (curve 3-4 in Fig.70A), a mixture of the working gas with water vapor is created, the temperature of the mixture being dependent on the fuel flow. The mixture is taken by a expander E1 and expanded to the temperature corresponding to the partial
4185 pressure of water vapor (curve 4-1 in Fig.70A), vapors condensing in the vertical plate separator 15.49, where the partial pressure of the vapor approaches zero. If this temperature is higher than the temperature of the cold source Tc, the inert gas undergoes further expansion in the expander E2, then it is introduced into the densifier D. This is mounted in a tank 15.50, filled with liquid, the temperature of which is maintained near Tc with the help of the heat exchanger HE, in which the liquid is recirculated
4190 with the help of the pump P. Through the isothermal compression of the working gas (curve 1 -2 in Fig.70A), part of the resulting heat is given to the environment, the other part being taken over by the rarefiers Ro and Rh to produce additional mechanical work and introduce residual heat into the oxygen and hydrogen reservoirs, where each of these gases is at its evaporation temperature, to produce the evaporation of these gases. The composition of the two adjacent engines also contains the compressors
4195 and expanders necessary to carry out the adiabatic jumps, corresponding to a Carnot cycle, or the heat exchangers to carry out isobaric or isochoric transformations, corresponding to Ericsson and Stirling cycles, respectively.
The engine is also equipped with devices for regulating gas flows, with devices for regulating isothermal speeds, with the possibility of changing the average temperature of the cycle, as well as with
4200 devices for regulating the speed of adiabatic devices. This type of engine is less polluting than any other heat engine in the state of the art.
On the same principle, engines can be made, having any operating cycle, that use hydrogen as fuel. Fig. 70B shows the T-s diagram of an engine operating after a Rankine cycle (curve 1 -2-3-4-5-1 ). This can be a thermal power plant, operating on fossil fuels, which can be transformed, after some
4205 modifications, into an ecological power plant, which only consumes hydrogen from the electrolysis of water, or produced by other less polluting methods). To obtain the heat necessary for its operation, an engine similar to the one in Fig. 70 is used, which consists of an engine whose working gas is an inert gas (when it is advantageous, two engines can be used, operating with different gases, which together achieves the desired cycle), which uses hydrogen as a fuel and pure oxygen as an oxidizer, both stored
4210 in a liquid state. The energy used to liquefy these gases, by the method described in the invention, is contained in the respective gases and is released entirely in the processes in the engine. The thermodynamic transformations through which the inert gas passes are drawn on the T-s diagram in linedot mode (Acad ISO 10W100), on the curve 1 -6-7-8-4-3-2-1. All isentropic transformations in this cycle are performed by compressors and dynamic turbines, isothermal transformation 3-2 by a densifier, whose
4215 thermal sponge is washed by boiler water at boiling temperature, curve 7-8 is an isobaric process obtained by burning hydrogen, and the curves 4-3 and 2-1 are the heat exchanges that take place in heat exchangers, to bring the water to the boiling temperature and to superheat the vapors obtained. If a suitable heat source is available (for example, solar energy obtained in CSP systems), the two engines can produce mechanical energy independently, simultaneously, or sequentially, depending on available
4220 resources and consumption needs. In this case, the inert gas goes through the 5-6-7-8-5 cycle, and if, beforehand, taking advantage of the existence of a cold source (for example, during the night, when the consumed energy also has a low price), it was stored isothermally a working gas (e.g. air) at high pressure, expanding it through a rarefier whose thermal sponge is washed by water from the steam turbine condenser, an appreciable amount of additional mechanical energy can be obtained.
4225 ICAES storage systems using isothermals are also ideal for storing energy provided by offshore wind turbines. The simplest solution is to install large metal tanks, with a diameter of several meters and a height equal to the depth of the sea in the area, mounted on surfaces on the seabed, properly arranged. Due to their considerable weight, these tanks are sufficiently stable to be used as a platform support for isothermals with their associated installations, piping systems, laboratory and maintenance facilities, each
4230 platform serving a suitable number of turbines.
Fig.71 shows a complex renewable energy storage facility provided by a multitude of sources: wind energy, solar energy, gravity energy, sea wave energy and submarine current energy. The storage plant can be located offshore, not very far from the coast and in areas where the water depth is not very high, on the surface of natural or artificial (reservoir) lakes. The devices intended to capture the energies
4235 available in that location are the wind turbines 1 , the lifting installation 2, the solar panels 5, the installations for harnessing the energy of the waves 6 and 9, the axial turbines 10 driven by the energy of the deep currents. The energy generated by these installations during the periods when the wind speed falls within the limits allowed for the safe operation of the wind turbines, during the periods when the solar radiation exceeds the minimum threshold necessary for the operation of the solar panels, during the
4240 periods when the height of the waves exceeds the lower limit necessary for the operation of the installations 6 and 9 and during periods when the speed of the submarine currents is sufficient for the profitable operation of the axial turbines 10, is converted into electrical energy to be stored in electric batteries, or is transmitted by submarine cables to the shore power grids for use, or stored until a reasonable request arises. Another way of energy storage, existing in the state of the art, is to use the
4245 captured energy to compress the atmospheric air, with the help of quasi-isothermal compressors, to pressures greater than 20-30 MPa and store it in tanks located on the ocean floor, to send it directly to the supply network at the time of demand.
The present invention adopts the CAES solution of energy storage by compressing atmospheric air, but uses, for compression and for expansion, the isothermalizers 7 described in this invention, and for
4250 storage it uses metal tanks of different sizes, which also can receive other destinations. One of these destinations has been described in the previous paragraphs, where the reservoirs are dimensioned in such a way that they can be used, with the help of the lifting plant, for a gravity storage, the stored energy being equal to that required to lift a quantity of water, equal to that storable in the available reservoirs, from the bottom of the sea to its surface. The same type of tank (whose weight, when containing only air,
4255 is approximately equal to the weight of the amount of water it displaces when submerged), can be used, whether the air inside is compressed or not, for to valorize the Archimedean energy of the waves, in installations 9, similar to those in the state of the art. Also, another functionality of this type of tank was previously described, the tank being used for onshore transportation of compressed air in the storage process (and possibly other materials), avoiding the use of submarine electrical cables and keeping for
4260 the isothermalizer only the compression function, the expansion being done on the coast.
Compressed air can also be stored in large tanks, with a vertical axis, or slightly inclined towards the direction of the prevailing currents, with a diameter of several meters and a height equal to the depth of the sea in the respective area, mounted on properly arranged surfaces on the seabed, 4 p. Due to their considerable weight, these tanks are stable enough to be used as supports for the location of turbines
4265 driven by the energy of marine currents, or for anchoring floating installations (installation 6 for capturing wave energy). Also, if they are properly sized and if they are helped by intermediate support points 11 , which use Archimedean forces (light tanks, with thinner walls, but which could displace large volumes of water), these cylinders, become pillars of support, can bear the weight of a platform 5p, on which the solar panels 5 can be placed, with their related installations, the lifting installation 2, the isothermalizer 7, 4270 the pipe system, facilities for laboratories and for maintenance work. Another use that can be transferred to these cylinders is that of supporting wind turbines. In this way, the pillars that support the turbines are shorter and better reinforced, they are protected from the stressful action of waves and sea currents, they can be located at greater distances from the coast and at higher heights.
Fig.72 shows a new type of wind turbine, with sliding blades, which offer a series of advantages
4275 in terms of construction simplicity, capture power and multiple placement possibilities. The turbine consists of two vertical pillars 16.1 , on each of them a drive device consisting of the drive wheels 16.2 (the version with gear wheels is represented in the figure) mounted on the hubs 16.2.1 (Fig.72A), which drive a well-stretched 16.2.2 link chain, or an elastic belt. The shaft of one of the drive wheels is coupled to the shaft of a user device, for example an electric generator. Between the two chains, the horizontal
4280 support bars 16.4 are mounted, parallel to each other, on which, by means of the bushings 16.5, the turbine blades are mounted, whose shape and dimensions are chosen taking into account the location of the turbine and the characteristics of the air currents from that area. Each bushing is provided with blade locking and rotating devices. The width of the blades is chosen small enough to allow their simultaneous rotation, both around the support bar and around their axis, to allow finding the best positioning,
4285 depending on the direction of the wind. Also, as a safety measure, in a high wind situation, the blades can be rotated, or allowed to rotate freely, to position themselves parallel to the direction of the wind, in the position of minimum resistance. Also, each support bar is provided at one of the ends with a rotating device 16.3, composed of a hub 16.3.1 and two fins, fixed perpendicular to the surface of the hub, in continuation of two of the radii of the circle representing the cross section (Fig.72Ba), located in such a
4290 way that when the blade support bar reaches the upper position, the operating rod 16.3.3 acts on one of the fins, causing in this way the 90° rotation around the longitudinal axis, of the whole bar, with the vanes mounted on it, positioning them in such a way that they continue to capture the air currents also during the phases when the bars have a downward movement. Maneuvering rod 16.3.3 in the lower position ensures the successive return of all vanes to their initial position as they pass through the rod. In the
4295 configuration in Fig.72Bb, the two vertical planes on which the blades move are spaced from each other, in order to use the wind energy more efficiently. Configurations where the blades move in a horizontal plane are also feasible.
A process that allows the capture of wind energy from a larger area than that provided by the surface of the turbine exposed to the air current consists in the placement of fixed guide vanes (16.6.1 of
4300 Fig.72B) in front of the turbine, which also have the role of accelerating current speed. Depending on the characteristics of the prevailing winds, these vanes can be sized and positioned in such a way that the captured energy is maximum.
The turbine in Fig.73 has a similar construction, but the support bars of the blades are mounted on some circular discs 16.7, located directly on the shaft of the electric generator. For high power
4305 turbines, the positioning of the blades 16.6 in relation to the wind direction changes continuously so that maximum power is provided to the turbine shaft. In the configuration in the figure, applicable for small- sized wind turbines, with a single blade for each of the radial directions, we chose the solution with 90° rotation around the longitudinal axis of each blade, using the device 16.3 described previously, positioning it in this way in the position where the interaction with the air currents, in the direction of
4310 return, is minimal. This position of the blades is also recommended for turbines usable for harnessing the energy provided by submarine currents (or for flowing water courses), or for those used to capture wave energy where, after each blade leaves the active phase, it rotates by 90 ° around its longitudinal axis, to oppose a minimum resistance to the water current (the configuration in the figure, where 8 is the surface of the river). In the case of turbines mounted on the surface of the sea, to capture wave energy, the
4315 rotation of the blades must take into account the fact that, immediately after the passage of a wave, in the immediate vicinity of the water surface, submarine currents arise in the opposite direction to the movement of the wave, which can be exploited if the change of the angle made by the blades with the direction of the current is done with some delay. Also, a turbine of the type shown in Fig.72B.a can be mounted, where the movement of the blade support chain can be done both horizontally and vertically.
4320 For flowing water courses, in the portions without significant level difference, the turbine shaft is mounted on the surface of the river, and the blades immersed in the water are parallel to the flow direction. Fixed vanes (16.6.1 from Fig.73) can also be added to this type of turbine to direct the current of air or water.
Another use that can be given to these types of turbines is the protection of banks and hydrotechnical constructions, by reducing the effect exerted by waves and submarine currents. Also, in
4325 the case of solid bodies moving at high speeds through a fluid (such as ships, submarines, tanks of the type shown in Fig. 68A, trains on rails, road vehicles, aircraft, etc.) part of the effect of braking exerted by the frictional forces between the fluid and the surface of the mobile object, can be converted by a set of turbines 16.6, mounted directly on the surface (16.22 of Fig.73A) of the mobile object, in semi-cylindrical housings created especially for this purpose. Depending on the configuration of the external surfaces of
4330 the mobile object (surfaces that are designed in such a way that the energy captured by the turbines is as large as possible), the layout, number and dimensions of the installed turbines are chosen. Several turbines can be grouped on the same axis, and mechanical (belts, gears, etc.) or electrical couplings are installed between the axes, the total energy being transmitted to an electric generator or a hydraulic motor. In the configuration of Fig.73B, the hydraulic turbines 16.6 are mounted on a mobile chain placed
4335 on a system of rollers 16.25 that take the captured energy, similar to that used by the wind turbines in Fig.72B. The turbine system is mounted on a light casing (16.23 of Fig.73B), built around the movable object 16.24. In both configurations, the return movement of the vanes is made after the rotation of each vane around its axis, so that the resistance opposed to the movement of the movable body is minimal.
The systems in Fig.74 and Fig.75 are intended to capture wave energy based on the pressure
4340 difference between two locations on the sea surface, located differently in relation to the position of the waves. Maximum efficiency is achieved when the distance between them is equal to the distance between the crest and the bottom of a wave. A similar construction can be made when the system operates based on the pressure difference at the same location at different times. As a rule, the energy captured with a single system is quite small, but through the construction of large assemblies, obtained by
4345 joining a large number of systems, the captured energy becomes significant. These systems become attractive when the respective assemblies are arranged in such a way as to constitute light floating platforms 16.8, on which solar panels, wind turbines with a vertical axis can be placed, or according to the invention, thermal energy capture systems, based on the difference in temperature between layers located at different depths, compressed air storage tanks from ACAES systems with isothermalizers, for
4350 offshore wind turbines, warehouses, power stations, etc.
The system in Fig.74 is composed of a large number of tubes 16.9, with a preferably rectangular section, which hermetically close a certain volume of air, placed next to each other, and one after the other, so that, together with a series of elements of joint, to form a floating platform. Above this platform is mounted another structure 16.8, used as a platform for storage and for mounting devices 16.10 with
4355 various destinations, made in such a way as to allow the circulation of the liquid vertically and to ensure the buoyancy of the whole assembly, together with all the machines mounted on it (if the load is very high, floating tanks are installed from place to place to support the assembly). The weight of the entire assembly is calculated in such a way that the platform composed of the active tubes 16.9 is submerged at a well-established depth h below sea level, the tubes being all the time covered by a layer of water with a
4360 height greater than h (equality being recorded only at the bottom of the wave). The system by which the platform 16.9 is fixed to the platform 16.8 must allow this depth to be adjusted according to the height of the waves 6.11. For this, the volume of gas in the sealed enclosures of the platform 16.9 must provide a lifting force equal to the weight of the platform. The length of a single tube is at least half the length of the most frequent waves at that location, so that when the tubes are horizontal and submerged at depth h,
4365 the maximum water pressure differences across their ends are as large as possible. For waves of height H, at one end the gas pressure will be patm + pgh, and at the other end patm + pg(H+h). For situations where the length or direction of wave propagation changes, the possibility of horizontal rotation of the tube system must be provided, in such a way that the maximum pressure difference can be obtained for most of the tubes. In this way, the pressure difference between the two ends of the tube is maximum,
4370 regardless of the height of the waves.
The active tubes are divided into two approximately equal compartments, separated in the middle of the tube, by a wall, or by an intermediate compartment 16.17, of smaller width (to obtain larger volumes, the intermediate compartment can have a larger diameter, or can be coupled with an external gas tank). These tubes are fixed to the platform 16.8, perpendicular to the prevailing direction of wave
4375 propagation. The possibility of extending the tubes through additional sections, similar in construction, must also be foreseen. Both compartments are provided at the ends with movable pistons 16.12, provided with rods 16.14 and with sealing gaskets, as well as with stoppers to avoid the piston exiting the tube. In the configuration in Fig.74, was chose a variant in which, between the piston and the bottom of the compartment, a bellows 6.13 is mounted with the length at least equal to that of the compartment and
4380 with the largest average diameter. In this case, the diameter of the piston is slightly smaller than that of the tube, the seals are missing, there is a thin layer of water between the walls of the tube and those of the bellows, and the piston is equipped with rollers that rotate on the walls of the tube. An easier configuration can be obtained if the tube is replaced by a few rails for the piston rollers to run on and intermediate rings mounted along the bellows. In both configurations, the rotation of one of the rollers
4385 (motor roller) can convert the pressure exerted by the displacement of the waves into rotational mechanical energy and can be collected to be stored, or it can be transformed into electrical energy and transmitted to a distribution network. Another method of capturing the energy of the back-and-forth gas flows consists in the installation of Wells turbines.
The two compartments (respectively, bellows) are filled with air with a pressure equal to the
4390 average pressure reached by the sea water at the depth at which the bellows are located (patm + pg(H/2+h)). When this pressure is achieved at both ends of the tube (the crest of the wave is above the intermediate compartment), the volume of air in the tube must ensure, for the upstream piston, a position close to the intermediate compartment, the respective section being filled with water, and for the downstream piston, a position close to the end of the tube, the respective section being filled with air. The
4395 compartments communicate with each other as well as with the partition compartment, the pressure inside them being the same at all times. Starting from this moment, the displacement of the waves causes the displacement of the two pistons in the same direction, an additional volume of water enters the downstream section, while from the upstream section, water is evacuated. When the bottom of the wave reaches above the intermediate compartment, the upstream piston reaches close to the end of the tube
4400 and the downstream piston reaches close to the intermediate compartment. These back-and-forth movements are repeated with the passage of each wave, without changing the amount of gas in the sealed enclosures. In the configuration in Fig. 74, the rods of the two pistons also cross the partition compartment, being coupled, in this area, with a mechanism 16.15 (common, or one for each rod) that transforms the energy of the two alternative movements into rotational mechanical energy , or directly into
4405 electricity. The system must be equipped with a load regulator, which regulates the travel speeds of the pistons, adjusted in such a way as to ensure the maximum travel lengths (maximum collected power).
The system in Fig.75 is similar, but the two compartments are separate tubes, mounted vertically, side by side, on the entire surface below the platform 16.8 located at the depth h, perpendicular to the surface of the platform. The rods 16.14 of the pistons 16.12 move vertically, and their lower end moves
4410 sealed inside a tube 16.18 with hydraulic liquid, constituting the piston of a hydraulic pump. In the variant in the figure, the bellows 16.13 in the tubes are evacuated, to create a pressure multiplier equal to the ratio between the surface of the piston in contact with the sea water and the surface of the piston at the opposite end of the rod, in contact with the hydraulic liquid. Under the action of these pistons, the liquid is pushed to/from the reservoir 16.19, passing through the hydraulic motors 16.20 and transferring the
4415 captured energy into rotational mechanical energy, or directly into electrical energy.
Tto collect the captured energy, other energy conversion and storage systems can be also implemented.

Claims

Claims
1. Technical process applied cyclically to the gas in a closed enclosure, during a cycle, or a limited time interval in this cycle, gas that changes its volume and/or pressure due to the action of some forces from outside the enclosure, characterized by the fact that, the gas from the enclosures is forced to perform a strictly isothermal transformation, and this type of transformation is obtained thanks to the action of some regulation systems of the actuation devices of the mobile components that generate the modifying forces, regulation systems which, through commands sent to these actuation devices, control the isothermal velocity, with which the volume of the enclosure and the pressure of the gas in the enclosure change, control whereby the difference between the average temperature of the gas in the enclosure and the average temperature of the elements with which it comes into contact remains at a rigorously constant value throughout the mentioned duration.
2. Thermodynamic machine with cyclic operation, designed for the application of the process of claim 1 , which we will call an isothermalizer, comprising:
- one or more closed enclosures provided with devices for the introduction at the beginning and evacuation at the end of each cycle of the gas in/out of the enclosure
- mobile components, with direct action on the gas state variables in the enclosures
- actuation devices for these mobile components, provided with adjustment devices in addition to which there may be other components:
- liquid and/or solid bodies introduced into the enclosure to exchange heat with the working gas
- devices for the continuous or intermittent introduction/evacuation into/from the enclosures of a liquid and/or an aqueous foam,
- devices for the introduction/evacuation of an additional gas flow rate from the enclosures, after the end of the intake phase,
- transducers for measuring the characteristics of the gas in the enclosures, of the heat transfer agent and of the machine components characterized by the fact that, during a period of time within a cycle, between the end of the intake phase and the beginning of the discharge phase, the gas in each enclosure is subjected to a strictly isothermal transformation through which it changes its pressure in a monotonically increasing manner (in which case the machine is a compressor, which we will call a densifier), or monotonically decreasing (in which case the machine is an expander, which we will call a rarefier), and this type of transformation is obtained due to the action of the regulating devices on the actuation of those mobile components (solid or liquid) of the machine that change the volume of the enclosure and the pressure of the gas in the enclosure by: changing the speed of the piston, or/and the rotation of the rotor, or/and the flow of the liquid piston, or/and changing the input and output flow of the circulated heat transfer agent, or/and the change of the input and output flow of the additional gas, correlation of all these changes being carried out in such a way that the difference between the average temperature of the gas in the enclosure and the average temperature of the elements with which it comes into contact (walls of the enclosure, elements mounted in the enclosure, the heat transfer agent, aqueous foam) remains at a value rigorously constant throughout the mentioned duration.
3. Technical process according to claim 1 , characterized in that, before and after the isothermal transformation that takes place in the isothermalizer enclosure at the working temperature Tiz, they take place in the same enclosure, or outside it, at the beginning and at the end of the cycle, during small fractions of the cycle, through the control of the system controller exerted on the same actuation devices, or on some devices introduced for this purpose, rapid thermodynamic transformations, almost adiabatic, through which the temperature of the working gas quantity is brought to the Tjz temperature, and after the isothermal transformation, the temperature of this amount of gas is changed, up to a desired temperature (Fig.l B)
4. Isothermalizer with positive displacement according to claim 2, characterized in that the keeping of the constant average gas temperature is carried out by forcing, throughout the cycle, through the kinematic links of the constructive elements and/or through preset commands transmitted to the actuation device with variable instantaneous speed/angular velocity of the piston/rotor, a vjz variation which was previously determined theoretically and/or experimentally
5. Isothermalizer according to claim 2, characterized in that the keeping of the constant average temperature is carried out with the help of movable elements operated by devices whose speed and position is imposed by an automatic controller, which throughout the cycle, receives and processes the signals received in real time, from measuring devices installed inside and outside the respective enclosure, determines the required position of the movable elements corresponding to the measured state variables and issue issues commands to the actuators to correct the deviations
6. Isothermalizer according to claim 5, characterized in that the automatic adjustment device receives signals from the moving element position sensor and from the pressure sensors installed in the working room (Fig. 2)
7. Isothermalizer according to claim 5, characterized in that the automatic adjustment device receives signals from a series of temperature sensors mounted on the inner and outer surfaces of the device, estimates (zonal and/or total) the amount of heat given by the working gas to its ambient environment and that accumulated by the thermal sponges, as well as the heat given to the outside environment and sends commands accordingly to the actuation devices of the movable components
8. Isothermalizer according to claim 5, characterized in that the automatic adjustment device receives signals from a series of temperature, pressure, flow and position sensors mounted inside and outside the isothermalizer and outputs signals to the adjustment devices of the flow rate of the heat-transfer fluids, to control, zonally and/or totally, the amount of heat exchanged by the working gas with its ambient environment, in order to obtain a rigorous isothermal transformation of the gas in its enclosure
9. Laboratory apparatus intended for the study of the behavior of gases and vapors in closed enclosures with variable volume, characterized in that it has the configuration of an isothermalizer according to claim 2, but the commands transmitted to the mobile organs are chosen by the operator, depending on the problem studied, then, based on the collected signals in real time, the controller device calculates and records the resulting trajectory according to the transmitted commands
10. Technical process for the design and realization of isothermalizers according to claim 2, intended for operation in conditions where the environmental parameters and the input and output characteristics of the working gas vary in a narrow band, characterized in that, the determination of the optimal isothermal trajectory for the given conditions, or of another desired variation curve T(t), is done with a prototype of an isothermalizer equipped with high-performance control and execution devices, then a new prototype is made, in which a number of these devices are replaced with devices simpler to make and cheaper (as far as possible with simple kinematic devices), the series of simplifying prototypes can continue until the best performance is achieved
11. Technical process for making isothermalizers according to claim 2 characterized by the fact that, when certain operating conditions are changed, the shape of the isothermal trajectory (or the optimal one) does not change, but changes the temperature difference between the gas and its ambient environment, which leads to a change in the same ratio of the duration of each cycle, a change that can be obtained by varying the voltage/current of the supply of DC motors, or the frequency of the current with which the AC motors are supplied
12. Isothermalizer with solid piston according to claim 2, characterized in that the solid piston is actuated by the pressure of a liquid agent supplied by a variable flow rate hydraulic motor, and has a telescopic rod composed of several segments (13.31 of Fig.3b), these segments entering the kinematic chain successively, at certain values of the gas volume/pressure, changing the surface on which the working fluid pressure is exerted, and consequently, the displacement speed of the piston
13. Isothermalizer with positive displacement according to claim 2, characterized by the fact that, between the walls and the movable element of the machine, a deformable device called deformable thermal sponge (12.6 of Fig.2) is inserted, characterized by a variable geometry, being made of solid components whose total surface in contact with the gas in the cylinder is large and which is composed of a series of elements that, due to their elasticity, or the kinematic links between them, change their shape and/or position, depending on the position of the piston/rotor, causing a change in the minimum volume in which this sponge can be fitted, but keeping almost unchanged the contact surface of the sponge with the gas inside the isothermalizer
14. Isothermalizer with piston, with alternative movement according to claim 13, characterized in that its thermal sponge is similar to bath sponges, made of elastic materials (various polymers, elastomers, various types of rubber, meshes, or elastic metal foams), where, since the manufacturing phase, measures are taken to increase the mechanical resilience and the heat transfer coefficient
15. Piston isothermalizer, with alternative displacement according to claim 13, characterized in that its thermal sponge is made up of one or more helical springs, with a circular or rectangular section, arranged between the piston and the cylinder cover (Fig. 5), or between flat plates parallel to them (Fig.9), the springs may have coils with different sections and different diameters, some of the springs with a small diameter can be mounted inside the ones with a larger diameter
16. Piston isothermalizer, with alternative displacement, according to claim 13, characterized in that its thermal sponge is made up of helical springs, elastic cords, bellows and other types of elastic elements mounted between the piston and the cover (Fig. 6, Fig. 7, Fig.8, Fig.9), on which are fixed heat-absorbing flat plates, parallel to the surface of the piston
17. Alternative displacement isothermalizer according to claim 15, characterized in that, on both sides of the flat plates of its thermal sponge, fins or other vertical elements are fixed, with fixed length, or with variable length, through various telescoping methods (Fig. 8, 10)
18. Alternative displacement isothermalizer according to claim 13, characterized in that, at the periphery of the horizontal plates are mounted on their contour (5.11 c, figure 8), and/or inwards, vertical short walls, to retain on the plates a certain amount of liquid, from that existing in the enclosure, or from that in the cooling/heating circuit, with the aim of reducing the dead volume and/or to remove the heat accumulated by the sponge
19. Alternative displacement isothermalizer according to claim13, characterized in that its thermal sponge is made by alternating flat metal plates and curved, elastic metal plates (Fig.12, 13 si 14), plates which have, when the piston is at the top dead center (TDC), approximately the same surface, and through which perforations are made for the convective circulation of the gas in the enclosure
20. Alternative displacement isothermalizer according to claim 13, characterized in that its thermal sponge is made of horizontal flat metal plates, mounted on a mechanical structure with variable geometry, fixed between the piston and the cylinder cover, which deforms through the movement of the piston without altering the parallelism of the plates
21. Alternative displacement isothermalizer according to claim 20, characterized in that its thermal sponge (Fig.15) is made of flat horizontal metal plates (5.1 1 ), mounted on horizontal supports (5.20) fixed in foldable carry-supports (5.19) in the form of the narrow blades, the support-carriers having the two ends fixed by joints that allow tilting movements, one end of which is fixed, by means of a movable joint, to a flange (5.18) rigidly fixed to the piston, and the other end being fixed, also by a movable joint, to a pivoting arm, the pivoting arm having a guide roller (5.16) attached to it, which can run on a rail, or in a channel (5.17) in the cylinder head, the displacement of the piston leading to the pivoting of the supportcarriers , accompanied by a change in the distance between the horizontal plates
22. Alternative displacement isothermalizer according to claim 20, characterized in that its thermal sponge (Fig. 16) is made of flat horizontal metal plates (5.11 ), mounted on folding supports (5.25) which are fixed in/on carry-supports made of rods, bars, pipes, or narrow blades, 5.22, 5.23, 5,24, the whole structure having the shape of harmonics or bellows, whose height and/or width changes with the displacement of the piston.
23. Technical process by which a prior art piston compressor can be transformed into a densifier according to claim 2, thus increasing its energy efficiency, characterized in that this compressor is fitted with a thermal sponge according to claim 13, a drive device according to claims 4, or 5 and/or a cooling/lubricating system, after performing a set of preparatory dimensional changes to its components, which allow these changes
24. Densifier according to claim 2, characterized in that when the pressure in the enclosure reaches a set point, an suction valve is opened for the penetration of a liquid piston, the gas in the enclosure being further compressed, then, after reaching the desired pressure, it is exhausted without changing its pressure, through this valve, if it is at the highest elevation of the enclosure, or through another valve properly located (Fig.17, Fig.18)
25. Piston isothermalizer according to claim 2, characterized by the fact that it contains an electronic processor that stops the piston periodically, at the top or the bottom dead center, for a period of time, so that the thermal sponge (alone or together with the piston) is extracted from the enclosure, in order to cool/heat it quickly, or for the purpose of temporary storage, in order to store the thermal energy it has absorbed (Fig.18A)
26. Densifier/rarefier according to claim 2, characterized in that the working enclosure(s) of the machine is/are included in a cooling/heating system which also contains a variable flow and pressure pump for continuously or intermittently circulating the heat-transfer fluid (under liquid form, aqueous foam, or liquid mixed with surfactants) and an external heat exchanger, the introduction of the agent into the cylinder being done by free flow, or by spraying
27. Densifier/rarefier according to claim 26, characterized in that, in the configuration of the sponge, there are zones for collecting the liquid mixed with surfactant, zones where working gas is injected from the outside, at a pressure equal to or higher than that in the cylinder, for foam regeneration
28. Isothermalizer according to claim 26, characterized in that it has an electronic processor that stops/slows down the movement of the piston, at certain predetermined time intervals, for predetermined time durations, or set by a control system that receives signals from a set of temperature sensors, leaving in operation, at the working flow rate, or at an increased one, the cooling system and the lubrication system
29. Isothermalizer according to claim 26, characterized in that the portion of the heat transfer liquid circuit located inside the cylinder is made of one or more tubular helical springs, in the walls of which are mounted devices for spreading the heat-transfer liquid inside the enclosures, springs that change their height depending on the position of the piston
30. Isothermalizer according to claims 20 and 26, characterized in that the portion of the heat-transfer fluid circuit located inside the cylinder is mounted on/in the skeleton supporting the horizontal plate system, or on similar independent support structures specially mounted for this purpose
31. Isothermalizer according to claim 2, characterized in that it has the same construction as a diaphragm compressor (Fig. 11 ), in which a thermal sponge made of elastic elements or a mechanically deformable assembly is inserted, as well as a controlled transfer system of heat between the gas inside the enclosure and its ambient environment
32. Isothermalizer according to claim 2, characterized in that its cylinder is composed of two segments with different diameters (Fig.19), in the main cylinder 5.1 (whose inner diameter is equal to the outer diameter of the piston) a thermal deformable sponge composed of the horizontal metal plates 5.11 is mounted (and its volume in the fully compressed state is only a fraction less than the volume of the cylinder corresponding to this position of the piston), so that when the piston moves up to the top cover, it pushes the entire amount of gas into the second segment 5.1s of the cylinder, provided with a dense thermal sponge with a large absorption surface, after which the discharge valve 5.5e opens, and the cylinder is flooded with the liquid on the discharge line
33. Isothermalizer according to claim 2, equipped with a thermal sponge made of horizontal plates 5.11 , characterized in that it has one or more vertical bellows 12.28 (Fig. 20), made of elastic, resistant materials, such as rubber, in the wall of which , at the level of each space bordered by two horizontal plates, one or more spraying systems 12.29, of the sprinkler type, are mounted
34. Isothermalizer with liquid piston, according to claims 1 and 2, characterized by the fact that a thermal sponge composed of horizontal plates (Fig. 21 ) is mounted inside the cylinder, on the whose periphery (or on a portion thereof) access paths are made for the access of the liquid agent above the previous plate (skirts 7.3b in Fig.21 ), in such a way that the liquid penetrates almost simultaneously into all the spaces bounded by the horizontal plates, to create closed gas pockets, at almost equal pressures at different elevations, each such gas bag being provided with its own holes for gas inlet/outlet, thus becoming a small liquid piston compressor/expander (elementary isothermalizer)
35. Liquid piston isothermalizer, according to claims 2 and 34, characterized by the fact that it is made by alternating, in the same enclosure (Fig. 22), two types of elementary isothermalizers:
- the first type (7d from Fig.22) consists of two close horizontal plates, where the liquid comes from the main column, the compressed gas is transferred, as its pressure increases, through check valves mounted in the upper wall (7.10 from Fig.22 ) in the upper elementary isothermalizer, of the second type, where it is forced to pass through a layer of liquid with which it exchanges thermal energy, and the lower wall is equipped with check valves (7.1 1 from Fig.22) for liquid supply of sprinklers , or nozzles, from the lower elementary isothermalizer of the second type
- the second type (7c from Fig.22) consists of two more spaced horizontal plates, it does not have access to the main column of the liquid piston, it has valves mounted on the upper wall for the intake and for spraying the liquid from the isothermalizer of the first type located above his, it has valves for the gas coming from the lower elementary isothermalizer of the first type mounted on the lower wall, and in a side wall it has check valves through which the initial intake and final exhaust of the working gas is made
36. Liquid piston isothermalizer with vertical cylinder, according to claim 2, characterized in that, in the upper part of the cylinder, a non-deformable thermal sponge is mounted, with a large absorption surface (5gs from Fig. 23), and in the lower part a deformable thermal sponge, and in this lower part, in a first stage, a solid piston compresses the gas and reduces the volume of the deformable sponge, and in the second stage, when the piston reaches a certain position, one or more valves are opened in the piston (5.2s), allowing the penetration of a liquid agent , which still functions as a piston and as a cooling agent of the upper thermal sponge ( Fig .23 and 23A)
37. Isothermalizer with liquid piston made according to claim 2, as well as for other types of thermodynamic devices with a vertical axis, whose enclosure (Fig. 24) is divided into several mini- isothermalizers through the horizontal metal walls 7.17 and through the vertical wall 7.18, characterized by the fact that the vertical wall is provided, at the base of each compartment, with slots for the penetration of the liquid piston into the mini-isothermalizers, characterized by the fact that its thermal sponge is composed of horizontal plates, one in each compartment, between these plates and the upper wall of the compartment being fixed telescopic fins 5.10, equal in length, which change their length simultaneously and equally, due to the vertical movement of some horizontal plates 5.11 , with a slightly smaller surface than of the lower wall, plates resting each on a plate 7.21 fixed on a movable vertical rod 7.20, the displacement of the rod being determined by a mechanism coupled with the mechanisms that actuate the devices for adjusting the flow of the liquid piston, in such a way that the horizontal plates moves at the same speed as the surface of the liquid in the respective compartment moves, until the maximum compression of the telescopic fins, each compartment of the isothermalizer communicating through a pipe 7.23, closed by a valve 7.22, with a tank under constant pressure 7r, the opening of the valves 7.22, when the desired pressure is achieved, causing the replacement of the compressed gas in the isothermalizer with liquid from the tank, so that, after opening the gas inlet valve 12.12 and the liquid outlet paths and after withdrawing the vertical rod, the isothermalizer becomes suitable for a new cycle
38. Isothermalizer with liquid piston made according to claim 37, characterized in that in the gas containers under the horizontal plates are isothermalizers, provided with a light, non-deformable thermal sponge and are connected to each other by flexible pipes, and one of the containers is connected to an external isothermalizer which, together with the containers, forms a gas piston isothermalizer, compressing the gas inside the containers, the instantaneous pressure in this gas piston isothermalizer being correlated with that in the liquid piston isothermalizer, in so that the intake and discharge of the gas
39. Liquid piston isothermalizer made according to claim 37 (Fig.24C), characterized in that the isothermalizers 7.18s with telescopic fins are provided with a liquid piston inlet valve 7.18v, which opens only after its lower wall 7.18p, which is, in this phase, also a solid piston, reaches its upper position, and the gas containers below this piston are liquid piston isothermalizers 7.18i, equipped with a light, non- deformable thermal sponge, in which the liquid enters in the first phase of compression, causing the compression of the telescopic fins of the elemental thermal sponges 7.18s and the gradual increase in pressure in the upper isothermalizers 7.18s, until the solid piston is blocked, after which the gas pressure rises until it reaches the value of the pressure in the storage tank and are successively opened the valves 7.22.2 from the vertical pipes 7.23, pipes which make the connection between each mini-isothermalizer and the tank 7r
40. Double-effect liquid piston isothermalizer, made according to the process of claims 1 , 2 and 3, characterized in that it is composed of two single-effect liquid piston isothermalizers (Fig. 25), so that the hydraulic motor 7m which ensures the movement of the two pistons at isothermal speed, alternately aspirates and discharges the liquid from/into the enclosures of the two isothermalizers, through the controlled closing/opening of some valves 7v.
41. Process for manufacturing the thermal sponges of the isothermalizers made according to claim 2, characterized in that light materials, with high thermal conductivity, plated with thin layers of metal are used.
42. Apparatus for determining the mode of variation during a cycle, of the average temperature in a closed enclosure with variable volume, made according to the method of claim 1 , characterized in that, during that cycle, the instantaneous pressure of the gas in the said enclosure is measured and the position of the mobile devise, with the help of pressure and position transducers, and based on this pair of data, the instantaneous average temperature of the gas in the closed enclosure and how it varies over time is determined, with the help of gas laws; the device can be reversed, in the sense that it can be equipped with a controller in the memory of which the desired trajectory for a specific application is entered and based on the signals collected from the transducers, it issues commands to the executive bodies to ensure the realization of this trajectory
43. Liquid piston isothermalizer with vertical cylinder, according to claim 2, characterized in that, in the upper part of its cylinder, a thermal sponge (7gs from Fig. 26, 26a, 26b) with a large absorption surface is mounted, separated from the lower part by a wall (7.2s) in which a valve is mounted that allows the opening, on external command, of a communication path, in this lower part being mounted one or more thermal sponges (7g) with a high capacity of elastic deformation, made of plates elastic metal, with corrugations of different radii of curvature, or elastic metal plates rolled in the form of rolls of variable diameter, or of strips and other small elastic elements, placed in sealed deformable bags (7.14), together with a liquid fraction that removes the dead volume of these bags when fully compressed, the inflatable bags being each provided with a lower opening through which they can communicate with the layer of liquid, the opening being closed by an externally controlled valve and with an upper hole (7.2o) through which each bag communicates, also through valves, with the upper part of the cylinder, the isothermalizer operating in the first phase by moving the liquid piston until the gas is discharged from the inflatable bags, and in the second phase, after equalization at external order of the pressures in the two enclosures and the recovery of the mechanical energy stored in the elastic elements of the thermal sponges, by further moving the liquid piston
44. . Liquid piston isothermalizer according to claim 43, characterized in that the deformable bags in the lower compartment contain vertical metal plates with different radii of curvature, which when subjected to external pressure created by the liquid piston, flatten to the point of overlap.
45. Liquid piston isothermalizer with vertical cylinder, according to claim 43, characterized in that a single bag, or a series of bags in the form of mattresses or tubes is mounted in the lower compartment (Fig. 26a), the bags being filled with a mixture of elastic and inelastic elements, so that in the free state it occupies most of the volume of this compartment, and when the pressure is maximum it occupies as little of this volume as possible.
46. . Liquid piston isothermalizer with vertical cylinder, according to claim 2, characterized by the fact that a sponge with a large heat absorption surface is mounted in its cylinder in the upper part (7gs from Fig. 27), and in the lower one a thermal sponge in the form of a tubular pipe (7.16) with double walls, in which sprinklers are mounted, and the liquid piston of the isothermalizer circulates through the double wall of the pipe, from where a part penetrates both into the gas layer from inside the tube, as well as in the space between the pipe and the cylinder walls, the other part entering, together with the excess liquid evacuated from the enclosure, into the pipes of a loop cooling circuit
47. Liquid piston isothermalizer with a vertical cylinder, according to claim 2, characterized by the fact that in its cylinder are mounted (Fig. 27a), a sponge with a large heat absorption surface (7gs from Fig. 27a), in the upper part, and in the lower one, thermal sponges made of concentric vertical cylinders (7.3v), arranged at greater distances in the central part of the densifier, but increasingly closer towards its periphery, fixed alternately on the upper wall and on the lower part of the device, in such a way as to create for the liquid piston that enters the cylinder through the central area, a path as long as possible, the sponge plates being provided with holes (7.3o) made in the upper part of the vertical plates to maintain the same pressure throughout the cylinder and with other holes made at lower elevations, to accentuate upward convective currents
48. Rotary isothermalizer, according to claims 1 and 2, characterized by having the same construction as a rotary compressor, or a rotary pump of the state of the art for which, additional sealing measures between gas enclosures with pressure differences between them are taken and to which is attached a controlled heat transfer system between the gas enclosure and its environment and/or which is operated by systems that enforces the isothermal angular velocity
49. Rotary isothermalizer, according to claim 48, characterized in that it is carried out on the configuration of a rotary compressor with a blade in rotor (Fig.28) to which a liquid/aqueous foam spraying system with nozzles mounted in the wall of the stator and/or rotor is attached, the fluid flow being controlled by the control valves and being also circulated through an external heat exchanger
50. Rotary isothermalizer, according to claim 48, characterized in that the liquid supplying the sprinklers is taken from a tank in which this isothermalizer is fitted
51. Rotary isothermalizer, according to claim 48, characterised by the fact that a thermal sponge consisting of cylindrical metal sheets of different diameters, with values between the diameter of the rotor and that of the stator is installed between the rotor and the stator, each of these sheets having an opening along one of the generators, to allow the blade to move alternately, the rotor centerline being moved towards the stator centerline, in the plane containing them, with a distance equal to the combined thickness of all these sheets, without leaving any spaces for gas leaks from the high pressure area to the low pressure area (Fig.29)
52. Rotary isothermalizer, according to claim 48, characterised by the attachment to a rolling piston compressor of a thermal sponge according to claim 51 , of a cooling system according to claim 49 and/or of an angular velocity control system (f ig.32B)
53. Rotary Isothermalizer, according to claim 48, characterized by the fact that it is realized by attaching to a rotor vane compressor a cooling/heating system according to claim 49 (Fig.33) and a system adjusting the angular velocity and heat transfer fluid flows, so as to obtain the isothermal trajectory
54. Rotary isothermalizer, according to claim 48, characterized in that in the rotor of the apparatus, cavities are made in the space between the blades, where pipes equipped with sprinklers and other elements of a solid thermal sponge are fitted (Fig.34 and Fig .35)
55. Rotary isothermalizer, according to claim 48, characterized by the fact that in the pipe through which the compressed gas is discharged, mounted at the highest level of the device (Fig. 35), a layer of liquid is permanently kept, through which it is removed, in the time of discharge, the dead volume of the apparatus and through which the pressures of the gas in the apparatus and of the gas in the upstream device (e.g. a storage tank) are equalized
56. Double-effect piston isothermalizer, according to claim 2, characterized by being made by joining two isothermalizers in the same cylinder (Fig.36), and their common piston is operated by means of profiled cams (6.14) fitted in one of the enclosures, by means of springs (6.16) mounted in the other enclosure and by means of telescopic carry-supports (5.22, 5.23, 5.24) mounted in both enclosures
57. Rotary isothermalizer according to claim 48, characterized by having the same construction as any gear pump (Fig.37A), to which a controlled heat transfer system between the gas in the enclosure and its environment is attached and to which it applies technical procedure in claim 55
58. Rotary isothermalizer according to claim 48, characterized by having the same construction as any cam pump (Fig.37B), to which a controlled heat transfer system between the gas in the enclosure and its environment is attached and to which it applies technical procedure in claim 55
59. Rotary isothermalizer according to claim 48, which has the same construction as any state-of-the-art liquid ring compressor, characterized in that thermal sponges are mounted in the spaces between the rotor blades.
60. Rotary isothermalizer according to claim 48, which has the same construction as any state-of-the-art screw compressor (Fig.38A), characterized in that a thermal sponge is inserted between its spirals (6.18 and 6.19 in Fig.38A), consisting of rectangular metal plates (6.21 in Fig.38A), of a width equal to or less than the height of the main spirals and of a length approximately equal to that of these spirals, which in the unstressed state have the same shape and the same radii of curvature as the main spirals, which are supported on the fixed spirals cover and are spaced apart by rectangular elastic metal slats, with the same width as the main spirals and with a much shorter length (6.20 in Fig.38C), made and arranged in such a way that in all cross sections perpendicular to the spirals, in which the distance between the main spirals is minimal, occupy all the free space between the spirals, forming a sealing plug which separates watertight the compressor regions with different pressures as well as a controlled system for heat transfer between the gas inside the appliance and its environment
61. Isothermalizer according to claim 48, characterized in that it has the same construction as a peristaltic compressor (Fig. 38D), in which, in the peristaltic tube(s) constructed of deformable and at the same time elastic materials, a thermal sponge formed of rectangular metal plates is introduced, this metal plates being rectangular in shape, slightly wider than the diameter of the tube not deformed and the length approximately equal to that of the tube, metal plates which, in the unstressed state, have the same radius of curvature as the peristaltic tube support (6.21 in Fig.38C, or 5.14 in Fig.38E) and are spaced by elastic springs, or by rectangular elastic metal slats, with the width equal to that of the main plates and much smaller in length (6.20 in Fig.38C), made and arranged in such a way that in the cross-section from the position in which they are tensioned by the peristaltic roller to occupy the whole area of the section of the peristaltic tube, forming a sealing plug which separates watertight the two compressor regions with different pressures, the system being provided with a controlled heat transfer system between the gas inside the apparatus and its environment
62. Rotary screw isothermalizer according to claim 48, characterized in that its rotors have variable pitch, decreasing towards the exit and that it is provided with systems for carrying out heat exchange by spraying with liquid, or by introducing foam, or oil with drops in suspension, the inlet and outlet flows in/from each enclosure being determined by the automatic control of the valves on the liquid circulation pipes, and the sprinklers for dispersing the heat transfer liquid are mounted in the device casing and/or in the metal body of the rotors, being powered from a channel that passes through their axis.
63. Isothermalizer according to claim 2, hereinafter called a gas piston isothermalizer, characterized by the fact that it is composed of a first stage consisting of one or more solid or liquid piston compressors/expanders, and a second stage, consisting of from a reservoir (8.2I from Fig.39, Fig.39A) equipped with a thermal sponge cooled/heated in continuous flow, with a volume significantly larger than that of first-stage compressors, a reservoir in which the difference between the temperature of the gas and that of the heat transfer agent (and , by default, of the sponge) is maintained at a constant value by adjusting the coolant flow rate and and the flow rate of the gas from/to the first stage, the tank gas pressure increasing/decreasing with each cycle of the first stage
64. Isothermalizer with gas piston, according to claim 63, characterized in that the first stage is composed of a rotary blower C1 that pushes the gas into a tank R, from where it is sucked by a densifier C2 that pushes gas in the second stage with the pressure and the temperature of the gas in its tank 8.2
65. Gas piston isothermalizer, according to claim 63, characterized in that its first stage is composed of a quasi-isentropic compressor/expander (C1 in Fig. 39A6), which discharges/suctions the working gas into/from a tank with constant pressure p compressed gas with constant temperature Tiz, from where it is taken/where it is discharged by one or more isothermalizers, which discharge/absorb the gas to/from the second stage in which a thermal sponge is mounted, the gas temperature being maintained at a constant Tjz value, and the pressure has a variable value pr
66. Gas piston isothermalizer, according to claim 63, characterized in that its first stage consists of two isentropic compressors/expanders (C1 and 03 in Fig.43) which together would ensure the increase/decrease of the gas temperature from the value Tamb/Tiz to the value Tiz/Tamb and an increase/decrease in pressure from the value po/pr to the value pr/p0 (where pr is the gas pressure of the second stage), between the two compressors/ expanders being an isothermalizer interposed, of whose valves are controlled by an automatic system, system which, based on the pressure signal from the second stage, ensures at the entrance/exit to/from the second isentropic compressor/expander, that pressure that ensures at its exit/entrance, an equal pressure with the pressure at that moment from the second stage
67. Gas piston isothermalizer, according to claim 63, characterized in that the tank of second stage (8.2i of Fig .39, Fig.4O), together with its thermal sponge and its cooling/heating system, is inserted into a tank (8.2 from Fig.39) of larger dimensions, which in its lower part contains a heat transfer liquid, and in its upper part, above the level of the inner tank (8.2i), contains a working gas layer, which communicates through a pipe with the gas from the inner tank, as well as a thermal sponge (8gs from Fig.39) for cooling/heating this upper gas bag, the liquid from the main tank (8.2) being transported, keeping a constant level in the tank 8.2 , by a hydraulic pump, through an heat exchanger (HE)
68. Gas piston isothermalizer, according to claim 63, characterized in that the gas in the inner tank (8.2i in Fig.4O, Fig.40A) is cooled/heated by a system composed of a deformable metal band running on a roller system mounted in the main tank (8.2 of Fig .40) at the boundary between the two tanks, or in both tanks (Fig.40A), in such a way that the metal band transports thermal energy between the gas in the inner tank and the liquid in the main tank, the openings through which the band passes from one tank to the other being sealed as well as possible and the fluid level in the inner tank being maintained constant, at the lowest possible level, with the help of a pump
69. Gas piston isothermalizer, according to claim 63, characterized in that the second stage of the transformation takes place in a reservoir (8.2i of Fig.41 ) which is, at the same time, the primary of a plate heat exchanger, and the secondary of the exchanger contains a fluid that absorbs thermal energy from the primary and transports it to another heat exchanger, where it gives it to another medium, or it contains a refrigerant with vapor at the saturation limit, in which case this secondary becomes the evaporator/condenser of a heat pump
70. Gas piston isothermalizer, according to claim 63, characterized by being composed of two isothermal transformation systems, each with an isothermalizer (8.1 din Fig.42) and an tank (8.2i) with its own cooling/heating system (Fig.42), the two (8.2i) tanks together constituting a heat exchanger
71. Gas piston isothermalizer, according to claim 63, characterized in that its second stage (tank 8.2 of Fig .42A) contains a thermal sponge composed of closely spaced vertical plates or/and of thin, deformable metal wires, which is cooled/heated by a (8.8a) sprinkler system and/or a foam generator system (8.9), the liquid accumulated at the bottom of the tank being kept at a constant level by means of a pump (8.6) which discharges this liquid through a cooling/heating system containing a heat exchanger and one or more coolers/heaters with bubbles or with gas layers, gas originating from the first compression/expantion stage
72. Isothermalizer with gas piston and heat accumulation (Fig.43, Fig.44), according to claim 63, characterized in that the second stage is equipped with a thermal sponge made by mounting on the upper cover of the tank a large number of vertical plates, bars, or vertical wires, single, or in bundles (Fig. 44), deformable, or fixed, metallic, or made of other materials, and through the distribution system to introduce, with a minimum pressure difference, the heat transfer liquid, to drain slowly on the vertical elements of the thermal sponge, towards the lower cover, from where it is taken up by the recirculation pump
73. Process for thermal insulation of isothermalizers according to claim 2, of the tanks according to claim 71 and 72, as well as of other state-of-the-art devices, characterized in that the insulation layers are arranged around the respective device on a series of structures (frames, nets, plates, etc) mounted around the object to be insulated in such a way as to form a circulation channel for a heat transfer fluid (Fig.45), formed by successive layers of fluid with a progressively increasing/decreasing temperature, from the medium to the surface of the device, which takes most of the heat emitted by the device and introduces it into a series of useful applications
74. Process of active thermal insulation of devices and tanks according to claim 73, characterized in that the insulating structure (Fig. 45C) is composed of registers of tubular pipes 15.5, made of insulating materials, placed in parallel planes, the exit hole in the last register of each plane being connected to the entrance hole 15.6 of the first register in the next plane, the space not occupied by the registers being occupied by heat-insulating materials
75. Process of active thermal insulation of devices and tanks according to claim 73, characterized in that the insulating structure (Fig. 45G) is composed of insulating plates made either of compact materials, having a large number of small diameter perforations, which achieve communication paths between successive layers of fluid parallel to each other, located at a short distance from each other, on peripheral spacer elements (on all four sides), in the form of sticks 15.7, to create sealed volumes from the outside, occupied by the heat transfer fluid, either from spongy materials with open cells
76. Process for active thermal insulation of devices and tanks according to claim 73, characterized in that the insulating structure (Fig. 45H) is composed of vertical insulating plates, the spaces between the plates being traversed, in a horizontal direction, by an air flow generated by a blower, and in the vertical direction, of thin liquid flows, circulated by a pump
77. Gas-gas or gas-liquid heat exchanger, made with isothermalizers according to claim 2, characterized in that, before being introduced into the exchanger, the gases (respectively, the gas) are isothermally compressed in a densifier D (Fig .46), up to a pressure PM, and at the exit they are expanded, also isothermally, in a rarefier R, up to the initial pressure Po
78. Gas-liquid heat exchanger according to claim 77, characterized in that the liquid agent is also used as a heat transfer fluid, cooling/heating the sponge of the densifier/rarefier, ensuring the isothermal trajectory by changing the speed of the piston and by controlled changing the flow of the cooling liquid, extracted from the liquid pipe
79. Heat recuperator made with isothermalizers according to claim 2, characterized in that the heat recovery is done in a liquid-atmospheric gas heat exchanger HE, then it is stored in the thermal sponge of a gas piston isothermalizer with heat accumulation
80. Waste gas heat recovery facility, made with isothermalizers according to claim 79, characterized in that the hot gases, after passing through a pre-filter-treatment facility F (Fig. 51 ), are isothermally compressed, at the exit temperature of filter, to an economical pressure in a Diz densifier, then they are compressed adiabatically in a compressor C, which brings the gas to the pressure of the next stage, after which it is introduced into a gas piston densifier with energy storage, according to of claim 72
81. Waste heat recovery plant from the combustion gases of a CT thermal plant (Fig. 51 A) intended for the production of the thermal agent for the heating facilitys, made with isothermalizers according to claim 2, characterized in that the hot gases are taken with temperature Tm from the exit from the boiler by a densifier D with a solid or liquid piston, which also has the role of sucking the gas required for combustion from the ambient environment, which compresses it isothermally, up to a pressure pm, chosen economically, the temperature of the thermal sponge being kept constant with the help of an HE liquid-hot gas heat exchanger, mounted at the boiler outlet, then the gases are taken over by a expander (gas turbine) which transforms this thermal energy into mechanical energy, to set in motion an electrical power generator, or is stored in thermally insulated tanks, or in a densifier with heat accumulation according to claim 72
82. Heat pumps for the transfer of thermal energy from a cold source to a hot source, made with isothermalizers according to claim 2 ( Fig .52 and Fig.53), characterized in that the two isothermalizers are each immersed in a a heat-transfer liquid bath, in which there is also a pump that circulates the liquid from the bath through the isothermalizer enclosure, exchanging thermal energy with its thermal sponge and through one heat exchanger each, the one of the densifier being mounted in the hot source, and that of the rarefier in the cold source
83. Heat pumps according to claim 82 (Fig.53B and Fig.53D), characterized in that one of the heat sources is the atmospheric air, the other is the air from the ambient environment, and the working gas from the two isothermalizers is the air sucked from room atmosphere and discharged to the same atmosphere after processing
84. Heat pumps according to claim 82 (Fig.53F and Fig.53G), characterized in that one of the heat sources is the atmospheric air, the other is the air from the ambient environment, and the working gas from the two isothermalizers is the air sucked from the outside environment and repressed in the same atmosphere
85. Heat pumps according to claim 82 (Fig. 54), having the same configuration as the pumps in claim 84, but in containers 15.20 and/or 15.21 the heat exchangers HE1 and/or HE2 are no longer mounted, but one Riz1 rarefier, respectively Riz2, each coupled with a Diz1 or Diz2 densifier. The rarefier Riz1 is mounted in enclosure 15.18 (together with the densifier Diz, the rarefier Riz2 is mounted in enclosure 15.21 , the densifier Diz1 in enclosure 15.18.1
86. Heat pumps according to claim 82, characterized in that the air drawn in from the outside is introduced in the first phase into a calciner for its thermodynamic sterilization
87. Heat pump made with isothermalizers according to claim 2 (Fig.55 and Fig.55C), characterized in that it is composed of a Diz densifier, mounted in a liquid enclosure 15.25 metallic, provided with cooling fins, which suck air from the enclosure and compresses it isothermally and from an adiabatic expander E, mounted in a gas mixer 15.24, which expands this gas adiabatically, or by a rarefier that expands it isothermally, the replacement of the air sucked in by the heat pump being done by the active insulation of the partition wall
88. Heat pump made with isothermalizers according to claim 2, for the production of cold in refrigeration and air conditioning installations in buildings, characterized in that its rarefier absorbs heat from the liquid coming from a liquid-ambient heat exchanger, and the densifier gives it to the liquid in a thermal energy storage tank, to be used as a heat source for a heat engine, hot water for various processes, domestic hot water for local use, or to be delivered to a network
89. Air treatment and conditioning system, made with heat pumps according to claim 82, characterized in that it has a thermodynamic calcinator for air sterilization, and the isothermalizers in the composition exchange their thermal energy with the latent energy from the condensation and evaporation of a refrigerant
90. System for sterilizing and cooling/heating the air in one or more enclosures, according to claim 82, characterized in that it is composed of two loops connected to each other, one for air, the other for the working gas, one operating after one cycle Carnot, the other operating according to an inverted Carnot cycle
91. System for treating gas from one or more enclosures, according to claim 2, characterized in that the intake and exhaust of gases and liquids from thermodynamic machines is done through wide holes, which can have the same section as the cylinder of the respective device (Fig. 60A), orifices that are opened and closed by operating a multi-way valve, and enclosures may be created in the valve body for temporary fluid storage
92. Isothermalizer according to claim 91 , characterized in that the gas intake and exhaust from the devices is done through a 3-way valve (Fig. 60B), the compressed gas being introduced into a cavity in the valve ball, and the discharge being done by the hydraulic liquid from the discharge pipe, which communicates directly with the constant pressure gas tank
93. System for cooling, respectively heating the air in a room, according to claim 82, characterized in that the working gas loop works according to a Stirling or Ericsson cycle (Fig. 61 )
94. Gas liquefaction systems, characterized in that it is made using isothermal densifiers and rarefiers according to claim 2, together with polytropic compressors and expanders, heat exchangers and/or other thermodynamic devices
95. System for gas liquefaction according to claim 94, characterized in that it works according to a Siemens cycle, the liquefied gas being sucked in and isothermally compressed (curve 1 -2 in Fig.62A), in a Diz1 densifier (Fig.62), up to a pressure P2 that corresponds to an entropy s2, slightly higher than the entropy of the critical point, then, the gas is expanded adiabatically (curve 2-3 in Fig.62A), in a turbine T, up to a pressure below the saturation curve of vapors, close to the atmospheric pressure Pa and at a temperature below the critical point, the discharge being done in a condenser where, by extracting the latent heat with a heat pump (curve 3-4 in Fig.62A), the gas is completely liquefied.
96. System for the liquefaction of gases according to claim 95, characterized in that the condenser of the installation is the secondary 15.21 of a plate heat exchanger (Fig.62), through whose primary 15.22 circulates a heat-transfer fluid, recirculated by a pump 15.27 through a tank 15.25 in which the Riz2 rarefier is mounted, a heat pump operating in a Carnot cycle (curve 2'-5'-4'-3' in Fig.62A), or similar, and the densifier Diz2 of this heat pump pumps the heat extracted as well as that originating from the mechanical work consumed, to another heat-transfer fluid, at the ambient temperature, or at a temperature different from it, the liquefaction flow control being performed by an additional system for cooling the gas in the condenser, composed of a Diz3 densifier, which uses the liquefied gas itself as a cooling liquid and an expansion valve
97. A method of adjusting the operating characteristics of thermal engines, of refrigeration systems, of energy storage facilities, of liquefaction facilities and other types of facilities which have in structure isothermalizers according to claim 2, characterized in that it uses controllers and other devices to change during operation, the “isothermal velocities” and compression/expansion ratios of compressors and expaders in the composition
98. Method for the increase efficiency of state-of-the-art internal combustion engines, operating after a closed or open cycle, characterized in that the first phase of this cycle is an isothermal compression at minimum cycle temperature (usually equal to atmospheric temperature), performed by a densifier according to claim 2, which replaces those phases of these cycles, in which the thermal energy of the gas in the system is released to the environment at temperatures above the minimum temperature, without modifying the other phases (Fig.63D)
99. Internal combustion engine according to claim 98, characterized in that it is composed of a densifier according to claim 2, an adiabatic compressor with an adjustable compression ratio, a combustion chamber which is at the same time an expander and in which the compressed air from downstream apparatus is heated, which use any type of fuel and in which the useful volume varies by the displacement of a movable piston, displacement which takes place during combustion, the products of combustion being introduced into a turbine (or other type of expander), by moving the combustion chamber piston in the opposite direction, at the outlet of the turbine the gases having the pressure and temperature close to the atmospheric ones ( Fig .63E)
100. A quasi-isothermal combustion chamber for internal combustion engines according to claim 98, characterized in that it is located in the cylinder of a piston expander (Fig.63F), that the fuel supply starts since the gas suction phase and lasts throughout the displacement of the piston to TDC, the combustion is triggered immediately after the suction valve/drawer is closed and the combustion continues throughout the fuel intake, and the fuel flow and piston speed are regulated by a controller to achieve an isothermal expansion.
101. Energy storage system, characterized in that it is performed using isothermal densifiers and rarefiers according to claim 2, together with polytropic compressors and expanders, heat exchangers and/or other thermodynamic devices from the state of the art.
102. Energy storage system according to claim 101 (Fig.64), characterized in that the mechanical energy available at a certain time is used for adiabatic compression of a working gas (which can even be the atmospheric air) up to the temperature Ts, after which the compressed gas is brought to the storage temperature (which can even be the atmospheric temperature) in a heat exchanger and stored in a tank under constant pressure, while the heat-transfer fluid that took over the thermal energy difference is stored at temperature Ts, in a thermally insulated tank, for as when an energy demand arises, the compressed gas to be expanded in a isothermal rarefier Riz1 , and the stored thermal energy is extracted with a heat engine, equipped with isothermal densifiers and rarefiers, whose hot source is the stored agent (whose temperature decreases progressively) and the cold source is the environment, or a reservoir from which the Riz1 isothermal rarefier draws its energy necessary for expansion
103. Energy storage system according to claim 102 (Fig.65), characterized in that the compression process in the storage phase consists of three distinct steps: an isentropic compression from ambient temperature to a temperature T^Tu+AT, where Tn is the thermal sponge temperature at that time, an isothermal compression in a densifier at the temperature Tiz and an isentropic compression from the temperature Tiz to the maximum temperature Tm (from the heat exchanger entrance)
104. Energy storage system according to claim 101 , characterized in that the available energy is taken up by a hybrid system operating after a reversed Carnot cycle, or an equivalent one (Stirling, Ericsson) (Fig. 66), which consumes this energy to transfer heat from the tank Rr to the tank Rd, in which the isothermalers D.iz and R.iz are immersed, progressively increasing the temperature and pressure differences with respect to the ambient environment of the gas in the two isothermalizers D.iz and R.iz, as well as the temperature differences with respect to the ambient environment of all the elements that make up the devices in the system, of the heat transfer materials stored in the system tanks and of the gas in the tanks R5, R6 (when they exist) that store the thermal energy recovered by the Rec layers of active insulation of the Rr and Rd tanks and of the additional tanks that store the heat-transfer agent R1 , R2, R3 and R4, process that results in the formation of two finite sources of thermal energy, one hot, the other cold, sources that will be used at the time of an external demand for mechanical energy, by the hybrid system that reverses its sense of operation, becoming a thermal engine that it produces mechanical energy by going through the processes of the storage phase in the opposite direction
105. Energy storage system according to claim 101 , characterized in that it is composed of:
- a storage tank R under constant pressure (Fig. 67) - a mechanical energy storage loop, which in turn is composed of a densifier D.iz2 mounted in an Rr tank with heat-transfer agent with Tiz2 temperature kept constant, an adiabatic compressor C2 and an adiabatic expander T2, which ensures the temperature jump of the aspirated gas from the ambient temperature to the isothermal compression temperature T,z2 and vice versa,
- an thermal energy storage loop, which in turn consists of a heat pump with rarefier R.izp mounted in the same tank Rr and with the densifier D.izp mounted in another tank Rd with heat transfer agent with temperature Tiz1 , variable in increasing direction, from ambient temperature to maximum and an adiabatic compressor Cp and an adiabatic expander Tp,
- a mechanical power supply loop composed of a rarefier R.iz1 mounted in the Rd tank, an adiabatic compressor C1 and an adiabatic expander T1 , which starts in the recovery phase, to expand at the Tiz1 temperature, variable in decreasing direction the gas of tank R, until full use of the thermal energy stored in the tank Rd.
106. Energy storage system for small applications according to claim 101 , characterized in that it is composed of a single reversible isothermalizer (with one or more stages) coupled with one or more tanks under constant pressure, as well from tanks for hydraulic fluid, coupled to a hydraulic pump/motor, of a system for controlling the gas temperature in the storage tanks and possibly of an electric generator to introduce the excess energy into the network
107. System for storing mechanical energy in the form of compressed gas at high pressure, using densifiers according to claim 2, made on the surface of a sea or a lake, characterized in that it is combined with a gravity storage system (Fig.68) composed from a skeleton 15.29 on which the lifting installations 15.30 are mounted, a foundation 15.33 buried in the ground, in which a positioning dwelling 15.34 is made, with the shape and dimensions of the lower part of a tank for gas storage, the tank being provided with a valve 15.32 for filling with liquid and a valve 15.27 for the release of air, the volume of the tank being slightly greater than the volume required for the empty tank to float on the surface of the accumulation, so that it accumulates gravitational energy, equal to the weight of the volume of water that can be stored in the tank, multiplied by the depth of that reservoir, energy it releases if it sinks to the bottom of the reservoir after it has been completely filled with water
108. Underwater transport installation with low energy consumption (Fig. 68A), characterized in that it consists of a cargo container, a set of tanks for storing compressed gas, according to claim 107, dimensioned in such a way as to ensure buoyancy of the container and a handling tank 15.44, compartmentalized by a wall 15.47, in one of the compartments being air under pressure, and the other compartment can admit water through the valve 15.47 up to the level that balances the weight of the transporter with the weight of the volume of water displaced, the tank being provided with a compressor, or a pump, which can act on the liquid level, generating in this way, an ascending or descending Archimedean force, all these components being protected by a hydrodynamically shaped casing, provided with a running gear that can move on some rails, or on a cable system
109. Thermal engine with internal combustion, which uses hydrogen as fuel, made with the help of isothermalizers according to claim 1 (Fig. 70), characterized by the fact that it consists of a closed circuit in which there is an inert gas under pressure, consisting of a expander D, mounted in a reservoir 15.50 of heat-transfer fluid, followed by an adiabatic compressor, then a combustion chamber 15.48 under constant pressure, supplied by working gas, oxygen and hydrogen pipes, the hydrogen pipe being common with that of the working gas 15.53, from an expander E1 , which discharges into a liquid separator 15.49 with vertical plates, for the separation of water vapor resulting from the combustion of hydrogen, and from an expander E2, which adiabatically reduces the pressure to the pressure in the expander, a tank 15.51 with liquid hydrogen, a tank 15.52 with liquid oxygen, two thermal engines, in which the working gas is helium, both having the rarefier mounted in the tank 15.50, and the expander in the hydrogen tank, respectively in the oxygen one , an liquid- atmospheric air heat exchanger with forced circulation HE that maintains a constant temperature in the tank 15.50 and devices for regulating gas flows, for regulating isothermal speeds, with the possibility of changing the average temperature of the cycle, as well as with devices for regulating the speed of the adiabatic devices
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