WO1998014738A1 - Centrifugal heat transfer engine and system - Google Patents
Centrifugal heat transfer engine and system Download PDFInfo
- Publication number
- WO1998014738A1 WO1998014738A1 PCT/US1997/017482 US9717482W WO9814738A1 WO 1998014738 A1 WO1998014738 A1 WO 1998014738A1 US 9717482 W US9717482 W US 9717482W WO 9814738 A1 WO9814738 A1 WO 9814738A1
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- WO
- WIPO (PCT)
- Prior art keywords
- heat transfer
- heat
- rotatable
- rotor
- heat exchanging
- Prior art date
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Classifications
-
- F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
- F25—REFRIGERATION OR COOLING; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT PUMP SYSTEMS; MANUFACTURE OR STORAGE OF ICE; LIQUEFACTION SOLIDIFICATION OF GASES
- F25B—REFRIGERATION MACHINES, PLANTS OR SYSTEMS; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT PUMP SYSTEMS
- F25B3/00—Self-contained rotary compression machines, i.e. with compressor, condenser and evaporator rotating as a single unit
Definitions
- the present invention relates to a method and apparatus for transferring heat within diverse user environments, using centrifugal forces to realize the evaporator and condenser functions required in a vapor-compression type heat transfer cycle.
- air conditioning systems in commercial operation use the reversible heat transfer cycle, described above.
- air conditioning systems transfer heat from one environment (i.e., an indoor room) to another environment (i.e., the outdoors) by cyclically transforming the state of a refrigerant (i.e. working fluid) while it is being circulated throughout the system.
- a refrigerant i.e. working fluid
- the state transformation of the refrigerant is carried out in accordance with a vapor-compression refrigeration cycle, which is an instance of the more generally known "reversible adiabatic heat transfer cycle" .
- the refrigerant in its saturated vapor state enters a compressor and undergoes a reversible adiabatic compression.
- the refrigerant then enters a condenser, wherein heat is liberated to its environment causing the refrigerant to transform into its saturated liquid state while being maintained at a substantially constant pressure.
- the refrigerant passes through a throttling (i.e. metering) device, wherein the refrigerant undergoes adiabatic throttling.
- the refrigerant enters the evaporator and absorbs heat from its environment, causing the refrigerant to transform into its vapor state while being maintained at a substantially constant pressure. Consequently, as a liquid or gas, such as air, is passed over the evaporator during the evaporation process, the air is cooled.
- a liquid or gas such as air
- the vapor- compression refrigeration cycle deviates from the ideal cycle described above due primarily to the pressure drops associated with refrigeration flow and heat transfer to or from the ambient surroundings.
- a number of working fluids can be used with the vapor-compression refrigeration cycle described above.
- Ammonia and sulfur dioxide were important refrigerants in the early days of vapor-compression refrigeration.
- azeotropic refrigerants such as R-500 and R-502, are more commonly used.
- Halocarbon refrigerants originate from hydrocarbons and include ethane, propane, butane, methane, and others. While it is a common practice to blend together three or more halogenated hydrocarbon refrigerants such as R-22, R125, and R-290, near-azeotropic blend refrigerants suffer from temperature drift.
- Hydrocarbon based fluids containing hydrogen and carbon are generally flammable and therefore are poorly suited for use as refrigerants. While halogenated hydrocarbons are nonflammable, they do contain chlorine, fluorine, and bromine, and thus are hazardous to human health.
- the main refrigerants in use are the halogenated hydrocarbons, e.g., dichlorodifluoromethane (CCL2F2) , commonly known as R-12 refrigerant.
- CCL2F2 dichlorodifluoromethane
- R-12 refrigerant R-12 refrigerant.
- chlorofluorocarbons, (CFCs) CFCs
- hydrochlorofluorocarbons hydrochlorofluorocarbons
- Hydrofluorocarbons, (HFCs) contain hydrogen, fluorine, and carbon.
- one end of the tube assembly functions as a condenser, while the other end thereof functions as an evaporator.
- means are provided for directing separate streams of gas or liquid across the condenser and evaporator assemblies for effecting heat transfer operations with the ambient environment.
- a primary object of the present invention to provide an improved method and apparatus for transferring heat within diverse user environments using centrifugal forces to realize the evaporator and condenser functions required in a vapor-compression type heat transfer cycle, while avoiding the shortcomings and drawbacks of prior art apparatus and methodologies.
- a further object of the present invention is to provide such apparatus in the form of a centrifugal heat transfer engine which, by eliminating the use of mechanical compressors, reduces the introduction of heat into the system by the internal moving parts of conventional motor driven compressors, and energy losses caused by refrigeration lubricants used to lubricate the moving parts thereof.
- a further object of the present invention is to provide a centrifugal heat transfer engine that contains the refrigerant within a closed system in order to avoid leakage, yet being operable with a wide range of refrigerants.
- a further object of the present invention is to provide a centrifugal heat transfer engine having a rotor structure with a closed, fluid circulating system that contributes to a dynamic balance of refrigerant flow.
- a further object of the present invention is to provide a centrifugal heat transfer engine having a rotor structure embodying a fluid circulation system which, when rotated direction in a first direction, has a first portion that functions as a condenser and a second portion that functions as an evaporator to provide a refrigeration unit, and when the direction of the rotor structure is reversed, the first portion functions as an evaporator and the second portion functions as a condenser to provide a heating unit.
- a further object of the present invention is to provide a centrifugal heat transfer engine that either condenses or evaporates a chemical refrigerant as it is passed through a plurality of helical passageways which are part of its rotor structure .
- a further object of the present invention is to provide a centrifugal heat transfer engine which provides a simple apparatus for carrying out a refrigeration cycle without the necessity for compressors or other internal moving parts that introduce unnecessary heat into the refrigerant.
- a further object of the present invention is to provide a centrifugal heat transfer engine which does not require refrigerant contamination with an internal lubricant, and thus permits the refrigerant to function at optimum heat transferring quality.
- a further object of the present invention is to provide a centrifugal heat transfer engine having a temperature responsive torque-controlling system in order to maintain the angular velocity of the rotor structure within prespecified operating range, and thus maintain the flow of refrigerant through the fluid circulating system of the rotor structure.
- a further object of the present invention is to provide such a centrifugal heat transfer engine with a rotatable structure containing the self-circulating fluid circuit having a bidirectional throttling device placed between the condenser section and the evaporator section of the fluid circuit.
- a further object of the present invention is to provide such a bidirectional throttling device for controlling the flow rate of liquid refrigerant into the evaporization length of the evaporator section of the rotor structure, and the amount of pressure drop between the liquid pressurization length and the evaporization length during a range of axial velocities (RPM) of the rotor structure.
- RPM axial velocities
- a further object of the present invention is to provide such a centrifugal heat transfer engine, in which the optimum axial velocity is arrived at and controlled by a torque controlling system responsive to temperature changes detected in the ambient air or liquid being treated using an array of temperature sensors.
- a further object of the present invention is to provide such a centrifugal heat transfer engine with a spiral passage along the shaft of the rotor structure in order to cause vapor-compression as it draws the heavy refrigerant vapor from the evaporator to the condenser in both clockwise and counterclockwise directions of rotation.
- a further object of the present invention is to provide such a centrifugal heat transfer engine with a rotor structure having heat transfer fins in order to enhance heat transfer between the circulating refrigerant and the ambient environment during the operation of the engine.
- a further object of the present invention is to provide such a centrifugal heat transfer engine, in which the closed refrigerant flow circuit within the rotor structure is realized as spiraled tubing assembly having spiraled tubular sections which are both held in position by structural supports anchored to the shaft and connected to spiraled tubes.
- a further object of the present invention is to provide such a centrifugal heat transfer engine, in which the rotor structure is constructed as a solid assembly and the closed refrigerant flow circuit, including its spiral return passageway along the axis of rotation, is formed therein.
- Another object of the present invention is to provide a novel heat transfer engine which can be used to transfer heat within a building, home, automobile, tractor-trailer, aircraft, freight train, maritime vessel, or the like, order to maintain one or more temperature control functions.
- An even further object of the present invention is to provide a novel heat transfer engine, wherein heat can be transferred without the use of a vapor-compression cycle.
- the apparatus of the present invention is provided in the form of a reversible heat transfer engine.
- the heat transfer engine comprises a stator, port connectors, a heat exchanging rotor, torque generator, temperature selector, a plurality of temperature sensors, a fluid flow rate controller, and a system controller.
- the stator housing has primary and secondary heat transfer chambers, and a thermal isolation barrier disposed therebetween.
- the primary and secondary heat transfer chambers each have inlet and outlet ports and a continuous passageway therebetween.
- a first port connector is provided for interconnecting a primary heat exchanging circuit to the heat ports of the primary heat transfer chamber, so as to permit a primary heat exchanging medium to flow through the primary heat exchanging circuit and the primary heat exchanging chamber during the operation of the heat transfer engine.
- a second port connector is provided for interconnecting a secondary heat exchanging circuit to the inlet and outlet ports of said secondary heat transfer chamber, so as to permit a secondary heat exchanging medium to flow through the secondary heat exchanging circuit and the secondary heat transfer chamber during the operation of the reversible heat transfer engine, while the primary and secondary heat exchanging circuits are in substantial thermal isolation of each other.
- the heat exchanging rotor is rotatably supported within the stator housing about an axis of rotation and having a substantially symmetrical moment of inertia about the axis of rotation.
- the heat exchanging rotor has a primary heat exchanging end portion disposed within the primary heat transfer chamber, a secondary heat exchanging end portion disposed within the secondary heat transfer chamber, and an intermediate portion disposed between the primary and secondary heat exchanging end portions.
- the heat exchanging rotor contains a closed fluid circuit symmetrically arranged about the axis of rotation and has a return portion extending along the direction of the axis of rotation.
- the primary heat exchanging end portion of the rotor is disposed in thermal communication with the primary heat exchanging circuit, and the secondary heat exchanging end portion of the rotor is disposed in thermal communication with the secondary heat exchanging circuit.
- the intermediate portion of the rotor is physically adjacent the thermal isolation barrier so as to present a substantially high thermal resistance to heat transfer between the primary and secondary heat exchanging chambers during operation of the heat transfer engine.
- a predetermined amount of a heat carrying medium is contained within the closed fluid circuit of the heat exchanging rotor.
- the heat carrying medium is characterized by a predetermined heat of evaporation at which the heat carrying medium transforms from liquid phase to vapor phase, and a predetermined heat of condensation at which the heat carrying medium transforms from vapor phase to liquid phase.
- the direction of phase change of the heat carrying liquid is reversible.
- the function of the torque generator is to impart torque to the heat exchanging rotor and cause the heat exchanging rotor to rotate about the axis of rotation.
- the function of - li the temperature selector is to select a temperature to be maintained along the primary heat exchanging circuit.
- the function of the temperature sensor is to measure the temperature of the primary heat exchanging medium flowing through the inlet and outlet ports of the primary heat exchanging chamber, and for measuring the temperature of the secondary heat exchanging medium flowing through the inlet and outlet ports of the primary heat exchanging chamber.
- the function of the fluid flow rate controller is to control the flow rate of the primary heat exchanging medium flowing through the primary heat exchanging chamber and the flow rate of the secondary heat exchanging medium flowing through the secondary heat exchanging chamber, in response to the sensed temperature of the heat exchanging medium at either the inlet or outlet port in either the primary or secondary heat exchanging chambers and to satisfy the temperature selector setting.
- the function of the torque controller is to control the torque generating means in response to the sensed temperature of the heat exchanging medium at either the inlet or outlet port in either the primary or secondary heat exchanging chambers and the selected operating temperature setting.
- Fig. 1 is a schematic representation of the first illustrative embodiment of the heat transfer engine of the present invention, showing the fluid-carrying rotor structure thereof being rotated about its shaft by a torque generator controlled by a system controller responsive to the temperatures measured from a plurality of locations about the system;
- FIG. 2A an elevated side view of the fluid-carrying rotor structure of the first illustrative embodiment of Fig. 1, shown removed from the stator portion thereof, and with indications depicting which fluid carrying tube sections carry out the condenser and evaporator functions respectively, when the rotor structure is rotated in the direction shown;
- Fig. 2B a top view of the fluid-carrying rotor structure of the first illustrative embodiment of the Fig. 1, shown removed from the stator portion thereof, with indications depicting the location of the throttling device and rotor shaft coil penetrations;
- Fig. 3 an elevated side view of the fluid-carrying rotor structure of the first illustrative embodiment of Fig. 1, shown removed from the stator portion thereof, with indications depicting which fluid carrying tube sections carry out the condenser and evaporator functions, respectively, when the rotor structure is rotated in the direction shown;
- Fig. 4A is an elevated side view of the rotatable support shaft of the rotor structure of the first illustrative embodiment of Figs. 1 and 2, showing the spiraled passageway extending therealong and shaft end bearing surfaces machined in the shaft core material ;
- Fig. 4B is an elevated cross-sectional side view of the rotatable support shaft of Fig. 4A, shown inserted into its shaft cover sleeve and welded thereto with a bead of weld formed around the circumference thereof;
- Fig. 5 is an elevated cross-sectional longitudinal view of the rotatable support shaft of the rotor structure of the first illustrative embodiment of Fig. 1;
- Figs. 6A and 6B are cross-sectional views of the rotatable support shaft of the rotor structure of the first illustrative embodiment taken along lines 6A-6A and 6B-6B, respectively, of Fig. 5, showing the manner in which the end portions of the spiral coil structure are connected to the spiraled passage formed along the rotatable support shaft of the rotor structure of the first illustrative present invention;
- Fig. 7A is a first elevated side view of a support element used to support a section of the fluid-carrying spiraled tube portion of the rotor structure of the first illustrative embodiment of the present invention
- Fig. 7B is a second elevated side view of the support element shown in Fig. 7A;
- Fig. 7C is an elevated axial view of one spiral turn of the fluid-carrying spiraled tube portion of the rotor structure of the first illustrative embodiment of the present invention shown in Fig. 1;
- Fig. 8A is schematic representation of the heat transfer engine of the first illustrative embodiment of the present invention installed within a heat transfer system, wherein the primary and secondary heat exchanging chambers of the stator are operably connected to the primary and secondary heat exchanging circuits of the system, respectively, so that the primary and secondary heat transferring portions of the rotor structure are in thermal communication with the same while the heat transfer engine is operated in its cooling mode;
- Fig. 8B is schematic representation of the heat transfer engine of the first illustrative embodiment of the present invention installed within a heat transfer system, wherein the primary and secondary heat exchanging chambers of the stator are operably connected to the primary and secondary heat exchanging circuits of the system, respectively, so that the primary and secondary heat transferring portions of the rotor structure are in thermal communication with the same while the heat transfer engine is operated in its heating mode;
- Fig. 9 is a graphical representation of the closed-loop operating characteristic of the heat transfer engine of the present invention (i.e., with the primary and secondary heat exchanging portions of the rotor in thermal communication with primary and secondary heat exchanging circuits of a heat transfer system) , showing the ideal rate of heat exchange from the primary portion of the rotor to the secondary portion thereof, as a function of angular velocity of the rotor about its axis of rotation;
- Figs. 10A, 10B and IOC collectively, show a flow chart illustrating the steps of the control process carried out by the temperature-responsive system controller of the heat transfer engine of the present invention, operated in either its cooling or heating mode;
- Fig. 11A is a schematic representation of the rotor structure of the heat transfer engine of Fig. 1, showing the physical location of the liquid and gaseous phases of refrigerant within the rotor structure thereof when the heat transfer engine is at rest prior to entering the cooling mode;
- Fig. 11B is a schematic representation of the rotor structure of the heat transfer engine of Fig. l, showing the physical location of the liquid, gaseous and vapor phases of refrigerant within the rotor structure thereof during the first few revolutions thereof during the first stages of start up operation in its cooling mode;
- Fig. 11C is a schematic representation of the rotor structure of the heat transfer engine of Fig. 1, showing the physical location of the liquid, homogeneous fluid, vapor and gaseous phases of refrigerant within the rotor structure thereof during the second stage of start up operation in its cooling mode;
- Fig. 11D is a schematic representation of the rotor structure of the heat transfer engine of Fig. 1, showing the physical location of the liquid, homogeneous fluid, vapor and gaseous phases of refrigerant within the rotor structure thereof when vapor compression begins within the centrifugal heat transfer engine during the third stage of start up operation in its cooling mode;
- Fig. HE is a schematic representation of the rotor structure of the heat transfer engine of Fig. 1, showing the physical location of the liquid, homogeneous fluid, vapor and gaseous phases of refrigerant within the rotor structure thereof during the fourth stage of start-up operation in its cooling mode;
- Fig. 11F is a schematic representation of the rotor structure of the heat transfer engine of Fig. 1, showing the physical location of the liquid, homogeneous fluid, vapor and gaseous phases of refrigerant within the of the rotor structure of the heat transfer engine of Fig. 1 rotor structure thereof as vapor compression occurs during the fifth stage of start-up operation in its cooling mode;
- Fig. 11G is a schematic representation of the rotor structure of the heat transfer engine of Fig. 1, showing the physical location of the liquid, homogeneous fluid, vapor and gaseous phases of refrigerant within the rotor structure as superdeheating and condensation begin during the sixth stage of start-up operation in its cooling mode;
- Fig. 11F is a schematic representation of the rotor structure of the heat transfer engine of Fig. 1, showing the physical location of the liquid, homogeneous fluid, vapor and gaseous phases of refrigerant within the rotor structure as superdeheating and condensation begin during the sixth stage of start-up operation in
- HH is a schematic representation of the rotor structure of the heat transfer engine of Fig. 1, showing the physical location of the liquid, homogeneous fluid, vapor and gaseous phases of refrigerant within the rotor structure thereof during the seventh stage of start up operation in its cooling mode;
- Fig. HI is a schematic representation of the rotor structure of the heat transfer engine of Fig. 1, showing the physical location of the liquid, homogeneous fluid, vapor and gaseous phases of refrigerant within the rotor structure during the eight (i.e., steady-state) stage of operation in its cooling mode;
- Fig. 12A is a schematic representation of the rotor structure of the heat transfer engine of Fig. 1, showing the physical location of the liquid and gaseous phases of refrigerant within the rotor structure thereof when the centrifugal heat transfer engine is at rest prior to entering its heating mode;
- Fig. 12B is a schematic representation of the rotor structure of the heat transfer engine of Fig. 1, showing the physical location of the liquid, gaseous and vapor phases of refrigerant within the rotor structure thereof during the first few revolutions thereof during the first stages of start up operation in its heating mode;
- Fig. 12C is a schematic representation of the rotor structure of the heat transfer engine of Fig. 1, showing the physical location of the liquid, homogeneous fluid, vapor and gaseous phases of refrigerant within the rotor structure thereof during the second stage of start up operation in its heating mode;
- Fig. 12D is a schematic representation of the rotor structure of the heat transfer engine of Fig. 1, showing the physical location of the liquid, homogeneous fluid, vapor and gaseous phases of refrigerant within the rotor structure when vapor compression begins within the centrifugal heat transfer engine during the third stage of start up operation in the heating mode;
- Fig. 12E is a schematic representation of the rotor structure of the heat transfer engine of Fig. 1, showing the physical location of the liquid, homogeneous fluid, vapor and gaseous phases of refrigerant within the rotor structure thereof during the fourth stage of start-up operation in its heating mode;
- Fig. 12F is a schematic representation of the rotor structure of the heat transfer engine of Fig. 1, showing the physical location of the liquid, homogeneous fluid, vapor and gaseous phases of refrigerant within the rotor structure thereof as vapor compression occurs during the fifth stage of start-up operation in its heating mode
- Fig. 12G is a schematic representation of the rotor structure of the heat transfer engine of Fig. 1, showing the physical location of the liquid, homogeneous fluid, vapor and gaseous phases of refrigerant within the rotor structure as superdeheating and condensation begin during the sixth stage of start-up operation in its heating mode;
- Fig. 12H is a schematic representation of the rotor structure of the heat transfer engine of Fig. 1, showing the physical location of the liquid, homogeneous fluid, vapor and gaseous phases of refrigerant within the rotor structure thereof during the seventh stage of start up operation in the heating mode;
- Fig. 121 is a schematic representation of the rotor structure of the heat transfer engine of Fig. 1, showing the physical location of the liquid, homogeneous fluid, vapor and gaseous phases of refrigerant within the rotor structure thereof during the eight (i.e., steady-state) stage of operation in the heating mode;
- Fig. 13 is an elevated, partially cut-away view of a roof-mounted air-conditioning system, in which the centrifugal heat transfer engine of the first illustrative embodiment is integrated with conventional air return and supply ducts that extend into and out of structural components of a building;
- Fig. 14A is an elevated cross-sectional view of the centrifugal heat transfer engine of the second illustrative embodiment of the present invention, showing its fluid- carrying rotor structure rotatably supported in a precasted stator housing having primary and secondary fluid input and outport ports connectable to primary and secondary heat exchanging circuits, respectively, so that heat exchanging fluid cyclically flowing therethrough passes over a multiplicity of turbine blades affixed to the rotor structure and imparts torque thereto in order to maintain the angular velocity thereof in accordance with its temperature- responsive controller;
- Fig. 14B is an elevated end view of the centrifugal heat transfer engine of Fig. 14A, showing flanged fluid conduit connections for connection to primary and secondary heat exchanging circuits;
- Fig. 15A is an elevated transparent side view of the rotor structure of the heat transfer engine shown in Figs. 14A and 14B, removed from its stator housing, showing spiraled geometric similarities between the primary and secondary heat transfer portions of the heat transfer engine of first illustrative embodiment shown in Fig. 1 and the primary and secondary heat transfer portions of the heat transfer engine of the second illustrative embodiment shown in Fig. 14A and 14B;
- Fig. 15B is an elevated exploded view of the fluid- circulating rotor structure of the second illustrative embodiment shown in Figs. 14A and 14B, removed from its stator housing, showing how the precasted rotor disc structures are joined together to provide an integral structure within which a self-circulating closed fluid circuit is formed and how the suction shaft screw and throttling device orifice are inserted into the rotor shaft assembly;
- Fig. 15C is an elevated side view of the spiraled suction screw and throttling device orifice of the rotor structure of the heat transfer engine of the second illustrative embodiment
- Fig. 15D is a side view of the threaded port cap and gasket being fitted on the charging end of the rotor structure of the heat transfer engine of the second illustrative embodiment of the present invention
- Fig. 15E is an elevated end view of a vaned rotor disk of the second illustrative embodiment, showing a spiraled portion of the fluid carrying circuit formed therein and the turbine vane slots machined in the surfaces thereof;
- Fig. 15F are two elevated views of a turbine vane of the heat transfer engine of the second illustrative embodiment, showing the vane base and illustrating a possible blade surface conFiguration;
- Fig. 15G is an elevated side view of a vaned rotor disc of the rotor of the heat transfer engine of Figs. 14A and 14B, showing its turbine vanes, and a machined fluid passageway portion formed in the rotor structure thereof;
- Fig. 15H is an elevated end view of the first end rotor disk of the secondary heat transfer portion of the rotor shown in Fig. 15B, showing its spiraled portion of the fluid carrying circuit formed therein;
- Fig. 151 is an elevated, side view of the first rotor end disc of the secondary heat transfer portion of the rotor shown in Fig. 15B;
- Fig. 15J is an elevated end view of the first rotor end disc of the primary heat transfer portion of the rotor of Fig. 15B, showing its spiraled portion of the fluid carrying circuit formed therein;
- Fig. 15K is an elevated side view of the first rotor end disc of the primary heat transfer portion of the rotor of Fig. 15B, showing its spiraled portion of the fluid carrying circuit formed therein;
- Fig. 15L is an elevated transparent side view of the fluid-carrying rotor structure of the second illustrative embodiment of the heat transfer engine hereof, shown removed from the stator portion thereof with the closed fluid carrying circuit embedded within a heat conductive, solid- body rotor structure;
- Fig. 16A is a schematic representation of the rotor structure of the heat transfer engine of Figs. 14A and 14B, showing the physical location of the liquid and gaseous phases of refrigerant within the rotor structure thereof when the heat transfer engine hereof is at rest prior to entering its cooling mode;
- Fig. 16B is a schematic representation of the rotor structure of the heat transfer engine of Figs. 14A and 14B, showing the physical location of the liquid, gaseous and vapor phases of refrigerant within the rotor structure during the first few revolutions thereof during the first stages of start up operation in the cooling mode;
- Fig. 16C is a schematic representation of the rotor structure of the heat transfer engine of Figs. 14A and 14B, showing the physical location of the liquid, homogeneous fluid, vapor and gaseous phases of refrigerant within the rotor structure during the second stage of start up operation in the cooling mode;
- Fig. 16D is a schematic representation of the rotor structure of the heat transfer engine of Figs. 14A and 14B, showing the physical location of the liquid, homogeneous fluid, vapor and gaseous phases of refrigerant within the rotor structure when vapor compression begins within the heat transfer engine during the third stage of start up operation in its cooling mode;
- Fig. 16E is a schematic representation of the rotor structure of the heat transfer engine of Figs. 14A and 14B, showing the physical location of the liquid, homogeneous fluid, vapor and gaseous phases of refrigerant within the rotor structure during the fourth stage of start-up operation in its cooling mode
- Fig. 16F is a schematic representation of the rotor structure of the heat transfer engine of Figs. 14A and 14B, showing the physical location of the liquid, homogeneous fluid, vapor and gaseous phases of refrigerant within the rotor structure as vapor compression occurs during the fifth stage of start-up operation in its cooling mode;
- Fig. 16G is a schematic representation of the rotor structure of the heat transfer engine of Figs. 14A and 14B, showing the physical location of the liquid, homogeneous fluid, vapor and gaseous phases of refrigerant within the rotor structure as superdeheating and condensation begin during the sixth stage of start-up operation in its cooling mode;
- Fig. 16H is a schematic representation of the rotor structure of the heat transfer engine of Figs. 14A and 14B, showing the physical location of the liquid, homogeneous fluid, vapor and gaseous phases of refrigerant within the rotor structure during the seventh and steady-state state of start up operation in its cooling mode;
- Fig. 161 is a schematic representation of the rotor structure of the heat transfer engine of Figs. 14A and 14B, showing the physical location of the liquid, homogeneous fluid, vapor and gaseous phases of refrigerant within the rotor structure during the eighth state stage of operation, at an angular velocity exceeding steady-state, in its cooling mode;
- Fig. 17 is a schematic diagram of a heat transfer system, in which the heat transfer engine of the second illustrative embodiment is arranged so that the rotor structure thereof is rotated by fluid (water) flowing through the secondary heat exchanging fluid circuit, while the angular velocity thereof is controlled using a pump and flow control valve controlled by the temperature-responsive system controller;
- Fig. 18 is a schematic diagram of a heat transfer system, in which a turbine-based heat transfer engine of the present invention is arranged so that the rotor structure thereof is rotated by an electric motor in direct connection with the rotor, while water from a cooling tower is circulated through the primary heat exchanging circuit;
- Fig. 19 is a schematic diagram of a heat transfer system, in which the primary heat exchanging chamber of a first turbine-based centrifugal heat transfer engine hereof is connected to the secondary heat exchanging chamber of a second turbine-like heat transfer engine hereof, whereas the primary heat transfer chamber of the secondary turbine-like heat transfer engine is in fluid communication with a cooling tower while the secondary heat exchanging chamber of the second turbine-like heat transfer engine is in fluid communication with fluid supply circuit;
- Fig. 19 is a schematic diagram of a heat transfer system, in which the primary heat exchanging chamber of a first turbine-based centrifugal heat transfer engine hereof is connected to the secondary heat exchanging chamber of a second turbine-like heat transfer engine hereof,
- 20 is a schematic diagram of a hybrid heat transfer engine, in which the primary heat transfer portion of the rotor is realized as coiled structure mounted on a common shaft and contained within a primary heat transfer chamber of the coiled heat transfer engine of the first illustrative embodiment, whereas the secondary heat transfer portion of the rotor is realized as a turbine-like finned structure mounted on the common shaft and contained with a secondary heat transfer chamber of the turbine-like heat transfer engine of the second illustrative embodiment, shown operated in its cooling mode;
- Fig. 21 is schematic diagram of the hybrid heat transfer engine of Fig. 20, wherein the primary heat transfer portion thereof functions as an air or gas conditioning evaporator while the secondary heat transfer portion functions as a condenser in an open loop fluid cooled condenser, driven by an electric motor connected directly to the rotor shaft by way of a magnetic torque converter;
- Fig. 22 is a schematic diagram of a heat transfer system of the present invention embodied within an automobile, wherein the rotor of the heat transfer engine is rotated by an electric motor driven by electrical power supplied through a power control circuit, and produced by the automobile battery recharged by an alternator within the engine compartment;
- Fig. 23 is a schematic diagram of a heat transfer system of the present invention embodied within an refrigerated tractor trailer truck, wherein the rotor of the heat transfer engine is rotated by an electric motor driven by electrical power supplied through a power control circuit and produced by a bank of batteries recharged by an alternator within the engine compartment;
- Fig. 24 is a schematic diagram of a heat transfer system of the present invention embodied within an aircraft equipped with a plurality of heat transfer engines of the present invention, wherein the rotor of each heat transfer engine is rotated by an electric motor driven by electrical power supplied through voltage regulator and temperature control circuit, and produced by an onboard electric generator;
- Fig. 25 is a schematic diagram of a heat transfer system of the present invention embodied within a refrigerated freight train equipped with a plurality of heat transfer engines of the present invention, wherein the rotor of each heat transfer engine is rotated by an electric motor driven by electrical power supplied through voltage regulator and temperature control circuit, and produced by an onboard pneumatically driven electric generator;
- Fig. 26 is a schematic diagram of a heat transfer system of the present invention embodied within a refrigerated shipping vessel equipped with a plurality of heat transfer engines of the present invention, wherein the rotor of each heat transfer engine is rotated by an electric motor driven by electrical power supplied through voltage regulator and temperature control circuit, and produced by an onboard pneumatically driven electric generator;
- Fig. 27A is an elevated side view of the fluid carrying rotor structure of an alternative of the heat transfer engine of the present invention, shown removed from the stator portion thereof, wherein the spiraled return passageway is shown extending outside the rotor shaft, along the direction of rotor rotation; and
- Fig. 27B is an elevated side view of the rotatable support shaft of the rotor structure of Fig. 27A, showing that no portion of the fluid carrying circuit is machined or embodied in the support shaft of this embodiment of the rotor structure.
- a first illustrative embodiment of the centrifugal heat transfer engine is shown.
- this embodiment of the heat transfer engine comprises a rotatable structure (i.e., "rotor") realized as a spiral coiled tubing assembly, that is rotatably supported by a stationary structure (“stator”) .
- rotor rotatable structure
- stator stationary structure
- coiled centrifugal heat transfer engine this embodiment of the heat transfer engine shall be referred to as the coiled centrifugal heat transfer engine.
- reversible centrifugal heat transfer engine 1 comprises a number of major system components, namely: a stator housing 2; primary port connection assembly 3 ; secondary port connection assembly 4 ; heat-exchanging rotor 5; a heat carrying medium 6; torque generator 7 ; temperature selection unit 9 ; temperature sensors 9A through 9D; primary and secondary fluid flow rate controllers 10A and 10B; and temperature-responsive system controller 11.
- a stator housing 2 primary port connection assembly 3
- secondary port connection assembly 4 heat-exchanging rotor 5
- heat carrying medium 6 heat carrying medium 6
- torque generator 7 torque generator 7
- temperature selection unit 9 temperature sensors 9A through 9D
- primary and secondary fluid flow rate controllers 10A and 10B primary and secondary fluid flow rate controllers
- the stator housing comprises primary and secondary heat transfer chambers 13 and 14, and a thermal isolation barrier 15 disposed therebetween.
- the primary heat transfer chamber shall hereinafter and in the claims shall indicate the environment within which the temperature of a fluid (i.e. gas or liquid) contained therein is to be maintained by way of operation of the heat transfer engine hereof.
- Primary heat transfer chamber 13 has inlet and outlet ports 16A and 16B, and secondary heat transfer chamber 14 has inlet and outlet ports 16C and 16D.
- Primary port connection assembly 3 is provided for interconnecting a primary heat exchanging circuit 20 (e.g., ductwork) to the inlet and outlet ports of the primary heat transfer chamber, so as to permit a primary heat exchanging medium 21, such as air or water, to flow through the primary heat exchanging circuit and the primary heat exchanging chamber during the operation of the heat transfer engine, while the primary and secondary heat exchanging circuits are in substantial thermal isolation of each other.
- a primary heat exchanging circuit 20 e.g., ductwork
- secondary port connection assembly 4 is provided for interconnecting a secondary heat exchanging circuit 22 to the inlet and outlet ports of the secondary heat transfer chamber, so as to permit a secondary heat exchanging medium 23 to flow through the secondary heat exchanging circuit and the secondary heat transfer chamber during the operation of the heat transfer engine, while the primary and secondary heat exchanging circuits are in substantial thermal isolation of each other.
- heat exchanging rotor 5 is rotatably supported within the stator housing 2 about an axis of rotation 25 and has a substantially symmetrical moment of inertia about the axis of rotation.
- the heat exchanging rotor has a primary heat exchanging end portion 2A disposed within the primary heat transfer chamber 13, a secondary heat exchanging end portion 2B disposed within the secondary heat transfer chamber 14, and an intermediate portion 2C disposed between the primary and secondary heat exchanging end portions 2A and 2B.
- the heat exchanging rotor 5 contains a closed fluid circuit 32 symmetrically arranged about the axis of rotation and has a return portion 26A extending along the direction of the axis of rotation.
- the primary heat exchanging end portion 2A of the rotor is disposed in thermal communication with the primary heat exchanging circuit 20, whereas the secondary heat exchanging end portion 2B of the rotor is disposed in thermal communication with the secondary heat exchanging circuit 22.
- the intermediate portion 2C thereof is physically adjacent the thermal isolation barrier 15.
- stator structure 2 is realized as a pair of rotor support elements 27A and 27B mounted upon a support platform 28 in a spaced apart manner.
- a predetermined amount of a heat carrying medium 6, such as refrigerant, is contained within the closed fluid circuit 32 and 26A of the rotor.
- the heat carrying medium is characterized by three basic thermodynamic properties: (i) its predetermined heat of evaporation at which the heat carrying medium transforms from liquid phase to vapor phase; and (ii) its predetermined heat of condensation at which the heat carrying medium transforms from vapor phase to liquid phase; and (iii) direction reversibility of phase change of the heat carrying liquid.
- suitable refrigerants for use with the heat transfer engine hereof include fluid refrigerants having a liquid or gaseous state during applicable operating temperature and pressure ranges.
- a refrigerant When selecting a refrigerant, the following consideration should be made: compatibility between the refrigerant and materials used to construct the closed fluid flow passageway; chemical stability of the refrigerant under conditions of use; applicable safety codes (e.g., non-flammable refrigerants made be required) ; toxicity; cost factors; and availability.
- applicable safety codes e.g., non-flammable refrigerants made be required
- the double spiral-coil geometry with the spiral return path along the rotor central axis has been discovered to be the preferred geometry of the present invention.
- the double spiral coil geometry is shown embodied in a rotor structure of one form or another.
- the function of the torque generator 7 is to impart torque to the heat exchanging rotor 5 in order to rotate the same about its axis of rotation at a predetermined angular velocity.
- the torque generator may be realized a variety of ways using known technology. Electric, hydraulic and pneumatic motors are just a few types of torque generators that may be coupled to the rotor shaft 29 and be used to controllably impart torque thereto under the control of system controller 11.
- the function of the temperature selecting unit 9 is to select (i.e., set) a temperature which is to be maintained along at least a portion of the primary heat exchanging circuit 20.
- the temperature selecting unit 9 is realized by electronic circuitry having memory for storing a selected temperature value, and means for producing an electrical signal representative thereof.
- the temperature sensors 9A, 9B, 9C, and 9D located at inlet and outlet ports 16A, 16B, 16C and 16D may be realized using any state of the art temperature sensing technology.
- the function of such devices is to measure the temperature of the primary heat exchanging medium 21 flowing through the inlet and outlet ports of the primary heat exchanging chamber 13, and the secondary heat exchanging medium 23 flowing through the inlet and outlet ports of the secondary heat exchanging chamber 14, and produce electrical signals representative thereof for use by the system controller 11 as will be described in greater detail hereinafter.
- the function of the primary and secondary fluid flow rate controllers 10A and 10B is to control the rate of flow of primary and secondary heat exchanging fluid within the primary and secondary heat exchanging circuits, respectively.
- the function of the primary fluid flow rate controller 10A is to control the rate of heat flow between the primary heat exchanging portion of the rotor and the primary heat exchanging circuit passing through the primary heat exchanging chamber of the stator housing.
- the function of the secondary fluid flow rate controller 10B is to control the rate of heat flow between the secondary heat exchanging portion of the rotor and the secondary heat exchanging circuit passing through the secondary heat exchanging chamber of the stator housing.
- the fluid flow rate controllers are controlled by the temperature responsive system controller 11 of the engine.
- Primary and secondary fluid flow rate controller 10A and 10B may be realized in a variety of ways depending on the nature of the heat exchanging medium being circulated through primary and secondary heat exchanging chambers 13 and 14 as the rotor is rotatably supported within the stator. For example, when the primary heat exchanging medium is air ported from the environment in which the air temperature is to be maintained, then primary fluid flow controller 10A may be realized by an air flow control valve (e.g., damper), whose aperture dimensions are electromechanically controlled by electrical control signals produced by the system controller.
- an air flow control valve e.g., damper
- primary fluid flow controller may be realized by an water control flow valve, whose aperture dimensions are electromechanically controlled by electrical control signals produced by the system controller.
- the function of the primary fluid flow rate controller is to control the flow rate of the primary heat exchanging medium flowing through the primary heat exchanging chamber in response to the sensed temperature of the heat exchanging medium at either the inlet or outlet port in either the primary or secondary heat exchanging chambers, and the temperature selected by temperature selection unit. Greater details with regard to this aspect of the control process will be described hereafter.
- the secondary fluid flow rate controller 10B may be realized in a manner similar to the primary fluid flow rate controller 10A.
- the primary and secondary heat exchange fluids are different in physical state (e.g., the primary heat exchange fluid can be air, while the secondary heat exchange fluid is water, and vice versa) .
- the function of the secondary fluid flow rate controller is to control the flow rate of the secondary heat exchanging medium flowing through the secondary heat exchanging chamber, in response to the sensed temperature of the heat exchanging medium at either the inlet or outlet port in either the primary or secondary heat exchanging chambers and the temperature selected by temperature selection unit.
- the system controller 11 of the present invention has several other functions, namely: to read the temperature of the ambient operating environment measured by way of temperature sensors 9, 9A, 9B, 9C, and 9D; and in response thereto, generate suitable control signals which directly control the operation of torque generator 7 ; and indirectly control the angular velocity of the heat exchanging rotor, relative to the stator; and control the fluid flow rate of the primary and secondary heat exchanging fluids 21 and 23 flowing through the primary and secondary heat exchanging chambers 13 and 14, respectively.
- the need to control the angular velocity of the heat exchanging rotor, and the flow rates of the primary and secondary heat exchanging fluids will be described in detail hereinafter with reference to the thermodynamic refrigeration process of the present invention.
- the reversible heat transfer engine of the present invention has two modes of operation, namely: a heating mode which is realized when the heat exchanging rotor is rotated in a first predetermined direction of rotation; and a cooling mode which is realized when the rotor is rotated in a second predetermined direction of rotation.
- a heating mode which is realized when the heat exchanging rotor is rotated in a first predetermined direction of rotation
- a cooling mode which is realized when the rotor is rotated in a second predetermined direction of rotation.
- the enclosure i.e., stator
- heat exchanging rotor 5 of the first illustrative embodiment is realized as a length of tubing 32 symmetrically coiled around support shaft 29 extending along the axis of rotation of the rotor.
- the tubing assembly 36 and 37 has a double spiral-coil geometry, and the support shaft contains a spiral return passage 33 formed therethrough with an inlet opening 34 and an outlet opening 35.
- the spiral- coiled tubing assembly has a first spiral tubing portion 36, a second spiral tubing portion and bi-directional metering device 38 disposed therebetween.
- the ends of the first and second spiral tubing portions 36 and 37 are attached to both the inlet 52 and outlet 53 openings of the spiral return passage 33 along the rotor shaft and creates the closed fluid circulation circuit within the heat transfer structure.
- the function of the bi-directional metering device 38 is to control (1) the rate of flow of liquid refrigerant into the second spiral tubing portion 36 and (2) the amount of pressure drop between the secondary and primary tubing portions during a preselected range of rotor angular velocities (RPM) .
- RPM rotor angular velocities
- the optimum rotor angular velocity is arrived at and controlled by the system controller in response to temperature changes in the air or liquid being treated by the heat transfer engine of the present invention.
- the reason the throttling device 38 is bidirectional is to allow for refrigerant flow reversal when the direction of rotor rotation is reversed when switching from the cooling mode to the heating mode of the heat transfer engine.
- the rotor shaft 29 comprises a central shaft core 40 of solid construction enclosed within as cylindrical tube cover 41.
- a charging port 42 is provided along the end of the central tube in order to provide access to refrigerant inside the closed (i.e., sealed) self-circulating fluid circulation circuit (i.e., system) .
- central shaft core 40 has a spiraled passage 33 formed about the outer surface thereof, and is enclosed within tube cover 41, thereby creating a spiral shaped passageway 33 from one end of the rotor shaft to the other end thereof.
- Figs. 4A central shaft core 40 has a spiraled passage 33 formed about the outer surface thereof, and is enclosed within tube cover 41, thereby creating a spiral shaped passageway 33 from one end of the rotor shaft to the other end thereof.
- a pair of holes 44 are drilled through cylindrical tube cover 41 into the spiraled passageway 33 at the ends of the central shaft 29A and 29B. These holes allow the first and second end portions of double-coil tubing assembly to interconnect with the ends of the spiral rotor shaft, and thus form the closed fluid circulation circuit within the rotor structure.
- the rotor of the first illustrative embodiment also includes a plurality of tubing support brackets 45A, 45B, 45C and 45D for support of the spiraled tubular sections thereof in position about its central shaft.
- each of these tubing support brackets comprises shaft attachment means 45 extending from the rotor shaft 29, and tubing support element 46 for supporting a selected portion of the tubing assembly spiraled about the rotor shaft.
- These tubing support brackets may be made from any suitable material such as metal, composite material, or other functionally equivalent material.
- the tubing used to realize the rotor of the first illustrative embodiment may vary in inner diameter as the diameter of the tubing around the central shaft varies.
- the exterior surface of the rotor tubing is finned, while the internal surface thereof is rifled as this construction will improve the heat transfer function of the rotor.
- the heat transfer engine hereof is shown installed in an environment 50 through which the primary heat exchanging circuit 20 passes in order to control the temperature thereof while the engine in operated in its cooling mode. While the medium within this illustrative environment will typically be ambient air, it is understood that other mediums may be temperature maintained in different applications.
- the closed fluid flow circuit of rotor is arranged according to the first conFiguration.
- Fig. 10B the heat transfer engine hereof is shown installed in the same environment 50 shown in Fig. 10B, while the engine is operated in its heating mode.
- the closed fluid flow circuit of rotor is arranged once again according to the first rotor conFiguration.
- the direction of the rotor rotation is clockwise about the +z axis of the stator reference system when the engine is operated in its heating mode.
- the rotor 52 is realized as a solid body having first and second end portions 2A and 2B of truncated-cone like geometry, connected by a central cylindrical portion 2C extending about an axis of rotation.
- a closed fluid flow circuit 26 having essentially the same geometry as rotor 5 of the first illustrative embodiment is embodied (or embedded) within the solid rotor body. As such, this embodiment shall be referred to as the embedded rotor embodiment of the present invention.
- the closed fluid circuit of rotor 52 symmetrically extends about its rotor axis of rotation.
- bi-directional metering device 38 is realized within the central portion of the rotor body, as shown.
- one end of the rotor has an access port 95 and 96, (e.g., a removable screw cap) for introducing refrigerant into or removing refrigerant from the closed fluid flow circuit.
- the fluid flow circuit may be realized in the solid body of the rotor in a variety of ways. One way is to produce a solid rotor body in two symmetrical half sections using injection molding techniques, so that respective portions of the closed fluid flow circuit are integrally formed therein. Thereafter, the molded body halves can be joined together using appropriate gaskets, seals and fastening techniques. Advanced composite materials, including ceramics, may be used to construct the rotor body.
- the rotor may be realized by assembling a plurality of rotor discs, each embodying a portion of the closed fluid flow circuit. Details regarding this alternative embodiment will be described in greater detail hereinafter.
- the direction of rotation of the spiral tubing along the closed fluid flow circuit is essential.
- the portion of the fluid flow circuit along the rotor shaft i.e., the rotor axis
- the portion of the fluid flow circuit extending outside of the rotor shaft is bisected by the bi-directional metering device into a first outer fluid flow path portion and second outer fluid flow path portion .
- the end section of these outer fluid flow path portions away from the metering device connect with the end sections of the inner fluid flow path, to complete the closed fluid flow path within the heat exchanging rotor.
- the first outer fluid flow portion extends spirally about the +z axis in counter-clockwise (CCW) direction from the first end portion of the shaft to the metering device, and then continues to extend spirally about the +z axis in a counter-clockwise (CCW) from the metering device to the second end portion of the rotor shaft; and looking from the point of origin of the reference system down the +z axis, the inner fluid flow path extends spirally about the +z axis in a clockwise(CW) direction.
- the second possible conFiguration as shown in Figs.
- the first outer fluid flow portion extends spirally about the +z axis in a counter-clockwise (CCW) direction from the first end portion 26 of the shaft to the inlet of the fluid flow tube 84 as shown in FIG. 17A, and then continues to extend spirally about the +z axis in counter-clockwise (CCW) from the fluid flow tube device to the second end portion of the rotor shaft; looking from the point of origin of the reference system down the +z axis, the inner fluid flow path extends spirally about the +z axis in a counter-clockwise direction (CCW) .
- CCW counter-clockwise
- the direction of shaft rotation will be different for each heat transfer mode (e.g., cooling mode or heating mode) selected by the system user.
- the function of the throttling device of the present invention is to assist in the transformation of liquid refrigerant into vapor refrigerant without impacting the function of the rotor within the heat transfer engine hereof.
- this system component i.e., the metering device
- this fluid flow passageway has an inner cross-sectional area that is smaller than the smallest inner cross-sectional area of the evaporator section of the rotor.
- a properly designed metering device will operate in a bi-directional manner (i.e., in the cooling or heating mode of operation).
- the function of the metering device is to provide the necessary pressure drop between the condensor and evaporator functioning portions of the heat transfer engine hereof, and allow sufficient Superheat to be generated across the evaporator functioning portion of the rotor.
- the metering device should be designed to provide optimum fluid flow characteristics between the primary and secondary heat transfer portions of the rotor.
- the metering device can be easily realized by welding (or brazing) a section of hollow tubing between the primary and secondary heat exchanging portions, having an inner diameter smaller than the inner diameter of the primary and secondary heat exchanging portions.
- the ends of the small reduced diameter tubing section can be flared so that the inner diameter of this small tubing section are matched to the inner diameter of the tubing from which the primary and secondary heat exchanging portions are made.
- tubing of the primary and secondary heat exchanging portions can be continuously connected by welding or brazing process and that the metering device can be realized by crimping or stretching the tubing adjacent the connection, to achieve the necessary reduction in fluid flow passageway.
- the closed fluid passageway is realized within a solid-body rotor structure suitable for turbine type application where various types of fluid are used to input torque to the rotor during engine operation.
- the metering device can be easily realized by welding (or brazing) a section of hollow tubing between the primary and secondary heat exchanging portions, having an inner diameter smaller than the inner diameter of the primary and secondary heat exchanging portions, as shown in Fig. 18.
- a plurality of metering devices of the type described above can be used in parallel in order to achieve the necessary reduction in fluid flow passageway, and thus a sufficient pressure drop thereacross the primary and secondary heat exchanging portions of the rotor.
- the condenser functioning portion of the rotor would terminate in a first manifold-like structure, to which the individual metering devices would be attached at one end.
- the evaporator portion of the rotor would terminate in a second manifold-like structure, to which the individual metering devices would be attached at their other end.
- a reiterative design procedure is used to design and construct the metering device so that system performance specifications are satisfied by the operative engine construction. This design and construction procedure will be described below.
- the first step of the design method involves determining the system design parameters which include, for example: the Thermal Transfer Capacity of the system measured in BTUs/hour; Thermal Load on the system measured in BTU/hour; the physical dimensions of the rotor; and volume and type of refrigerant contained within the rotor (less than 80% of internal volume) .
- the second step involves specifying the design parameters for the metering device which, as described above, include primarily the smallest cross-sectional area of the fluid passageway between the first and second heat exchanging portion of the rotor. According to the method of the present invention, it is not necessarily to calculate the metering device design parameters using a thermodynamic or other type of mathematical model.
- an initial value for the metering device design parameters i.e., the smallest cross- sectional area of the fluid passageway
- the next step of the design method involves attaching infra-red temperature sensors to the inlet and outlet ports of the evaporator-functioning portion of the rotor, and then connecting these temperature sensors to an electronic (i.e., computer-based) recording instrument well known in the temperature instrumentation art.
- the heat transfer engine is operated under the specified thermal loading conditions for which it was designed.
- the primary design parameter i.e., smallest cross-sectional area
- T h i s condition is detected using the following design criteria.
- T eo is not greater than T e ⁇ by 6 degrees, then there is not enough Superheat being generated at the evaporator, or the angular velocity of the rotor is too low. If this condition exists, then the rotor angular velocity is increased to Wmax and recheck T e ⁇ and T e ⁇ . Then if T eo is not greater than T e ⁇ by 6 degrees, then the smallest cross- sectional area (e.g., diameter) through the metering device is too large and a reduction therein is needed. If this condition is detected, then the engine is stopped. The metering device is modified by reducing the cross-sectional area of the metering device by an incremental amount.
- the modified engine is then restarted and T e ⁇ and T eo remeasured to determine whether the amount of the Superheat produced across the evaporator is adequate. Thereafter, the reiterative design process of the present invention is repeated in the manner described above until the desired amount of Superheat is produced within the rotor of the production prototype under design.
- the design parameters of the metering device are carefully measured and recorded, and the metering device at which this operating condition is achieved is used to design and construct "production models" of the heat transfer engine.
- the design model of the heat transfer engine requires infra-red temperature sensors for Superheat monitoring purposes.
- Figs. 8A, 8B, and 10A to IOC the temperature-response control process of the present invention will be described for both the cooling and heating modes of the centrifugal heat transfer engine.
- a complex distribution of centrifugal forces are automatically generated and act upon the molecules of refrigerant contained within the closed circuit. This causes the refrigerant to automatically circulate within the closed circuit in a cyclical manner from the first end portion of the rotor, to the second end portion thereof, and then back to the first end portion along the spiral fluid flow path of the support shaft.
- the engine is operated in its cooling mode, and the spiral tubing section 36A of the rotor within the primary heat exchanging chamber functions as an evaporator while the spiral tubing section 37A within the secondary heat exchanging chamber functions as a condenser.
- the overall function of the rotor in the cooling mode is to transfer heat from the primary heat exchanging chamber to the secondary heat exchanging chamber under the control of the system controller.
- the refrigerant contained within the closed fluid circuit automatically circulates therewithin in a cyclical manner from the second end portion of the rotor, to the first end portion thereof, and then back to the second end portion along the spiral fluid flow path of the support shaft.
- the engine is operated in its heating mode, and the spiral tubing section of the rotor within the primary heat exchanging chamber 36A functions as a condenser, while the spiral tubing section 37A within the secondary heat exchanging chamber functions as an evaporator.
- the overall function of the rotor in the heating mode is to transfer heat from the secondary heat exchanging chamber to the primary heat exchanging chamber under the control of the system controller.
- the fluid velocity of the refrigerant within the rotor is functionally dependent upon a number of factors including, but not limited to, the angular velocity of the rotor relative to the stator, the thermal loading upon the first and second end portions of the rotor, and internal losses due to surface friction of the refrigerant within the closed fluid circuits.
- design factors such as the number of spiral coils, the heat transfer quality of materials used in their construction, the diameter of the spiral coils, the primary heat transfer surface area, the secondary heat transfer surface area, and the rotor angular velocity, and horsepower can be varied to alter the heat transfer capacity and efficiency of the centrifugal heat transfer engine.
- the heat exchanging rotor In order to cool the ambient environment (or fluid) to the selected temperature set by thermostat 9, the heat exchanging rotor must transfer, at a sufficient flow rate, heat from the primary heat exchanging chamber to the secondary heat exchanging chamber, from which it can then be liberated to the secondary heat exchanging circuit and thus maintain the selected temperature in a controlled manner. Similarly, to heat the ambient environment (or fluid) to the selected temperature set by the thermostat, the heat exchanging rotor must transfer, at a sufficient flow rate, heat from the secondary heat exchanging chamber to the primary heat exchanging chamber, from which it can then be liberated to the primary heat exchanging circuit and maintain the selected temperature in a controlled manner.
- each of the ports in the primary or secondary heat exchanging chambers of the heat transfer engine has installed within its flowpath, a temperature sensor 9A through 9D operably connected to the temperature-responsive system controller 11.
- the function of each of these port-located temperature sensors is to measure the temperature of the liquid flowing through its associated fluid inlet or outlet port as it passes over and/or through the end portions of the rotor.
- thermostat 9 or a like control device provides a means for setting a threshold or target temperature that is to be maintained within the primary heat exchanging chamber as the primary and secondary heat exchanging fluids are caused to circulate within the primary and secondary heat exchanging chambers, respectively.
- the primary function of the system controller is to manage the load-reduction operating characteristics of the heat transfer engine. In the illustrative embodiments, this is achieved by controlling (1) the angular velocity of the rotor within prespecified limits during system operation, and (2) the flow rate of the primary and secondary heat exchange fluids circulating through the primary and secondary heat exchange chambers of the engine, respectively.
- rotor-velocity and fluid flow-rate control is achieved by maintaining particular port-temperature constraints (i.e., conditions) on a real-time basis during the operation of the system in its designated mode of operation.
- these temperature constraints are expressed as difference equations which establish constraints (i.e., relations) among particular sensed temperature parameters.
- constraints i.e., relations
- the temperature control differential is AQ ,
- the temperature selector 9 is limited by the design capacity of the particular heat transfer engine at hand. As shown in Fig. 9, if the RPM 2) exceeds S) H , the refrigeration effect begins to decrease for one of two reasons: (1) the load has diminished to a point where no heat is available to be transferred in functional quantities; and (2) the weight of the liquid refrigerant in the liquid pressurization length by centrifugal forces exceeds pressurizing forces exerted on the refrigerant by the liquid pressurization lengths spiraled structure.
- Optimum operating conditions for the heat transfer engine are between S) L and S) H , and Q L and Q H . The intersections indicated are dictated by thermal capacity, refrigerant type and volume, and application, and are located by operational calibration.
- these temperature constraints of the system control process are maintained by the system controller during cooling or heating modes, respectively. These temperature constraints depend on the ambient reference temperature Tl set by thermostat 9 , and the temperatures sensed at each port of the first and secondary heat exchanging circuits of the system. The process by which the system controller controls the rotor velocity and fluid flow rates in the primary and secondary heat exchanging chambers will be described in detail below.
- the system control program of the illustrative embodiment is shown in the form of a computer flow diagram.
- the system controller executes the control program in a cyclical manner in order to automatically control the rotor velocity and fluid flow rates within prespecified operating conditions, while achieving the desired degree of temperature control along the primary heat exchanging circuit.
- the plurality of data storage registers associated with the system controller 11 are periodically read by its microprocessor. Each of these data storage registers is periodically (e.g., 10 times per second) provided with a new digital word produced from its respective A/D converter associated with the temperature sensor (9A, 9B, 9C, 9D) measuring the sensed temperature value.
- the data storage registers associated with the system controller are updated with current temperature values measured at the input and output ports of the primary and secondary heat exchanging chambers of the system.
- the primary and secondary fluid flow rate controllers are controlled to allow fluid flow rates up to about 10 percent (10%) of the maximum flow rate.
- the angular velocity of the rotor is controlled by the microprocessor performing the following rotor-velocity control operations represented by the following rules: if ⁇ r ⁇ ⁇ a - ⁇ t ⁇ 2 "F , then
- an increase in the rate of primary heat exchanging fluid through the primary heat exchanging chamber affects the refrigeration cycle by increasing the rate and amount of heat flowing from the primary heat transfer portion of the rotor to the secondary heat transfer portion thereof, as illustrated by the heat transfer loop in Fig. ⁇ A.
- PFR heat exchange fluid flow
- the microprocessor After performing the operations at Blocks E, F and G, the microprocessor reads once again the temperature values in its temperature value storage registers, and then at Block J determines whether there has been any change in mode (e.g., switch from the cooling mode to the heating mode). If no change in mode has been detected at Block J, then the microprocessor reenters the control loop defined by Blocks E through H and performs the operations specified therein to control the angular velocity of the rotor ⁇ and the flow rates of the primary and secondary fluid flow-rate controllers, PFR and SFR
- Block J in Fig. 10B the microprocessor determines whether the mode of the heat transfer engine has been changed (e.g. , from the cooling mode to the heating mode) then the microprocessor returns to Block C in Fig. 10A and then proceeds to Block K.
- the microprocessor controls the torque generator (e.g., motor) so that the rotor is rotated in the CW direction up to about 10% of the maximum design velocity 2) H , while the primary and secondary fluid flow rate controllers are controlled to allow fluid flow rates up to about 10 percent (10%) of the maximum flow rate.
- the torque generator e.g., motor
- an increase in the rate of secondary heat exchanging fluid through the secondary heat exchanging chamber affects the refrigeration cycle by increasing the rate and amount of heat flowing from the secondary heat transfer portion of the rotor to the primary heat transfer portion thereof, as illustrated by the heat transfer loop in Fig.8B.
- SFR heat exchange fluid flow increase
- the microprocessor After performing the operations at Blocks L, M and N, the microprocessor reads once again the temperature values in the temperature value storage register of the system controller, and at Block P determines whether there has been any change in mode (e.g. , switch from heating mode to cooling mode) . If no change in mode has been detected at Block P, then the microcontroller reenters the control loop defined by Blocks L through N and performs such operations in order to control the angular velocity of the rotor and the flow rates of the primary and secondary fluid flow-rate controllers. If at Block P in Fig.
- the microprocessor determines that the mode of the heat transfer engine has been changed (e.g., from the heating mode to the cooling mode) then the microprocessor returns to Block C in Fig. 10A and then proceeds to Block D.
- the speed at which the microprocessor traverses through this control loops described above will typically be substantially greater than the rate at which the temperature values may change as indicated by the data values in the temperature storage registers.
- the system controller can easily track the thermodynamics of the heat transfer engine of the present invention.
- the parameters (Wmax, W in, PFRmax, PRFmin, SFRmax, SFRmin) employed in the control process described above may be determined in a variety of ways.
- the parameters (W H , W L , PFRmax, PFRmin, SFRmax, and SFRmin) employed in the control process described above may be determined in a variety or ways.
- W H rotor RPM
- PFRmax, PFRmin, SFRmax, and SFRmin W H (rotor RPM) is primarily determined by the strength of materials used to construct the rotor, and, secondly, at an RPM where Q H is realized.
- Q H is found by acquiring the temperature of the fluid entering the primary heat transfer portion and the temperature of the fluid leaving the primary heat transfer portion. The lowest of the two temperature is subtracted from the highest temperature and the sum is the fluid temperature difference.
- the fluid temperature difference multiplied by the specific heat of the fluid being used equals the BTU per poind that particular fluid has absorbed or dissipated.
- W L is determined when the RPM is reduced to a point where no appreciable net refrigeration affect is taking place.
- PFRmax can be gallons per minute (GPM) for liquids or cubic feet per minute (CFM) for gasses.
- GPM gallons per minute
- CFM cubic feet per minute
- water entering the primary heat transfer portion at a temperature of 60 °F and leaving the primary heat transfer portion at 50 °F has a temperature difference of 10°F.
- Water has a specific heat of 1 BTU per pound at temperatures between 32 °F and 212 °F. Therefore, water recirculated at 100 gallons per minute, having a temperature difference of 10 °F is transferring 60,000 BTU per hour. Five tons of refrigeration and 60,000 BTUH heating.
- Air entering the primary heat transfer portion at a temperature of 60 °F and leaving the primary heat transfer at 50 °F has a temperature difference of 10 °F and contains 22 BTU per pound (dry air and associated moisture). Air at 60 °F and 50 percent relative humidity also contains approximately 22 BTU per pound (dry air and associated moisture) .
- the sensible heat ratio of the 60 °F air in the above example is .46 and the sensible heat ratio of the 50 °F air is .73.
- the 60 °F air contains mostly latent heat, about 11.88 BTU latent heat and 10.12 BTU sensible heat.
- the 50 °F air contains most sensible heat, about 5.94 BTU latent heat and 16.06 BTU sensible heat.
- the net refrigeration affect is the difference between 11.88 BTU and 5.94 BTU, or 5.94 BTU per pound of recirculated air has been transferred from the air into the primary heat transfer portion. In that condition, the air contains 13.01 cubic feet of air per poind. The air contracts slightly during cooling, about .19 cubic foot per pound of dry air. And, if 2,000 cubic feet of air are recirculated per minute, the net refrigeration affect will be 544,788.24 BTU per hour, or 4.57 tons of refrigeration.
- PFRmax would be 2000 CFM and SFRmax will equal PFRmax because of the lack of heat being introduced into the self-circulating circuit from internal motor windings and the heat of compression caused by reciprocating compressors.
- the range between PFRmin and PFRmax, and SFRmin and SFRmax is determined by physical aspects of a particular installation. Physical aspects can range from total environmental load reduction control system to a simple on-off control circuit.
- FIG. HA to HI the refrigeration process of the present invention will now be described with the heat transfer engine of the present engine being operation in its cooling mode of operation.
- each of these drawings schematically depicts, from a cross-sectional perspective, both the first and second heat exchanging portions of the rotor.
- This presentation of the internal structure of the closed fluid passageway throughout the rotor provides a clear illustration of both the location and the state of the refrigerant along the closed fluid passageway thereof.
- Fig. HA the rotor is shown at its rest position, which is indicated by the absence of any rotational arrow about the rotor shaft.
- the internal volume of the closed fluid circuit is occupied by about 65% of refrigerant in its liquid state.
- the entire spiral return passageway along the rotor shaft is occupied with liquid refrigerant, while the heat exchanging portions of the rotor are occupied with liquid refrigerant at a level set by gravity in the normal course.
- the portion of the fluid passageway above the liquid level in the rotor is occupied by refrigerant in a gaseous state.
- the closed fluid passageway is thoroughly cleaned and dehydrated prior to the addition of the selected refrigerant to prevent any contamination thereof.
- the rotor is rotated in a counterclockwise (CCW) direction within the stator housing of the heat transfer engine.
- CCW counterclockwise
- the primary heat transfer portion will perform a liquid refrigerant evaporating function
- the secondary heat transfer portion performs a refrigerant vapor condensing function.
- the liquid refrigerant within the spiraled passageway of the shaft begins to flow into the secondary heat transfer (i.e., exchanging) portion of the rotor and occupies the entire volume thereof.
- a very small portion (i.e., about one coil turn) of the primary heat transfer portion is occupied by refrigerant vapor as it passes through the throttling (i.e., metering ) device, while the remainder of the primary heat transfer portion of the rotor and a portion of the spiraled passageway of the shaft once occupied by liquid refrigerant is occupied with gas.
- the boundary between the length of liquid refrigerant and length of gas (or refrigerant vapor) in the rotor is, by definition, the "Liquid Seal" and resides along the primary heat transfer portion of the rotor shaft at this early stage of start-up operation.
- the Liquid Seal is located between the condensation and throttling processes supported within the rotor.
- the Liquid Seal has two primary functions within the rotor, namely: during start-up operations, to occlude the passage of refrigerant vapor, thereby forcing the vapor to condense in the secondary heat transfer portion (i.e., condenser) ; and, more precisely, during steady state operation the Liquid Seal resides at a point along the length of the secondary heat transfer portion where enough refrigerant vapor has condensed into a liquid by absorbing "Latent Heat", thereby occupying the total internal face area of the passageway.
- Solid Heat is defined herein as the heat absorbed by (into) the liquid refrigerant (homogeneous fluid) during the evaporization process, as well as the heat discharged from the gaseous refrigerant during the condensation process.
- Liquid refrigerant contained in the first one half of the secondary heat transfer portion between the rotor shaft and the point of highest radius (from the center of rotation) is effectively moved and partially pressurized by centrifugal force, and the physical shape of the spiraled passageway, outwardly from the center of rotation into the second one half of the secondary heat transfer portion.
- Liquid refrigerant contained in the second one half of the secondary heat transfer portion between the point of highest radius (from the center of rotation) and the throttling device (i.e., metering) is affectively pressurized (against flow restriction caused by the throttling device and Liquid Seal) by the physical shape of the spiraled passageway and centrifugal force.
- Thermal Load or “Thermal Loading” as used here shall mean the demand of heat transfer imposed upon the heat transfer engine of the present invention in a particular mode of operation.
- Liquid refrigerant is pressurized due to (i) the distribution of centrifugal forces acting on the molecules of the liquid refrigerant therein as well as (ii) the pressure created by the liquid refrigerant being forcibly driven into the secondary heat transfer portion against the Liquid Seal and the metering device flow restriction. As shown in Fig.
- the Liquid Seal moves towards the secondary heat transfer portion, and refrigerant flow into the primary heat transfer portion is restricted by the throttling device and the refrigerant stacks up in the secondary heat transfer portion. Very little refrigerant flows into the primary heat transfer portion, and no refrigeration affect has yet taken place.
- the small amount of vapor in the primary heat transfer portion will gather some "Superheat” which will remain in the vapor and gaseous refrigerant within the primary heat transfer portion, as a result of the Liquid Seal.
- Superheat shall be defined as a sensible heat gain above the saturation temperature of the liquid refrigerant, at which a change in temperature of the refrigerant gas occurs (sensed) with no change in pressure.
- the rotor continues to increase in speed in the CCW direction.
- the Liquid Pressurization Length of the refrigerant begins to create enough pressure within the secondary heat transfer portion to overcome the pressure restriction caused by the throttling device and thus liquid begins to flow into the primary heat transfer portion of the rotor.
- the Liquid Seal has moved along the rotor shaft towards the secondary heat transfer portion.
- refrigerant beyond the metering device and into about the first spiral coil of the primary heat transfer portion is in the form of a "homogeneous fluid" (i.e., a mixture of liquid and vapor state) while a portion of the first spiral coil and a portion of the second one contain refrigerant in its homogeneous state.
- a homogeneous fluid i.e., a mixture of liquid and vapor state
- the term "homogeneous fluid” shall mean a mixture of flash gas and low temperature, low pressure, liquid refrigerant experiencing a change-in-state (the process of evaporization) due to its absorption of heat.
- the length of refrigerant over which Evaporization occurs shall be defined as the Evaporization Length of the refrigerant, whereas the section of the refrigerant stream along the fluid flow passageway containing gas shall be defined as the Superheat Length, as shown.
- the homogeneous fluid entering the primary heat transfer portion "displaces" the gas therewithin, thereby pushing it downstream into the spiraled passageway of the rotor shaft. Throttling of liquid refrigerant into vapor absorbs heat from the primary heat transfer portion of the rotor, imparting "Superheat" to the gaseous refrigerant. A "cooler” vapor created by the process of throttling enters the primary heat transfer portion and begins to absorb more Superheat. Refrigerant gas and vapor are compressed between the homogeneous fluid in the primary heat transfer portion and the Liquid Seal in the spiraled passageway of the rotor shaft.
- the Liquid Line shall be defined as the point where the homogeneous fluid ends and the vapor begins along the length of the primary heat transfer portion. Therefore, the liquid line illustrated in Figs. HC to HF can occupy a short length of the primary heat transfer portion as a mixture of homogeneous fluid and a very dense vapor which extends downstream to the Superheat length. The exact location along the primary heat transfer portion will vary depending on the quantity of homogeneous fluid, which is in proportion to the amount of heat being absorbed and the Thermal Load (i.e., heat transfer demand) being imposed on the heat transfer engine in its mode of operation. The Liquid Line is not to be confused with the Liquid Seal .
- the quantity of refrigerant vapor within the primary heat transfer portion of the rotor continues to increase due to the increased production of flash gas from throttling of liquid refrigerant.
- the Liquid Seal has moved towards the end of the rotor shaft and the secondary heat transfer portion inlet thereof. Also, during this stage of operation, the flow of heat (i.e., Superheat) from the primary heat transfer portion is still trapped behind the Liquid Seal in the spiraled passageway of the rotor shaft.
- the Superheat Heat from the primary heat transfer portion is unable to pass onto the secondary heat transfer portions primary and secondary heat transfer surfaces, and thus optimal operation is not yet achieved at this stage of engine operation.
- some heat may transfer into the rotor shaft from the refrigerant vapor if the shaft temperature is less that the temperature of the refrigerant vapor; and some heat may transfer into the refrigerant vapor if the refrigerant vapor temperature is less than that of the rotor shaft.
- the rotor shaft and its internal spiraled passageway is a systematic source of primary and secondary Superheat transfer surfaces where heat can be either introduced into the vapor or discharged from the vapor.
- the rotor At the stage of operation shown in Fig.HF, the rotor is approaching its steady-state angular velocity, and is shown operating in the CCW direction of operation at what shall be called "Threshold Velocity". As shown, the remaining liquid refrigerant in the rotor shaft is now completely displaced by refrigerant vapor produced as a result of the evaporization of the liquid refrigerant in primary heat transfer portion of the rotor. Consequently, Superheat produced from the primary heat transfer portion is permitted to flow through the spiraled passageway of the rotor shaft and into the secondary heat transfer portion, where it can be liberated by way of condensation across the secondary heat transfer portion.
- the Liquid Seal is no longer located along the rotor shaft, but within the secondary heat transfer portion of the rotor, near the end of the rotor shaft. Vapor compression begins to occur in the last part of the primary heat transfer portion and along the spiraled passageway of the rotor.
- the pressure of the liquid refrigerant in the Liquid Pressurization Length has increased sufficiently enough to further increase the production of homogeneous fluid in the primary heat transfer portion. This also causes the quantity of liquid in the secondary heat transfer portion to decrease "Pulling" on the flash gas and vapor located in the spiraled passageway in the rotor shaft, and in the primary heat transfer portion downstream from the homogeneous fluid.
- the homogeneous fluid is evaporating absorbing heat within the primary heat transfer portion of the rotor for transference and systematic discharge from the secondary heat transfer portion.
- the vapor within the primary heat transfer portion can contain more Superheat by volume than the gas with which it is mixed.
- the increased volume in dense vapor in the primary heat transfer portion provides a means of storing Superheat (absorbed from the primary heat exchanging circuit) until the vapor stream flows into the secondary heat transfer portion of the rotor where it can be liberated to the secondary heat exchanging circuit by way of conduction.
- the heat transfer engine of the present invention is operated at what shall be called the "Balance Point Condition", the refrigeration cycle of which is illustrated in Fig. 17A and 17B.
- the refrigerant within the rotor has attained the necessary phase distribution where simultaneously there is an equal amount of refrigerant being evaporated in the primary heat transfer portion as there is refrigerant vapor being condensed in the secondary heat transfer portion of the rotor.
- the Superheat that has "accumulated" in the refrigerant vapor during the start up sequence shown in Figs. HA through HF begins to dissipate from the DeSuperheat Length of the refrigerant stream along the secondary heat transfer portion of the rotor.
- the density of the refrigerant gas increases, and vapor compression occurs as the Superheat is carried by the refrigerant gas from the Superheat Length of the primary heat transfer portion to the DeSuperheat Length in the secondary heat transfer portion by the spiraled passageway in the rotor shaft.
- the spiraled passageway of the rotor shaft has a greater compressive affect on the vapor therein at this stage of operation.
- the spiraled passageway of the shaft is pressurizing the Superheated gas and dense vapor against the Liquid Seal in the secondary heat transfer portion.
- pressurization of liquid refrigerant in the secondary heat transfer portion of the rotor pushes the liquid refrigerant through the throttling device at a higher pressure, sufficiently enough, which causes a portion of the liquid refrigerant to "flash" into a gas, thereby, reducing the temperature of the remaining homogeneous fluid (i.e., liquid and dense vapor) entering the primary heat transfer portion thereof.
- the liquid refrigerant portion of the homogeneous fluid evaporates, creating sufficient vapor pressure therein that it displaces vapor downstream within the primary heat transfer portion into the spiraled passageway of the rotor shaft.
- This vapor pressure enhanced by vapor compression caused by the spiraled passageway in the rotor shaft, pushes the same into the secondary heat transfer portion of the rotor, where its Superheat is liberated over the DeSuperheat Length thereof.
- the Liquid Seal tends to remain near the same location in the secondary heat transfer portion, while the Liquid Line tends to remain near the same location in the primary heat transfer portion.
- the temperature and pressure of the refrigerant in the secondary heat transfer portion of the rotor is higher than the refrigerant in the primary heat transfer portion thereof.
- the rate of heat transfer from the primary heat exchanging chamber of the engine into the primary heat transfer portion thereof is substantially equal to the rate of heat transfer from the secondary heat transfer portion of the engine into the secondary heat exchanging chamber thereof.
- the throttling process of the present invention can be described in terms of the three sub-processes which determine the condition of the refrigerant as it passes through the throttling device of the engine in either of its rotational directions. These sub-processes are defined as the Liquid Length, the Bubble Point, and the Two Phase Length. For purposes of clarity, the suprocesses of the throttling process will be described as they occur during start-up operations and steady-state operations.
- the Liquid Length begins at the inlet of the throttling device and continues to the Bubble Point.
- the Bubble Point exists at point inside (or along) the throttling device, (i) at which the Liquid Length (liquid refrigerant) is separated or distinguishable from the Two Phase Length (foamy, liquid and vapor refrigerant) and (ii) where enough pressure drop along the restrictive passage of the throttling device has occurred to cause a portion of the liquid refrigerant to evaporate (a single bubble) and reduce the temperature of the surrounding liquid refrigerant (two phase, bubbles and liquid) for delivery into the evaporator section of the rotor.
- the Latent Heat given up by the liquid refrigerant during its change in state at the Bubble Point is contained within the bubbles produced at the Bubble Point. Heat absorbed by these bubbles in the evaporator section of the rotor is Superheat.
- the Bubble Point can exist anywhere along the throttling devices length depending on the amount of thermal load imposed on the heat transfer engine.
- the Liquid Length extends over that portion of the throttling device containing pure liquid refrigerant up to the Bubble Point.
- the Two-Phase Length extends from the Bubble Point into the evaporator inlet of the rotor and (foamy, liquid and vapor refrigerant) .
- the Condensation Length and Evaporation Length each contain an equal amount of liquid refrigerant. This is because the amount of heat entering the primary heat transfer portion of the rotor is equal to the amount of heat leaving the secondary heat transfer portion thereof.
- the first reason is that the primary heat transfer portion of the rotor has a higher rate of heat transfer by virtue of the higher-than-design temperature difference existing between the homogeneous fluid in the primary heat transfer portion of the rotor and the air or liquid passing over the primary heat transfer surfaces.
- the second reason is that the increase in the throttling process lowers the temperature and pressure of the homogeneous fluid entering the primary heat transfer portion of the rotor.
- the additional liquid refrigerant in the secondary heat transfer portion of the rotor reduces the available internal volume needed for adequate vapor-to-liquid condensation.
- the centrifugal heat transfer engine is "Over Loaded". In such cases, a larger rotor should be used for the application.
- the centrifugal heat transfer engine has more liquid refrigerant in the primary heat transfer portion than is contained by the secondary heat transfer portion.
- the accumulation of liquid refrigerant in the primary heat transfer portion is due the low rate of heat transfer in the primary heat transfer portion.
- the temperature and pressure of the refrigerant in the secondary heat transfer portion can be increased by reducing the rate of flow of the heat exchanging fluid circulating through the secondary heat exchanging chamber. Such a decrease in fluid flow causes an increase in temperature and pressure of the refrigerant in the primary heat transfer portion which, in turn, causes an increase in temperature and pressure of the refrigerant in the primary heat transfer portion.
- the increase in temperature and pressure of the refrigerant in the primary heat transfer portion increases the amount of heat (BTU) per pound that a hydrocarbon refrigerant is capable of absorbing, to an optimum saturation temperature and pressure.
- BTU heat
- the industry design standard is 95 degrees Fahrenheit condensing temperature.
- Such a controlled decrease in fluid flow shall be referred to as "Secondary Pressure Stabilization”.
- Such a controlled decrease in fluid flow can increase the engines coefficient of performance (COP, or BTU/WATT) of the heat transfer engine.
- a similar increase or decrease in the primary heat exchanging fluid flow shall be referred to as "Primary Pressure Stabilization".
- the RPM of the rotor can be reduced causing a reduction in the refrigeration affect to satisfy a lesser load demand.
- This type of operation, or mode is called Load Reduction Control (or Unloading) .
- Thermal Loading is where the rotor RPM is increased to satisfy a higher load demand.
- the location of the Liquid Seal is affected by the amount of load being exerted on the evaporization process.
- Liquid pressurization begins at the Liquid Seal and occurs inside the spiraled condenser section along the Liquid Pressurization Length up to the inlet of the throttling (i.e., metering) device inlet.
- the liquid refrigerant is forced toward the central axis of rotation by the spiraled shape of the Liquid Pressurization Length in the condenser functioning section of the rotor.
- centrifugal forces produced during rotor rotation causes the liquid pressure to gradually increase along the Liquid Pressurization Length, providing a continuous supply of higher pressure (condensed) liquid refrigerant to the inlet of the throttling device where the Liquid Length begins.
- centrifugal forces within the rotor increase the weight of the liquid refrigerant contained in the spiraled Liquid Pressurization Length and cause the liquid refrigerant therewith to pressurize against the flow restricting pressure drop produced by the fluid flow geometry of the throttling device, thereby completing the refrigeration cycle of the centrifugal heat transfer engine.
- Fig. HH the heat transfer engine of the present invention is shown operating just below its an "optimum" (low load) operating condition, whereas in Fig. HI, the heat transfer engine is shown operated excessively beyond its "optimum” operating condition.
- optimum operating condition used above is not to be equated with the term “Balance Point” operating condition. Rather “optimum operating condition” is a point of operation where the amount of liquid refrigerant in the primary heat transfer portion is slightly higher than the amount of liquid refrigerant in the secondary heat transfer portion.
- This operating point is considered optimum as the lower temperature refrigerant in the primary heat transfer portion is capable of containing more heat (i.e., BTU per pound) than the higher pressure and temperature liquid refrigerant contained in the secondary heat transfer portion of the rotor. Consequently, during engine operation, the flow rate of heat exchanging fluid within the secondary heat exchanging chamber of the engine is reduced at times by the system controller, as this increases the temperature of the secondary heat transfer portion (i.e., during the cooling mode), and thereby increasing the "rate" of heat flow from the secondary heat transfer portion of the rotor (particularly on large capacity engines) into the secondary heat exchanging fluid circulating through the secondary heat exchanging chamber. If the thermal load on the engine is further reduced beyond that shown in Fig.
- the spiraled passageway in the rotor shaft prevents a condition where the Liquid Pressurization Length is starved of liquid refrigerant.
- This safety measure is provided by the fact that at least sixty five percent of the total internal volume of the rotor is occupied by refrigerant, and that quantities of refrigerant exceeding the internal volume of the primary heat transfer portion and extending into the spiraled passageway in the rotor shaft are rapidly moved into the secondary heat transfer portion (by way of the rotating spiraled passageway along the rotor shaft) , thereby rapidly replenishing the Liquid Pressurization Length thereof.
- the Liquid Seal has moved nearer to the throttling device, and even though the Liquid Seal is located in the secondary heat transfer portion, the Liquid Pressurization Length is still pressurizing the liquid refrigerant.
- the heat transfer engine is shown operated at a point of operation where the "load” has diminished sufficiently to cause the liquid refrigerant within the rotor to "accumulate” in the primary heat transfer portion thereof.
- the system controller of the engine should be reacting to a reduction in temperature in the primary heat exchanging chamber, thereby reducing the RPM of the rotor.
- the flow rate controller associated with the primary heat exchanging chamber should be starting to reduce the flow rate of heat exchanging fluid circulating within the secondary heat exchanging chamber.
- the rotor RPM would be further decreased in order to reduce the refrigeration affect. In turn, this would increase the "overall system pressure", causing the ambient temperature about the primary heat exchanging portion to increase, thereby preventing the formation of ice (or accumulation of process fluid) on the primary and secondary heat transfer surfaces thereof.
- FIG. 12A to 121 the refrigeration process of the present invention will now be described with the heat transfer engine of the present engine being operation in its heating mode of operation.
- each of these drawings schematically depicts, from a cross-sectional perspective, both the first and second heat exchanging portions of the rotor.
- This presentation of the internal structure of the closed fluid passageway throughout the rotor provides a clear illustration of both the location and the state of the refrigerant along the closed fluid passageway thereof.
- Fig. 12A the rotor is shown at its rest position, which is indicated by the absence of any rotational arrow about the rotor shaft.
- the internal volume of the closed fluid circuit is occupied by about 65% of refrigerant in its liquid state.
- the entire spiral return passageway along the rotor shaft is occupied with liquid refrigerant, while the heat exchanging portions of the rotor are occupied with liquid refrigerant at a level set by gravity in the normal course.
- the portion of the fluid passageway above the liquid level in the rotor is occupied by refrigerant in a gaseous state.
- the closed fluid flow passageway is thoroughly cleaned and dehydrated prior to the addition of the selected refrigerant to prevent any contamination thereof.
- the rotor is rotated in a clockwise (CW) direction within the stator housing of the heat transfer engine.
- the primary heat transfer portion will perform a liquid refrigerant evaporating function
- the secondary heat transfer portion performs a refrigerant vapor condensing function.
- the liquid refrigerant within the spiraled passageway of the shaft begins to flow into the secondary heat transfer (i.e., exchanging) portion of the rotor and occupies the entire volume thereof.
- a very small portion (i.e., about one coil turn) of the primary heat transfer portion is occupied by refrigerant vapor as it passes through the throttling (i.e., metering ) device, while the remainder of the primary heat transfer portion of the rotor and a portion of the spiraled passageway of the shaft once occupied by liquid refrigerant is occupied with gas.
- the Liquid Seal resides at a point along the length of the secondary heat transfer portion where enough refrigerant vapor has condensed into a liquid thereby occupying the total internal face area of the passageway.
- the Liquid Seal moves towards the secondary heat transfer portion, and refrigerant flow into the primary heat transfer portion is restricted by the throttling device and the refrigerant stacks up in the secondary heat transfer portion. Very little refrigerant flows into the primary heat transfer portion, and no refrigeration affect has yet taken place. The small amount of vapor in the primary heat transfer portion will gather some Superheat which will remain in the vapor and gaseous refrigerant within the primary heat transfer portion, as a result of the Liquid Seal.
- the rotor continues to increase in speed in the CW direction.
- the Liquid Pressurization Length of the refrigerant begins to create enough pressure within the secondary heat transfer portion to overcome the pressure restriction caused by the throttling device and thus liquid begins to flow into the primary heat transfer portion of the rotor.
- the Liquid Seal has moved along the rotor shaft towards the secondary heat transfer portion. The homogeneous fluid entering the primary heat transfer portion "displaces" the gas therewithin, thereby pushing it downstream into the spiraled passageway of the rotor shaft.
- the amount of refrigerant vapor in the primary heat transfer portion increase due to increased throttling and increased "Flash" gas entering the same.
- the effect of this is to increase the quantity of homogeneous fluid entering the primary heat transfer portion of the rotor.
- the Liquid Seal has moved even further along the rotor shaft towards the secondary heat transfer portion. Also, less liquid refrigerant occupies the spiraled passageway of the rotor shaft, while more homogeneous fluid occupies the primary heat transfer portion of the rotor.
- the direction of heat flow is from the primary heat transfer portion to the secondary heat transfer portion (i.e., in the form of Superheat).
- this heat flow is trapped behind the Liquid Seal in the spiraled passageway of the shaft.
- the quantity of refrigerant vapor within the primary heat transfer portion of the rotor continues to increase due to the increased production of flash gas from throttling of liquid refrigerant.
- the Liquid Seal has moved towards the end of the rotor shaft and the secondary heat transfer portion inlet thereof. Also, during this stage of operation, the flow of heat (i.e..
- Superheat from the primary heat transfer portion is still trapped behind the Liquid Seal in the spiraled passageway of the rotor shaft. Consequently, the Superheat from the primary heat transfer portion is unable to pass onto the secondary heat transfer portions primary and secondary heat transfer surfaces, and thus optimal operation is not yet achieved at this stage of engine operation.
- some heat i.e., Superheat
- the rotor is approaching its steady-state angular velocity, and is shown operating in the CW direction of operation at its "Threshold Velocity".
- the remaining liquid refrigerant in the rotor shaft is now completely displaced by refrigerant vapor produced as a result of the evaporization of the liquid refrigerant in primary heat transfer portion of the rotor. Consequently, Superheat produced from the primary heat transfer portion is permitted to flow through the spiraled passageway of the rotor shaft and into the secondary heat transfer portion, where it can be liberated by way of condensation across the secondary heat transfer portion.
- the Liquid Seal is no longer located along the rotor shaft, but within the secondary heat transfer portion of the rotor, near the end of the rotor shaft. Vapor compression begins to occur in the last part of the primary heat transfer portion and along the spiraled passageway of the rotor.
- the pressure of the liquid refrigerant in the Liquid Pressurization Length has increased sufficiently enough to further increase the production of homogeneous fluid in the primary heat transfer portion. This also causes the quantity of liquid in the secondary heat transfer portion to decrease "Pulling" on the flash gas and vapor located in the spiraled passageway in the rotor shaft, and in the primary heat transfer portion downstream from the homogeneous fluid.
- the homogeneous fluid is evaporating absorbing heat within the primary heat transfer portion of the rotor for transference and systematic discharge from the secondary heat transfer portion into the heat exchanging fluid circulating through the primary heat exchanging chamber.
- the vapor within the primary heat transfer portion can contain more Superheat by volume than the gas with which it is mixed.
- the increased volume in dense vapor in the primary heat transfer portion provides a means of storing Superheat (absorbed from the primary heat exchanging circuit) until the vapor stream flows into the secondary heat transfer portion of the rotor where it can be liberated to the secondary heat exchanging circuit by way of conduction.
- the heat transfer engine of the present invention is operating at what shall be called the "Balance Point Condition".
- the refrigerant within the rotor has attained the necessary phase distribution where simultaneously there is an equal amount of refrigerant being evaporated in the primary heat transfer portion as there is refrigerant vapor being condensed in the secondary heat transfer portion of the rotor.
- the secondary heat transfer portion is adding heat to the primary heat transfer chamber.
- the Superheat that has "accumulated" in the refrigerant vapor during the start up sequence shown in Figs.l2A through 12F begins to dissipate from the DeSuperheat Length of the refrigerant stream along the secondary heat transfer portion of the rotor.
- the density of the refrigerant gas increases, and vapor compression occurs as the Superheat is carried by the refrigerant gas from the Superheat Length of the primary heat transfer portion to the DeSuperheat Length in the secondary heat transfer portion by the spiraled passageway in the rotor shaft.
- the spiraled passageway of the rotor shaft has a greater compressive affect on the vapor therein.
- the spiraled passageway of the shaft is pressurizing the Superheated gas and dense vapor against the Liquid Seal in the secondary heat transfer portion.
- pressurization of liquid refrigerant in the secondary heat transfer portion of the rotor pushes the liquid refrigerant through the throttling device at a sufficiently higher pressure, which causes a portion of the liquid refrigerant to "flash" into a gas, thereby, reducing the temperature of the remaining homogeneous fluid (liquid and dense vapor) entering the primary heat transfer portion thereof.
- the liquid refrigerant portion of the homogeneous fluid evaporates which creates sufficient vapor pressure therein that it displaces vapor downstream within the primary heat transfer portion into the spiraled passageway of the rotor shaft. This vapor pressure, enhanced by vapor compression caused by the spiraled passageway in the rotor shaft, pushes the same into the secondary heat transfer portion of the rotor, where its
- the Liquid Seal tends to remain near the same location in the secondary heat transfer portion, while the Liquid Line tends to remain near the same location in the primary heat transfer portion.
- the temperature and pressure of the refrigerant in the secondary heat transfer portion of the rotor is higher than the refrigerant in the primary heat transfer portion thereof.
- the rate of heat transfer to the primary heat exchanging chamber of the engine from the secondary heat transfer portion thereof is substantially equal to the rate of heat transfer from the primary heat transfer portion of the engine into the secondary heat exchanging chamber thereof.
- the heat transfer engine of the present invention is shown operating just below its an optimum (low load) operating condition.
- Fig. 121 the heat transfer engine is shown operated excessively beyond its "optimum" operating condition.
- the Liquid Seal is located in the secondary heat transfer portion, and even though the Liquid Seal has moved nearer toward the throttling device, the Liquid Pressurization Length is still pressurizing the liquid refrigerant.
- the demand for heat by the system controller during this state of operation has diminished sufficiently to cause the liquid refrigerant within the rotor to "accumulate" in the primary heat transfer portion thereof.
- the system controller of the engine should be reacting to an increase in temperature in the primary heat exchanging chamber, reducing the RPM of the rotor, and the flow rate controller associated with the primary heat transfer chamber should be starting to reduce the flow rate of the heat exchanging fluid circulating within the secondary heat exchanging chamber.
- the heat transfer engine of the first illustrative embodiment is shown installed on the roof of a building or similar structure, as part of an air handling system which is commonly known in the industry as a Roof-Top or Self-contained air conditioning unit, or air handler.
- the heat transfer engine functions as a roof-top air conditioning unit which can be operated in its cooling mode or heating mode.
- air conditioning shall include the concept of cooling and/or heating of the air to be "temperature conditioned", in addition to the conditioning of air for human occupancy which includes its temperature, humidity, quantity, and cleanliness.
- the air handling unit comprises an supply air duct 60 and an return air duct 61, both penetrating structural components of a building.
- the rotor of the centrifugal heat transfer engine is rotated by a variable-speed electric motor 62.
- the angular velocity of the rotor is controlled by a torque converter or magnetic clutch 63.
- the primary heat transfer portion of the rotor 68 functioning as the evaporator during the cooling mode, is insulated from the secondary heat transfer position functioning as the condenser.
- a fan 64 rotated by a variable speed motor 65, is provided for moving atmospheric air over the secondary heat transfer portion of the rotor.
- a blower wheel 66 inside a blower housing rotated by a variable speed motor 67 is provided for moving air over the primary heat transfer portion of the rotor creating air circulation in the primary heat exchange circuit.
- the air temperature at the inlet of the secondary heat exchanging chamber 14 is sensed by a temperature sensor located in the air flow upstream of the secondary heat transfer portion 69, whereas the air temperature at the outlet thereof is sensed by a temperature sensor located in the air flow downstream from the secondary heat transfer portion 69.
- the air temperature at the inlet of the primary heat exchanging chamber 13 is sensed by a temperature sensor located in the air flow upstream of the primary heat transfer portion 68, wherein the air temperature at the outlet thereof is sensed by a temperature sensor located downstream from the primary heat transfer portion 68.
- a simple external on/off thermostat switch 9 can be used to measure temperature Tl and thus start motors 62, 65 and 67 during the heating or cooling mode of operation.
- the function of the air supply duct 60 is to convey refrigerated (i.e., cooled/conditioned) air from the primary heat transfer portion of the rotor, into the structure (e.g., space to be cooled) , whereas the function of the air return duct 61 is to convey air from the structure back to the primary heat transfer portion for cooling.
- the direction of the rotor is reversed by torque generator 62, and the function of the air supply duct is to convey heated air from the primary heat transfer portion of the rotor, into the structure (e.g., space to be heated), whereas the function of the air return duct 61 is to convey air from the structure back to the primary heat transfer portion for heating.
- the heat transfer engine of the second illustrative embodiment 70 comprises a stator housing 71 within which a turbine-like rotor 72 is rotatably supported.
- the rotor is realized as solid rotary structure having a turbine-like geometry.
- a closed self-circulating fluid-carrying circuit 73 is embodied.
- the closed fluid carrying circuit has spiraled primary and secondary tubular heat transfer passageways, and a metering device which will be described in greater detail.
- these passageways are molded and/or machined in substantially similar disks of different diameters that are stacked and fastened together to form a unity structure.
- heat transfer fins are added to each of the disks in order to (1) increase the secondary heat transfer surface areas thereof and (2) provide a means of systematic fluid circulation.
- the stator assembly 70 comprises a pair of split-cast housing halves 71A and 71B which are machined to form the fluid flow circuit, and bolted together with bolts 74.
- the stator housing has primary and secondary heat exchanging chambers 75 and 76, within which the primary and secondary portions of the heating exchanging rotor are housed.
- flanged fluid piping couplings i.e., port connections
- 77A and 77B and 78A and 78B are provided to the input and output ports of the primary and secondary heat exchanging chambers of the stator housing, respectively, as shown in Figs. 14A, 14B and 20.
- Conventional fluid carrying pipes with flanged fittings can be easily connected to these flanged port connections.
- the flow of heat exchanging fluid into the input ports of the primary and secondary heat exchanging chambers of the stator housing will be such that each such fluid flow imparts torque to the rotor shaft in a cooperative manner, to perform positive work.
- the angular velocity of the rotor can be controlled in a number of different ways depending on the application at hand. Referring now to Figs. 15A through 15L, the structure of the rotor of the second illustrative embodiment will be described in greater detail.
- the primary heat exchanging portion of the rotor comprises a first set of rotor disks 80A having radially varying outer diameters and a second set of rotor disks 80B having radially uniform outer diameters.
- the secondary heat exchanging portion of the rotor comprises a first set of rotor disks 81A having radially varying outer diameters and a second set of rotor disks 81B having radially uniform outer diameters.
- Fig. 15A, 15B, and 15C the primary heat exchanging portion of the rotor comprises a first set of rotor disks 80A having radially varying outer diameters and a second set of rotor disks 80B having radially uniform outer diameters.
- each of these rotor disks has a central bore 82 of substantially the same diameter, and a small section of the fluid flow circuit (i.e., passageway) 83 machined, molded or otherwise formed therein.
- the exact geometry of each section of fluid flow passageway within each rotor disc will vary from rotor disk to rotor disk. However, these sections of fluid flow passageways combine over the length of the rotor to form the greater portion of the closed fluid flow circuit 83 embodied within the rotor structure of the second illustrative embodiment.
- the central bearing structure 80 of the rotor comprises an assembly of subcomponents, namely: an outer cylindrically-shaped bearing sleeve 81 for rotational support within a suitable support structure provided within the stator housing; an inner fluid flow cylinder 82 of substantially cylindrical geometry adapted to be received within bearing sleeve 81, having first and second disc-receiving collars 83 and 84 of reduced diameter adapted for receipt by inner rotor disc 85 and 86, respectively; a pair of thrust plates 87 and 88 having inner central bores with diameters slightly greater than the outer diameter of the inner fluid flow cylinder; and a inner fluid flow tube 89 having a inner bore 90 extending along its entire length, and a spirally-extending flange 91 formed on the exterior surface thereof, for directing return refrigerant.
- the central portion of the rotor functions not only as a rotor bearing structure, but also as (i) the refrigerant metering (i.e., throttling) device of the rotor and (ii) a fluid flow return passageway.
- the refrigerant metering (i.e., throttling) device of the rotor i.e., throttling
- a fluid flow return passageway i.e., a fluid flow return passageway.
- the end ost turbine disks 92 and 93 have machined within their plate or body portion , a section of fluid flow passageway 82 which extends from a direction substantially perpendicular to the rotor axis of rotation, to a direction substantially co-parallel with the rotor axis.
- These sections of closed fluid flow circuit allow refrigerant to flow continuously from the linear portion thereof to the spiral portions thereof.
- each end turbine disk is provided with a charging port 94 and which is in fluid communication with its central bore 82.
- turbine disc 92 and 93 have exterior threads 95 which are received by matched interior threads on charging port caps 96A and 96B which can be easily screwed onto and off the charging ports of these rotor discs.
- a seal 97 is provided between each charging port cap and its end rotor disc, as shown.
- each turbine disc set, 80A and 81A carry a plurality of turbine-like fins 99 for purpose of imparting torque to the rotor when heat exchanging fluid flows thereover while flowing through the heat exchanging chambers of the engine.
- the shape of these fins will be determined "by their function.
- the fins will be have 3-D surface characteristics which aid in imparting hydrodynamically generated torque to the rotor during engine operation.
- each fin has a base portion 100 which is designed to be received within a mated slot 101 formed in the outer end surface of each rotor disc.
- Various types of techniques may be employed to securely retain these turbinelike fins within their mounting slots.
- the section of fluid flow passageway machined in the planar body portion of each rotor disk will vary in geometrical characteristics, depending on the location of the rotor disc along the rotor axis.
- the fluid flow passageway 83 in each rotor disk extends about the center of the rotor disc.
- rotor discs 85 and 86 are structurally different than the other discs comprising the heat exchanging portions of the rotor of the second illustrative embodiment.
- inlet and outlet rotor discs 85 and 86 are machined so that during the cooling mode, refrigerant in vapor state, is transported from the first heat exchanging portion of the rotor to the second heat exchanging portion thereof by way of the spiraled passageway 102, and during the heating mode, vapor refrigerant is transported in the reverse flow direction through the central portion of the rotor.
- the section of fluid passageway in rotor disks 85 and 86 must extend radially inward towards enlarged central recesses 91A and 91B respectively, which are adapted to receive the end of cylindrical flanges 83 and 84 of fluid flow cylinder 80 shown in Fig. 15B.
- inlet and outlet rotor disks 85 and 86 have central bores 82 which are aligned with the central bore of the other rotor disks in the rotor structure.
- the inner fluid flow cylinder 80 has an axial bore machined, or otherwise drilled and formed, along its longitudinal extent. Also, fluid flow openings 103 and 104 are formed in the cylindrical flange structures 83 and 84, respectively, extending from the end portions of the inner fluid cylinder.
- the inner diameter of the axial bore 105 formed through outer fluid flow cylinder 82 is about 0.002 inches smaller than the outer diameter of the inner fluid flow tube 89 which carries the spirally extending flange 91.
- a thin, annular-shaped fluid flow channel 102 is formed therebetween along the entire length thereof.
- the fluid flow openings 103 and 104 in the flanges of outer fluid flow cylinder 82 are aligned with the terminal portions of the section of the fluid flow passageway in inlet and outlet rotor discs 85 and 86 (i.e., at the circumferential edge of circular recess 91A and 9IB formed in these disc sections) .
- the annular-shaped fluid flow channel 102 places the portion of the fluid flow circuit along the first heat exchanging portion of the rotor in fluid communication with the portion of the fluid flow circuit along the second heat exchanging portion of the rotor.
- the section of fluid flow passageway 90 passing through the inner fluid flow tube 89 functions as a bidirectional throttling (i.e., metering) device within the rotor, as it serves to effectively restrict the flow of refrigerant passing therethrough by virtue of its length and inner diameter characteristics.
- throttling i.e., metering
- the length and inner diameter dimensions of the linear flow passageway through the inner fluid flow tube can be selected so that the required amount of throttling is provided within the closed fluid circuit during engine operation.
- the linear length of the throttling channel is about four (4) inches
- the diameter of throttling channel will need to be about 0.028 inches.
- the total refrigerant charge required can be as little as 1.5 pounds of liquid refrigerant for small capacity systems, to hundreds of pounds of liquid refrigerant for larger capacity systems.
- the rotor structure described above can be made using virtually any number of rotor disks. It is understood, however, that the number of rotor disks used will depend, in large part, on the thermal load requirements (tonnage in BTUH) which must be satisfied in the application at hand.
- Fig. 15A shows the assembled rotor structure of the second illustrative embodiment removed from within its stator.
- This figures shows the secondary heat transfer portion, primary heat transfer portion, the rotor shaft 80, the rotor fins 99, and charging ports 95 and 96 of the rotor.
- the assembly of the rotor structure of the second illustrative embodiment may be achieved in a variety of ways. For example, once assembled in their proper order and configuration, the rotor disks can be welded together and thus avoiding the need for pressure/liquid-seals (e.g., gaskets) , or bolted together and thus requiring the need for seals or gaskets.
- portions of the rotor structure may be realized using casted parts which can be assembled together using welding and/or bolting techniques well known in the art.
- FIGs. 16A to 161 the refrigeration process of the present invention will now be described with the heat transfer engine of the second illustrative embodiment in its cooling mode of operation.
- each of these drawings schematically depicts, from a cross- sectional perspective, both the first and second heat exchanging portions of the rotor.
- This presentation of the internal structure of the closed fluid flow passageway throughout the rotor provides a clear illustration of both the location and the state of the refrigerant along the closed fluid flow passageway thereof.
- the heat transfer engine turbine of the second illustrative embodiment accomplishes a refrigeration affect through the sub-processes of throttling, evaporization, superheating, vapor compression, desuperheating, condensation, liquid seal formation and liquid pressurization in the same order except using a the turbine-like rotor structure described above.
- Fig. 16A the rotor is shown at its rest position, which is indicated by the absence of any rotational arrow about the rotor shaft.
- the internal volume of the closed fluid circuit is occupied by about 65% of refrigerant in its liquid state.
- the entire spiral return passageway along the rotor shaft is occupied with liquid refrigerant, while the heat exchanging portions of the rotor are occupied with liquid refrigerant at a level set by gravity in the normal course. No throttling of liquid into refrigerant vapor occurs at this stage of operation.
- the portion of the fluid passageway above the liquid level in the rotor is occupied by refrigerant in a gaseous state.
- the closed fluid flow passageway is thoroughly cleaned and dehydrated prior to the addition of the selected refrigerant to prevent any contamination thereof.
- the rotor is rotated in a clockwise (CW) direction within the stator housing of the heat transfer engine.
- the liquid refrigerant within the spiraled passageway of the shaft begins to flow into the secondary heat transfer (i.e., exchanging) portion of the rotor and occupies substantially the entire volume thereof.
- throttling of liquid refrigerant into vapor refrigerant begins to occur across the throttling channel bore 90 inside the rotor.
- the rotor continues to increase in angular velocity in the CW direction.
- the Liquid Pressurization Length of the refrigerant begins to create enough pressure within the secondary heat transfer portion of the rotor to overcome the pressure restriction presented by the throttling channel, and thus liquid refrigerant begins to flow into the primary heat transfer portion of the rotor.
- the Liquid Seal has moved along the rotor shaft towards the secondary heat transfer portion of the rotor thereof.
- refrigerant beyond the throttling channel and extending into about the first spiral of fluid flow passageway within the primary heat transfer portion is in the form of a homogeneous fluid (i.e., a mixture of refrigerant in both its liquid and vapor state) .
- the homogeneous fluid entering the primary heat transfer portion of the rotor "displaces" the gaseous refrigerant therewithin, thereby pushing it downstream into the spiraled passageway of the rotor shaft.
- Sufficient throttling of liquid refrigerant into vapor occurs causing a sufficient temperature drop in the primary heat transfer portion of the rotor and thus causing transfer of Superheat into the gaseous refrigerant.
- Refrigerant gas and vapor are compressed between (i) the homogeneous fluid in the primary heat transfer portion and (ii) the Liquid Seal formed along the spiraled fluid flow passageway of the rotor shaft.
- a Liquid Line is formed in where the homogeneous fluid ends and the vapor begins along the length of the primary heat transfer portion.
- the Liquid Line can occupy (i.e., manifest itself along) a short length of the primary heat transfer portion as a mixture of homogeneous fluid and a very dense vapor which extends downstream to the Superheat Length.
- the exact location of the Liquid Line along the primary heat transfer portion of the rotor will vary depending on the quantity of homogeneous fluid therein, which will be proportional to the amount of heat being absorbed and the thermal load imposed on the primary heat transfer portion of the rotor.
- the direction of heat flow (i.e., in the form of Superheat) is from the primary heat transfer portion of the rotor to the secondary heat transfer portion thereof.
- this heat flow is trapped behind the Liquid Seal formed along the spiraled passageway of the rotor shaft.
- the quantity of refrigerant vapor within the primary heat transfer portion of the rotor continues to increase due to the increased production of flash gas from throttling of liquid refrigerant across the throttling channel.
- the Liquid Seal has moved towards the end of the rotor shaft and the secondary heat transfer portion inlet thereof. Also, the flow of heat (i.e., in the form of Superheat) from the primary heat transfer portion is still trapped behind the Liquid Seal in the spiraled passageway of the rotor shaft. Consequently, the Superheat from the primary heat transfer portion of the rotor is unable to pass onto the secondary heat transfer portion of the rotor. Consequently, optimal operation is not yet achieved at this stage of engine operation.
- some heat may transfer into the rotor shaft from the refrigerant vapor if the shaft temperature is less that the temperature of the refrigerant vapor; and some heat may transfer into the refrigerant vapor if the refrigerant vapor temperature is less than that of the rotor shaft.
- the rotor shaft and its internal spiraled passageway provide primary and secondary Superheat transfer surfaces where heat can be either absorbed into or discharged from the vapor stream circulating within the closed fluid flow circuit of the rotor.
- Heat produced by friction from the rotor shaft bearings is absorbed by the refrigerant vapor along the length of the rotor shaft and can add to the amount of Superheat entering the secondary heat transfer portion.
- This additional Superheat further increases the temperature difference between the Superheated vapor and the secondary heat transfer surfaces of the secondary heat transfer portion. In turn, this increases the rate of heat flow from the Superheated vapor within the rotor, and thus enhances the heat transfer locations required to achieve steady state operation.
- Vapor compression has begun to occur in the tail end of the primary heat transfer portion and along the spiraled passageway of the rotor.
- the pressure of the liquid refrigerant along the Liquid Pressurization Length has increased sufficiently enough to further increase the production of homogeneous fluid in the primary heat transfer portion of the rotor.
- This also causes the quantity of liquid in the secondary heat transfer portion to decrease the "Pulling Effect" on the flash gas and vapor located in the spiraled passageway in the rotor shaft, as well as in the primary heat transfer portion of the rotor downstream from the homogeneous fluid.
- the pulling affect on the flash gas enhances vapor compression taking place along the spiraled passageway of the rotor shaft.
- the homogeneous fluid is evaporating absorbing heat within the primary heat transfer portion of the rotor for transference and systematic discharge from the secondary heat transfer portion.
- the vapor within the primary heat transfer portion of the rotor can contain more Superheat by volume than the gas with which it is mixed.
- the increased volume in dense vapor in the primary heat transfer portion provides a means of storing Superheat (absorbed from the primary heat exchanging circuit) until the vapor stream flows into the secondary heat transfer portion of the rotor where it can be liberated to the secondary heat exchanging circuit by way of conduction.
- the heat transfer engine of the present invention is shown operating at what shall be called the "Balance Point Condition" (i.e. , steady-state condition) .
- the refrigerant within the rotor has attained the necessary phase distribution where simultaneously there is an equal amount of refrigerant being evaporated in the primary heat transfer portion as there is refrigerant vapor being condensed in the secondary heat transfer portion of the rotor.
- the heat transfer engine is operating along the linear portion of its operating characteristic, shown in Fig. 9.
- 16A through 16F begins to dissipate from the DeSuperheat Length of the refrigerant stream along the secondary heat transfer portion of the rotor.
- the density of the refrigerant gas increases while vapor compression occurs as a result of Superheat being carried by the refrigerant gas from the Superheat Length along the primary heat transfer portion to the DeSuperheat Length along the secondary heat transfer portion via the spiraled passageway of the rotor shaft.
- the Superheat is dissipated in the secondary heat transfer portion of the rotor and compressed vapor in the secondary heat transfer portion thereof begins to condense into liquid refrigerant, a denser vapor remains.
- the spiraled passageway of the rotor shaft has a greater compressive affect on the vapor therein at this stage of operation.
- the spiraled passageway of the shaft pressurizes the superheated gas and dense vapor against the Liquid Seal formed in the secondary heat transfer portion of the rotor.
- pressurization of liquid refrigerant in the secondary heat transfer portion of the rotor pushes the liquid refrigerant through the throttling device at a sufficiently higher pressure, which causes a portion of the liquid refrigerant to "flash" into a gas.
- the liquid refrigerant portion of the homogeneous fluid evaporates creating sufficient vapor pressure therein which displaces vapor downstream within the primary heat transfer portion, into the spiraled passageway of the rotor shaft.
- the Liquid Seal tends to remain near the same location in the secondary heat transfer portion of the rotor, while the Liquid Line tends to remain near the same location in the primary heat transfer portion thereof.
- the temperature and pressure of the refrigerant in the secondary heat transfer portion of the rotor is higher than the refrigerant in the primary heat transfer portion thereof.
- the rate of heat transfer from the primary heat exchanging chamber of the engine into the primary heat transfer portion thereof is substantially equal to the rate of heat transfer from the secondary heat transfer portion of the engine into the secondary heat exchanging chamber thereof.
- a heat transfer system according to the present invention is shown, wherein the rotor of the heat transfer engine thereof 70 is driven (i.e., torqued) by fluid flow streams 95A flowing through the secondary heat exchanging circuit 95B of the system.
- heat liberated from the secondary heat exchanging portion 94 of the rotor is absorbed by a fluid 95A from pump 97A and a typical condenser cooling tower 97.
- cooling tower 97 is part of systematic fluid flow circuit in a cooling tower piping system where heat is exchanged with the cooling tower and consequently with the ambient atmosphere.
- the heat transfer engine 70 is "pumping" a fluid 96A, such as water, through a typical closed-loop tube and shell heat exchanger 98 and its associated piping 96B and flow control valve 98A.
- This heat transfer system is ideal for use in chilled-water air conditioning systems as well as process-water cooling systems .
- the fluid flow rate controller in primary heat exchanging circuit 96B is realized as a flow control valve 98A which receives primary heat exchanging fluid 96A by way of the primary heat exchanging portion 93 of the heat exchanging engine 70.
- the system controller 11 generates suitable signals to control the operation of the flow control valves (i.e., by adjusting the valve flow aperture diameter during engine operation) .
- the secondary fluid flow rate controller is realized as a flow rate control valve 97B designed for controlled operation under the control of system controller 11. ;
- Fig. 18 a modified embodiment of heat transfer system of Fig. 17 is shown.
- FIG. 19 another embodiment of a heat transfer system according to the present invention is shown, wherein two (or more) turbine-like heat transfer engines 125 and 127 are connected in a cascaded manner.
- the primary heat transfer portion of heat transfer engine 125 is in thermal communication with the secondary heat transfer portion of heat transfer portion 127, while the primary heat transfer portion of the rotor of engine 127 is in thermal communication with a closed chilled water loop flowing through the primary heat exchanging chamber thereof, and the secondary heat transfer portion of the rotor of engine 125 is in thermal communication with a closed process-water loop flowing through the secondary heat exchanging chamber thereof.
- the rotor of heat transfer engine 125 is driven by electric motor 126 coupled there by way of a first torque converter, while the rotor of heat transfer engine 127 is driven by electric motor 128 coupled therebetween by way of a second torque converter.
- a hybrid- type heat transfer engine has a secondary heat transfer portion 129 adapted from the heat transfer engine of the first embodiment and a secondary heat transfer portion 130 adapted from the heat transfer engine of the second embodiment.
- the function of the primary heat transfer portion is to serve as an air cooled condenser, whereas the function of the secondary heat transfer portion is to serve as an evaporator in a closed-loop fluid chiller.
- rotational torque is imparted to the rotor of the hybrid engine by allowing fluid to flow over the primary heat transfer vanes of the primary heat transfer portion 130 thereof.
- the hybrid- type heat transfer engine has a secondary heat transfer portion 129 adapted from the heat transfer engine of the first embodiment and a secondary heat transfer portion 130 adapted from the heat transfer engine of the second embodiment.
- the function of the primary heat transfer portion is to serve as a ga or air conditioning evaporator, whereas the function of the secondary heat transfer portion is to serve as a condenser in an open loop fluid cooled condenser. As shown in Fig.
- rotational torque is imparted to the rotor of the hybrid engine by an electric motor 134 connector to the rotor shaft 135 by a magnetic torque converter 133, whereas allowing fluid to flow over the primary heat transfer vanes of the primary heat transfer portion 130 thereof.
- a heat transfer engine of the present invention IS embodied within an automobile.
- the rotor of the heat transfer engine is rotated by an electric motor driven by electrical power which is supplied through a power control circuit, and produced by the automobile battery that is recharged by an alternator within the engine compartment of the automobile.
- a heat transfer engine of the present invention is embodied within a refrigerated tractor trailer truck.
- the rotor of the heat transfer engine is rotated by an electric motor driven by electrical power which is supplied through a power control circuit and produced by a bank of batteries recharged by an alternator within the engine compartment of the truck.
- a plurality of heat transfer engines of the present * invention are embodied within an aircraft.
- the rotor of each heat transfer engine is rotated by an electric motor.
- the electric motor is driven by electrical power which is produced by an onboard electric generator and supplied to the electric motors through voltage regulator and temperature control circuit.
- a plurality of heat transfer engines of the present invention are embodied within a refrigerated freight train.
- the rotor of each heat transfer engine is rotated by an electric motor driven by electrical power.
- the electric power is produced by an onboard pneumatically driven electric generator, and is supplied to the electric motors through a voltage regulator and temperature control circuit.
- a plurality of heat transfer engines of the present invention are embodied within a refrigerated shipping vessel.
- the rotor of each heat transfer engine is rotated by an electric motor driven by electrical power.
- the electric power is produced by an onboard pneumatically driven electric generator, and is supplied to the electric motors through a voltage regulator and temperature control circuit.
- the spiraled return passageway of the closed fluid-carrying circuit is realized along the support shaft of the rotor structure.
- the tubing associated with spiraled return passageway can be wrapped about the support shaft 29 along the axis of rotation of the rotor structure in a direction consistent with the direction that the spiraled passageway would normally extend about the axis of rotation of the rotor to achieve self-propelled circulation of the heat carrying fluid through the rotor structure.
- bidirectional metering device 38 can be incorporated into externally-wrapped spiraled return passageway as shown in Fig. 27A.
- bi- directional metering device 38 can be eliminated from the spiraled return passageway.
- the number of turns of tubing of the spiraled suction return passageway is only restricted by the diameter of the tubing and the length of the support shaft 29.
- the diameter of the support shaft can be any diameter suitable to the application at hand.
- Figs. 27A and 27B are not limited to rotor structures realized using coils of tubing as shown, for example, in Fig. 1. Rather, the closed fluid circuit design shown in Fig. 27A can also be realized within a solid-body structure constructed from machined discs, as shown in Fig. 15A, using a split housing design, or by any other suitable construction technique available to those skilled in the art.
- Such alternative self-circulating fluid circuit designs can be realized in various types of rotor structures, and such rotor structures can be rotatably supported by diverse types of stator structures, as will become apparent to those skilled in the art having had the benefit of reviewing technical disclosure hereof.
- each embodiment is designed using 3-D computer workstation having 3-D geometrical modelling capabilities, as well as mathematical modelling tools to develop mathematical models of each engine hereof using equation of energy, equations of motion and the like, well known in the fluid dynamics and thermodynamics art.
- simulation of proposed system designs can be carried out on the computer workstation, performance criteria established, and design parameters modified to achieve optimal heat transfer engine designs based on the principles of the present invention disclosed herein.
- cooling or heating fluid e.g., air
- air cooling or heating fluid
- the air flow can be easily directed over the primary heat exchanging portion of the rotor in order to condense moisture in the air stream, and thereafter directed over the secondary heat exchange portion of the rotor in order to re-heat the air for redistribution (reentry) into the conditioned space associated with the primary heat exchanging fluid circuit.
- both the coiled heat transfer engine and the embedded-coil (i.e., turbine line) heat transfer engine turbine of the present invention can be cascaded is various ways, utilizing various refrigerants and fluids, for various capacity and operating temperature requirements.
- Digital or analog type temperature and pressure sensors may be used to realize the system controllers of such embodiments.
- electrical, pneumatic, and/or hydraulic control structures can also be can be.used to realize such embodiments of the present invention.
- the heat transfer engines hereof can be readily modified to operate with heat carrying fluids that transfer heat between the primary and secondary portions of the rotor without undergoing phase-transformation.
- the bi-directional trotting device (of the various illustrated embodiments) can be removed and replaced with a non-restrictive tubing or conduit section, thereby enabling the heat carrying medium to flow between the primary and secondary heat transfer portions of the rotor without experiencing a pressure increase (or decrease) otherwise required for phase- transformation in vapor-compression type refrigeration cycles.
- modified heat transfer engines of the present invention are simple to manufacture, operate and repair and are capable of reliably carrying out transfer (and exchanging) functions in diverse types of systems.
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- Engineering & Computer Science (AREA)
- Physics & Mathematics (AREA)
- Mechanical Engineering (AREA)
- Thermal Sciences (AREA)
- General Engineering & Computer Science (AREA)
- Engine Equipment That Uses Special Cycles (AREA)
- Heat-Exchange Devices With Radiators And Conduit Assemblies (AREA)
- Centrifugal Separators (AREA)
Abstract
Description
Claims
Priority Applications (3)
Application Number | Priority Date | Filing Date | Title |
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EP97945340A EP1012508A4 (en) | 1996-10-01 | 1997-09-30 | Centrifugal heat transfer engine and system |
AU46567/97A AU4656797A (en) | 1996-10-01 | 1997-09-30 | Centrifugal heat transfer engine and system |
CA002270987A CA2270987C (en) | 1996-10-01 | 1997-09-30 | Centrifugal heat transfer engine and system |
Applications Claiming Priority (2)
Application Number | Priority Date | Filing Date | Title |
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US08/725,648 US5906108A (en) | 1992-06-12 | 1996-10-01 | Centrifugal heat transfer engine and heat transfer system embodying the same |
US08/725,648 | 1996-10-01 |
Publications (2)
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WO1998014738A1 true WO1998014738A1 (en) | 1998-04-09 |
WO1998014738A9 WO1998014738A9 (en) | 1998-11-05 |
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PCT/US1997/017482 WO1998014738A1 (en) | 1996-10-01 | 1997-09-30 | Centrifugal heat transfer engine and system |
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US (8) | US5906108A (en) |
EP (1) | EP1012508A4 (en) |
AU (1) | AU4656797A (en) |
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WO (1) | WO1998014738A1 (en) |
Cited By (1)
Publication number | Priority date | Publication date | Assignee | Title |
---|---|---|---|---|
US10030961B2 (en) | 2015-11-27 | 2018-07-24 | General Electric Company | Gap measuring device |
Families Citing this family (51)
Publication number | Priority date | Publication date | Assignee | Title |
---|---|---|---|---|
US5906108A (en) * | 1992-06-12 | 1999-05-25 | Kidwell Environmental, Ltd., Inc. | Centrifugal heat transfer engine and heat transfer system embodying the same |
US6964176B2 (en) | 1992-06-12 | 2005-11-15 | Kelix Heat Transfer Systems, Llc | Centrifugal heat transfer engine and heat transfer systems embodying the same |
US6505475B1 (en) | 1999-08-20 | 2003-01-14 | Hudson Technologies Inc. | Method and apparatus for measuring and improving efficiency in refrigeration systems |
US7086242B2 (en) * | 2001-07-13 | 2006-08-08 | Ebara Corporation | Dehumidifying air-conditioning apparatus |
US20050137251A1 (en) * | 2002-03-18 | 2005-06-23 | Aaron Garzon | Dexanabinol and dexanabinol analogs regulate inflammation related genes |
US7010936B2 (en) * | 2002-09-24 | 2006-03-14 | Rini Technologies, Inc. | Method and apparatus for highly efficient compact vapor compression cooling |
US8463441B2 (en) | 2002-12-09 | 2013-06-11 | Hudson Technologies, Inc. | Method and apparatus for optimizing refrigeration systems |
US6829833B2 (en) * | 2003-03-14 | 2004-12-14 | Thomas Langman | Tool guide |
ITMI20031021A1 (en) * | 2003-05-21 | 2004-11-22 | Whirlpool Co | REFRIGERATOR WITH VARIABLE DIMENSION EVAPORATOR. |
US7491037B2 (en) * | 2005-08-05 | 2009-02-17 | Edwards Thomas C | Reversible valving system for use in pumps and compressing devices |
GB2436075B (en) * | 2006-03-17 | 2009-04-15 | Genevac Ltd | Evaporator and method of operation thereof |
US20070271938A1 (en) * | 2006-05-26 | 2007-11-29 | Johnson Controls Technology Company | Automated inlet steam supply valve controls for a steam turbine powered chiller unit |
US7637031B2 (en) * | 2007-06-26 | 2009-12-29 | Gm Global Technology Operations, Inc. | Evaporator core drying system |
AT505532B1 (en) * | 2007-07-31 | 2010-08-15 | Adler Bernhard | METHOD FOR THE CONVERSION OF THERMAL ENERGY OF LOW TEMPERATURE IN THERMAL ENERGY OF HIGHER TEMPERATURE BY MEANS OF MECHANICAL ENERGY AND VICE VERSA |
US20100307156A1 (en) | 2009-06-04 | 2010-12-09 | Bollinger Benjamin R | Systems and Methods for Improving Drivetrain Efficiency for Compressed Gas Energy Storage and Recovery Systems |
US8225606B2 (en) | 2008-04-09 | 2012-07-24 | Sustainx, Inc. | Systems and methods for energy storage and recovery using rapid isothermal gas expansion and compression |
US8479505B2 (en) | 2008-04-09 | 2013-07-09 | Sustainx, Inc. | Systems and methods for reducing dead volume in compressed-gas energy storage systems |
US8240140B2 (en) | 2008-04-09 | 2012-08-14 | Sustainx, Inc. | High-efficiency energy-conversion based on fluid expansion and compression |
US8359856B2 (en) | 2008-04-09 | 2013-01-29 | Sustainx Inc. | Systems and methods for efficient pumping of high-pressure fluids for energy storage and recovery |
US7802426B2 (en) | 2008-06-09 | 2010-09-28 | Sustainx, Inc. | System and method for rapid isothermal gas expansion and compression for energy storage |
US8250863B2 (en) | 2008-04-09 | 2012-08-28 | Sustainx, Inc. | Heat exchange with compressed gas in energy-storage systems |
US8037678B2 (en) | 2009-09-11 | 2011-10-18 | Sustainx, Inc. | Energy storage and generation systems and methods using coupled cylinder assemblies |
US8474255B2 (en) | 2008-04-09 | 2013-07-02 | Sustainx, Inc. | Forming liquid sprays in compressed-gas energy storage systems for effective heat exchange |
WO2009126784A2 (en) | 2008-04-09 | 2009-10-15 | Sustainx, Inc. | Systems and methods for energy storage and recovery using compressed gas |
US7958731B2 (en) | 2009-01-20 | 2011-06-14 | Sustainx, Inc. | Systems and methods for combined thermal and compressed gas energy conversion systems |
US8448433B2 (en) | 2008-04-09 | 2013-05-28 | Sustainx, Inc. | Systems and methods for energy storage and recovery using gas expansion and compression |
US8677744B2 (en) | 2008-04-09 | 2014-03-25 | SustaioX, Inc. | Fluid circulation in energy storage and recovery systems |
WO2010105155A2 (en) | 2009-03-12 | 2010-09-16 | Sustainx, Inc. | Systems and methods for improving drivetrain efficiency for compressed gas energy storage |
JP2012528296A (en) * | 2009-05-29 | 2012-11-12 | パーカー−ハニフイン・コーポレーシヨン | Vapor compression cooling system with pumped loop drive |
US8104274B2 (en) | 2009-06-04 | 2012-01-31 | Sustainx, Inc. | Increased power in compressed-gas energy storage and recovery |
US8452459B2 (en) * | 2009-08-31 | 2013-05-28 | Fisher-Rosemount Systems, Inc. | Heat exchange network heat recovery optimization in a process plant |
US20110061832A1 (en) * | 2009-09-17 | 2011-03-17 | Albertson Luther D | Ground-to-air heat pump system |
WO2011056855A1 (en) | 2009-11-03 | 2011-05-12 | Sustainx, Inc. | Systems and methods for compressed-gas energy storage using coupled cylinder assemblies |
US8171728B2 (en) | 2010-04-08 | 2012-05-08 | Sustainx, Inc. | High-efficiency liquid heat exchange in compressed-gas energy storage systems |
US8191362B2 (en) | 2010-04-08 | 2012-06-05 | Sustainx, Inc. | Systems and methods for reducing dead volume in compressed-gas energy storage systems |
US8234863B2 (en) | 2010-05-14 | 2012-08-07 | Sustainx, Inc. | Forming liquid sprays in compressed-gas energy storage systems for effective heat exchange |
US8495872B2 (en) | 2010-08-20 | 2013-07-30 | Sustainx, Inc. | Energy storage and recovery utilizing low-pressure thermal conditioning for heat exchange with high-pressure gas |
US8754558B2 (en) | 2010-10-06 | 2014-06-17 | Ramiro Casas | Kinetic energy to electric power converter |
US9270149B1 (en) * | 2010-10-06 | 2016-02-23 | Ramiro Casas | Kinetic energy to electric power converter |
US8578708B2 (en) | 2010-11-30 | 2013-11-12 | Sustainx, Inc. | Fluid-flow control in energy storage and recovery systems |
WO2012158781A2 (en) | 2011-05-17 | 2012-11-22 | Sustainx, Inc. | Systems and methods for efficient two-phase heat transfer in compressed-air energy storage systems |
US8935933B1 (en) * | 2011-07-14 | 2015-01-20 | Ronald Koelsch | Battery operated transfer refrigeration unit |
US20130091836A1 (en) | 2011-10-14 | 2013-04-18 | Sustainx, Inc. | Dead-volume management in compressed-gas energy storage and recovery systems |
DE102012203695A1 (en) * | 2012-03-08 | 2013-09-12 | Siemens Aktiengesellschaft | Electric machine with a dual-circuit cooling |
WO2013155491A1 (en) * | 2012-04-12 | 2013-10-17 | Lightsail Energy Inc. | Compressed gas energy storage system |
US9242525B2 (en) * | 2013-09-30 | 2016-01-26 | Herbert S Kobayashi | Rotating air conditioner and method |
JP5747968B2 (en) | 2013-10-07 | 2015-07-15 | ダイキン工業株式会社 | Heat recovery type refrigeration system |
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US10578323B2 (en) | 2017-03-22 | 2020-03-03 | General Electric Company | Systems for dehumidifying air and methods of assembling the same |
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Citations (5)
Publication number | Priority date | Publication date | Assignee | Title |
---|---|---|---|---|
US348061A (en) | 1886-08-24 | schupp | ||
US2813698A (en) * | 1954-06-23 | 1957-11-19 | Roland L Lincoln | Heat exchanger |
US3025684A (en) * | 1959-06-23 | 1962-03-20 | Robert S Mclain | Refrigerating machine |
US3397739A (en) * | 1964-05-18 | 1968-08-20 | Sibany Mfg Corp | Heat exchange apparatus |
US5493868A (en) * | 1993-11-09 | 1996-02-27 | Sanyo Electric Co., Ltd. | Air conditioning apparatus usable for wide-range source voltage |
Family Cites Families (42)
Publication number | Priority date | Publication date | Assignee | Title |
---|---|---|---|---|
US1315282A (en) | 1919-09-09 | Howabb h | ||
US541575A (en) * | 1895-06-25 | Machine | ||
US182528A (en) * | 1876-09-26 | Improvement in carpet-stretchers | ||
US503611A (en) | 1893-08-22 | Siegfried marcus | ||
US1063636A (en) | 1911-10-13 | 1913-06-03 | Gen Electric | Method of and apparatus for forging or compressing fluids. |
US1145226A (en) | 1914-05-19 | 1915-07-06 | John C Bertsch | Rotary refrigerating apparatus. |
US1223919A (en) | 1914-08-05 | 1917-04-24 | Walter J Wilson | Rotary pump. |
US1352107A (en) | 1915-08-11 | 1920-09-07 | James H Wagenhorst | Pump or compressor |
US1204061A (en) | 1915-10-07 | 1916-11-07 | Bernard Plekenpol | Refrigerating apparatus. |
US1446727A (en) | 1919-01-25 | 1923-02-27 | Laurence K Marshall | Refrigerating apparatus |
US1589373A (en) * | 1919-11-11 | 1926-06-22 | Savage De Remer Corp | Refrigerating apparatus |
US1537937A (en) * | 1921-02-23 | 1925-05-19 | Savage De Remer Corp | Refrigerating apparatus |
GB182528A (en) * | 1921-03-24 | 1922-06-26 | Jay Grant Deremer | Refrigerating apparatus |
US1635523A (en) | 1926-03-22 | 1927-07-12 | Nat Pump & Compressor Company | Compressor |
US1889817A (en) | 1927-10-28 | 1932-12-06 | Audiffren Marcel | Rotary refrigerating machine |
DE541575C (en) | 1930-04-26 | 1932-01-13 | Bbc Brown Boveri & Cie | Compressor without stuffing box, especially for refrigeration machines |
US1969999A (en) | 1931-03-30 | 1934-08-14 | Joseph W Cuthbert | Compressor unit |
US2156628A (en) | 1936-04-30 | 1939-05-02 | Siemens Ag | Compression refrigerating apparatus |
US2111750A (en) | 1937-02-13 | 1938-03-22 | John F Carlson | Air conditioning machine |
US2229500A (en) | 1938-12-21 | 1941-01-21 | Bertram J Goldsmith | Refrigeration system and apparatus |
US2331878A (en) | 1939-05-25 | 1943-10-19 | Wentworth And Hull | Vane pump |
US2324434A (en) | 1940-03-29 | 1943-07-13 | William E Shore | Refrigerant compressor |
US2333208A (en) * | 1942-10-01 | 1943-11-02 | George T Spear | Drainpipe conveyer |
US2522781A (en) | 1946-06-06 | 1950-09-19 | Exner Hellmuth Alfredo Arturo | Centrifugal refrigerating machine |
US2440593A (en) | 1946-10-23 | 1948-04-27 | Harry B Miller | Radial vane pump mechanism |
US2670894A (en) | 1950-10-20 | 1954-03-02 | Borg Warner | Compressor |
US2609672A (en) | 1951-05-04 | 1952-09-09 | Ind Patent Corp | Unitized centrifugal refrigerating machine |
US2643817A (en) | 1952-11-22 | 1953-06-30 | Vadim S Makaroff | Compressor |
US2811841A (en) | 1953-11-13 | 1957-11-05 | Gen Electric | Refrigerator apparatus |
US2805558A (en) | 1954-12-20 | 1957-09-10 | Gen Electric | Refrigerating apparatus including rotating heat exchangers |
US2969743A (en) | 1956-12-01 | 1961-01-31 | Emanuel Di Giuseppe E Roberto | Rotary slidable-vane machines |
US3001384A (en) | 1957-06-14 | 1961-09-26 | William H Anderson | Space coolers |
US2969021A (en) | 1958-04-16 | 1961-01-24 | Acc Emanuel Di G E R Emanuel & | Automatic device for adjusting the output of rotary hydraulic machines |
US3026021A (en) | 1960-03-31 | 1962-03-20 | Acc Emanuel Di G E R Emanuel & | Slidable vane rotary compressor |
US3025694A (en) * | 1960-09-19 | 1962-03-20 | Harry F George | Latching and locking mechanism |
US3189262A (en) | 1961-04-10 | 1965-06-15 | William H Anderson | Space coolers |
US3098602A (en) | 1962-09-12 | 1963-07-23 | Keith R Torluemke | Thermal centrifugal compressor |
US3948061A (en) * | 1974-10-29 | 1976-04-06 | George B. Vest | Centrifugal refrigeration unit |
FR2333208A1 (en) * | 1975-11-28 | 1977-06-24 | Europ Propulsion | Domestic or industrial building heater - effecting mechanical operations necessary for thermodynamic cycle almost reversibly, with high efficiency |
US5168726A (en) * | 1991-08-21 | 1992-12-08 | York Charles L | Centrifugal refrigeration system |
US6964176B2 (en) * | 1992-06-12 | 2005-11-15 | Kelix Heat Transfer Systems, Llc | Centrifugal heat transfer engine and heat transfer systems embodying the same |
US5906108A (en) * | 1992-06-12 | 1999-05-25 | Kidwell Environmental, Ltd., Inc. | Centrifugal heat transfer engine and heat transfer system embodying the same |
-
1996
- 1996-10-01 US US08/725,648 patent/US5906108A/en not_active Expired - Fee Related
-
1997
- 1997-09-30 WO PCT/US1997/017482 patent/WO1998014738A1/en not_active Application Discontinuation
- 1997-09-30 AU AU46567/97A patent/AU4656797A/en not_active Abandoned
- 1997-09-30 EP EP97945340A patent/EP1012508A4/en not_active Withdrawn
- 1997-09-30 CA CA002270987A patent/CA2270987C/en not_active Expired - Fee Related
-
1999
- 1999-05-21 US US09/317,142 patent/US6321547B1/en not_active Expired - Fee Related
- 1999-05-24 US US09/317,055 patent/US6334323B1/en not_active Expired - Fee Related
-
2001
- 2001-08-03 US US09/922,214 patent/US20020092316A1/en not_active Abandoned
-
2002
- 2002-10-04 US US10/265,652 patent/US7010929B2/en not_active Expired - Fee Related
-
2003
- 2003-02-18 US US10/370,035 patent/US6948328B2/en not_active Expired - Fee Related
- 2003-02-25 US US10/374,763 patent/US7093454B2/en not_active Expired - Fee Related
-
2006
- 2006-08-14 US US11/503,855 patent/US20070144192A1/en not_active Abandoned
Patent Citations (5)
Publication number | Priority date | Publication date | Assignee | Title |
---|---|---|---|---|
US348061A (en) | 1886-08-24 | schupp | ||
US2813698A (en) * | 1954-06-23 | 1957-11-19 | Roland L Lincoln | Heat exchanger |
US3025684A (en) * | 1959-06-23 | 1962-03-20 | Robert S Mclain | Refrigerating machine |
US3397739A (en) * | 1964-05-18 | 1968-08-20 | Sibany Mfg Corp | Heat exchange apparatus |
US5493868A (en) * | 1993-11-09 | 1996-02-27 | Sanyo Electric Co., Ltd. | Air conditioning apparatus usable for wide-range source voltage |
Non-Patent Citations (1)
Title |
---|
See also references of EP1012508A4 |
Cited By (1)
Publication number | Priority date | Publication date | Assignee | Title |
---|---|---|---|---|
US10030961B2 (en) | 2015-11-27 | 2018-07-24 | General Electric Company | Gap measuring device |
Also Published As
Publication number | Publication date |
---|---|
EP1012508A1 (en) | 2000-06-28 |
US6334323B1 (en) | 2002-01-01 |
US20040040324A1 (en) | 2004-03-04 |
CA2270987C (en) | 2004-04-27 |
US6948328B2 (en) | 2005-09-27 |
US20030217566A1 (en) | 2003-11-27 |
CA2270987A1 (en) | 1998-04-09 |
US20030145616A1 (en) | 2003-08-07 |
EP1012508A4 (en) | 2001-10-24 |
AU4656797A (en) | 1998-04-24 |
US20070144192A1 (en) | 2007-06-28 |
US20020092316A1 (en) | 2002-07-18 |
US5906108A (en) | 1999-05-25 |
US6321547B1 (en) | 2001-11-27 |
US7010929B2 (en) | 2006-03-14 |
US7093454B2 (en) | 2006-08-22 |
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