GB2309492A - Heat engine with liquid working fluid, and constant volume, constant entropy, and constant pressure cycle stages - Google Patents

Abstract

A working fluid, which may be carbon dioxide, is heated at constant volume (isochorically). It may then expand at constant entropy, and be cooled at constant pressure. The fluid may be substantially saturated liquid before it is heated. Its heat may be transferred to or from heat pumps, which may use ammonia as working fluid. The working fluid may be contained in three working chambers. The or each chamber may have a moveable boundary, and releasable locking means therefor. Said boundary may be a piston. The engine may be used in power generating plant, and may use low grade - e.g., geothermal or waste - heat. The figure shows three cylinders 52, 53, and 54, and a heat pump 39 comprising compressor 40, evaporator 43, condenser 44, and pressure drop valve 42, 45 is a heat input; 46 a heat output. 51 and 52 are standpipes or equalization reservoirs.

Description

A HEAT ENGINE The present invention relates to a heat engine suitable for the exploitation of low grade (i.e. low temperature) heat energy from readily accessible and therefore cost effective sources.

Possible sources of low grade heat include geothermal energy and waste heat from power generating plant such as gas turbines. A heat source of less than about 3000C can be described as a low grade heat source although an embodiment of the invention is particularly suitable for heat sources at even lower temperatures such as geothermal heat sources below 1000C.

Any attempt to do this encounters two major problems. The first of which is the thermodynamic efficiency limitation imposed upon engines operating over narrow temperature ranges. Consider that ((Th - Tl) - Th} = ((Qi - Qo)- Qi} = Carnot thermal efficiency and further that {(Qo/Tl) - (Qi/Th)} has to be unity (where Th is the upper temperature, Tl is the lower temperature, Qi is the heat in and Qo is the heat out).

From this it is clear that heat engines operating over narrow temperature ranges will always exhibit a low or even very low thermodynamic efficiency. It should also be clear from the aforesaid chat these conditions cannot be avoided, which is why all schemes so far proposed have failed. Plant according to an embodiment of the invention allows for a cost effective exploitation of low grade heat.

The second major problem consists of a number of subproblems. The first is the problem of the heat source.

Geothermal energy is easily accessible in certain regions such as northern Italy with its steam springs. However, that steam is not of a high enough temperature nor pressure and it is also contaminated. In other regions boreholes are required which are at least 5 km deep and thus very expensive. Further, what can be obtained from these is often only hot water, or a mix of water and water vapour which is always contaminated and of low pressure.

Consequently in order to exploit geothermal energy it is desirable to have plant incorporating a cycle which allows not only the exploitation of heat sources at temperatures in the region of 25"C to 40 C; but which also makes the purity or composition of the heat source irrelevant for the actual power generating plant or heat engine.

The next variation on the second major problem refers to the second law of thermodynamics which clearly states that no heat source can be fully exploited in the sense of complete heat energy conversion to mechanical energy. In any such process some heat has to be rejected to a sink.

Since heat flows naturally only from hot to cold, it follows that the sink has to be of a lower temperature still.

Furthermore, not only must the above conditions be met, but there is also an additional requirement for all heat engines. This is that during the work phase of the cycle of the heat engine sufficient expansion of the working fluid must occur, to ensure movement of a working element such as a piston or turbine and thus exploitation of the energy imparted to the working fluid.

There is a final requirement, which is true for all engines, that all operations have to be cyclic, ie, an arrangement in which the machine in question returns to its original condition. For heat engines in particular it is a condition that the working fluid itself fits into a cycle, ie, that the working fluid has to return to its original state. If that is not the case, there cannot be an engine.

The present invention in its various aspects is defined in the appended claims to which reference should now be made.

Preferred embodiments of the invention will now be described, by way of example, with reference to the accompanying figures in which: Figure 1 is a block diagram for a heat engine suitable for use with a heat engine embodying the invention; Figure 2 is a block diagram illustrating the general layout of a second heat engine system layout; Figure 3 is a block diagram illustrating the general layout of a third heat engine layout; Figure 4 is a pressure-enthalpy diagram for a cycle suitable for the recovery of low grade heat; Figure 5 is a pressure-enthalpy diagram for an alternative cycle suitable for the recovery of low grade heat; Figure 6 is a diagrammatic representation of the elements required for exploiting the work done in the cycles of figures 4 and 5;; Figure 7 is a diagrammatic illustration of a cylinder/piston arrangement suitable for use with the cycles of figures 4 and 5 to drive a turbine using a hydraulic circuit Figure 8 is a block diagram illustrating a heat engine for exploiting the work done in the cycles of figures 4 and 5; Figure 9 is pressure-enthalpy diagram for a cycle suitable for use as the heat pump of the embodiment of the invention of figure 8; Figures 10 to 12 illustrate a connecting rod arrangement for use with the cylinder of figure 6 to drive a crankshaft; Figures 13 to 21 are a diagrammatic illustration of the operation of the cylinder and connecting rod/crankshaft arrangement of figures 6,10 to 12 in operation.

Tables 1 to 4 are vapour tables of the properties of carbon dioxide and ammonia reproduced by permission of the American Society of Heating, Refrigerating and Air Conditioning Engineers, Atlanta, Georgia from the 1989 ASHRAE Fundamentals Handbook (SI).

Figure 1 shows an equivalence box 4 symbolising an engine having a heat input 1 to which heat (Qin) is supplied from the source of low grade heat. The engine includes a work output 2 indicating work (Wout) taken from the engine, and a heat output 3 indicating heat (Qout) to be rejected.

Figure 2 shows a heat engine system 8 incorporating recovery of heat from the heat rejected by the heat engine. A heat engine equivalence box 64 having a heat input 11 to which heat (Qin) is supplied from the source of low grade heat, a work output 62 and a heat output 63.

A second equivalence box 5 represents a heat engine in reverse, ie a refrigerator or heat pump. The heat input for this is the heat output 63 (Qout) of the engine 64.

The heat output 6 of the heat pump 5 forms a second heat input to the heat engine 64 and supplies heat Qdelivered thereto. Thus in the plant layout shown in figure 2 the heat input required per unit work output is lower than in the plant of figure 1 due to the action of the heat pump 5 which recovers the rejected heat from the engine 64.

Although this plant arrangement reduces the amount of heat rejected from the system there is still, of course, some heat rejected. In the system of figure 2 the heat loss (Qout) from the heat outlet 3 of the system of figure 1, has now been replaced with a requirement for a work input 7 to the heat pump 5. An equivalence box 8 shown in dotted outline represents the overall engine corresponding to the combined heat engine and pump system.

Although no initial heat has been rejected to the environment, not all heat has been converted into work since the compressor of the heat pump 5 requires a work input. That means that the rejected heat loss (Qout) has been replaced with a Work input loss (Win). Since all mechanical work degrades ultimately back into heat, the entropy of the environment still increases correspondingly even though no immediately apparent heat has been rejected to the environment.

The recovery of waste heat using a heat pump provides a permanent low temperature heat sink for a narrow temperature range engine. That is to say that the heat sink is the evaporator of the heat pump. This results in a more efficient use of a low grade heat source by increasing the temperature difference between the heat source and the heat sink of the heat engine.

Figure 3 shows an alternative configuration of the plant elements of figure 2. The arrangement of figure 3 has the same elements as Figure 2 and consequently they have been given the same reference numbers, The main difference between the systems of figures 2 and 3 is that in figure 3 the heat input 61 from the heat source into the engine is indirect i.e. it is fed into the evaporator of the heat pump 5 which also receives the waste heat from the heat output 63 of the heat engine 64. Then the entire quantity of heat required by the engine is pumped by the refrigerator's compressor to the engine via the condenser of the refrigerator.

Thus the low grade heat source comes into contact with a low temperature heat receptacle which also forms the low temperature heat sink for heat rejected by the heat engine.

The temperature of the heat receptacle/sink (i.e. the evaporator of the heat pump 5) with which the heat source exchanges heat is lower than the temperature of the working fluid rejected from the engine 64. Using a heat pump to recover the waste heat (Qout) from the engine output 63 as well as to receive the heat input (Qin) provides a low temperature heat sink for the rejection of heat from the engine 64 of the engine system, and for the heat input 61. This provision of a low temperature heat sink increases the efficiency of recovery of heat from the low temperature heat source forming the heat input 61. A further advantage of this system layout is that it keeps the heat source medium separate from the working fluid of the heat engine 64 and thus eliminates contamination of the cycle working fluid.As discussed above this is particularly important for the exploitation of geothermal energy where the thermal fluid (ie heat source) may be a mixture of liquid and vapour which can cause degradation of the heat engine if used as the power generating working fluid. Furthermore, the use of a heat pump for this purpose is possible since such devices exhibit excellent efficiencies over narrow temperature ranges.

For ease of reference the description will from here describe the situation corresponding to the plant of Figure 3 as a "fullqin situation" or simply "fullqin".

This means that the entire quantity of heat to be provided to the power generating engine 64 is obtained via the evaporator to the compressor and then to the condenser of the heat pump. In other words, whenever "fullqin" is mentioned, this refers to not only the recovery of waste heat from the heat engine 64 via the evaporator of a heat pump 5, but also to the reception of heat from an external heat source through the evaporator.

Having dealt with the principles of waste heat recovery for the engine as well as that of waste heat recovery plus "fullqin" facility, reference will now be made to the cycle which forms the basis for the plant.

Figures 4 and 5 show a pressure-enthalpy diagram representing the cycle used for generating work from low grade heat. The dotted line 10 represents the saturated liquid line with the portion of the diagram above and to the left of the saturated liquid line 10 representing the purely liquid phase and the area below and to the right of saturated liquid line 10 representing the mixed vapour/liquid phase.

Heat (Qin) is supplied to the cycle working fluid to heat it at constant volume in the liquid phase from hl to h2.

The working fluid is then allowed to expand substantially isentropically (i.e. at constant entropy) from h2 to h3.

This isentropic expansion is the portion of the cycle which provides a work output. The working fluid is then cooled at constant pressure from h3 to hl and it is here that heat (Qout) is rejected.

In the cycle of figure 4 expansion occurs right into the vapour dome (the region in the pressure-enthalpy diagram which shows a mix of liquid and vapour, limited to the left by the saturated liquid line, to the right by the saturated vapour line and at the top by the critical point). In Figure 5 however expansion of the liquid ends before vaporisation. Otherwise both are identical cycles.

All sorts of working fluids can be used for the engine and for the heat pump. However, with the means available, the most suitable substances identified are carbon dioxide for the engine and ammonia for the heat pump. The latter due to its very high enthalpy during evaporation and condensation, the so called latent heat. Therefore, in order to avoid confusion about the two working fluids for the engine and its heat pump, the fluid for the engine will be referred to as carbon dioxide and that for the heat pump as ammonia unless the distinction is clear.

very high pressures are obtained at h2 as a result of the heating of the carbon dioxide in the liquid phase. Thus, the cycle can convert heat energy from a low grade heat source at small temperature difference into a large potential difference of pressure which can be exploited isentropically.

Note here that this cycle is asymmetric when compared with a Rankine type cycle in which both the work input and the work output (Win and Wout respectively) are both isentropic and the heat input and output (Qin and Qout respectively) are both isobaric.

The cycle of figures 4 and 5 has cycle stages which are successively isochoric, isentropic and isobaric.

To analyse this cycle consider the tables 1 to 3 of the thermodynamic properties of carbon dioxide as applied to a sample cycle having carbon dioxide as the working fluid and in which the carbon dioxide is heated from about 200C to about 500. The figures are based on experimental data produced and published by ASHRAE (American Society for Heating, Refrigeration and Air Conditioning Engineers);; hl = 552 kJ/kg T1 = 20.00 C = Tl h2 = 615 kJ/kg T2 = 53.44 C = Th h3 = 581.39 kJ/kg T3 = 26.00 C = Tout Plow = 65.8 Bar Phigh = 255 Bar (approximately) Th - Tl h2 - hl --- = ------- = p = 10.24% (Carnot efficiency) Th h2 hl/T1 = h2/Th The cycle is balanced Thermodynamically this efficiency is certainly mediocre, but consider that Heat input = Qin = (h2 - hl) = 63 kJ/kg, Work output = (h2 - h3) = 33.61 KJ/kg and Rejected heat = (h3 - hl) = 29.39 kJ/kg and therefore (Work - Qin) = 53.35% fuel efficiency.

It can therefore be seen that the cycle having an isochoric (ie at constant volume) heat input stage, an isentropic Work output stage, and an isobaric heat rejection stage for a working fluid in liquid state for at least most of the cycle results in an effective heat source exploitation via the conversion of low grade heat into high pressure, despite the severe thermodynamic efficiency restriction of operating with a low temperature difference.

The cycle described above results in a number of additional advantages due to the high density of a working fluid in liquid form.

The first is a very small volumetric requirement for even large quantities of working fluid per work production device.

The second is that a working fluid in liquid state possesses the highest possible thermal conductivity and in consequence relatively small heat exchangers are required.

The third advantage, and as far as work is concerned the most significant, is that employing large quantities of working fluid permits the effective use of small expansion ratios. This will be demonstrated for the carbon dioxide cycle discussed above.

Carbon dioxide in the cycle as described with the parameters mentioned has a density of 800 kg/m3 between hl and h2. At h3, the point of maximum expansion, the density has been reduced to 694.97 kg/m3. Thus the expansion ratio is (694.97-1 - 800-1) which equals 1.15, ie, 15%.

This may not seem much but consider this expansion ration with a cylinder and piston arrangement to extract work.

Let the cylinder dimension be 6m in diameter and the height be 7m. Thus the cylinders initial volume will be 197.92m3. With a density of 800 kg/m3 the cylinder will contain 158,366 kg of working fluid carbon dioxide. The expansion of 1.15:1 results in a final volume of 227.8m3, ie, displacement of 29.9m3 which of course results in a new cylinder height of 8.06m which means a piston stroke of (8.06m - 7m) i.e. 1.06m, which is more than sufficient to turn a crankshaft. This calculation assumes zero piston thickness.

The fourth advantage resulting from this is a very low piston speed (1.06m/s in the example above) which means extremely low friction losses and consequently negligible heat being added to the carbon dioxide. Thus the expansion during the work stage of the cycle is near adiabatic.

The fifth advantage resulting from the described cycle with a piston/cylinder arrangement to extract work is that upon the return stroke of the piston no work is lost due to piston friction, nor will work be lost by the piston having to work against or upon the working fluid.

The reason being that the piston returns to bottom dead centre (BDC) by virtue of the carbon dioxide in the cylinder contracting as it cools. Note that a pressure low of 65.8 Bar will cause problems if the back of the piston is exposed to ambient pressure. Steps to avoid this are discussed later in this document.

At this point in the discussion of the preferred embodiments it is important to consider the definition of work. So far work has been defined in thermodynamic terms as (h2 - h3). More definitions will be introduced in stages below , resulting in the term "finalwork" which is the work available after all losses. For now the term "realwork" will be defined and its importance in reference to the working fluid discussed.

Please note that upon expansion of the working fluid the friction of the piston will be converted into heat and this heat will be added to the working fluid which will therefore increase in enthalpy and thus in entropy. This of course results in reduced work output. So, to define realwork, it is ((h2 - h3) x (1 - piston friction)) where piston friction is expressed as a fraction.

Consequently the rejected quantity of heat increases and this will be called "realreject" which will be defined in the next paragraph.

The actual heat to be rejected is not (h3 - hl) but (real h3 - hl) = realreject.

Where real h3 = (h2 - realwork).

The reason for the importance of the above is that the cycle's effectiveness and therefore its value is derived from its (Work - fuel) or (Work - Qin) ratio and not its thermodynamic efficiency. Consequently there will be a final efficiency which will be defined as (finalwork t final Qin). Now, if a heat pump is used for waste heat recovery, then the greater the amount of waste heat taken into the heat pump's evaporator, the less the amount of heat required from the actual heat source. Thus the greater will be the efficiency of the heat to work conversion in terms of work done over heat required from an external source.

A brief analysis shall now be made of the basic engine of figure 1 embodying the invention without waste heat recovery or full Qin via a heat pump. That is to say that in addition to a natural heat source there will also be a natural heat sink. Figure 6 illustrates the cycle referred to in Figures 4 and 5. The positions hl, h2, and h3 are Tlow, Thigh and Tout respectively. A heat exchanger 31 is connected to the heat source and supplies the heat input Qin to the cycle working fluid.A second heat exchanger 34 is connected to the heat sink for rejecting heat Qout (h3 - hl). Heated fluid from the heat source is fed into the first heat exchanger 31 via an inlet 32 and, having heated the working fluid of the cycle and been cooled thereby, leaves the heat exchanger 32 via an outlet 33; ie the heat exchanger 31 transfers heat from the hot medium to the working fluid.

The temperature of the heat source thermal fluid (ie the heating fluid) at the inlet 32 is Thi and at the outlet 33 is Tho. The temperature of the cycle working fluid at the beginning of the constant volume heating (hl) is Tci = Tl while the temperature at the end of the constant volume heating (h2) is Tco = Th. Since T1 is the minimum temperature for the cycle working fluid and Th is the maximum temperature for the working fluid the Thi and Tho are be established in reference to the heat exchangers.

Considering the analysis of pages 10 and 11 it can be seen that there is a very large pressure difference between Phigh at 255 Bar and Plow at 65.8 Bar. The most effective type of heat exchanger able to withstand these high pressures is of the tube and shell design which has the advantage of being cheap, but the disadvantage of being limited to a maximum transfer effectiveness of 80 .

Assuming the effectiveness of the counter-current heat exchanger 31 to be 80% requires a heating medium the temperature of which has to be lowered from about 620C to 28 C in order to raise the temperature of carbon dioxide from 200C to 53.440C.

In the second counter-current heat exchanger 34 of Figure 6 the temperature of the carbon dioxide has to be lowered from 26"C ( as discussed above due to "realreject" resulting from system losses the actual temperature would be slightly greater and thus ease the task) to 20"C, ie from Tout to Tlow. This is done by rejecting the waste heat into a coolant which enters the second heat exchanger 34 via an inlet 35, and leaves the heat exchanger 34 via an outlet 36. Thus in the second cooling heat exchanger 34, the coolant inlet 35 has an input temperature Tci and the coolant output 336 has the output temperature Tco. Allowing for the heat exchangers effectiveness as 80%, it will be found that if, using figures calculated on pages 10 and 11 (Tout = 26"C and Tl = 20"C) the temperature for Tci is 18.5"C and for Tco the temperature is 24"C.

However, there is a drawback to the simple cycle of figure 6. First, a heat source of about 620C (or higher) is needed and geothermally these are easily available only in a limited number of locations ( for example, northern Italy and Arizona). Second, a natural heat sink is required at a temperature of 18.5"C (or lower) and in locations such as northern Italy or Arizona this is the case only in winter, early spring and late autumn. This means that in summer there is no heat sink and thus the engine cannot be operated. Thus for summer operation in warm locations the plant can incorporate the heat pump as described above and shown in figures 2 and 3. A heat pump can also be incorporated for use generally.

The working chamber of the heat engine is a piston/cylinder arrangement. The cylinders are provided with piston arrestors so as to be able to heat the working fluid at constant volume during the first stage of the power generating cycle.

The cylinders used as the working chamber for the working fluid of the cycle of figure 6 should be about 2m diameter and 10m long resulting in a piston stroke of 1.5m. This is at a ratio of expansion of 1.15:1, where the initial volume would be 2m diameter and 10m length while the final volume would be 2m diameter and 11.5m length. The work would be about 600 MJ per cylinder and due to the batch operation required (discussed below) there should be between 6 and 12 cylinders for a 1.5 GW plant.

If the use of turbomachinery is required for a particular purpose then hydraulic turbines can be used through which a hydraulic working fluid is pumped. The pumping action and work would then be produced by the piston. All that is required for this is a hydraulic circuit. The cylinder would be a two chamber arrangement separated by a lockable free floating piston. In one chamber there is the thermodynamic working fluid carbon dioxide; in the other a hydraulic working fluid, eg water. Obviously during expansion (Work) pumping/delivery of the hydraulic working fluid would be provided and during the return of the piston (Qout) suction/reception of the hydraulic working fluid into the cylinder would occur. During the heat input stage the piston does not move and nothing happens.

A possible cylinder arrangement for driving a hydraulic circuit is illustrated in Figure 7. A similar piston arrangement having a single chamber with a lockable piston can be connected to a crankshaft so as to drive a crankshaft directly.

Considering the arrangement of figure 7, a cylinder 114 has a free floating piston 117 and piston arrestors 118 between two fluid chambers 115,116. One of the chambers 116 contains carbon dioxide forming the working fluid of the cycle described above and the other chamber 115 contains a hydraulic fluid, e.g. water. Movement of the piston 117 in response to changes in the working fluid result in movement of the hydraulic fluid which is pumped by the movement of the piston 117 via a flow control valve 112 through an outlet 111 to the turbo machinery to be driven. The hydraulic fluid returns to the cylinder chamber 115 via a return pipe 113.

The cylinder 114 includes piston arrestors liB for locking the piston into position during the heat input stage of the working fluid cycle so as to ensure that the working fluid is heated at constant volume. These arrestors 118 can be of whatever design is suitable and whatever location is desirable as long as the piston is held in place during the heat supply stage of the working fluid cycle.

The working fluid chamber 116 includes inlet and outlet valves 121,122 for the working fluid. The working fluid is connected via these valves to the heat pump as described above and shown in figure 8. The inlet and outlet valves to the cooler/heat pump's evaporator are closed while the inlet and outlet valves to the heat source/heat pumps condenser are open. Thus the condition is Qin and hot working fluid is pumped into the cylinder through inlet 121 while cold (after Qout) working fluid is pumped out of the cylinder through outlet 112. Thus the thermodynamic working fluid is circulated through a heat exchanger, ie, the heat pump's condenser in the case of waste heat recovery or fullqin.

The advantages of converting the reciprocating piston motion into the rotary motion of a hydraulic turbine or any other hydraulic engine are that the equipment required is cheap, reliable and that the cost for a gearbox is eliminated. Another advantage is that the cylinders operate independently of each other. WHY? and this is of great advantage as shall be shown later on when describing a crankshaft connection. Further, hydraulic turbines are devices of great conversion efficiency. However, the introduction of various valves will cause losses, particularly if non-return valves are used.

The construction of the hydraulic circuit has not been discussed as these are well known.

The cylinders would need to be able to withstand a pressure high of about 300 Bar and at least 3 large tube and shell heat exchangers are required. The cost for the circulation pumping equipment (to cool/heat the thermal working fluid) is to be added as well as the cost for hydraulic turbines and pipes or crankshaft and gearbox.

Compared with the cost of a fossil fuel heat engine of equivalent output this will be minute - and the heat input can be arranged to be free.

Figure 8 illustrates plant having groups of working cylinders which operate in a manner such that they complement each other (i.e. batch operation) to take advantage of the heat rejected by one cylinder to heat the working fluid of another cylinder and so as to have a more continuous work output than is the case for a single cylinder system.

The plant of figure 8 includes at least one batch of three cylinders 53,54,55 and a heat pump 39 having a condensor 44 and evaporator 43. Valves 47, 48 control the flow of working fluid (i.e. carbon dioxide)through each of the cylinders 53,54,55 and the heat pump evaporator 43 and condensor 44. The flow of working fluid is controlled such that a given moment a first cylinder 53 is connected to the evaporator 43, a second cylinder 54 is connected to the condensor 44, and a third cylinder 55 is disconnected from the heat pump 41. The connections are then cyclically changed such that the cylinders 53, 54, 55 are each connected in turn to the condensor 44, disconnected, and the evaporator 43, . The low grade heat source is connected to the heat pump evaporator 43 via heat source input 45 and return 46.

In operation and as shown in figure 8, the working fluid in the cylinder 53 connected to the evaporator 43 is being cooled, the working fluid in the cylinder 54 connected to the condensor is being heated, and the working fluid in the disconnected cylinder 55 is doing work against a piston.

The heating and cooling process of the thermal working fluid needs to be considered in three aspects. The first is cooling, the second is heating and the third is the avoidance of waterhammer during the two processes.

Considering the plant of figure 8 and its operation in more detail.

The plant includes three cylinders 53, 54, 55 and a heat pump 39 consisting of a compressor 40, an evaporator 43, a condenser 44 and a pressure drop valve 42. The working fluid in the heat pump 39 is ammonia and that in the cylinders is carbon dioxide. Heat inputs and outputs 45,46 supply a heat transfer medium from the external heat source, and the return of that medium to that source or simply the rejection of that medium to the environment.

The cylinder 53 has two inlet and outlet flow control valves 47. There are also distribution valves 48 for controlling the connections of the cylinder 53 to the heat pump. Each cylinder 53, 54, 55 has a similar valve arrangement.

In the plant configuration shown in figure 8 the cylinder 53 is rejecting heat to the heat pump evaporator coolant fluid(ie Qout is occurring) and the condition of the working fluid in this cylinder 53 is changing from h3 to h2 (see figures 4,5,7). Simultaneously, the cylinder 54 is receiving heat from the condensing coolant fluid in the heat pump condensor 44 (ie Qin is occurring) and the condition of the working fluid in this cylinder 54 is changing from hl to h2 (see figures 4,5,7). In other words waste heat rejection and waste heat recovery are to occur simultaneously.

The plant includes pumps 49,50 for circulating the working fluid.

For the rejection of waste heat, i.e. the transformation from h3 to hl in the cycle of figures 4,5,7 the following will occur: working fluid will be pumped from the cylinder 53, via the inlet/outlet valves 47, across the distribution valves 48, by the pump 49, into and out of the evaporator 43. The cooling of the cylinder working fluid will cause the contraction of the working fluid and thus the piston within the cylinder 53 will return to bottom dead centre (BDC) . Thus waste heat is recovered via an artificial heat sink (ie a sink maintained at a temperature lower than ambient) formed by the coolant fluid in the condensor 44 which is permanently at a lower temperature than the Tout and Tlow temperatures of the power generating cycle working fluid.

In addition an artificial heat receptacle (at the same temperature as the sink) comprising the coolant fluid in the evaporator 43 is provided for the actual heat source entering the system at inlet 5.

The quantity of heat absorbed by the evaporator 43 consists of the rejected heat (real reject) plus the heat from the external heat source. The coolant of the heat pump is further raised in temperature due to the action of the compressor 40 which converts work into heat and thus reduces the heat requirement from the external heat source further.

Simultaneously with the above, the carbon dioxide working fluid of cylinder 54 is to be heated, ie Qin is to be done to cause the change from hl to h2. This is done by circulating the working fluid with the pump 50 through the condenser 44 and back into the cylinder 54. This process corresponds essentially to that for cooling.

Here it is to be noted that the pumps 49,50 produce heat by their work just as the compressor does, which of course reduces the requirement upon the external heat source 35 still further. The heat produced by these pumps does not come via a pressure difference, but through the turbulence of entering the condenser/evaporator, the viscosity based friction loss of the carbon dioxide and the turbulence when passing the valves. The effect of the pressure drop valve 42 for the heat pump can be neglected for practical purposes. It is just that, a pressure reduction valve, ie an expansion valve.

A discussion of what happens thermodynamically to the carbon dioxide in the course of the cooling and heating operation follows.

During cooling of the cycle working fluid there will be an amount of carbon dioxide in the hl condition in the dead space between the valves 47 and the evaporator 43.

This comes into contact and mixes with the carbon dioxide in the h3 condition once the cylinder inlet and outlet valves 7 are open. This results in losses which mean that the actual heat transfer temperature differences are slightly lower than the theoretical values estimated above.

Consider that the carbon dioxide h3 condition in the cylinder 53 at the beginning of the cooling stage of the cycle is as follows: T equals 26"C, h equals 581.39 kJ/kg and P equals 65.8 Bar while the density is 694.97 kg/m3. Conversely the carbon dioxide in the hl condition is as follows: T equals 20"C, h equals 532 kJ/kg, P equals 65.8 Bar and p = 800kg/m3. Thus, with the exception of the pressure, the two conditions of the two quantities of the carbon dioxide are not in equilibrium.

Upon mixing, these two conditions will equalize having regard to their respective quantities and masses, under the laws of physics. This means that a temperature drop for the carbon dioxide from the cylinder 53 will occur with a corresponding temperature rise for the ammonia contained within the cooling system (ie evaporator 43).

The same of course happens with the other quantities. Thus the quantity of heat to be transferred into the ammonia in the evaporator from the two quantities of carbon dioxide (from dead space and cylinder 53) remains the same as that for the single quantity from the cylinder while the temperature range narrows from the previous 26"C to 20"C to a new range of less than 26"C to 20"C.

The same process occurs in reverse for the heat input to the heating fluid in the cylinder 54 (Qin). Consider the carbon dioxide in the cylinder 54 and the dead space between the valves 47 and condensor 44. The cylinder is at hl, ie T= 200C, h = 552 kJ/kg, P = 65.8 Bar while p = 800 kg/m3. Conversely the carbon dioxide within the heating system is in h2 condition, that is T = 53.44"C, h = 615 kJ/kg, P = 255 Bar while p = 800 kg/m3. Thus only the density is equal while the remaining qualities will come into equilibrium once the corresponding inlet/outlet valves 54 for the cylinder 54 are open, and contact and mixing occur. In consequence the total amount of heat to be transferred by the condensor's ammonia into the carbon dioxide working fluid of the cylinder 54 will remain the same, but the starting temperature will be higher than 20"C. The starting pressure and enthalpy will also be higher where starting refers to the beginning of the heat transfer process. The exact difference in magnitudes of the theoretical state of the working fluid and the actual state simply depends upon the ratio of che mass of carbon dioxide within the cylinder 54 to the mass of carbon dioxide within the heating system dead space.

It should be noted here that it is necessary for the cylinder inlet and outlet valves 47, to open simultaneously otherwise there would be damage to the pumps 49,50. This is particularly the case for the circulation pump 50 on the heating side. Obviously if the inlet/outlet valves for the Qin cylinder 54 do not open simultaneously the pump 50, would have to cope with an immediate pressure difference of (255 Bar - 65.8 Bar) = 189.2 Bar.

Although the pressure ratio is only 3.87:1, a sudden surge at that pressure could damage the pump 50 if not destroy it completely.

During the heat supply to the cylinder working fluid (Qin operation) the pressure will rise due to the added heat at constant volume/density. This pressure rise will not affect the pump nor its Workin requirement since the pump 50 merely circulates carbon dioxide between cylinder 54 and condenser 44. Thus the pump continuously experiences the same pressure, enthalpy and temperature condition of the carbon dioxide on its suction, as well as its delivery side. Therefore the workload for the pump stems from the turbulence caused by the heat exchanger and valves to the carbon dioxide as well as the viscosity of the carbon dioxide.

After heat has been rejected by the working fluid of cylinder 53, heat has to be added, ie after the process of h3 to hl the process of hl to h2 has to take place (see figures 4,5,7). That is to say that the plant of figure 8 has to be reconfigured such that cylinder 53 is connected to the condensor 44 so as to take the place of the cylinder 54 and receive heat from the condensing coolant. This is done by switching the valves 47 of the cylinder position 53 in such a way that they connect with the condenser 44 (see figure 8 where this connection is shown in dotted lines). Thus flow of carbon dioxide occurs along the dotted lines to the distribution valves 48 into the condenser 44 and then back to the cylinder 53.

A complementary reconfiguration occurs for the connections of the cylinders 54,55 after work has been done in cylinder 55 (h3 to hl is now required) and the working fluid has been heated in cylinder 54 (h2 to h3 is now required) . Here the flow of carbon dioxide occurs along the dotted lines into the evaporator 43 and back for the cylinder 55 and the cylinder 54 is disconnected from the heat pump.

The plant of figure 8 includes standpipes or equalisation reservoirs 51,52. These are required as the carbon dioxide pumps operate continuously but due to the switching of the valves there is a definite risk of waterhammer which will occur if the valve timing is slightly out.

The standpipes or equalisation reservoirs 51,52 can be of the type used in hydraulic installations such as hydro stations operating Francis turbines. Their exact location, size and number in the plant circuit of figure 8 depends upon the actual layout and capacity of the plant circuit. The standpipe/equalisation reservoirs 51,52 shown in figure 8 do not necessarily denote or indicate the precise or only possible location. They are simply shown for the sake of completeness and to illustrate that they are required.

From the description above it is apparent that the cycle processes concerned will involve a batch operation.

Further, the heat transfer will take several seconds.

The work from a cylinder in the form of piston movement will consequently be extracted slowly as well, the duration of a piston stroke necessarily corresponding with the heat transfer time. This is of great advantage since, as a result, the engine friction (piston and crankshaft) will be extremely low. Slow piston speed implies very low piston friction losses, thus adding very little heat to the working fluid carbon dioxide and consequently leading to near adiabatic expansion.

Furthermore, very low speeds and very low friction loss translate into a correspondingly high mechanical efficiency.

To illustrate this consider a cylinder with a final post expansion height of 6.9m at h3, an initial height of 6m at hl and h2 and a diameter of 6m (assume the piston thickness to be zero). The density of the carbon dioxide is 800 kg/m3, so the cylinder will contain 135,716.8 kg of working fluid. After expanding to a density of 694.97 kg/m3 the occupied volume will be 6m diameter and 6.9m height. Thus the piston will have moved a distance of 0.9m which will also be its stroke, sufficient to turn a crankshaft. Assume now that the heat transfer process requires 10 seconds, then the piston velocity is a mere 9cm/sec, ie 0.09 m/s. The crankshaft will be correspondingly slow.With speeds such as these a friction loss of 1% would be excessive, but for the sake of this illustration a loss of 1% will be assumed for both the piston and the crankshaft in the calculations which follow. For the moment assume there to be only 4 kJ/kg carbon dioxide available as work.

Thus the work per cylinder will be 678.58 MJ over 10 seconds which equates to 67.85 MW/cylinder. Thus a machine having twenty cylinders producing work at any given moment would produce 1.357 GW and require sixty cylinders in total. The cylinders would be in batches of three; one for work, one for Qin and one for Qout, ie, (20x3 cylinders = 60 cylinders).

The operation of the heat pump 39 will be discussed with reference to figure 9. This shows a Pressure/enthalpy diagram for ammonia. The enthalpy of the ammonia is labelled qph so as to avoid confusion with the carbon dioxide (power cycle working fluid) enthalpy of figures 4 and 5. The critical point of the vapor dome has been denoted as "cp". Thus the curve to the left of cp denotes the saturated liquid line and the curve to the right of cp denotes the saturated vapour line. Along the pressure axis of the diagram there are the points P2 and Pl where P2 equals the condenser pressure of 30 Bar while P1 equals the evaporator pressure of 4 Bar. Thus the pressure ratio is 7.5:1.

Please note that qph4 and qphl are set as equal in respect of enthalpy, the reason being that there is very little enthalpy difference between them and that what there is, is negated by the expansion valve of the heat pump 41. There is also position qph2 which indicates the inlet enthalpy and density of the ammonia for the compressor. The heat absorption of the heat pump is therefore qph2 - qphl.

The relevant factors for determining the actual work requirement of the heat pump are the isentropic efficiency of the compressor 40 and the heat pump losses which include the resistance offered from qph2 to qph3, the resistance offered by the turbulence occurring when the ammonia enters the condenser and leaves the evaporator, and the friction of the ammonia itself whilst passing through the tubes and pipes of the heat pump.

The cycle of figures 4 and 5 is asymmetric insofar as Qin occurs isochorically (no piston movement - piston fixed by piston arrestors during heating), Work occurs isentropically (piston moves to top dead centre - TDC), and Qout occurs isobarically (piston moves to bottom dead centre - BDC).

This means that the piston cannot move during Qin and consequently during Qin the piston cannot be connected rigidly to a crankshaft since stalling of the engine would then occur during Qin. Therefore it is necessary to connect the piston to a crankshaft so as to allow the disconnection of the piston from the crankshaft assembly during the Qin phase.

A solution is to split the conventional connecting rod between crankshaft and piston into three separate but matching units (see figures 10, 11, 12). In Figure 10 there is the crank 81 of the crankshaft. There is further a crankshaft rod 82, a slider rod 83, and a piston rod 84. Contact blocks or buffers 85,86 having contact surfaces 85a,86a are provided on the rods 83,84.

The functioning of the split connecting rod is as follows: from the beginning to the end of the work stroke the contact block of the piston rod pushes against the contact block of the slider rod which is linked with the crankshaft rod. Thus the slider rod and the crankshaft rod simulate a piston rod during the work phase.

During the Qout phase the whole arrangement returns in unison, but note that the piston rod returns because the piston returns to BDC due to the contraction of the carbon dioxide being cooled and not because the slider rod is pushing against the piston rod, ie there is no load either upon the piston rod nor upon the slider rod, during the return stroke (Qout stage of cycle).

After the piston returns to BDC it is locked into its location in order for Qin to occur isochorically (i.e. at constant volume). The crankshaft continues to rotate and thus the slider rod moves away from the piston rod, which is static in conjunction with the pistons (see figures 10,12) The contact blocks at positions 85, 86 are made of a material which is soft enough to absorb a sudden impact should inaccurate timing methods be employed, but they are also strong enough to withstand the load encountered.

Alternatively they can also be spring loaded or even proper mechanical linkages can be employed which can be automatically actuated via a break/close mechanism which is activated by either the crankshaft, the heat transfer arrangement or both.

However, actuating the expansion process (the removal of the piston arresters described earlier) and actuating the heat removal by devices of either mechanical, electrical (or both) nature via the crankshaft is the most accurate method and will eliminate shocks upon the contact blocks.

Figures 13 to 21 illustrate a cylinder 71 having a piston 72, piston arrestors 73 and in which the piston 72 is connected to a split connecting rod arrangement similar to that described above.

Referring to Figure 13 this shows the cylinder during Qin with the piston locked into its location. If the situation is the beginning of the Qin phase then the crank arrangement is moving away from the piston rod. If it is the end of the Qin phase the crank arrangement is moving towards the piston rod. The piston's location is BDC.

Figure 14 depicts the piston at TDC at the end of the work stroke/the beginning of the Qout phase during which the piston returns to BDC.

Figure 15 illustrates the piston/cylinder and crank assembly at the moment in which work begins. The piston arrestors have been removed, the carbon dioxide begins to expand, the piston shown at BDC begins to move to TDC and the piston rod and the crank assembly move in unison by the piston rod pushing against the slider rod.

Figure 16 depicts the piston close to TDC and consequently the crank shaft has nearly completed a 1800 half turn. Please note the arrows indicating motion and the direction of the motion.

Figure 17 shows the piston returning to BDC, ie, the Qout phase. The carbon dioxide is circulated through the cooler (the heat pump's evaporator) and therefore contracts. Consequently the piston returns to BDC due to that contraction of the carbon dioxide and not because it receives work input. Thus there is no load between the crank mechanism and the piston with its piston rod.

They merely move together along the same path.

In Figure 18 the piston has reached BDC and is locked into that location and Qin commences.

In Figure 19 the piston stays locked, Qin continues and the crankshaft rotates free. Thus piston rod and slider rod are separate, that is, the slider rod moves away from the piston rod.

In Figure 20 the piston is still locked, thus maintaining constant volume. The Qin phase is nearly completed and the slider rod moves towards the piston rod.

The Qin phase is completed in Figure 21 the crankshaft has completed its free 360" rotation and the slider rod's contact block makes contact with the contact block of the piston rod. The governing or control devices unlock the piston and the operation recommences.

The piston speeds are low and consequently so is the piston friction loss.

The back pressure for the piston must be equal to the pressure low of the cycle, otherwise problems could occur. This can be achieved in two ways.

The first is to maintain a pressure within the crankcase which is equal to the minimum pressure of the cycle.

The second is the use of pressure reduction devices.

These are essentially pressure intensifiers in reverse.

However, employing such devices is simultaneously advantageous as well as disadvantageous. First the advantage: The conversion of high pressure and low volume flow into low pressure and high volume flow will be equivalent to an increased expansion ratio. This will be particularly useful if the work is finally extracted with hydraulic turbines, or with crankshafts and with relatively small quantities of working fluid. The disadvantage is that there are two pistons moving, one on the thermodynamic working fluid side, and on the other side of this will be a hydraulic fluid pushing another piston. That other piston then does work either in the form of pumping or turning a crankshaft. However, the other piston moves more slowly still (some additional friction loss) whilst adding an oscillating mass.

So, in general, pressurising the inner volume of the crank case is easier and more reliable. Thus there will be a substantial pressure difference between the crankcase and ambient, viz, 65.8 Bar.

The difference between the cylinder pressure and ambient will be even greater, viz, 255 Bar at maximum pressure.

This can be coped with in a variety of ways which can be combined. The first is making the cylinder more massive, i.e. simply building everything very strong. The second is to provide an exo-skeleton, ie, an external structure which supports and strengthens simultaneously. These two options are easy and cheap to implement.

Thirdly, in addition to the aforesaid an outer casing can be built around the engine structure. Maintaining a pressure of - for instance - 30 Bar within that outer casing will result in the following: The pressure ratio for the crankcase innards and the cylinder at pressure low to the outside (the inside of the outer casing) will be 2.19:1, while for the cylinders at pressurehigh, the ratio will be 8.5:1. In turn, the outer casing which contains the engine will need to be able to withstand a pressure difference of 30:1. The approach to be used and the pressure ratios in the case of outer casings depend mainly on cost.

So far mention has only been made of cylinder/piston arrangements to extract work from the working fluid of the cycle. It is possible to extract work by the use of turbo machinery directly from the working fluid but in the present state of knowledge of turbomachinery the resultant machinery and particularly control mechanisms, would be too complex to be worthwhile.

The plant described above is most efficient in large capacity installations. The reason is that the small (or relatively small) expansion volumes and ratios require very large quantities of working fluid. This means that correspondingly large quantities of heat pump fluid have to be employed. Further the heat pump fluids operate in the vapour region in which they exhibit less advantageous heat transfer co-efficients than the working fluids, which operate ideally in liquid form, though the heat transfer co-efficients for vapours are still better than those for gases.

Claims (33)

CLAIMS:

1. A method of converting heat into work comprising a cycle including the steps of heating a substantially liquid working fluid at constant volume and allowing the heated working fluid to expand and do work.

2. A method according to claim 1 wherein the cycle is a three stage cycle comprising the steps of; heating the working fluid at substantially constant volume, expanding the heated working fluid at substantially constant entropy and cooling the expanded working fluid at substantially constant pressure.

3. A method according to any preceding claim wherein the working fluid is a substantially saturated liquid at the beginning of the heating at constant volume.

4. A method according to any preceding claim wherein the cycle working fluid is carbon dioxide.

5. A method according to any preceding claim including the step of rejecting heat from the cycle working fluid to a heat pump.

6. A method according to any preceding claim including the step of supplying heat to the cycle working fluid from a heat pump.

7. A method according to claim 6 wherein an external heat source supplies heat to the heat pump.

8. A method of converting heat into work comprising the steps of sequentially heating the working fluid contained in three separate chambers, wherein the fluid in each chamber is substantially liquid and is heated at constant volume and the heated working fluid is allowed to expand and do work.

9. A method according to claim 8 wherein the working fluid of each cycle goes through a three stage cycle comprising the steps of; heating the working fluid at substantially constant volume, expanding the heated working fluid at substantially constant entropy, and cooling the expanded working fluid at substantially constant pressure.

10. A method according to claim 8 or 9 wherein the working fluid of each cycle is a substantially saturated liquid at the beginning of the heating at constant volume.

11. A method according to any of claims 8 to 10 wherein the cycle working fluid is carbon dioxide.

12. A power generating plant operating according to the method of any of the preceding claims.

13. A method substantially as herein described with reference to the accompanying figures.

14. A heat engine for converting heat into work including a chamber for a liquid working fluid, a moveable boundary for the chamber, and releasable locking means for locking the moveable boundary in a predetermined position, and heat supply means for supplying heat to the working fluid in the chamber.

15. A heat engine according to claim 14 wherein the working fluid is contained within a sealed system.

16. A heat engine according to claim 14 or 15 wherein the working fluid chamber is a cylinder, the moveable boundary is a piston within the cylinder and the releasable locking means are actuable piston arrestors.

17. A heat engine according to any of claims 14 to 16 wherein the working fluid chamber has a working fluid inlet and a working fluid outlet, a first heat exchanger for exchanging heat between a heat source and the working fluid, a second heat exchanger for exchanging heat between a heat sink and the working fluid and control means for controlling the flow of working fluid to the first and second heat exchangers.

18. A heat engine according to claim 17 wherein the heat sink is the heat input of a heat pump.

19. A heat engine according to claim 18 wherein the heat pump includes an evaporator for heating the heat pump fluid and the second heat exchanger supplies heat from the cycle working fluid to the pump fluid in the evaporator.

20. A heat engine according to claims 17, 18 or 19 wherein the heat source includes the heat output of a heat pump.

21. A heat engine according to claim 20 wherein the heat pump includes a condensor for cooling the heat pump fluid and the first heat exchanger supplies heat from the heat pump fluid in the condensor to the cycle working fluid.

22. A heat engine according to any of claims 18 to 21 wherein an external heat source supplies heat to the heat pump.

23. A heat engine to any of claims 14 to 22 wherein the cycle working fluid is liquid carbon dioxide.

24. A heat engine for converting heat into work including a batch of three chambers for a liquid working fluid, each working fluid chamber being provided with a move able boundary and releasable locking means for locking the move able boundary in position, heat supply means for supplying heat to the working fluid in each chamber and control means for controlling the supply of heat to each chamber such that heat is supplied to each chamber of the batch in turn.

25. A heat engine according to claim 24 wherein each moveable chamber boundary is connectable to means for exploiting the work done in moving the boundary and the heat engine includes control means for connecting each moveable boundary of the batch in turn to the means for exploiting work.

26. A heat engine according to claims 24 or 25 for converting heat into work including a batch of three chambers for a liquid working fluid, each working fluid chamber being provided with a moveable boundary and releasable locking means for locking the moveable boundary in position, a heat sink for receiving heat rejected by the working fluid in each chamber and control means for controlling the rejection of heat from each chamber such that heat is rejected from each chamber of the batch in turn.

27. A heat engine according to claims 24 to 26 wherein, in operation, the flow of heat into and out of the chambers of the batch and the connection to the means for exploiting work are timed such that the three chambers of the batch are mutually out of phase and, at a given moment, the working fluid in a first chamber of the batch is being heated at constant volume, the working fluid in a second chamber of the batch is expanding and doing work on the moveable boundary which is being exploited by the work exploiting means, and the working fluid in a third chamber of the batch is rejecting heat to the heat sink.

28. A heat engine according to claim 27 including a heat pump having a condensor for cooling the heat pump fluid and an evaporator for heating the heat pump fluid, a first heat exchanger for exchanging heat between the heat pump fluid in the condensor and the working fluid of the first chamber of the batch to heat the working fluid in the first chamber, and a second heat exchanger for exchanging heat between the heat pump fluid in the evaporator and the working fluid of the third chamber to cool the working fluid in the third chamber.

29. A heat engine according to claim 29 including a third heat exchanger for exchanging heat between an external heat source and the heat pump fluid in the evaporator.

30. A heat engine according to any of claims 24 to 29 wherein each working fluid chamber is a cylinder, the moveable boundary is a piston within the cylinder and the releasable locking means are actuable piston arrestors.

31. A heat engine according to any of claims 24 to 30 wherein the working fluid is carbon dioxide.

32. Power generating plant including the heat engine of any of claims 14 to 31.

33. A heat engine substantially as herein described with reference to the accompanying figures.