Method and apparatus for power storage

Abstract

Cryogenic energy storage systems, and particularly methods for capturing cold energy and re-using that captured cold energy, are disclosed. The systems allow cold thermal energy from the power recovery process of a cryogenic energy storage system to be captured effectively, to be stored, and to be effectively utilized. The captured cold energy could be reused in any co-located process, for example to enhance the efficiency of production of the cryogen, to enhance the efficiency of production of liquid natural gas, and/or to provide refrigeration. The systems are such that the cold energy can be stored at very low pressures, cold energy can be recovered from various components of the system, and/or cold energy can be stored in more than one thermal store.

Claims

1. A cryogenic energy storage system comprising: a cryogenic storage tank for storing a cryogen; a pump in fluid communication with the cryogenic storage tank for compressing the cryogen from the storage tank; a first cold thermal store for storing cold thermal energy having a first pathway therethrough for conveying a first heat transfer fluid; the first pathway extending through the first cold thermal store between an input end and an output end; a power recovery system including a first heat exchanger having: a second pathway therethrough for conveying the first heat transfer fluid from the output end of the first pathway through the first heat exchanger to the input end of the first pathway, whereby a first closed loop circulation of the heat transfer fluid is provided through both the first cold thermal store and the first heat exchanger, and a third pathway therethrough for conveying the compressed cryogen through the first heat exchanger to transfer cold thermal energy from the compressed cryogen to the first heat transfer fluid; the power recovery system further including one or more expansion turbines for expanding the cryogen conveyed by the third pathway to generate power; a co-located liquefier including a second heat exchanger having: a fourth pathway therethrough for conveying the first heat transfer fluid from the output end of the first pathway through the second heat exchanger to the input end of the first pathway, whereby a second closed loop circulation of the heat transfer fluid is provided through both the first cold thermal store and the second heat exchanger, and a fifth pathway therethrough for conveying a first working fluid through the second heat exchanger to transfer cold thermal energy from the first heat transfer fluid to the first working fluid for contributing to converting the first working fluid into new cryogen; and the co-located liquefier further including a sixth pathway for conveying the new cryogen to the cryogenic storage tank; a control valve that controls flows of the first heat transfer fluid between the second and fourth pathways and the first pathway to control the first and second closed loop circulations of the first heat transfer fluid through the first cold thermal store and one or both of the first and second heat exchangers; and a circulation pump located along the first pathway and operating in conjunction with the control valve for circulating the first heat transfer fluid along the first pathway through the first cold thermal store as well as from the output end of the first pathway along one or both of the second and fourth pathways that return the first heat transfer fluid to the input end of the first pathway.

2. The cryogenic energy storage system of claim 1, wherein the first heat transfer fluid is conveyed at a gauge pressure of less than 4 bar.

3. The cryogenic energy storage system of claim 1, wherein the control valve controls (a) the first closed loop circulation through both the first cold thermal store and the first heat exchanger and (b) the second closed loop circulation through both the first cold thermal store and the second heat exchanger.

4. The cryogenic energy storage system of claim 1, wherein a portion of the first pathway between the first cold thermal store and the output end of the first pathway provides for conveying at least a portion of the first heat transfer fluid from the first cold thermal store to the second and fourth pathways, and the control valve provides a connection from the second and fourth pathways to the input end of the first pathway for conveying the first heat transfer fluid to the first cold thermal store.

5. The cryogenic energy storage system of claim 1, wherein the second heat exchanger comprises one element of a cold box within the liquefier.

6. The cryogenic energy storage system of claim 1, wherein the first heat transfer fluid is a gaseous fluid.

7. The cryogenic energy storage system of claim 6, wherein the first heat transfer fluid is dry air, or dry nitrogen.

8. The cryogenic energy storage system of claim 1, wherein the first heat transfer fluid is a liquid.

9. The cryogenic energy storage system of claim 8, wherein the first heat transfer fluid comprises methane, methanol, propanol or propane or a water-glycol mix.

10. The cryogenic energy storage system of claim 1, wherein the cryogen comprises liquid air.

11. The cryogenic energy storage system of claim 1, wherein the co-located liquefier includes an air compressor.

12. The cryogenic energy storage system of claim 1, wherein the first cold thermal store has a variable geometry design.

13. The cryogenic energy storage system of claim 1, wherein the first cold thermal store comprises a containment vessel containing a storage media, wherein the first pathway is such that the first heat transfer fluid is in direct contact with the storage media.

Description

BRIEF DESCRIPTION OF THE DRAWINGS

(1) Embodiments of the present invention will now be described with reference to the figures in which:

(2) FIG. 1 shows the typical generating regime of an energy storage device;

(3) FIG. 2 shows a typical state of the art cryogenic energy storage system (CES) as disclosed in U.S. Pat. No. 6,920,759 B2 and E. M. Smith 1977 (ibid);

(4) FIG. 3 shows a typical state of the art high pressure regenerator as disclosed in U.S. Pat. No. 6,920,759 B2;

(5) FIG. 4 shows a schematic of a cryogenic energy storage system according to a first embodiment of the present invention;

(6) FIG. 5 shows a second embodiment of the current invention;

(7) FIG. 6 shows the reduction in the energy required to liquefy a kg of air for different quantities of cold recycle;

(8) FIG. 7 shows a third embodiment of the current invention;

(9) FIG. 8 shows a fourth embodiment of the current invention;

(10) FIGS. 9 to 11 show fifth, sixth and seventh preferred embodiments of the current invention;

(11) FIG. 12 shows an eighth embodiment of the current invention;

(12) FIG. 13 shows a ninth embodiment of the current invention; and

(13) FIG. 14 shows a tenth embodiment of the current invention.

DETAILED DESCRIPTION OF THE INVENTION

(14) FIG. 4 shows a schematic of a first simplified embodiment of the present invention. The system of FIG. 4 is similar to that shown in FIG. 2 except that the regenerator shown in FIG. 2, in which the process gasses flow directly through the regenerator, is replaced in FIG. 4 by a thermal store with an intermediate very low pressure heat transfer loop.

(15) During the power recovery (discharge) phase of the cycle, the cryogenic fluid is first transferred from the storage tank 100 to the pump 105 where the fluid is compressed to high pressure, of at least 30 bar, more preferably at least 50 bar, and more typically 100 to 200 bar. Cold thermal energy is transferred via a heat exchanger, referred to as the evaporator 302, to a heat transfer fluid that circulates between the evaporator and thermal store 301 via a very low pressure heat transfer loop 104. The heat transfer loop typically consists of connecting pipework, a circulation pump and associated instruments and control valves. The heat transfer fluid may be dry air, dry nitrogen or a suitable other gas or refrigerant fluid such as methane, methanol, propanol or propane.

(16) During the storage phase of the cycle, both heat transfer loops can be shut down to isolate the thermal store from the environment and minimise heat losses. Similarly, the cryogenic storage tank can be isolated from the environment to the extent that any boil-off gas can be vented from the cryogenic fluid safely but heat leak into the tank can be minimised.

(17) During the liquefaction (charging phase) of the cycle, ambient air is first compressed in a compressor 104 to at least 10 bar and more typically 40 to 60 bar. Typically, impurities such as CO.sub.2, water and hydrocarbon contaminants are removed from the high pressure gas. Cold thermal energy is then transferred from the thermal store to the high pressure gas via a very low pressure heat transfer loop 104 to a heat exchanger, referred to as the chiller 303. The now cold gas is then expanded through a valve 101 to produce liquid air that is stored in the cryogenic storage tank 100. Typically the composition of the cold air after the expansion valve is a mixture of liquid and gaseous air. The gaseous air is often returned to the chiller to provide additional cooling of the high pressure air before expansion.

(18) Further embodiments are described in more detail as follows.

(19) The second embodiment of the current invention is shown in FIG. 5. Considering the discharge (power recovery) part of the cycle first, cryogenic fluid, which could be liquid air, liquid nitrogen or a similar low temperature fluid is transferred first from the cryogenic storage tank 409 to the cryogenic pump 410. This is commonly achieved by a gravity feed from the storage tank but may be assisted by an intermediate low pressure boost pump. The cryogenic fluid is compressed by the cryogenic pump 410 to high pressure of at least 50 bar and more commonly 100 to 200 bar. Thermal energy is added to the cryogenic fluid from a heat transfer fluid in the evaporator 413 to raise the temperature of the cryogenic fluid to typically 0 to 20? C. During the heating process, the cryogenic fluid will undergo a change of phase from liquid to gas if the process is operated below the critical pressure, or more typically will expand remaining in the supercritical state if the process is operated above the critical pressure for the cryogenic working fluid. Thermal energy is transferred to the evaporator 413 to provide heating of the cryogenic fluid via a very low pressure circulation loop connecting the evaporator 413 to the thermal store 411 known as the high grade cold store. A heat transfer fluid is circulated through the high grade cold store 411 and evaporator 413, via the circulation pump 419 and control valve 412. The heat transfer fluid in this embodiment will typically be a gas such as dry air or dry nitrogen as few liquids are available that will remain in the liquid state over the temperature range of ?180? C. to 10? C. which the process will typically operate. The now heated cryogenic fluid will then be further heated in a super heater 414, using thermal energy from the ambient environment or more typically low grade waste heat from a co-located process, such as condensate water from a thermal power station. The heated cryogenic fluid is then expanded first through a high pressure turbine 415 and then through a low pressure turbine 417. The gas is usually re-heated between the turbine stages using a heat exchanger referred to as the re heater 416 using ambient or low grade waste heat. Additional turbine stages with inter-stage reheaters may be added to improve the efficiency of the process, three or four being the preferred number of stages.

(20) On completion of the power recovery process, the cold high grade cold store 411 and cryogenic storage tank are isolated to minimise thermal losses during the storage phase of the cycle. A relief valve is commonly included on the storage tank to allow safe venting of boil off gas from the stored cryogenic liquid.

(21) During the liquefaction (charging) phase of the cycle, the following process is used, which is a two expander variant of the Claude cycle and is commonly used on modern air liquefaction plants. The Claude cycle is a refrigeration cycle where part of the refrigeration cooling is provided through the expansion of a gas through an adiabatic expansion engine, such as a turbine. In the two expander variant used here, the process gas, which could be ambient air, nitrogen or a similar gas, is first compressed to typically 5 to 10 bar in the main air compressor 401. Impurities, such as CO.sub.2, water and hydrocarbons that may freeze in the cold parts of the plant are removed in the air purifier 402, typically a regenerative adsorber. The high pressure air is then further compressed in a recycle air compressor 403 to typically 25 to 70 bar. After the recycle air compressor 403, the flow is divided into two streams. The first stream is further compressed by the compressor stages of first a warm turbo expander 405 and then a cold turbo expander 404 to typically 40 to 60 bar. Preferably, the gas is compressed to above the critical point of the gas, or 39.2 bar for air. This stream then enters a multi pass heat exchanger, or more commonly a series of multi pass heat exchangers commonly referred to as a cold box 406. A cold box is generally embodied as an insulated (typically metal) box, filled with high performance insulation materials such as perlite. The low temperature heat exchange processes occur within the cold box via components such as the multi pass heat exchangers, phase separator, and turbine stages of the warm and cold turbines. The gas is cooled in the cold box 406 and further divided. One stream is expanded in the expander stage of the cold turbo expander 404 to the product delivery pressure, typically 1 to 6 bar. In expanding the gas, the gas is substantially cooled preferably to the point where liquid droplets are just starting to form. The partially wet vapours from the cold turbo expander 404 are then introduced to a phase separator 407. The second cold box stream is further cooled in the cold box 406 and then expanded in an expansion valve 408, often referred to as a Joules Thompson (JT) valve. The cold gas substantially condenses to liquid during the expansion process and enters the phase separator 407. The liquid product is separated from the vapour phase in the phase separator 407 and stored in the cryogenic storage tank 409 for later use. The vapour phase is returned to the cold box 406 which provides cooling for the high pressure gas stream.

(22) The second stream separated after the recycle air compressor 403 is first cooled in the cold box 406 and then expanded in the warm turbo expander 405 to reduce the pressure and temperature of the gas. This now cool gas stream is then introduced to the cold box 406 to provide cooling.

(23) In this embodiment of the present invention, additional cooling of the process gas is provided by recycling the cold energy recovered during the power recovery phase of the cycle and stored in the high grade cold store 411. This is achieved by circulating the heat transfer fluid through the high grade cold store 411 using the circulation pump 419 such that the heat transfer fluid is cooled to close to the condensation temperature of the process gas. The heat transfer fluid is introduced close to the bottom of the cold box 406 heat exchanger where it is warmed by the process gas and provides cooling of the process gas.

(24) The present inventors have found that effective recovery, storage and recycle of the cold energy released during the power recovery process can more than halve the energy required to re-liquefy the process gas. This is illustrated in FIG. 6 which shows the reduction in the energy (in kWhrs) required to liquefy a kg of air (the preferred process fluid) for different quantities of cold recycle, expressed in kJ/kg of working fluid. Modern liquefiers typically operate at a specific work requirement of 0.43 kWhrs per kg of liquid air produced. Calculations by the inventors have indicated that the specific work to liquefy the process gas can be reduced to less than 0.2 kWhrs/kg if more than 72% of the cold energy released during the power recovery process can be recovered.

(25) In a further embodiment of the present invention shown in FIG. 7, the cycle is modified to operate without the circulation pump 419. During power recovery, the very low pressure gaseous exhaust of the low pressure turbine 417 is used to evaporate and heat the working fluid in the evaporator heat exchanger 413. The resultant cold very low pressure stream can then be directly introduced to the high grade cold store 411 to recover and store the cold thermal energy. During liquefaction, the process is modified such that the very low pressure outlet of the warm turbine 405 is ducted after warming in the cold box 406 to the high grade cold store. The warm very low pressure gas stream is cooled in the high grade cold store 411 and the resultant very low pressure cold gas stream is introduced to the bottom of the cold box 406 to enable the cold energy to be recycled and used to reduce the energy cost of liquefaction. The flow of the cold store 411 is controlled by two control valves 412 that control the connection to the power recovery and liquefier parts of the cycle. These modifications to the design have the advantage of reducing the complexity and therefore cost of the system in that the circulation pump 419 is no longer required. There will be a small penalty in efficiency in that both the warm turbine 405 and low pressure turbine 417 will suffer higher back pressure from the high grade cold store 411. However, if the high grade cold store 411 is optimised to operate at a low pressure loss over the range of exhaust gas flow rates by sizing the flow area of the store such that the velocities and hence pressure losses in the store are low, the back pressure will be small.

(26) In a further embodiment shown in FIG. 8, an additional cold store 701 is added to the system. The component 420 shown in FIG. 8 is referred to as the Power Island and includes the components shown in the dashed box 420 in FIG. 7. The inventors have noticed that in the embodiment of FIG. 7, where the low pressure turbine 417 exhaust is used to provide the evaporation heating of the working process fluid during power recovery, only around two thirds of the available cold energy can be recovered for cold recycle. This is because the enthalpy released in heating the high pressure cryogen from the liquid state to ambient temperature is higher than the enthalpy required to cool air from ambient temperature to the temperature of the high pressure cryogenic liquid as follows:

(27) TABLE-US-00002 Enthalpy released in heating air at 100 bar 333 kJ/kg from ?170 to 10? C. Enthalpy required to cool air at 1 bar 182 kJ/kg from 10? C. to ?170? C.

(28) The inventors have realised that by adding an additional medium grade cold store 701 after the evaporator 413 it is possible to capture the cold energy not captured by the evaporator 413. The energy is captured using a closed loop circuit thermally linked to the power recovery circuit via main heat exchanger 703. The very high pressure working fluid typically leaves the evaporator 413 at ?90 to ?120? C. As such, gases such as air or nitrogen could be used as the working fluid in the medium grade cold store closed loop circuit, or alternatively liquid refrigerants such as methanol or propanol could be used as the working fluid in the medium grade cold store closed loop circuit. Such liquids are suitable for operating between these temperatures and ambient temperature. The use of liquid refrigerants is preferred as the energy cost in pumping a liquid through the medium grade cold store 701 and around the associated pipework and cold box 406 heat exchangers will be considerably lower than for a gas. In addition, the capital cost of the circulation pump 702, valves and pipework is lower for a liquid heat transfer circuit compared to a gas heat transfer circuit. The working fluid, such as a liquid refrigerant, is circulated using a pump 702. During power recovery, the circulation pump 702 is used to circulate the refrigerant liquid through the main heat exchanger 703, through the medium grade cold store 701 and through the cold box 406 heat exchangers. The refrigerant liquid is introduced at a point in the cold box 406 at equilibrium with the temperature of the medium grade cold store 701 to ensure optimal thermal efficiency.

(29) The hexagonal numbered boxes in FIG. 8 show positions in the system at which the gauge pressures and temperatures are as follows:

(30) TABLE-US-00003 Gauge Pressure Temperature 1 Low Pressure ?196? C. to ?177? C..sup. 2 Very High Pressure ?185? C. to ?170? C..sup. 3 Very High Pressure ?120? C. to ?90? C..sup. 4 Very High Pressure 10 to 20? C. 5 Very Low Pressure ?40 to 10? C. 6 Very Low Pressure ?185? C. to ?170? C..sup. 7 Very Low Pressure ?120? C. to ?90? C..sup. 8 Low Pressure 10 to 40? C. 9 Medium Pressure 10 to 40? C. 10 High Pressure 10 to 40? C. 11 Very Low Pressure 10 to 30? C. 12 Very Low Pressure ?185? C. to ?170? C..sup. 13 Low Pressure ?190 to ?170? C. 14 High Pressure ?190 to ?170? C. 15 Low Pressure ?194 to ?175? C.

(31) Equivalent positions in FIGS. 5, 7 and 9 to 11 are at equivalent temperatures and pressures.

(32) A further embodiment is shown in FIG. 9. This embodiment is the same as that of FIG. 8 except that the cold thermal energy stored in the medium grade cold store 701 is used to pre cool the warm turbine using a warm turbine cooler 901, instead of being used to cool the cold box 406. This embodiment has the advantage of removing one flow stream from the cold box 406, thus simplifying the design. The warm turbine cooler 901 can be a simple shell and tube, plate fin or similar heat exchanger.

(33) A further embodiment is shown in FIG. 10 in which the preferred embodiment is modified slightly such that the cold turbine exhaust is used to provide the cold recovery flow to the high grade cold store during the charging (liquefaction) phase of the cycle, rather than the warm turbine flow. The flow to the high grade cold store can be taken from the cold turbine exhaust directly, or preferably (as shown in FIG. 10) after the phase separator. This embodiment is preferred where the cryogenic liquid is stored at ambient or near ambient pressure and the phase separator (407) pressure is at or near ambient pressure. A further embodiment is shown in FIG. 11, in which the warm turbine is replaced by a liquid refrigeration module 1101. The medium grade cold from the medium grade cold store 701 is used to cool the high pressure liquid in the refrigeration module 1101 prior to expansion cooling to improve the refrigeration performance of the circuit. This embodiment improves the overall efficiency of the process but has a higher capital cost. As such, this embodiment is preferred in markets and applications where efficiency is very important and a higher capital cost is economically viable.

(34) A further embodiment is shown in FIG. 12, which is the same as the embodiment of FIG. 7, except that a proportion of an exhaust stream from the thermal store outlet 411 is recirculated to pass through the evaporator 413 for a second time. Whilst this modification is shown in the context of FIG. 7, it could be used in any of the embodiments shown in FIGS. 8 to 11, or those disclosed hereafter.

(35) The recirculation increases the flow rate of the fluid passing through the evaporator 413 to a level sufficient to capture all the cold thermal energy from the high pressure cold stream of cryogen. Accordingly, the modification negates the requirement for the medium grade cold store, 701, shown FIG. 8.

(36) By increasing the flow rate of the exhaust stream passing through evaporator 413 (which is typically at a pressure of 1 bar), the enthalpy required to cool the stream from 10? C. to ?170? C., for example, can be matched to the enthalpy released in heating the cryogen in the high pressure stream from ?170? C. to 10? C. As explained above, this enables substantially all of the cold energy to be captured in evaporator 413 by the low pressure stream.

(37) The recirculated air mass flow rate is typically between 0.5 and 1 times the exhaust stream flow, and more commonly ?0.8 when the liquid stream is pumped to 100 bar.

(38) A circulation pump 702 is utilised to balance the system and ensure that the exhaust stream flow rate is sufficiently high whilst the turbine back pressure is unaffected.

(39) This embodiment simplifies the system, removing the requirement for the medium grade cold store, main heat exchanger 703 and any secondary energy transferal medium. Among other advantages, this reduces capital costs.

(40) A further embodiment is shown in FIG. 13. This embodiment is the same as that shown in FIG. 7, except that a low grade thermal store 803 has been added to the system. The low grade thermal store could be used with any of the aforementioned embodiments.

(41) Low grade cold, typically between 0? C. and ?20? C., is produced in the expansion of the stream at the outlet of each turbine stage. The low grade cold thermal energy is captured by low grade exchangers 801. The low grade exchanger 801 can be simple shell and tube, plate fin or similar heat exchanger. In the embodiment of FIG. 13, two low grade exchangers 801 are shown, each located at the outlet stream of a turbine stage. However, more or fewer low grade exchangers 801 could be used. For example, the power island may include any number of turbine stages, and a low grade exchanger 801 may be provided for each turbine stage, or for only some of the turbine stages.

(42) A low grade cold pump 802 circulates a heat transfer fluid from the low grade heat exchangers 801 to the low grade thermal store 803 during the discharge phase of the energy storage system. The heat transfer fluid is typically a glycol water mix, although other fluids such as air, or refrigerant gases could be used.

(43) During the charging phase of the energy storage system, a second low grade cold pump 802 circulates the heat transfer fluid from the low grade thermal store 803 to both the recycle and main air compressor coolers (805, 804), whereby it is used to pre-cool the compressor inlet streams. The recycle and main air compressor coolers 805, 804 can be simple shell and tube, plate fin or similar heat exchangers.

(44) The inventors have found that by reducing the temperature of the compressor inlet stream, the gas becomes more dense and the necessary work to compress the gas per unit is reduced. The inventors have noted that in the charging phase, the bulk of the necessary work is done by the Recycle Air Compressor 403 and Main Air Compressor 401 in compressing the gas, therefore any reduction in the necessary work carried out by these compressors has a significant effect on the work done by the system.

(45) It will be appreciated that whilst the embodiment of FIG. 13 shows the low grade cold thermal energy being used to pre-cool the Recycle Air Compressor 403 and Main Air Compressor 401, it could be used to pre-cool just one of these compressors, or any number of different compressor. For example, one or more compressor coolers could be provided for the booster compressors 404, 405, driven by the warm and cold expander turbines.

(46) A further embodiment is shown in FIG. 14. This embodiment is similar to the embodiment shown in FIG. 13, but instead of capturing cold thermal energy from the power unit and using it to improve the efficiency of the liquefier, the embodiment in FIG. 14 effectively operates in reverse.

(47) Again, the embodiment of FIG. 14 is the same as that shown in FIG. 7, except that a hot thermal store 1002 has been added to the system. The hot thermal store could be used with any of the aforementioned embodiments

(48) Thermal energy, typically of temperature 60? C. to 90? C. is produced in the outlet streams of both the recycle and main compressors (1004, 1005). The thermal energy is captured by compressor intercoolers, stored in the Hot Thermal Store 1002, and discharged to provide superheating to each of the turbine stage inlet streams via Compressor Heat Super Heaters 1001.

(49) In the embodiment of FIG. 14, two compressor heater super heaters are shown, each located at the inlet stream of a turbine stage. However, more or fewer superheaters could be used. For example, the power island may include any number of turbine stages, and a superheater may be provided for each turbine stage, or for only some of the turbine stages.

(50) Moreover, it will be appreciated that whilst the embodiment of FIG. 14 shows the compressor intercoolers on each of the Recycle Air Compressor 403 and Main Air Compressor 401, an intercooler on just one of these compressors could be used, or on any number of different compressors. For example, one or more intercoolers could be provided for the booster compressors 404, 405, driven by the warm and cold expander turbines.

(51) A circulation pump 1003 circulates a heat transfer fluid from Compressor Intercoolers 1004, 1005 to the Hot Thermal Store 1002 during the energy store charge phase. The heat transfer fluid is typically a glycol water mix, although other fluids such as air, or refrigerant gases could be used. The Compressor Intercoolers 1004, 1005 can be a simple shell and tube, plate fin or similar heat exchanger.

(52) During the discharging phase of the energy storage system, a second pump 1003 circulates the heat transfer fluid from the Hot Thermal Store 1002 to the Compressor Heat Super Heaters 1001, super heating the inter stage turbine process stream.

(53) The inventors have found that the use of waste heat to superheat the turbine stream is an effective way of increasing the output from the Power Island, 420 with no added operational cost. Capturing the compressor heat allows the energy storage system to operate independently of sources of external waste heat to achieve higher efficiencies.

(54) The inventors have discovered that in most energy storage applications and markets, the rate of charging of the thermal stores and the rate of discharge of the thermal stores are often substantially different. This is because the time where energy prices are at a peak and the system is in discharge mode are considerably shorter than the times when energy prices are low and the device can be in charging mode. It is advantageous to use as much of the available charging time as possible as a smaller, cheaper charging system can be built. The consequence of this discovery is the rate of energy flow in and out of the thermal stores is substantially different between charging and discharge. All the embodiments of the current invention described above benefit significantly from the use of high grade and/or medium grade cold of a variable geometry design, as described in GB1013578.8. In particular, the or each thermal store may consist of a first thermal mass, a second thermal mass, and a third thermal mass, wherein the aspect ratios of at least two of the first, second and third thermal masses are different to one another, and an arrangement of conduits and valves configured to direct a heat transfer fluid to pass through a combination of one or more of the thermal masses. Alternatively, the or each thermal store may comprise a first thermal mass comprising solid particles, a second thermal mass comprising solid particles, a third thermal mass comprising solid particles, wherein the diameter of the particles in at least two of the first, second and third thermal masses are different to one another, and an arrangement of conduits and valves configured to direct a heat transfer fluid to pass through a combination of one or more of the thermal masses.

(55) It will of course be understood that the present invention has been described by way of example, and that modifications of detail can be made within the scope of the invention as defined by the following claims.