SYSTEM AND METHOD OF OPERATING SUCH A SYSTEM

Field of the invention

The present invention relates to a hybrid electricity storage and power generation system and a method of operating such a system, and in particular, one in which the electricity storage incorporates thermal energy storage.

Background to the Invention

A number of electrical energy storage systems have been developed for storing excess electrical energy so it can be returned when required, for example, at times of peak demand. Leading technologies for bulk power storage are Pumped hydro storage (PHS) systems and compressed air energy (CAES) systems. PHS requires specific geographical features, while current commercial CAES systems require underground storage caverns. Roundtrip efficiencies for PHS systems can be 70-85%, while current (diabatic) CAES systems are less than 55% efficient, although this is likely to increase to the 60-70% range for the newer adiabatic ACAES systems.

In recent decades thermal energy storage technology has developed significantly, particularly alongside solar thermal plants, where it can assist with variable solar power. A number of electricity energy storage technologies have now been proposed based around thermodynamic cycles that incorporate thermal energy storage.

One such system is a pumped heat electricity storage system, or PHES system, in which a working fluid is subjected to a thermodynamic heat pump cycle such that electrical energy is stored as thermal energy in thermal energy stores, or TES, during a charging mode, and whereby the working fluid is subjected to a thermodynamic heat engine cycle in which the thermal energy in the TES is retrieved as electrical energy in a discharging mode.

US2010/0301614 to Saipem discloses a PHES system based on TES using turbo- machinery and operating a closed Brayton cycle preferably using argon as the working gas. Using turbomachinery, the roundtrip efficiency depends upon a high turbine inlet temperature and so the proposed embodiments employ stores using porous refractory materials (e.g. porous bricks) storing heat (e.g. in the hot store) in excess of 1000°C. Factoring in real losses, however, overall efficiency is likely to be below 55%.

WO2009/044139 to Isentropic proposes an alternative PHES system based upon TES using reciprocating, piston based power machinery and operating a closed Brayton cycle again preferably using argon as the working gas. In this system, due to the higher efficiency of the reciprocating machinery, overall efficiency of over 70% is likely to be possible and store operating temperatures may be much lower. Since PHES systems require a certain number of hours to charge, and likewise can only deliver power for the time it takes for the hot and cold stores to discharge, their power generation capability is limited (e.g. 5-10 hours daily). Further, since the cycle involves two compression/expansion processes and two heat transfer processes on charge and also on discharge, a PHES system is very sensitive to the efficiency of each of the compression and expansion processes and the heat transfer processes, so that its successful implementation will be dependent upon a balance of (e.g. economic, cost, reliability, efficiency, longevity) factors.

The present invention is directed towards providing a novel hybrid electricity storage and power generation system based upon a PHES system.

Summary of the Invention

According to a first aspect of the present invention, there is provided a hybrid electricity storage and power generation system comprising:

(i) a pumped heat electricity storage PHES system comprising at least a first compressor, a first thermal energy store (TES) system, second power machinery, and a second thermal energy store (TES) system respectively arranged (i.e. in that order) within a first working fluid circuit and operable, using air as the working fluid, in:

• a charging mode that uses an open or closed, thermodynamic heat pump cycle to heat the first TES system and cool the second TES system, the first compressor and the second power machinery respectively compressing and expanding the air before entry to the first and second TES systems, respectively; and,

• a discharging mode that uses an open, thermodynamic heat engine cycle to cool the first TES system and to heat the second TES system, the first compressor being inactive and the second power machinery compressing the air before entry to the first TES system, whereby in this mode heated, pressurised air is produced from the (e.g. an outlet of the) first TES system; and,

(ii) a combustion turbine sub-system comprising a combustor and a downstream expansion turbine integrated with the first working fluid circuit by means of flow connectors; wherein the hybrid system is configured to be operable in the following generation modes: (a) a discharge generation mode in which the PHES system is operating in the above discharging mode, and the flow connectors are configured to direct the heated, pressurised air from the (e.g. outlet of the) first TES system to the combustor for combustion and then to the expansion turbine for subsequent expansion to produce electrical power; and,

(b) a normal generation mode in which the PHES system is not operating in the above charging or discharging modes, and the flow connectors are configured to direct heated, pressurised outlet air from the first compressor to the combustor for combustion and then to the expansion turbine for subsequent expansion to produce electrical power.

The hybrid system is based on a pumped heat electricity storage system or PHES system, in which a working fluid is subjected to a thermodynamic heat pump cycle such that electrical energy is stored as thermal energy in thermal energy store(s), or TES, during a charging mode, and whereby the working fluid is subjected to a thermodynamic heat engine cycle in which the thermal energy in the TES is retrieved as electrical energy in a discharging mode, which system may be a gas (only) based cycle (i.e. non- condensing) pumped heat electricity storage system i.e. where the fluid does not undergo a phase change.

In this invention, the two TES systems and additional power machinery effectively time shift the work of compression that would have occurred in a normal gas turbine cycle by storing energy as thermal energy using thermal energy storage TES. The system time shifts and reduces the work of compression that is required to return the gas to the state that it would have been post compression in a normal gas turbine cycle. In discharge generation mode, the second TES system (cold TES) cools the air as it passes through the TES so that the work of compression required to get to, say 18 bar, is much lower than the normal work of compression to get to 18 bar from ambient temperature. As the now compressed air passes through the first TES system (hot TES) it is now heated to very close to the correct temperature (and pressure) that it would have been prior to entering to enter the combustor in normal generation mode.

Since there is no compressed air storage requirement, this hybrid system is not location dependent. Further, because it is based upon a PHES system, in common with such systems, it is able to supply heated compressed air to the combustion turbine over an extended period without too much fluctuation in temperature or pressure.

Furthermore, the cold TES may be designed to go to cryogenic temperatures, but without the use of liquid cryogens ie all of the heat transfer is direct between dry air and he storage media. There are no phase changes involved in the process and the cold TES does not need to operate at high temperatures.

The hybrid system may have one main inlet or a plurality of inlets for use in the respective generation modes, an inlet being disposed upstream of the first compressor with respect to operation in the normal generation mode and the same, or a further inlet, being disposed upstream of the second TES system with respect to operation in the discharge generation mode, as well as at least one main outlet located downstream of the expansion turbine through which the hot exhaust gases exit in either generation mode. If the charging mode is configured for open cycle operation, the first fluid working circuit will require an inlet and an outlet.

Conveniently, the hybrid system will be configured such that the direction of the air flow simply reverses around the circuit between the charging mode and the discharging mode, in particular, so that it reverses direction through each limb containing a respective store. However, other configurations where the flow remains in the same directions in certain limbs, but additional by-pass circuits are used, may also be employed. The PHES system should preferably be configured such that a thermal front passes down lengthways along each store in one direction upon charging and the opposite direction for discharging (and is retained in the store), so as to preserve the quality of the thermal front.

The PHES system, in the charging mode, is preferably configured for closed cycle operation. The use of a closed cycle avoids the issue of freshly drawn air introducing moisture and carbon dioxide into the PHES system.

At least one dehumidifier may be disposed in the first working fluid circuit of the PHES system, for use in any open cycle, charging or discharging mode.

Such a dehumidifier will usually only be operated in the discharge generation mode during open cycle operation in order to dehumidify incoming air before it enters the second TES system, and hence should be located upstream of the second TES system, so as to prevent ice build-up therein. If charging is also conducted in an open cycle mode, the same (or a further) dehumidifier should ideally be provided in the first working fluid circuit (again) upstream, for example, of any sub-zero temperature parts of the circuit e.g. upstream of the second TES system and, if required, also the second power machinery. Usually, it is simplest for a dehumidifier to be provided at or near a hybrid system air intake that is being used as such within the first fluid working circuit during (open cycle) operation in the charging or discharging mode.

By dehumidifier is meant any suitable form of air drying/moisture removal apparatus. This may comprise a desiccator, for example, a desiccator wheel. The desiccator is optionally regenerated using waste heat, which may be from the hybrid system or, as described below, subsequent apparatus downstream of the hybrid system e.g. operating a bottoming cycle or other waste heat recovery process and which itself exhausts gas with residual lower grade heat. Regeneration is preferably conducted in real time while the desiccator is active.

Usually, the system is configured to start-up in a normal gas turbine mode.

In order to increase overall efficiency of power generation, the hybrid system may be configured to run a combined cycle where power generation comprises a Brayton cycle followed by a bottoming cycle such as a Rankine cycle. The hybrid system may comprise an exhaust outlet operatively linked to downstream apparatus configured to operate a bottoming cycle or other waste heat recovery process, so as to extract further energy from the hot exhaust gases exiting the hybrid system in a generation mode. Thus, the hybrid system may form part of a (derivative) combined cycle gas turbine CCGT plant, preferably, where the bottoming cycle is a (Rankine) steam cycle such that waste heat from the turbine exhaust is recovered by a heat recovery steam generator HRSG to power a steam turbine.

The first compressor may be selected from a positive displacement or dynamic compressor. Usually, a dynamic flow compressor will be preferred, particularly an axial flow compressor, although other dynamic compressors such as centrifugal or mixed-flow or turbo compressors may be employed.

In normal generation mode, the first compressor will usually be required to operate over a fixed pressure ratio at high temperatures (e.g. compressor outlet temperatures of up to 620°C, or up to 550°C, or 500°C) as in a typical gas turbine, and hence, a compressor of the same design as one based on a typical compressor stage of a gas turbine, for example an industrial (heavy duty) gas turbine, will be preferred, especially if required for pressurising air up to 30, more usually, 13-19 bar.

The second power machinery may comprise at least one expander and at least one compressor respectively arranged in parallel within the first working gas circuit, disposed between the first TES system and second TES system, each, for example, paired in series with an appropriate one-way valve. In this way, the at least one expander may be used

(together with a one-way valve) to expand the working gas during the charging mode and the at least one compressor may be used (together with a one-way valve) to compress the working gas during the discharging mode.

Alternatively, the second power machinery may comprise at least one reversible machine operable as an expander and a compressor during charging and discharging, respectively.

In that case, the second power machinery may comprise a reversible positive displacement machine. Preferably, a linear reciprocating machine e.g. piston based machinery is used, this being suited to offering high efficiency, fast response and simpler flow ducting for the air flows.

The PHES system stores energy in the charging mode according to a thermodynamic heat pump cycle, by charging both TES systems to produce a hotter first TES system with time and a cooler second TES system with time (relative to their starting temperatures) until the stores are part or fully charged, as required. To store energy in the TES systems in a charging mode, air from the first compressor is directed to flow successively in the PHES working fluid circuit through the first TES system for transfer of thermal energy to the first TES system, the second power machinery for expansion of the air, through the second TES system for transfer of thermal energy from the second TES system to the expanded air, before exiting the hybrid system or being recirculated to the first compressor. By "flow successively" is meant flow in that order through those components, but including the possibility that additional components of that type may be present (in series or in parallel) and that other minor components may be interspersed between those major components (for example, heat exchangers, dehumidifiers, filters, etc).

Similarly, the PHES system retrieves energy according to a thermodynamic heat engine cycle, by discharging the TES to cool the first TES system and to warm the second TES system, To retrieve energy from the TES systems in a discharging mode, air is directed to flow in the PHES working fluid circuit successively through the second TES system for transfer of thermal energy from the air to the second TES system, the second power machinery for compression of the air, through the first TES system for transfer of thermal energy from the first TES system to the compressed air, before being directed to the combustor for combustion and subsequent expansion in the expansion turbine so as to provide a power output.

The PHES system may be configured to operate with a maximum operating pressure (e.g. corresponding to the first compressor air outlet pressure during charging, or, the second power machinery outlet pressure upon discharging) of not more than 30 bar, or 25 bar, or 20 bar. Higher pressures of up to 35 bar or even 40 bar may be used but a number of issues start to become apparent. Above about 30 bar, it is likely that the final stage compressor blades will be subjected to temperatures over 620°C, which is difficult to manage from a material and design perspective and is likely to require active cooling of the blades. This is normal practice for turbine blades, but not for compressor blades.

If the stores are based upon direct thermal transfer from a gas to a gas permeable media contained within a large pressure vessel, then the vessel becomes more difficult and costly to build at these higher pressures, without an equivalent increase in energy density ie the energy density might increase by 25% while the structural requirements of the vessel might increase by 100%. Where the gas permeable media consists of hot pressure bearing channels located within the media then the temperature and pressure mean that creep becomes a significant issue. This requires more exotic steels and thicker walls, which all reduce the effectiveness of heat transfer.

The PHES system may be configured to operate with a maximum operating temperature (e.g. first TES system air inlet temperature) in the first TES system of not more than 650°C, or even not more than 500°C, or not more than 450°C. This will usually correspond to a first compressor air outlet temperature, during charging, of not more than 650°C, or not more than 500°C, or not more than 450°C, although it is possible for the compressor outlet temperature to be lower and for other sub-systems to provide a boost in temperature to air entering the first TES system.

The first TES system may be provided, at the end which receives outlet air from the first compressor, with an electrical heater configured to provide additional thermal energy to heat air passing through the store. This may raise the gas and hence the storage medium temperature by at least 50, or 70 or even by at least 100°C (subject to not exceeding the maximum operating temperature of the first TES system). The electrical heater may be configured to operate during the charging mode and, while drawing electrical power, may allow less fuel to be consumed in the combustion chamber during discharge generation mode. In this way, higher temperatures may conveniently be stored in the first TES system without an associated pressure rise that would increase stores cost.

The "maximum store temperature" may, however, also be raised by less direct methods e.g. further upstream.

In one embodiment, a pre-heater system is provided in the first working fluid circuit upstream of the first compressor with respect to the charging mode, for preheating air prior to it entering the first compressor in the charging mode. In this way, the PHES system may store heat in the first TES system at a higher temperature during the (preferably closed cycle) charging without changing the pressure ratio i.e. without any increase in peak pressure (i.e. improving the energy density of the system) such that, in the discharge generation mode, the system is able to provide a higher temperature pressurised gas such that less fuel needs to be supplied to the combustor. The maximum power produced during the discharge-generation mode will remain unchanged if the pressure and the peak combustion temperature do not change.

For increased efficiency, the pre-heater system is preferably configured to supply thermal energy derived from waste heat to the air. This may be waste heat available in real time or that has been stored and may originate either from the hybrid system, associated systems (e.g. downstream), or other separate equipment co-located on-site.

The pre-heater system may comprise at least one heat exchanger provided upstream of the first compressor with respect to the charging mode, which heat exchanger is configured in the charging mode to receive heat (in real time) from at least one heat exchanger located downstream of the first TES with respect to the charging mode. In this way, the energy density and efficiency of the PHES system may be improved and in the discharge generation mode, the PHES system is able to provide a higher temperature pressurised gas such that less fuel again needs to be supplied to the combustor.

The upstream heat exchanger will usually be configured to receive heat via a heat transfer fluid HTF circuit from the at least one heat exchanger located downstream of the first TES with respect to the (preferably closed cycle) charging mode, for preheating air entering the first compressor. The upstream and downstream heat exchangers may, however, transfer heat directly between them if configured so as to form a counter-current heat exchanger.

To achieve preheating with heat exchangers linked across the first working fluid circuit e.g. by an HTF circuit (with the correct thermal gradient across them), it will be appreciated that gas circulating downstream of the first TES system must be sufficiently hotter than that circulating upstream of the first compressor. A TES will usually be operated such that a thermal front is retained within, and moves backwards and forwards within the store with storage medium on the hot and cold sides of the thermal front respectively held at approximately the last gas inlet temperature on charging the store (from the hot end) and the last gas inlet temperature upon discharging the store (from the cold end). The latter temperature will therefore be the temperature exhibited by the gas exiting the first TES system during charging (i.e. the last "minimum store temperature" of the first TES, which may or may not correspond to the very initial uncharged (e.g. ambient) temperature) and will normally be higher than ambient). (Usually, once up and running, the store will operate between a maximum store temperature and minimum store temperature, with the thermal front confined to run between the two store ends, but not leaving the store.) The last gas inlet temperature when discharging the store (from the cold end) may selectively be raised, during the previous discharge mode, by choosing the degree, if any, at which to discard any of the waste heat generated by the second power machinery work of compression ( i.e "cold compressor"). The simplest set-up is to configure the (or all) heat exchangers located downstream of the first TES so that they are bypassed or inoperative (ie bypassed to avoid any pressure drop through the heat exchanger or inoperative so that no HTF flows through them and hence the heat exchanger has no cooling effect after it is raised to approximately the air temperature in the circuit) during the discharge/generation mode, and hence, so that all the (low grade) waste heat from the cold compressor becomes stored (at a higher "minimum store temperature") in the first TES system. In the subsequent charging mode, the heat exchanger downstream of the first TES system is then operative to transfer that heat (in effect, waste heat that was temporarily stored in the first TES), for example, via a HTF circuit, to the upstream heat exchanger.

Heating the inlet air prior to compression is counter-intuitive for a number of reasons. In normal operation of a gas turbine GT (combustion turbine), it is well-known that the power output of the GT falls as the air inlet temperature rises. This is because warmer air is less dense so that the overall mass flow rate through the compressor/combustor/expander falls. In addition, it is harder to compress a hotter gas so that the amount of work required for the compression increases with temperature. A normal rule of thumb is that for every degree of temperature rise, the power output of a GT drops by about 0.5%.

Adding heat to a closed cycle PHES system (and/or not discarding waste heat from it) is also counter-intuitive as the normal problem with closed cycles is that the losses (due to irreversibilities within the cycle) build up in the system creating a temperature rise which needs to be rejected to the atmosphere via heat exchangers. Adding additional heat to such a closed system will make this heat rejection issue much more significant. High efficiency heat exchangers that provide small temperature differences between the HTF flow and the air in the circuit will still provide a significant pressure drop to the air flow. They also require power for pumps and fans. Consequently, where there are large amounts of heat rejection occurring through heat exchangers it will have an impact on system efficiency.

Increasing the compressor work relative to the expander work (by only changing the compressor inlet temperature) improves both the efficiency and energy density of the system. It improves the efficiency of the system because it increases the work ratio (on charging) ie the work input (the difference between the compressor work and expander work (on charging)) divided by the total work of all processes (on charging). This has the effect of reducing the effect of losses for different parts of the system, but in particular the power machinery. Furthermore, being able to operate in an open cycle mode on discharge means that these heat additions to the system can be easily removed from the system without the requirement of additional heat exchangers. The energy density, or amount of energy stored per Kg of storage medium (or volume), is improved because the media is holding higher temperature heat in the same mass of medium.

Any suitable thermal energy store (TES) system may be used to store the thermal energy in the first TES system and second TES system (i.e. as "heat" and "cold" respectively i.e. relatively higher and lower temperature thermal energy).

The first TES system is preferably configured to withstand a maximum operating temperature not higher than 650°C, or 600°C, or ideally not more than 550°C, and hence, the store materials and construction may be designed to withstand these temperatures. The maximum operating temperature is usualy within the range of 450-650°C, more preferably 500-600°C. The second TES system is preferably configured to withstand a minimum operating temperature not lower than -160°C, or more preferably, not lower than -140°C.

Where the PHES is configured to operate in a closed cycle during the charging mode, it is highly preferred if the section of the first working fluid circuit between the first TES system and the second power machinery is at or near ambient temperature, for example, within 30°C, or even 20 °C of ambient temperature, and, or, likewise that the section of the first working fluid circuit between the second TES system and the first compressor is at or near ambient temperature, for example, within 30°C, or even 20 °C of ambient temperature. Preferably, that part of the circuit is also configured to be at or near (within 0.1 bar of) ambient pressure, that is the second power machinery is configured to operate over a pressure ratio so as to provide a suitable outlet pressure that generates that condition.

Usually, at least one heat exchanger is provided downstream of the first TES system (and of the second TES system), upon charging, in order to allow removal of any undesired heat from the PHES system.

In one embodiment, the first and/or second TES system comprises a direct TES system comprising at least one thermal energy store disposed within the working gas (PHES) circuit and through which the working gas has a flow path for direct exchange of thermal energy to a thermal storage medium contained within the thermal energy store. The thermal storage medium may comprise a gas-permeable (e.g. porous or particulate) storage mass through which the working gas can flow for direct heat transfer to or from the thermal storage medium. The gas-permeable storage mass may be provided with channels, pores or plenums allowing gas flow through the mass. Thus the storage mass may be particulate, e.g. a packed bed of solid particles through which the air passes exchanging thermal energy directly, or, it may comprise a solid matrix or monolith provided with channels or interconnecting pores extending therethrough.

In a preferred embodiment, the thermal storage medium in the store is a particulate medium disposed in respective, individually accessible, layers arranged successively downstream of each other and the TES is configured so that the flowpath of the working gas passing through the particulate layers can be selectively altered in response to the progress of the thermal transfer in each layer. This may enable the flow path to bypass inactive upstream or downstream layers of the storage media where thermal transfer is complete or minimal. In this way, the pressure drop across the storage media, especially in the case of a particulate medium, can be minimised allowing the use of smaller particulate so as to provide more efficient thermal transfer. Preferably, the leading edge and trailing edges of the thermal front are separately controlled and the moving of the working gas into a fresh layer (leading edge control) and the removal of the working gas from an exchanged layer (trailing edge control) is selectively controlled so as to control the thermal front.

The total cross-sectional area available for fluid flow in the first TES system should be selected dependent upon the mass flow rate of the air around the first working circuit, which depends on the power rating of the first compressor. Since a gas permeable, direct heat transfer store requires a slow flow for efficient heat transfer, a much larger cross- sectional area will be required to maintain the same mass flow rate as though the first compressor. In the case of the first TES system, however, these need to be built to withstand high operating pressures. Once the pressure in 1st TES store exceeds about 25, or even 15 bar, it may be more cost effective to store heat in separate modular stores than a single store of the same cross-section. Thus, the first TES system may comprise a modular system comprising a plurality of stores arranged in parallel within the first working circuit.

There is further provided a method of operating a hybrid electricity storage and power generation system, wherein the hybrid system comprises:

(i) a pumped heat electricity storage PHES system comprising at least a first compressor, a first thermal energy store (TES) system, second power machinery, and a second thermal energy store (TES) system respectively arranged (in that order) within a first working fluid circuit and operates using air as the working fluid:

• in a charging mode that uses an open or closed, thermodynamic heat pump cycle to heat the first TES and cool the second TES, the first compressor and the second power machinery respectively compressing and expanding the air before entry to the first and second TES, respectively; and,

· in a discharging mode that uses an open, thermodynamic heat engine cycle to cool the first TES and to heat the second TES, the first compressor being inactive and the second power machinery compressing the air before entry to the first TES, whereby in this mode heated, pressurised air is produced from (e.g. an outlet of) the first TES; and,

(ii) a combustion turbine sub-system comprising a combustor and a downstream expansion turbine integrated with the first working fluid circuit by means of flow connectors; wherein the hybrid system operates in both of the following modes:

(a) a discharge generation mode in which the PHES system is operating in the above discharging mode, and the flow connectors are configured to direct the heated, pressurised air from (e.g. an outlet of) the first TES to the combustor for combustion and then to the expansion turbine for subsequent expansion to produce electrical power;

(b) a normal generation mode in which the PHES system is not operating in the above charging or discharging modes, and the flow connectors are configured to direct heated, pressurised outlet air from the first compressor to the combustor for combustion and then to the expansion turbine for subsequent expansion to produce electrical power.

In one embodiment, the pre-heater system is provided in the first working fluid circuit upstream of the first compressor, with respect to the charging mode, and preheats the air before it enters the first compressor in the charging mode. Such pre-heating should preferably raise the air temperature by not more than 120°C, more preferably, by not more than 100°C or even not more than 75°C, and will usually produce a rise in temperature of at least 20°C, or 40°C.

For increased efficiency, the pre-heater system ideally uses waste heat to preheat the air.

The pre-heater system may comprise at least one heat exchanger provided upstream of the first compressor, and a heat exchanger located downstream of the first TES, with respect to the charging mode, and in that mode the heat exchanger upstream of the first compressor receives heat from the heat exchanger located downstream of the first TES. In that set-up, in the charging mode, the heat exchanger located downstream of the first TES system may receive heat that has been selectively stored in the first TES system (e.g. at a selectively raised temperature) during the previous discharge generation mode by selective operation of that heat exchanger in that mode.

During the previous discharge generation mode the air inlet temperature to the first

TES system may be selectively raised by selecting the degree to which the heat exchanger located downstream of the first TES system discards heat (e.g. generated by the second power machinery). Thus, the heat exchanger may be inoperative or fully or partially bypassed or, otherwise restricted in its activity, so that it discards more or less of the (low grade) waste heat that is inevitably generated by the second power machinery (especially turbo/axial flow machinery). Alternatively, during the previous discharge generation mode, the air inlet temperature to the first TES system may be selectively raised by supplying at least some heat to the heat exchanger located downstream of the first TES system from an external source; this may therefore allow injection of higher grade heat, e.g. higher grade waste heat from downstream or associated systems operating concurrently in the discharge generation mode.

Usually, the heated, pressurised air respectively produced from the outlet of the first TES and from the first compressor in the respective generation modes is of a similar pressure, for example, within 0.5 bar of each other and if no pre-heating has been used, also within a similar temperature, for example, within 30K of each other.

In a second aspect, there is provided a hybrid combustion turbine electricity storage and power generation system comprising:

(i) a combustion turbine based system comprising a first compressor, at least one flow controller, a combustor and an expansion turbine arranged respectively downstream of each other; and,

(ii) an energy storage system integrated with the combustion turbine based system by means of the at least one flow controller, the energy storage system comprising at least a first thermal energy storage TES system for removing and returning thermal energy to compressed air passing through it upon charging and discharging the TES system, respectively,

wherein the energy storage system is configured:-

• to store thermal energy in a charging mode in which air is compressed in the first compressor and passes through the first TES system so as to heat the store;

• to retrieve thermal energy in a discharging mode in which air passes back through the first TES system so as to cool the store;

wherein the hybrid system is configured to be operable in the following generation modes:-

(a) a normal generation mode in which the energy storage system is not operating in the above charging or discharging modes, and the flow connectors are configured to direct heated, pressurised outlet air from the first compressor to the combustor for combustion and then to the expansion turbine for subsequent expansion to produce electrical power; and,

(b) a discharge generation mode in which the energy storage system is operating in the above discharging mode, and the flow connectors are configured to direct heated, pressurised air from the first TES system to the combustor for combustion and then to the expansion turbine for subsequent expansion to produce electrical power; and,

wherein a pre-heater system is provided upstream of the first compressor with respect to the charging mode, and is configured in the charging mode to preheat air entering the first compressor so as to increase the temperature of air entering the first TES system.

Use of a pre-heater system to add heat at this upstream point during charging (e.g. by a substantially isobaric heat transfer), allows more heat to be stored in the TES system without a commensurate rise in pressure, which would add to TES system cost. In this way, the energy density and efficiency of the hybrid system may be improved and in the discharge generation mode, the system is then able to provide a higher temperature pressurised gas to the combustor such that less fuel needs to be supplied to the combustor (the expansion turbine power output need not change). The heat addition may conveniently be by means of a heat exchanger and, because that additional heat stored in the first TES system is discharged through the gas turbine during the discharge generation mode, it does not create a problem of waste heat build-up.

The reason for the improvement in efficiency is that the amount of work carried out per unit mass of gas processed by the compressor increases, which means that the losses associated with processing a certain mass of gas actually fall. Furthermore, the amount of heat in the storage media is related to the mass of gas processed and the increased work translates to a higher energy density in the thermal stores. As the mass flow through the first compressor will fall (due to less dense air), it also allows for a reduction in the size of any second power machinery.

The energy storage system may comprise a pumped heat electricity storage system. As described above, adding a pre-heater system to a PHES based hybrid GT system may markedly improve the energy density and efficiency of such a system.

The pre-heater system may have any of the features, or be used in any of the methods, as described above in relation to the first aspect. In particular, such pre-heating should preferably raise the air temperature by not more than 120°C, more preferably, by not more than 100°C or even not more than 75°C, and will usually produce a rise in temperature of at least 20°C, or 40°C.

Brief Description of the Figures

Specific embodiments of the present invention will now be described, by way of example only, with reference to the accompanying drawings, in which:

Figure 1 a is a schematic diagram of a conventional open cycle gas turbine (OCGT) system of the prior art;

Figure 1 b is a schematic diagram of a conventional combined cycle gas turbine (CCGT) system of the prior art;

Figure 2 is a schematic diagram of a conventional Pumped Heat Electricity Storage System (PHES) system of the prior art;

Figure 3 is a schematic diagram of a hybrid power generation and energy storage system according to the present invention;

Figure 4a is a schematic diagram showing the system of Figure 3 operating in normal power generation mode;

Figure 4b is a schematic diagram showing the system of Figure 3 operating in a charge mode;

Figures 4c is a schematic diagram showing the system of Figure 3 operating in a discharge and generation (combustion) mode;

Figures 5a and 5b are respective schematic diagrams of further hybrid power generation and energy storage systems according to the present invention;

Figure 6a is a schematic diagram of a of further hybrid power generation and energy storage system according to the present invention with a pre-heater system, and operating in a charging mode;

Figure 6b is a schematic diagram of the system of Figure 6a operating in a discharge and generation mode;

Figure 6c is a schematic diagram of the system of Figure 6a but modified to include a further heat input, and operating in a discharge and generation mode; and,

Figure 6d is a schematic diagram of the system of Figure 6c operating in a discharge and generation mode.

The present invention relates to a hybrid energy storage and power generation system based upon a PHES system that has been modified to incorporate combustion turbine functionality. Each of those technologies is briefly outlined below.

Figure 1 a shows a typical layout of a conventional prior art, open cycle, gas turbine (OCGT) 10 used for peaking power generation. This operates an open Brayton cycle and comprises an upstream compressor 1 1 normally directly coupled to a downstream turbine (expander) 14 and driving a generator 15 (e.g. connected to a transformer/grid). In a normal configuration the compressor, turbine and generator are all directly mechanically coupled on the same shaft by drive couplings (not shown).

The working gas flow circuit comprises compressor 11 , combustion chamber 12 and turbine 14 respectively arranged downstream of each other and combustion chamber 12 is separately supplied with natural gas 13. Air enters the compressor at ambient conditions (e.g. 288K, 1 bar) and is compressed up to a higher pressure and temperature (e.g. 702K, 16.6 bar). The hot high pressure air enters the combustion chamber where it is mixed with natural gas and caused to combust, heating the gas to a much higher temperature (e.g. 1486K, 16.6 bar). This air is then expanded back to atmospheric pressure in the turbine, which produces more power than the compressor absorbs, hence there is a net generation of power that can drive the generator 15. The cooled air is exhausted from the turbine well above ambient temperature (e.g. 816K, 1 bar).

Figure 1 a shows a simple open cycle gas turbine employing a top thermodynamic (Brayton) cycle to generate electricity. However, there are a number of known different variants on this simple cycle that can improve efficiency and these may involve steam injection, reheat, recuperation and/or intercooling. Furthermore, the pressure ratios and temperatures in the different stages may be different depending upon the manufacturer and/or varied as required. For example OCGT aeroderivative gas turbines normally operate at higher pressure ratios than industrial gas turbines in CCGT's.

Commonly, a top Brayton cycle is combined with a bottom thermodynamic cycle such as a Rankine (or Kalina) cycle, whereby further heat is recovered from the exhausted Brayton fluid and used to generate additional electricity in "combined cycle" power plants. 5 Figure 1 b shows a typical layout of a conventional prior art combined cycle gas turbine (CCGT) 30 used for power generation.

The initial section comprises a gas turbine that is similar to that used in the OCGT 10; however it normally operates so that the exhaust temperature is slightly hotter either by operating at a lower pressure ratio or by combusting to a higher turbine inlet temperature.

10 After the exhaust from the turbine 14, the hot high temperature exhaust gas (e.g. at 816K, 1 bar) enters a heat exchanger 16, in this case a heat recovery steam generator or HRSG, where it is cooled while heating a counterflow of water that is at high pressure. The water normally becomes superheated during the heat exchange process and is then expanded through steam turbine 17 to a lower pressure. This steam is then condensed in condenser

15 20 before being pumped back to a high pressure by water pump 19 to return to the heat exchanger 16. The condenser 20 is normally supplied with a cooling water flow from a river or the sea. Steam turbine 17 is normally directly coupled to water pump 19 by generator 18 and the expansion of the steam in the steam turbine 17 produces more power than the water pump 19 absorbs, resulting in a supplementary net production of power according to

20 a Rankine cycle.

Turning to consider a PHES system, this is ideally operable according to a Brayton cycle when discharging and an inverse Brayton cycle during charging. In the charging mode, the PHES system stores energy according to a thermodynamic heat pump cycle, by charging the TES to heat a first TES system and cool a second TES system; each TES

25 system may comprise one or more stores in series and/or parallel. In the discharging mode, the PHES system retrieves energy according to a thermodynamic heat engine cycle, by discharging the TES systems to cool the first TES system and to warm the second TES system. Separate power machinery or the same reversible power machinery may be used for the charging and discharging processes.

30 By way of example only, a prior art PHES system and its operation is now described with reference to Figure 2, which shows a PHES system 50 based on turbomachinery (e.g. axial flow machinery) and operating according to a closed Brayton cycle.

System 50 comprises first and second TES 140 and 160 and, in this example, turbo based machinery 1 13, 1 14, 1 15, 1 16 disposed within a working gas circuit. The TES 140,

35 160 respectively comprise a pressure vessel 142, 162 containing a thermal storage medium 143, 163 through which the working fluid passes for direct transfer of thermal energy to, or from, the storage medium where it is stored as sensible heat.

In a charge mode using electrical power (e.g. supplied from the grid via a motor/generator), the working gas circulates in a clockwise direction. Valves 122 and 123 are set so that gas can flow from TES 160 to TES 140 via compressor 1 13. Valves 124 and 125 are set so that gas can flow from TES 140 to TES 160 via expander (turbine) 1 15. Gas leaves the top of TES 160 and flows through heat exchanger 146 rejecting any waste heat to atmosphere, before being compressed to a higher temperature and pressure by compressor 1 13. The gas flows through TES 140 where it is cooled to near ambient and in the process transfers heat (the heat of compression) to the thermal media 143. The gas is further cooled to near ambient temperature by heat exchanger 145, which helps the system to reject waste heat. Gas exits heat exchanger 145 and is expanded from near ambient to a lower pressure and sub-ambient temperature in expander 115. The now cold gas enters TES 160 at the bottom and is warmed up as it passes through the thermal media 16, before gas exiting the TES 160 to pass around the system again. The work of the compressor 113 is greater than the work of the expander (turbine) 115 and hence this system requires work to be input to drive the process. This normally occurs via an electric motor and is how electrical energy is input to the system.

Charging therefore leads to heating of first TES 140 and cooling of second TES 160 relative to their respective starting temperatures such that, in effect, heat is "pumped" up a thermal gradient from the second TES to the first TES; these are often referred to as the cold store and hot store, respectively, since they respectively become colder and hotter during charging; however, whether or not the second store of a particular system operates at sub-ambient or sub-zero temperatures will depend on the working gas and pressure ratios selected. Charging may continue until the two stores are part or fully charged.

In a discharge mode, the working gas reverses the direction of flow through the TES circulating in an anti-clockwise direction. Valves 122 and 123 are set so that gas can flow from TES 140 to TES 160 via expander (turbine) 1 14. Valves 124 and 125 are set so that gas can flow from TES 160 to TES 140 via compressor 116. Gas leaves the top of TES 140 at the higher pressure and temperature (stored in the medium) and is expanded to a lower temperature and pressure by expander (turbine) 1 14. The gas is further cooled by heat exchanger 146 before it flows through TES 160 where it is cooled to sub-ambient temperatures and in the process transfers heat to the thermal media 163. The gas exits TES 160 and is compressed by compressor 1 16 to a higher temperature and pressure. It is then cooled to near ambient temperature by heat exchanger 145, which helps the system to reject waste heat. The gas exits heat exchanger 145 and enters TES 140 where it is now heated up to high temperatures. The now hot gas exits the TES 140 and can pass around the system again. The work of the compressor 116 is less than the work of the expander (turbine) 114 and hence this system generates power that can be converted to electricity. This normally occurs via an electric generator and is how electrical energy is output from the system.

It will be appreciated that this cycle involves two compression/expansion processes and two heat transfer processes on charge and also on discharge. As a result, the system is very sensitive to the efficiency of both the compression/expansion processes and the heat transfer processes.

Turning to consider the expansion/compression processes, it is well understood in the art that reciprocating compression and expansion processes can have high 'reversible adiabatic' efficiencies if correctly designed. This is not the case with axial flow machinery, where the process, while close to adiabatic, is highly irreversible. For example, if axial flow machinery is used then the normal measure of efficiency is polytropic efficiency. Current state of the art axial flow machinery have polytropic efficiencies as high as 93% and in most normal cases a figure of 90% is more achievable. An ideal PHES system using axial flow compressors and expanders with a 93% polytropic efficiency and ignoring all other losses will have a round-trip efficiency of only 73%. This is shown in the first line of the Table 2 below where H refers to the hot machinery (compressor 113 and expander (turbine) 114 in the Fig. 2 system) polytropic efficiency and C to the cold machinery (compressor 116 and expander (turbine) 115 in the Fig. 2 system) polytropic efficiency. The major additional real losses are electrical losses, mechanical losses, duct losses, pressure losses through the thermal stores and irreversible heat transfer within the thermal stores. When factoring in these other real losses, this efficiency will drop to below 55%.

The majority of these losses must be rejected from the system and this means that heat exchanger 145 must be of sufficient capacity to carry out this function.

Table 1 : PHES Round Trip cycle efficiency from Polytropic Efficiency (only)

Taking an example where the turbomachinery is assumed to have a polytropic efficiency of 90%, and the working fluid is assumed to be argon, which is a monatomic gas, these losses are evident as temperature rises across the machinery in the Table below. (With argon, it is possible to achieve a larger change in temperature for a given pressure ratio, but obviously only where a closed cycle is used.) Table 2 shows temperatures and pressures around the closed gas circuit in charging mode (clockwise) and discharging mode (anti-clockwise), by way of example only, for a selected pressure ratio of 1 :9, with both heat exchangers exchanging heat back to approaching close to ambient in both modes.

Table 2: Exemplary temperatures and pressures for Figure 2 PHES system

This table clearly shows the effect of low efficiency in that if you compress argon from 1 to 9 bar the temperature rise is from 298K to 799K. However, when you expand it back to the original temperature it is now much hotter ie 359K. The increase in temperature is directly linked to the irreversibility of the process and would usually be rejected from the system as waste heat at heat exchanger 146, for example. A lower polytropic efficiency will give a higher temperature rise.

By way of comparison, reciprocating machinery is expected to achieve higher efficiencies that are equivalent to 98% polytropic efficiency for both hot and cold machinery (with less waste heat rejection required). From this it can be seen that, including real losses, an overall efficiency for a PHES system of around 74% can be achieved.

Axial flow machinery, however, has some significant advantages over reciprocating machinery. It is relatively low cost, has very high power densities, normally only has one moving part (a single spinning shaft), is reliable and has low maintenance costs and can operate at very high temperatures. The turbine section ('expander') of a gas turbine can normally tolerate continuous gas temperatures of 1400-1600°C. From a thermodynamic perspective, this is extremely positive as it means heat engine cycles can be designed with very high peak temperatures. (A CCGT has the highest thermal efficiency of any power plant currently built because the dual combination of the gas turbine and steam turbine effectively correspond to a heat engine operating over a temperature range of 1600°C down to ambient.) Large industrial gas turbines are normally made up of axial flow compressors and axial flow turbines.

Axial flow compressors or turbines are dynamic compression/expansion devices and where they are referenced by way of example below it should also be understood that other dynamic/turbo machinery such as centrifugal compression/expansion machinery could be used.

Figure 3 shows a first hybrid power generation and energy storage system according to a first embodiment of the present invention. This embodiment is based on a PHES system operating using air as the working fluid, but the flow circuit is modified in the discharge mode so that instead of hot, pressurised gas exiting the hot store and expanding through the PHES turbine machinery, it is directed to a combustion chamber for combustion and subsequent expansion in a (higher temperature) turbine. While the charging mode operates as a closed cycle (this being highly preferred), the discharging mode operates as an open cycle and hence the apparatus is also modified in this regard as detailed below.

The PHES part of the system 70 again comprises first and second thermal stores 140 and 160 acting as the hot and cold stores, compressors 1 13,1 16 and expander (turbine) 1 15, 3 way valves 122,123, 124 and 125 and heat exchanger 145.

Additional components are an air inlet 251 to the system 70 and a dessicator/dehumidifier 250 located downstream of valve 122 between it and cold store 160, so that ambient air may be drawn in through dessicator/dehumidifier 250. Dessicator/dehumidifier 250 may or may not be in operation when air is drawn in through the air inlet 251. The dessicator/dehumidifier may be based upon a number of different known technologies that can remove moisture from an air flow (i.e. dessicate/dehumidify the flow). One example of the many solutions it to use a dessicant wheel that is regenerated from waste heat from the exhaust of the gas turbine or, where a bottoming cycle with an HRSG is present, the exit of the HRSG before it enters the stack (exhaust chimney). An alternative approach may involve the use of a chiller system to lower the inlet air temperature. Or a combination may also be used, so that the air may be chilled prior to passing a dessicant. As the dessicator/dehumidifier has an impact on storage efficiency, it is preferable to minimise the amount of energy that it uses.

A heat exchanger 146 is optionally located between the cold store 160 and first compressor 1 13 so that in the (clockwise) charging mode it is downstream of the cold store and upstream of the first compressor. It may or may not be in operation when air is drawn past. Turning to the combustion turbine apparatus that is integrated into the PHES system, there is a combustion unit 213 and a high temperature expander (turbine) 214 that exhausts through outlet 252. This exhaust is at high temperature and 252 may optionally be connected to a steam turbine via an HRSG to recover the energy in the exhaust. Alternatively, other known sub-systems for extracting further energy from the exhausted air of a gas turbine (Brayton top cycle) may similarly be incorporated into this hybrid system.

While Figure 3 depicts a PHES system 70 with valves 124 and 125 controlling access to two parallel limbs (for flow in opposite directions) and provided with a compressor 115 or turbine expander 116, respectively, each of which will usually be axial flow machinery intended to handle unidirectional flow, it is possible for the valves and two limbs to be removed, and replaced with a single limb containing reversible power machinery that can handle flow in both directions, acting as an expander for flow in one direction and as a compressor for flow in the reverse direction. In this location, reversible power machinery such as reciprocating machinery could be used because the lower temperatures in this part of the circuit mean that the density of the gas is higher and hence the mass flows through a certain size of machinery will be higher.

Figure 4a illustrates the operation of the embodiment in normal generation mode. Air is drawn in through air inlet 251 and valve 122 is set to divert the flow through heat exchanger 146, which is not in operation (although a bypass circuit around the exchanger could equally be used) and to compressor 113, which compresses the gas to a higher temperature and pressure. If compressor 1 13 is based on an industrial gas turbine then this is likely to be in the region of 13-18bar and 672K to 746K. Valve 123 directs the flow to combustor 213, where fuel is combusted to raise the temperature of the flow further and then expanded in expander (turbine) 214. The exhaust is likely to be in the temperature range 825K to 875K and is suitable for driving a 'bottoming' steam or other similar cycle. The exhaust exits the system through outlet 252.

All of the embodiments of the present invention are depicted in the Figures as using generic compressors or expanders (turbines). However, it should be understood that compressor 1 13 and expander (or turbine) 214 with combustor 213 may conveniently be based around conventional industrial gas turbine components. For example the compressor could be the compressor stage of a gas turbine, the combustor and turbine could also be from the same gas turbine, but configured to run independently. As has been mentioned, the exhaust from the expander (turbine) will also be at high temperature and may be run through an HRSG to generate additional energy, although the steam cycle part is not shown in the Figures for the sake of simplicity.

The three way valves are expected to allow flow between any two directions at any time, but to restrict flow through the third direction. The connections will be selected depending upon the mode of operation. Such valves may replaced with alternative valves, or removed, for example, if an open cycle is used (e.g. valve 122 could be removed if the circuit is converted at that point to an open circuit with a pair of respective open ducts (e.g. 5 inlet and outlet).

The following table gives some indicative temperatures and pressures around different parts of the circuit in normal generation mode (Fig. 4a):

Table 3: Exemplary temperatures and pressures for Figure 4a mode

10 Figure 4b shows the hybrid system of Figure 3 operating in a charging mode. Valve 122 is set so that gas can pass from TES 160 to heat exchanger 146. Dessicator/dehumidifier 250 is inactive and could be bypassed by a bypass circuit if required. Valve 123 is set so that gas can pass from compressor 1 13 to TES 140. Valve 124 is set so that gas can pass from heat exchanger 145 to expander (turbine) 1 15. Valve 15 125 is set so that gas can pass from expander (turbine) 1 15 to TES 160.

Air is drawn from the top of TES 160 and passes through optional heat exchanger 146. It is explained in a further embodiment that this heat exchanger may provide additional pre-heat to the air flow in certain applications. However, this may not always be the case and it may be beneficial to cool the flow by rejecting heat to ambient if the 0 ambient temperature is lower. The difference in temperature might be significant (eg 30-50 °C) if a dessicator wheel has been used during open cycle mode, which has added heat to the inlet flow. It would normally be beneficial to cool the flow to ensure that the post compression temperature (after compressor 113) remained below a certain level, for example 620°C (e.g. to remain within the compressor operational temperature range). The air then passes through compressor 1 13 which compresses the gas to a higher temperature and pressure. If compressor 1 13 is based on an industrial gas turbine, then this is likely to be in the region of 13-18bar and 672K to 746K if the inlet temperature is around 288K. It should be noted that some designs of industrial gas turbines operate at higher pressure ratios such as 30 bar, which will lead to higher temperatures post compression. Valve 123 directs the flow to TES 140 and the hot high pressure air passes through TES 140 and transfers heat from the air to the thermal media 143 heating the TES 140.

TES 140 comprises a thermally insulated vessel 142 and thermal storage media 143 which may be any suitable TES store. Thermal media 143 may comprise a packed bed of suitable thermal media such as high temperature concrete, ceramic components, refractory materials, natural minerals (crushed rock) or other suitable material such as alumina. Thermally insulated vessel 142 must be designed so that the high pressure flow (usually at between 12 and 30 bar and between 675K and 925K) can pass through the vessel transferring heat directly to/from the thermal media 143. As the media 143 is in the form of a packed bed with direct heat exchange to compressed gas, the thermally insulated vessel 142 will need to be an insulated pressure vessel. TES which are described as suitable for use in prior art PHES systems may equally be used in the hybrid system of the present invention.

The following table gives some indicative temperatures and pressures around different parts of the circuit operating in charging mode (Fig. 4b):

Table 4 : Exemplary temperatures and pressures for Figure 4b mode *[F is air that has been drawn into the cold store from ambient and is therefore at ambient ie 288K, while D is the result of heat exchange through a wall and would therefore be hotter than ambient e.g. 10 degree difference.]

As explained above, the efficiency of both compression/expansion processes and both thermal transfer processes needs to be maximised. In this respect, examples of thermal stores that may be especially suitable for removing/returning thermal energy directly at high temperatures of at least 500-650°C, and pressures up to 30 bar, are the thermal stores described in WO2012/127178 and WO2013/60650 to Isentropic Ltd.

WO2012/127178 describes a TES where the storage media is divided up into separate respective downstream sections or layers. The flow path of the heat transfer fluid through the layers can be selectively altered using flow controllers dependent upon the progression of the thermal front, so as to access only certain layers at selected times, so as to avoid pressure losses through inactive sections upstream or downstream of the sections where the thermal front is located and to control thermal front behaviour so as to maximise store utilisation. WO2013/60650 describes a similar valved, layered store with functionality allowing it to store thermal energy in a controllable manner.

Thus, in the present system each TES preferably has layered, individually access controlled storage, more specifically, a sensible heat store incorporating a gas-permeable, solid thermal storage medium into and out of which heat is transferred directly from the fluid, disposed in respective, downstream, individually access controlled layers (e.g. controlled by flow controllers or valves), in order to improve the efficiency of storage of heat up to temperatures of about 650°C. It should be noted that the flow velocity through such a layered (or continuous) bed may be as low as 0.5 m/s or even lower. Slower flows increase the time available for heat transfer, which is important as heat transfer is a time based process, and hence, improves the efficiency of the heat transfer processes in this hybrid system.

The size of the compressor 1 13 determines the mass flow rate of air passing through the TES 140. The greater the mass flow rate, the higher the power rating of the system ie a compressor using twice the power would generate twice the mass flow rate. For derivative machinery from gas turbines, it will be appreciated that the gas velocities leaving the compressor are very high and the mass flow rates are also very high. For example, they may be in the region of 500kg/s. This hot high pressure flow is preferably slowed down efficiently before passing through the store. This means that a high efficiency diffuser would normally be used to allow some or most of the dynamic pressure to be recovered as static pressure. Effusers may similarly be employed upon exiting the stores.

It will also be appreciated that the TES 140 may be quite large. For example, if using a mineral storage media, it is likely that the system will require 3-4000 tons of media per hour of charging. A 10m diameter thermal store that is 20m high might contain 2000 tons of storage media. So for 6 hours of storage it would require around 20,000 tons of media. This could be provided within one vessel that was 25m in diameter and 20m high, however the structural requirement for building large pressure vessels are significant and it is likely that using many smaller vessels will be both simpler and safer. Consequently, the TES 140 may be made up of multiple vessels, where the flow passes through them all in parallel. In this example, there may be 10 pressure vessels, each filled with 2000 tons of thermal media, which would provide the correct cross-sectional area for gas flowing through the storage media and also ensure a low pressure drop. There may be many, for example, 24 separate pressurised stores that make up TES 140.

In TES 140, the air is cooled by the particulate matter and exits the store close to ambient temperature. It passes through an optional heat exchanger, 145, where it exchanges heat with the surroundings and cools to a temperature very close to ambient. If there is no temperature difference then heat exchange 145 will not be in operation. The high pressure but cooled air exits heat exchanger 145 and passes via valve 124 to expander (turbine) 115, which expands the gas and cools it down to around 145K and atmospheric pressure ~1 bar. The air then passes via valve 125 into TES 160. The slow moving air now passes through a second thermal energy store (TES) 160. This store could be manufactured in a similar way to TES 140 but structurally it is likely to be simpler as it is designed to be at or very close to ambient pressure. On passage through TES 160 the gas receives heat from the packed bed material and exits the store at a temperature close to ambient, while the store is correspondingly cooled. In this way the air leaves the TES 160 and returns to compressor 1 13.

It is important to note that in this system the same compressor is used for both the charging cycle and for the normal power generation mode. The charging mode is conveniently operated as a closed circuit and the temperature of the air entering the TES 140 is very similar (e.g. within 20°C or less ) to that of the air entering the combustor 213 in normal generation mode. By using a closed circuit, any residual moisture and C02 in the circuit freeze out early in the process and allow the cycle to operate in a completely dry mode. Charging continues as long as required or until the thermal stores are 'full'.

Figure 4c shows the hybrid power generation and energy storage system of Figure 3 operating in a discharging and generation mode.

Valve 122 is set so that air can pass from dessicator/dehumidifier 250, which is active, to TES 160. Valve 125 is set so that gas can pass from TES 160 to compressor 1 16. Valve 124 is set so that gas can pass from compressor 1 16 to heat exchanger 145. Valve 123 is set so that air can pass from TES 140 to combustor 213.

Air is drawn in from the outside via inlet 251 and passes through dessicator/dehumidifier 250, which will dry the air flow prior to passing into TES 160. If no dessicator/dehumidifier is used then it will be necessary to have a de-icing strategy for the cold store as there is likely to be significant ice formation as the incoming air is cooled, which will have an impact on efficiency.

The now dry air passes through TES 160, where the thermal media 163 cools the air to 145K (for example) rewarming in the process. The air then passes through compressor 1 16 which compresses the gas to 16.6 bar and 353K (for example). The air then passes through heat exchanger 145, which rejects heat to ambient and cools the flow back towards ambient. The now high pressure cool air enters TES 140 and is heated by thermal media 143, which media in turn is cooled. The air leaves the TES 140 at a temperature and pressure similar to that which compressor 1 13 would normally generate in operation and enters combustor 213. In combustor 213, fuel is burnt to raise the temperature of the gas before it enters expander (turbine) 214.

The following table gives some indicative temperatures and pressures around different parts of the circuit in discharging and generation mode (Fig. 4c)

Table 5: Exemplary temperatures and pressures for Figure 4c mode

It will be noted that the "cold compressor" 1 16 is not able to return the gas upon discharge to point D at the same temperature (i.e. the 298K as received by "cold expander" 1 15 in the parallel limb during charging (D= 298K)), but rather, due to irreversibilities in the axial flow compressor 116, the gas exits it at D=353K more than 50 degrees warmer. In this embodiment, that waste heat is discarded at heat exchanger 145 to ambient such that upon exit from the heat exchanger the gas has cooled to within, say 10 degrees, of ambient (298K). In this way, during discharge, waste heat is discarded and the "discharged temperature" of the hot TES system 140 is effectively "reset" to near ambient/298K. This is in contrast to the later embodiments of Figure 6, as described below. Alternatively the heat rejection may occur on both charge and discharge so part of the waste heat is rejected on discharge and part of the waste heat is stored in the first TES to be rejected on charge. For the heat exchanger to perform equal heat rejection on charge and discharge, the temperature drop will be greater when the system is in discharge mode as there is a larger temperature difference with the ambient conditions ie drop temperature from 353K to 320K on discharge and from 320K to 298K on charge.

The compressors and expanders (turbines) may be connected electrically or mechanically, for example, by way of a clutch and/or by way of their own separate electrical motor and/or generator or by some other suitable means.

By way of example only, the PHES part of the hybrid system using air as the working fluid may operate according to a closed inverse Brayton heat pump cycle when charging and an open Brayton heat engine cycle when discharging.

The hot thermal store 140 may comprise an insulated storage vessel 142 that is designed to withstand higher pressures and any suitable gas permeable storage media as detailed above, for example, particulate thermal storage media 143. Particulate media will normally be some form of crushed mineral, but may be any other suitable material.

The cold thermal store 160 may comprise a low (near atmospheric) pressure, low temperature thermal store 160 that may comprise an insulated storage vessel 162 and any similarly any suitable thermal media 163; for example, particulate media may be used and may comprise a plastics material capable of maintaining integrity at lower temperatures and that need not necessary be suited to higher temperature use.

The cold TES need not operate at high temperatures, which means that it is possible to use a storage medium material that would not be suitable in a higher temperature TES, such as, for example, an organic material.

Organic materials normally have much higher heat capacities at low temperatures than minerals. They also have lower densities and lower thermal conductivity. This means that for a store that will store the same quantity of 'cold' it is likely to be larger and lighter than a store where the media is made from a mineral or similar material. Much of the cost of the thermal store can be related to the ability to support the mass of the storage media. Consequently, it is possible to build a lighter lower cost, but larger, thermal store using an organic material. The organic material may also be more resistant to spalling and cracking and any dust may be likely to cause minimal damage.

The organic materials may be either natural, such as natural rubber, or synthetic, such as polypropylene, or other similar low cost materials, including recycled/re-used waste materials such as the remains of car tyre rubber (e.g. in the form of crumb rubber). Car tyres are a combination of rubber mixed with fillers and steel and when recycled the steel is removed and the remaining material (comprising mainly rubber) can be used as the particulate heat storage structure.

The at least one organic material may be selected so as to be suitable for repeated use at temperatures below 100 deg C, or even at temperatures below 150 deg C.

Filters may be added to the circuit to ensure that the amount of particulate in the air flow is kept low or that the size of any particulates is below a certain threshold. For example there can be an issue with dust generated by certain types of crushed rock that may be used within a thermal store as they are thermally cycled between two different temperatures. Consequently, it is understood that filters may be added to the circuit as required by good engineering practice. These filters may be located within the thermal stores or within the duct work. Generally they will create lower losses if they are located within regions of the system where the gas velocities are low.

An example of a hybrid system with derivative components based around a Siemens v84.3 gas turbine might operate with the following variables, by way of example only:-

GENERATION

Hot Compressor

Compressor Efficiency 0.9

i Inlet Temperature K 288

Mass Flow kg/s 430

i Inlet Pressure bar 1

Outlet Pressure bar 16.6

Outlet Temperature K 702

Work of Compression J/s/kg 416392

Work of Compression MW 179

Hot Turbine and Combustion

i Mass flow Rate pre combustion kg/s 430

i Natural Gas added in Combustor kg/s 20 i Equivalent Turbine Inlet Temperature K 1486

Turbine Efficiency 0.88

Adjusted Gamma 1.32 i Exhaust Temperature K 816

Work of Expansion J/s/kg 795556

Work of Expansion MW 358 i Electrical and Mechanical Losses MW 9

Power Output for Normal Generation Mode MW 170

Table 6: Normal Generation Mode (Fig. 4a)

CHARGING

Hot Compressor

i Compressor Efficiency 0.9 i Inlet Temperature K 288

Mass Flow kg s 430 i Inlet Pressure bar 1

Outlet Pressure bar 16.6

Outlet Temperature K 702 i Work of Compression J/s/kg 416392 i Work of Compression MW 179

Temperature Entering Hot Store K 702

Temperature Exiting Hot Store K 320

Temperature pre heat exchange K 320 i Temperature post heat exchange K 298

Cold Turbine

Turbine Efficiency 0.9 i Inlet Temperature K 298 i Mass Flow kg/s 430 i Inlet Pressure bar 16.6

Outlet Pressure bar 1 i Outlet Temperature K 145

Work of Expansion J/s/kg 154012

Work of Expansion MW 66

Temperature Entering Cold Store K 145

Temperature Exiting Cold Store K 288

Temperature pre heat exchange K 288 i Temperature post heat exchange K 288 i Electrical and Mechanical Losses MW 7

Charging Power Input for Charging Mode MW 120

Table 7: Charging Mode (Fig. 4b)

DISCHARGING

Temperature Entering Cold Store

from dehumidifier K 288

Temperature Exiting Cold Store K 145

Cold Compressor

Compressor Efficiency 0.9

Inlet Temperature K 145

Mass Flow kg/s 430

Inlet Pressure bar 1

Outlet Pressure bar 16.6

Outlet Temperature K 353

Work of Compression J/s/kg 209286 i Work of Compression MW 90

Temperature pre heat exchange K 353

Temperature post heat exchange Temperature Entering Hot Store K 320

Temperature Exiting Hot Store K 702 i Hot Turbine and Combustion

Mass flow Rate pre combustion kg/s 430

Added Gas kg/s 20

Equivalent Turbine Inlet Temperature K 1486

Turbine Efficiency 0.88

Adjusted Gamma 1.32

Exhaust Temperature K 816

Work of Expansion J/s/kg 795556

Work of Expansion MW 358

Electrical and Mechanical Losses MW 13

Discharging +Generation Power Output MW 255

Table 8: Discharging and Generation Mode (Fig. 4c)

If properly designed the pressure drops through the system will be quite low relative to the system pressure. Hence, it is a reasonable approximation to refer to different elements in the high pressure part or low pressure as being around the same pressure. There will however be small differences between different parts depending upon the design. For example the pressure drop across a store might be of the order of 0.05bar.

It may be desirable, during the charge (closed) cycle, to vary the quantity of gaseous mass in the circuit. This variation is related to the changes in temperature of each store and in the geometry of each store. It is preferable to design the system so that a small amount of air can be vented from point F during the charge cycle. This can be achieved by fitting a non-return valve to the circuit at this point. The vented air will be dry and there is no need to add new air that might have moisture and C02. On discharging it is necessary to add air to the mass to replace this vented air, however when operating in open cycle mode this happens automatically.

Figures 5a and 5b are respective schematic diagrams of two further hybrid power generation and energy storage systems according to the present invention, comprising alternative modified versions of the system of Figure 3.

Figure 5a shows a hybrid system 90 which has been modified to include an electric heating element 270 installed at the hot end of the hot store 140. The efficiency of a CCGT is in the region of 60% and an OCGT 40%. Consequently, any heat added to the air post compression but before the combustor 213, is converted to electrical power at either a 60% or 40% efficiency and reduces the quantity of gas that is required. While a 40% efficiency is low, a 60% efficiency is very acceptable as there is minimal additional equipment that needs to be added for the additional power input.

In this hybrid system 90, such heat may be input during the charging mode. Thus, an electric element installed in the hot store at its hot end may be turned on during charge mode and the air from the compressor 1 13 is heated further, say to 875K from 675K. Hence, during charging, the media will be heated up to close to this higher temperature causing more energy to be stored, while the temperatures downstream around the circuit remain unchanged; upon discharge, the electric element is switched off and the additional heat is blown out of the circuit through the combustor and expansion turbine (i.e at a higher gas outlet temperature) thereby allowing a fuel saving.

Figure 5a also shows a reversible compressor/expander 1 15' fitted into the cold part of the circuit. This removes the requirement for three way valves 124 and 125. The compressor/expander could conveniently be a reciprocating machine where it can change function between compressor and expander by changing valve timing, quite rapidly if required.

Figure 5b shows a further hybrid system 190 which is similar to that of Figure 3 but has been modified to allow charging in an open cycle mode. Thus, three-way valve 122 is removed and the circuit made open at that point so that there are two ducts 251 ' and 251 " open to atmosphere. In charging mode, duct 251 ' acts as an inlet and 251 " acts as an outlet for open cycle charging and a dessicator/dehumidifier 250' is added. With suitable flow controllers this unit could be combined with dessicator/dehumidifier 250. In discharging mode, duct 251 " acts as an inlet to discharge the PH ES system (passing gas first through the dehumidifier downstream of that inlet 251 ") and, in normal generation mode, duct 251 ' acts as an inlet for normal generation mode i.e. through the combustion turbine sub-system, where a dehumidifier is not required downstream of the inlet.

Turning to consider the overall system efficiency of the Figure 3 hybrid system, the normal charging cycle in Figure 4a is the same process as a conventional OCGT, except that the compressor 1 13 and the expander (turbine) 214 may not be directly coupled on the same shaft. In this mode the losses of both the compressor 1 13 and expander (turbine) 214 are budgeted for within the generation cycle. Consequently, if a storage mode is added that effectively decouples the timing of the two processes then only the losses associated with the additional processes need to be budgeted. Hence, when carrying out an analysis from a storage perspective the efficiency of these two processes can be treated as lossless.

This proposal is fundamentally different from proposed TES systems where heat is stored in a TES and extracted from the TES via a Brayton cycle, for example in solar thermal plants. In these cases, the TES is acting as a replacement or supplementary heat source to a combustor - ie it either eliminates or reduces the amount of fuel required to generate power. In this invention, by contrast, the two TES's and additional power machinery are provided to time shift the work of compression that would have occurred in a normal gas turbine cycle. The system does this by reducing the work of compression that is required to return the gas to the state that it would have been post compression in a normal gas turbine cycle.

When discharging, the cold TES cools the air as it passes through the TES so that the work of compression required to get to, say 18 bar, is much lower than the normal work of compression to get to 18 bar from ambient temperature. As the now compressed air passes through the hot TES it is now heated to very close to the correct temperature (and pressure) that it would have been prior to entering to enter the combustor in normal generation mode.

Furthermore, the cold TES may be designed to go to cryogenic temperatures, but without the use of liquid cryogens ie all of the heat transfer is direct between dry air and he storage media. There are no phase changes involved in the process.

The cold TES does not need to operate at high temperatures, which means that it is possible to use material that would not be suitable in a higher temperature TES. In this way, heat storage apparatus is provided in which a low cost material may be used to provide heat storage on the cold side of the system (e.g. a material which is suitable for use in a store where the temperature does not rise beyond 200 deg C or beyond 100 deg C but which is unsuitable for use on the hot side of the system).

While an open cycle may be used in the charging mode in certain instances (as shown in Figure 5b), it does require at least one dehumidifier in active operation disposed at or near the air intake, which will tend to reduce the cycle efficiency.

The difficulty with using an open cycle for this process on charge is that condensation will occur in the cooler parts of the hot TES and, in addition, during the expansion process in the expander (turbine) 115, the remaining moisture and C0 ₂ will condense and then freeze. This is likely to cause blockage in the ductwork and/or damage to the expander (turbine). The use of a dessicator/dehumidifier is an energy consuming process, which only needs to occur when the system is running in open cycle mode. With the proposed configuration of closed cycle for charge and open cycle for discharge this energy penalty is only required for half of the cycle and hence these losses are halved. A further advantage of only using an open cycle on discharge is that any small trace elements of moisture in the air will be trapped within the cold store and the power machinery will only process with dry air.

This proposal differs from other proposed storage cycles as it still incorporates combustion but it is looking to time shift part of the work of a normal generation cycle ie part of the compression work - to a time when electricity may be lower cost and generation is not required. It should be understood that this is not a generation cycle and that the net recoverable energy will always be less than 100%.

Examples of uses of the system

For example, when the system is connected to a grid with a large element of renewables also attached the system might run in the following modes:

• charge mode during the night to charge on low cost electricity from nuclear power plants

• discharge plus generation in the morning period

· charge in the middle of the day when solar PV generation is at a peak or generation during the day when there is no sunshine

• discharge plus generation in the evening period

Alternatively there might be a failure of a major power plant on the grid and the system might need to operate in a more continuous power generation manner during the winter such as:

• charge mode during the night to charge on low cost electricity from nuclear power plants

• generation in the morning period and middle of the day

• discharge plus generation in the evening period

Table 9: PHES Round Trip cycle efficiency from Polytropic Efficiency (only)

The first row of the table above is the efficiency of running a pure PHES cycle, such as that shown in Figure 2, with turbomachinery, using state of the art polytropic efficiencies for the hot machinery (113 and 1 14) and cold machinery (115 and 1 16) of 93%. From this it can clearly be seen that the ideal round trip efficiency of the cycle is 73% and including other real losses is likely to be 55% or lower, as explained previously.

In the second row the efficiency of the hot (high temperature) compressor 1 13 and expander (turbine) 214 is set to 100%. The cold compressor 1 16 and expander (turbine) 1 15 is assumed to be a pair of axial flow machines with a polytropic efficiency of 93%. This has the effect of ignoring the losses from the hot machinery and consequently raises both the ideal roundtrip efficiency and the real roundtrip efficiency including losses. Allowing for the remaining losses, the figure is likely to be around 75% for normal charging/discharging operation. The hot compressor / expander (turbine) losses can be ignored as they are already budgeted for within the combustion cycle losses. Furthermore, the capital machinery for the combustion cycle can be budgeted for within cost of a generating plant ie the cost of the hot compressor and hot turbine are accounted for within part of the cost of the combustion plant. This means that with the addition of two thermal stores and two much smaller axial (or similar) flow compressors, or a reversible (e.g. positive displacement) power machine, a storage cycle can be incorporated within a combustion cycle with a high efficiency and low cost.

While different combinations or sizes of power machinery are possible, the advantage of this system is that it is possible to build a power plant with the generation efficiency close to that of a CCGT that also incorporates energy storage. An example of why this is useful is that in Germany on 18 ^th May 2014 renewables accounted for 74% of the actual demand. A large number of thermal plant were also kept operating to provide spinning inertia to the grid. Spinning inertia is used to help maintain grid frequency and to reduce the rate of change of grid frequency. Renewables (wind and solar) do not contribute any significant inertia and hence there is still a requirement to run thermal power plant. The amount of renewables that could have been generating was much higher, but some of the renewable generation was curtailed (deliberately turned off) as the system could not absorb the energy. Consequently the ability to build a power station that can be run in an 'energy storage mode' while also contributing inertia to the system (due to having rotating machinery) is highly desirable.

From a technical perspective an advantage of this system is that it can operate over a constant pressure ratio, so can easily use existing axial flow machinery and benefit from the proven reliability of such machines.

Figures 6a to 6d show a further hybrid system 100 according to the present invention which is similar to the system of Figure 3, but that has been modified to incorporate one example of preheat functionality. Preheating the gas prior to entry to the first compressor upon charging can secure additional benefits in terms of increased energy density and efficiency, especially where effectively "waste heat" is used, as will now be described. The adaptation with respect to the Figure 3 system is a preheater system 144/146/145, whereby the two heat exchangers 145 and 146 are connected together by a separate heat transfer fluid circuit 144 (dash-dot lines) so that they can exchange heat in a counterflow manner to allow heat transfer between those two points in the PHES working fluid circuit. This is a preferred embodiment although it would also be possible to transfer heat between the two heat exchangers in other ways, for example, by direct heat transfer if the two heat exchangers were incorporated in a single counter-current heat exchanger with their respective air flow ducts arranged in a co-aligned, counter-current flow set-up.

For example, during the charging mode, the flow from point F to point A can be heated by cooling the flow from point C to point D. In addition to this connection it is also expected that these heat exchangers 146,145 will have the ability to reject waste heat to atmosphere if required after the main heat exchange processes have taken place. This is easy to achieve by adding external heat exchangers to the circuit where necessary. The main purpose of heat exchanger 146, in this scenario, is to increase the temperature of the gas entering the compressor A so that it is above ambient.

To achieve preheating with heat exchangers linked across the first working fluid circuit by an HTF circuit, the gas circulating downstream of the first TES system must be sufficiently hotter than that circulating upstream of the first compressor. In a TES, storage medium on the hot and cold sides of the thermal front will be respectively held at approximately the last gas inlet temperature on charging the store (from the hot end) and the last gas inlet temperature upon discharging the store (from the cold end). The latter temperature will therefore be the temperature exhibited by the gas exiting the first TES system during charging (i.e. the last "minimum store temperature") of the first TES. The last gas inlet temperature from discharging the store (from the cold end) may selectively be raised, during the previous discharge mode, by choosing the degree, if any, at which to discard any of the waste heat generated by the second power machinery i.e "cold compressor". The simplest set-up is to configure the (or all) heat exchangers located downstream of the first TES so that they are bypassed or inoperative during the discharge/generation mode, and hence, so that all the (low grade) waste heat from the cold compressor becomes stored (at a higher "minimum store temperature") in the first TES system. In the subsequent charging mode, the heat exchanger downstream of the first TES system is then operative to transfer that heat (in effect, waste heat that was temporarily stored in the first TES) via the HTF circuit to the upstream heat exchanger.

Heating the inlet air prior to compression in a gas turbine is counter-intuitive for a number of reasons. In normal operation of a gas turbine the power output of the GT falls as the air inlet temperature rises. This because warmer air is less dense so that the mass flow rate through the compressor falls and also the amount of work required for the compression increases with temperature. A normal rule of thumb is that for every degree of temperature rise the power output of a GT drops by about 0.5%. Furthermore, adding heat to a closed cycle PHES system is also counter-intuitive as the normal problem with closed cycles is that the losses build up in the system creating a temperature rise which needs to be rejected to the atmosphere via heat exchangers. Adding additional heat to such a closed system will make this issue much more significant as there will be large quantities of additional heat exchange occurring. This amount of heat exchange is normally expensive to build and inefficient.

In this hybrid system we are looking to store energy, and in charging mode, the power input is the difference between the compressor work and the expander work. Increasing the compressor work relative to the expander work improves the power input, efficiency and energy density of the system. It improves the power input because even though the mass flow rate will drop the change between compressor and expander power is greater and hence power input increases even as the mass flow rate drops. It improves the efficiency of the system because it increases the work ratio ie the work input divided by the total work of all processes. This has the effect of reducing the rate of losses. Furthermore, being able to operate in an open cycle mode on discharge means that these heat additions to the system can be easily removed from the system without the requirement of additional heat exchangers. The largest impact is on energy density as it can improve the energy density of the system by as much as 50%.

Figure 6a) shows a system where in charging mode the waste heat stored in the hot thermal stores 140 (first TES system) is used to pre-heat the air entering the compressor, while being cooled by this air. In this way the air entering cold expander 1 15 is cooled to near ambient and the air entering hot compressor 1 13 is raised to a temperature above ambient.

Figure 6b) shows the system where in discharging mode the waste heat generated by cold compressor 116 (seen as a rise in temperature) is stored in the hot thermal store 140 by not operating heat exchanger 145. In this way, the warm air enters the bottom of the TES 140 and the heat is stored within the storage media. At the same time, gas is being discharged from the hot thermal store 140 at a higher temperature than in the Figure 3 system to the combustor, where its increased temperature means that less fuel is required to achieve the same power generation as the Figure 3 system.

Figure 6c) shows the system where in a discharging mode additional waste heat, for example from the exhaust stack of the steam turbine, is added to the air flow and to raise the temperature further. In this way heat at a higher temperature is stored in the thermal stores 140.

Figure 6d) shows the same system in charge mode, with the HTF circuit active once again to redistribute the heat upstream. The hotter air can be used to pre-heat the air entering the hot compressor 113, while being cooled by this air. This system allows the pre-heating to achieve higher temperatures than might be achieved with just the use of waste heat from the cycle. In this way the air entering cold expander 115 is cooled to near ambient and the air entering hot compressor is raised to a temperature well above ambient. The work of the compressor per unit mass processed is greatly increased by this mode of operation.

Some example temperatures and pressures at points in the gas flow path for Figures 6a, 6b, 6c and 6d, respectively, are given in the following tables:

Table 10: Charging Mode - Pre-heat with Waste Heat (Fig. 6a)

Figure 6b In use during Pressure Temperature

this process (bar absolute) (Kelvin)

A No

B 16.6 836

C 16.6 353

D 16.6 353

E 1 145

F 1 288 G 1 288

H 16.6 1486

1 1 816

Table 1 1 : Discharge with Generation - Store Waste Heat (Fig. 6b)

Table 13: Charging Mode - Pre-heat with Higher Grade Waste Heat (Fig. 6d) POWER CALCULATIONS

By way of example only, typical figures for a large gas turbine generation plant are used to quantify the effect of integrating PHES with an existing gas power plant in Table 14 below:-

5

Compensated

Pre-heating Pre-heating Variation

Ambient 343K 363K Mass Flow at

363k pre-heat

CCGT

2 Gas Turbines MW 340 340

1 Steam Turbine MW 170 170

Charging MW 240 31 1 336 267

Normal MW 510 510 510 510

Discharging plus MW 680 680 680 680 generation

Energy Density 100% 134% 147%

Table 14 - Modes of Operation for System of Fig. 3 and 5, with losses factored in

10

The table shows a CCGT that consists of two 170MW gas turbines connected to a steam turbine also of 170MW power output. This is equivalent to using two of the gas turbines listed in Tables 6, 7 & 8 (normal generation 170MW, charging 120MW, discharging 255MW per GT) with a 170MW steam turbine. Using two similar hot

15 compressors 113 and two similar turbines 214 a system could be built that would charge at 240MW, generate at 510MW and discharge (including generation) at 680MW.

Using a hybrid system such as that shown in Figure 6a would result in pre-heating of the compressor inlet to 343K and an increase in the power input to 31 1 MW if the mass flow remained constant. Pre-heating to 363K would increase this figure further to 336MW

20 again if the mass flow remained constant. The compensated column calculates the power input when the mass flow rate is reduced to compensate for the reduction in density as a result of the higher temperature (363K). It can be seen that even though the mass flow rate has dropped the power input has actually increased.

The change in energy density is even more dramatic. It can be seen that pre-

25 heating raises the energy density by almost 50% above the basic case. This means that for the same sized stores more than 50% extra energy can be stored. The other difference about this system is that the recovery in energy does not increase the power output of the system. The higher temperature in the TES 140 reduces the amount of gas required when the system is discharged.

In terms of pre-heating it is possible to build compressor blades that operate continuously at temperatures of up to 620°C with current technology. This means that the amount of pre-heating must take into account the peak compressor temperature that can be accommodated.

The present invention thus provides a hybrid combustion turbine power generation system with inbuilt energy storage based on a PHES system for flexible load management during power generation.

While the present invention has been described in detail with reference to certain preferred embodiments, other embodiments of the invention are possible so that the scope of the appended claims should not be limited to the description of the preferred embodiments. It should further be noted that features of the respective embodiments may be combined, where this is clearly possible, without departing from the scope of the invention.

In particular, it will be appreciated that the present invention provides an alternative means of supplying heated, pressurised gas for subsequent combustion and expansion in a Brayton top thermodynamic cycle, and hence, any bottoming cycle, or other energy/ heat recovery (e.g. co-generation) method known for extracting more energy from the exhausted Brayton fluid may similarly be employed in conjunction with the hybrid system of the present invention.