THERMOELECTRIC ENERGY

FIELD OF THE INVENTION

The present invention relates generally to the storage of electric energy. It relates in particular to a system and method for storing electric energy in the form of thermal energy in thermal energy storage.

BACKGROUND OF THE INVENTION

Base load generators such as nuclear power plants and generators with stochastic, intermittent energy sources such as wind turbines and solar panels, generate excess electrical power during times of low power demand. Large-scale electrical energy storage systems are a means of diverting this excess energy to times of peak demand and balance the overall electricity generation and consumption.

In an earlier patent application EP1577548 the applicant has described the idea of a thermoelectric energy storage (TEES) system. A TEES converts excess electricity to heat, stores the heat, and converts the heat back to electricity, when necessary. Such an energy storage system is robust, compact, site independent and is suited to the storage of electrical energy in large amounts. Thermal energy can be stored in the form of sensible heat via a change in temperature or in the form of latent heat via a change of phase or a combination of both. The storage medium for the sensible heat can be a solid, liquid, or a gas. The storage medium for the latent heat occurs via a change of phase and can involve any of these phases or a combination of them in series or in parallel.

All electric energy storage technologies inherently have a limited round-trip efficiency. Thus, for every unit of electrical energy used to charge the storage, only a certain percentage is recovered as electrical energy upon discharge. The rest of the electrical energy is lost. If, for example, the heat being stored in a TEES system is provided through resistor heaters, it has approximately 40% round-trip efficiency. The efficiency of thermoelectric energy storage is limited for various reasons rooted in the second law of thermodynamics. Firstly, the conversion of heat to mechanical work is limited to the Carnot efficiency. Secondly, the coefficient of performance of any heat pump declines with increased temperature difference between input level and output level. Thirdly, any heat flow from a working fluid to a thermal storage and vice versa requires a temperature difference in order to happen. This fact inevitably degrades the temperature level and thus the capability of the heat to do work.

It is noted that many industrial processes involve provision of thermal energy and storage of the thermal energy. Examples are refrigeration devices, heat pumps, air conditioning and the process industry. In solar thermal power plants, heat is provided, possibly stored, and converted to electrical energy. However, all these applications are distinct from TEES systems because they are not concerned with heat for the exclusive purpose of storing electricity.

It is known in the art that heat can be provided to the thermal storage unit through a heat pump. For example, a Stirling machine (for reference, see US patent 3080706, column 2, lines 22-30). Also, International Patent WO 2007/134466 discloses a TEES system having an integrated heat pump.

A heat pump requires work to move thermal energy from a cold source to a warmer heat sink. Since the amount of energy deposited at the hot side is greater than the work required by an amount equal to the energy taken from the cold side, a heat pump will "multiply" the heat as compared to resistive heat generation. The ratio of heat output to work input is called coefficient of performance, and it is a value larger than one. In this way, the use of a heat pump will increase the round-trip efficiency of a thermoelectric energy storage system. The round-trip efficiency is the amount of electricity provided from the storage divided by the amount of electricity provided to the storage.

US Patent 4089744 discloses a method of thermal energy storage by means of reversible heat pumping. Excess electrical output is stored in the form of sensible heat by using it to raise the temperature level of a heat storage fluid. In this scheme, the source of low level heat is stored hot water, which also serves as the working fluid in the heat pump and the turbine cycles. A thermodynamic analysis, such as the type of analysis shown in Figure 6, shows that the efficiency of schemes equivalent to that of US 4089744 is limited to about 50%. Thus, there is a need to provide an efficient thermoelectric energy storage having a round- trip efficiency of, preferably, greater than 55%.

DESCRIPTION OF THE INVENTION

It is an objective of the invention to provide a thermoelectric energy storage system for converting electrical energy into thermal energy to be stored and converted back to electrical energy with an improved round-trip efficiency. This objective is achieved by a thermoelectric energy storage system according to claim 1 and a method according to claim 7. Preferred embodiments are evident from the dependent claims.

According to a first aspect of the invention, a thermoelectric energy storage system is provided which comprises a hot storage unit which is in connection with a heat exchanger and contains a thermal storage medium, a working fluid circuit for circulating a working fluid through the heat exchanger for heat transfer with the thermal storage medium, and wherein the temperature difference between the working fluid and the thermal storage medium in the hot storage unit is minimized during heat transfer.

When the thermoelectric energy storage system is in a charging (or "heat pump") cycle, the thermodynamic machine includes a turbine, and when the thermoelectric energy storage system is in a discharging (or "turbine") cycle, the thermodynamic machine includes a compressor.

Preferably, the hot storage unit comprises at least two hot storage units, each hot storage unit is in connection with a heat exchanger and contains a thermal storage medium.

In a preferred embodiment, the heat exchanger or heat exchangers are common to both the charging and discharging cycles. However, it is also possible that there are separate heat exchangers for the charging and discharging cycles. Two or more heat exchangers utilized in series are preferably connected hydraulically.

Further, the thermal storage medium may be a liquid and a flow rate of the thermal storage medium may be modified such that the temperature difference between the working fluid and the thermal storage medium in each hot storage unit is minimized during heat transfer.

The thermal storage medium of the present invention may be a solid or a liquid. The particular embodiment illustrated in Figures 3 and 4 of the accompanying description shows a version wherein the thermal storage medium is a liquid.

In a preferred embodiment, a single working fluid circuit containing a single type of working fluid is utilized for both the charging and discharging cycles. However, it is also possible that there are separate working fluid circuits for the charging and discharging cycles. Further, each separate working fluid circuit may contain a different type of working fluid.

Preferably, the temperature of the thermal storage medium at entry and exit points of each connected heat exchanger is modified such that the temperature difference between the working fluid and the thermal storage medium in each hot storage unit is minimized during heat transfer.

Further, at least one of the hot storage units may contain a different type of thermal storage medium such that the temperature difference between the working fluid and the thermal storage medium in each hot storage unit is minimized during heat transfer.

In a preferred embodiment, the hot storage unit or units comprise a thermal storage medium for sensible heat storage and a phase change storage medium for latent heat storage, which are arranged such that the temperature difference between the working fluid and the thermal storage medium in each heat exchanger unit is minimized during heat transfer.

Preferably, the temperature difference between the working fluid and the thermal storage medium in each hot storage unit is less than 50 ⁰ C during heat transfer.

In a second aspect of the present invention, a method is provided for storing thermoelectric energy in a thermoelectric energy storage system, the method comprising charging a hot storage unit by providing heat via a heat exchanger to a thermal storage medium by compressing a working fluid, discharging the hot storage unit by expanding the working fluid heated via the heat exchanger from the thermal storage medium through a thermodynamic machine, and modifying the thermal storage media parameters to ensure the temperature difference between the working fluid and the thermal storage medium is minimized during charging and discharging.

Preferably, the step of modifying the thermal storage media parameters comprises modifying the flow rate of the thermal storage medium.

Further, the step of modifying the thermal storage media parameters may comprise modifying the initial temperature and final temperature of the thermal storage medium.

Preferably, the step of modifying the thermal storage media parameters comprises modifying the type of thermal storage medium.

BRIEF DESCRIPTION OF THE DRAWINGS

The subject matter of the invention will be explained in more detail in the following text with reference to preferred exemplary embodiments, which are illustrated in the attached drawings, in which:

Figure 1 shows a simplified schematic diagram of a thermoelectric energy storage system;

Figure 2 is an enthalpy-pressure diagram of the heat pump cycle and the turbine cycle in an exemplary TEES system;

Figure 3 is a schematic illustration of a cross-section through a heat pump cycle portion of a TEES system of the present invention;

Figure 4 is a schematic illustration of a cross-section through a turbine cycle portion of a TEES system of the present invention;

Figures 5a - 5f depict simplified enthalpy-temperature diagrams of the working fluids and thermal storage fluids in the heat exchangers during charging and discharging; Figure 6 shows an enthalpy-temperature diagram of the heat transfer from the cycles in a TEES system of the present invention;

Figure 7 shows an enthalpy-temperature diagram of the heat transfer from the cycles in an optimized scenario in a TEES system of the present invention;

For consistency, the same reference numerals are used to denote similar elements illustrated throughout the figures.

DETAILED DESCRIPTION OF PREFERRED EMBODIMENTS

Figure 1 depicts a schematic diagram of a TEES system 10 in accordance with the present invention which comprises a hot storage 12 and a cold storage 14 which are coupled to each other by means of a heat pump cycle system 16 and a turbine cycle system 18. The hot storage 12 contains a thermal storage medium, the cold storage 14 is a heat sink, and both the heat pump cycle and the turbine cycle contain a working fluid.

The heat pump cycle system 16 comprises, in the flow direction of the working fluid, an evaporator 20, a compressor train 22, a heat exchanger 24, and an expansion valve 26. The turbine cycle system 18 comprises, in the flow direction of the working fluid, a feed pump 28, a heat exchanger 30, a turbine 32, and a condenser 34. The heat exchangers 24, 30 in both the heat pump cycle system and the turbine cycle system are located to exchange heat with the hot storage 12. The evaporator 20 and the condenser 34 in the heat pump cycle system 16 and the turbine cycle system 18 respectively, are located to exchange heat with the cold storage 14.

The cold storage 14 is a heat reservoir at any temperature lower than the hot storage temperature. However, the cold storage temperature may be higher or lower the ambient temperature. In fact, the cold storage may be another heat sink such as cooling water or air from the ambient. In an alternative embodiment, the turbine and compressor train may be thermodynamic machines based on positive displacement such as reciprocating or rotary expanders or compressors.

The compressor train 22 may comprise one or several individual compressors with possible intercooling (not shown). The turbine 32 may comprise one or several individual turbines with possible reheating (not shown). Similarly, the evaporator 20, the condenser 34, the feed pump 28 and the expansion valve 26 may comprise one or multiple units.

In operation, the working fluid flows around the TEES system 10 in the following manner. The working fluid in the compressor 22 is initially in vapour form and surplus electrical energy is utilized to compress and heat the working fluid. The working fluid is fed through the heat exchanger 24 where the working fluid discards heat into the hot storage medium.

The compressed working fluid exits the heat exchanger and enters the expansion valve

26. Here the working fluid is expanded to the lower pressure of the evaporator. The working fluid flows from the expansion valve into the evaporator 20 where the working fluid is heated to evaporation. This is realized using available heat from the cold storage.

In the condenser 34, working fluid is condensed by exchanging heat with the cold storage 14. The condensed working fluid exits the condenser via the outlet and is pumped into the heat exchanger 30 at the hot storage via the feed pump 28. Here the working fluid is heated, evaporated, and overheated from the stored heat from the hot storage medium. The working fluid exits the heat exchanger 30 and enters the turbine 32 where the working fluid is expanded thereby causing the turbine to generate electrical energy.

The expansion valve 26, the evaporator 20, and the compressor 22 are in operation during a period of charging, or the "heat pump cycle". Similarly, the turbine 32, the condenser 34 and the feed pump 28 are in operation during a period of discharging or the "turbine cycle". The hot storage 12 is in operation at all times; during charging, storage, and discharging. These two cycles can be clearly shown in an enthalpy-pressure diagram, such as Figure 2.

The solid-line cycle shown in Figure 2 represents the heat pump cycle that is charging the hot storage and the heat pump cycle follows a counter-clockwise direction as indicated by the arrows. The working fluid is assumed to be water for this exemplary embodiment. The heat pump cycle starts in the evaporator at point A where steam is evaporated to form vapor using heat from the cold storage (transition A->B1 in Figure 2). In the next stage of the heat pump cycle, the vapor is compressed utilising electrical energy in two stages from point B1 to C1 and B2 to C2. Where compression occurs in two stages this is a consequence of the compressor train comprising two individual units. In between these two compression stages, the working fluid is cooled from point C1 to B2. The hot, compressed, overheated vapor exits the compression train at point C2 where it is cooled down to the saturation temperature at D1 , condensed at D2, and further cooled down to point D3. This cooling down and condensation is realized by transferring the heat from the working fluid into the hot storage thereby storing the heat energy. The cooled working fluid is returned to its initial low pressure state at point A via the expansion valve.

The dotted-line cycle shown in Figure 2 represents the Rankine turbine cycle that is discharging the hot storage and the cycle follows a clockwise direction as indicated by the arrows. The Rankine turbine cycle starts at point E, where the pump is utilized to pump the working fluid in its liquid state from point E to F1. Next, from point F1 to point G, the working fluid receives the heat from the thermal storage medium. In detail, the heat is transferred from the thermal storage medium to the working fluid causing the working fluid to heat up at F2, to boil at F3, and attain a certain degree of superheat at G. The superheated working fluid vapor at point G is expanded down to point H in a mechanical device such as a turbine to generate electricity. Following the expansion, the working fluid enters the condenser where it is condensed to its initial state at point E by exchanging heat with the cold storage.

The roundtrip efficiency of the complete energy storage process, that is the heat pump cycle and the Rankine turbine cycle, is calculated in the following manner; the work provided by the turbine expansion divided by the work used in the heat pump compressor:

(h _G -h _H )/(hc2-hB2 ⁺ hci-hBi) _!

where the letter h denotes the enthalpy of the corresponding point. For the exemplary conditions depicted in Figure 2, the roundtrip efficiency is 50.8%. It is not possible from the enthalpy-pressure diagram alone to judge if this is a particularly efficient TEES system, or how it could be improved in efficiency.

With reference to the TEES system illustrated in Figure 1 , the heat exchanger 24 in the heat pump cycle components 16 and the heat exchanger 30 in the turbine cycle components 18 may comprise several individual heat exchangers arranged in series, as illustrated in Figures 3 and 4, respectively. Figure 3 depicts a simplified schematic diagram of the heat pump cycle components 16 in a thermoelectric energy storage system 10 of the present invention. Here, three individual hot storage units x, y, z are arranged in series. Each hot storage unit x, y, z comprises a heat exchanger 36, 38, 40 in connection with a storage tank pair 42, 44, 46. Each storage tank pair comprises a cold tank and a hot tank wherein the flow of the thermal storage medium is from the cold tank to the hot tank via the associated heat exchanger. The three hot storage units in Figure 3 are denoted x, y and z from left to right in the diagram. In the present embodiment, the heat exchangers are counterflow heat exchangers, and the working fluid of the cycle is water.

In operation, the heat pump cycle components 16 of Figure 3 perform essentially in a similar manner as heat pump cycle components 16 of the TEES system described in respect of Figures 1 and 2. In addition, the working fluid flows through the further two separate heat exchangers. In the exemplary situation shown in Figure 3, in the direction of flow of the working fluid, the initial and final temperatures of the working fluid as it passes through heat exchanger 40 are 510 ⁰ C and 270 ⁰ C, through heat exchanger 38 are 270°C and 270 ⁰ C, and through heat exchanger 36 are 270°C and 100°C. Thus, an overall temperature drop of 410 ⁰ C is achieved.

The characteristics of the working fluid (shown as a solid line) and thermal storage medium (shown as a dashed line) of each of the three heat exchangers 36, 38, 40 and associated storage tank pair 42, 44, 46 during charging are shown in Figure 5 in the enthalpy-temperature graphs a), b) and c), respectively. The temperature of the thermal storage medium in each stage is increasing, whilst the temperature of the working fluid decreases only in stages a) and c).

Figure 4 depicts a simplified schematic diagram of the turbine cycle components 18 in a thermoelectric energy storage system 10 of the present invention. Here, the arrangement of three individual hot storage units x, y, z, arranged in series, are the same units shown in Figure 3. Again, each storage tank pair 42, 44, 46 comprises a hot tank and a cold tank, however the flow of the thermal storage medium is from the hot tank to the cold tank via the heat exchanger.

In operation, the turbine cycle components 18 of Figure 4 perform essentially in a similar manner as turbine cycle components of the TEES system described in respect of Figures 1 and 2. In addition, the working fluid flows through the further two separate heat exchangers. In the exemplary situation shown in Figure 4, in the direction of flow of the working fluid, the initial and final temperatures of the working fluid as it passes through heat exchanger 36 are 80 ⁰ C and 240 ⁰ C, through heat exchanger 38 are 240°C and 240 ⁰ C, and through heat exchanger 40 are 240°C and 490°C. Thus, an overall temperature increase of 410 ⁰ C is achieved.

When the heat pump cycle components 16 are in operation, then the working fluid conduit for the heat pump cycle is coupled to the hot storage units x, y, z. When the turbine pump cycle components 18 are in operation, then the working fluid conduit for the turbine cycle coupled to the hot storage units x, y, z, instead. In this way, the turbine cycle obtains thermal energy from the hot storage units that was deposited by the heat pump cycle.

The characteristics of the working fluid (shown as a solid line) and thermal storage medium (shown as a dashed line) of each of the three heat exchangers 36, 38, 40 and associated storage tank pairs 42, 44, 46 during discharging are shown in Figure 5 in the enthalpy-temperature graphs d), e) and f), respectively. The temperature of the thermal storage medium in each stage is decreasing, whilst the temperature of the working fluid increases only in stages d) and f).

Figure 6 shows the isobars, ie. lines of constant pressure, from Figure 5 a) - f) on a single temperature-enthalpy graph for a particular system embodiment. Further, the capital letters used are consistent with Figure 2. Thus, Figure 6 illustrates the heat transfer process at the three separate hot storage units x, y, z during the charging and discharging of the TEES system 10.

The solid line isobars C2 to D3 represent the heat pump cycle, the dotted line isobars F1 to G represent the Rankine turbine cycle, and the dashed line isobars X1 to X2, Y1 to Y2, Z1 to Z2 represent the thermal storage media in the three hot storage units x, y, z, respectively.

Heat can only flow from a higher to a lower temperature. Consequently, the characteristic isobars for the working fluid during cooling in the heat pump cycle have to be above the characteristic isobars for the thermal storage media, which in turn have to be above the characteristic isobars for the working fluid during heating in the turbine cycle. The slope of these characteristic isobars is defined by the product of the massflow (kg/s) and heat capacity (J/kg/K) of each thermal storage medium relative to the massflow of the working fluid. This product is different for each of the three heat transfer subsections; heating/cooling of liquid water in hot storage unit x, boiling/condensation in hot storage unit y, and providing/extracting heat to the supersaturation region in hot storage unit z.

The temperature profiles are stationary in time due to the sensible heat storage in the thermal storage media. Thus, whilst the volume of thermal storage media in each heat exchanger remains constant, the volume of hot and cold thermal storage media stored in the hot and cold tanks changes. Also, the temperature distribution in the heat exchangers remains constant.

Importantly, the present invention determines that the smaller the average temperature difference between the working fluid and the heat storage media during heat transfer, the greater the efficiency of the TEES system. In an enthalpy-temperature graph, this feature is observed as a relatively closer positioning of the characteristic isobars of the charging and discharging cycles, as shown in Figure 7.

The present invention determines that the thermal storage media may be the same or a different fluid in each hot storage unit x, y and z. Further, the present invention determines that the thermal storage media may be at a different temperature in each hot storage unit x, y and z. Also, the flow-rate of the thermal storage media within each hot storage unit may differ. Specifically, in order to achieve an optimized roundtrip efficiency of the TEES system various combinations of the thermal storage media, the initial and final temperature of the thermal storage media and the thermal storage media flow-rates may be utilized.

In the improved efficiency scenario illustrated in Figure 7, the flow-rate of the thermal storage medium through heat exchanger 38 of hot storage unit y is increased by a factor of three in comparison with the scenario in Figure 6. (It should be noted that the flow rate in heat exchanger 38, in Figure 6, was set to an arbitrary rate that was relatively larger than the flow rate in heat exchangers 36 and 40, but the flow rate was not optimized as in Figure 7.) A decrease in average temperature differences between the thermal storage medium and the working fluid during heat transfer in heat exchanger 38 of hot storage unit y can be noted. Consequently, a resultant TEES system design has a higher saturation temperature in heat exchanger 38 in the turbine cycle than before (denoted as F2' and F3' in Figure 7 in comparison with F2 and F3 in Figure 6). This equates to a temperature of 230 ⁰ C in Figure 7, in comparison with 200 ⁰ C in Figure 6. Consequently, the roundtrip efficiency of the TEES system in the embodiment of Figure 7 is 61.1% in comparison to an efficiency of 50.8% in Figure 2.

In others words, the present invention requires the temperature difference between the working fluid of the heat pump cycle and the heat storage media, as well as the temperature difference between the working fluid of the turbine cycle and the heat storage media to be relatively small (for example, smaller than 50 ⁰ C on average). This is achieved through modification of certain TEES parameters as specified above.

In a preferred embodiment of the present invention, the three thermal storage media are fluids. For example, these may be three different liquid sensible heat storage media such as water, oil, or molten salts. Also, in a preferred embodiment of the present invention, the heat exchangers are counterflow heat exchangers, having a minimal approach temperature 10 K (ie. the minimal temperature difference between the two fluids exchanging heat is 10 K) and the expansion device is preferably a thermostatic expansion valve.

In a further preferred embodiment, the heat at the boiling/condensation heat exchanger 38 is transferred to the latent heat of a phase transition of a storage medium enabling an even closer match of the temperature profiles in the boiling/condensation region. A preferred embodiment uses steam as the working fluid for both the heat pump cycle and the turbine cycle.

In an alternative preferred embodiment there is no cold storage reservoir, but evaporator and condenser instead use heat from the ambient as an (infinitely large) reservoir for the cold side of the heat pump cycle and the turbine cycle. The cold storage of Figure 1 , which is a second heat storage reservoir, has latent heat storage at temperatures around 100 ⁰ C at the cold side of the heat pump cycle and the turbine cycle. Because of the temperature dependence of the saturation pressure of working fluids such as water, such an additional heat storage reservoir may result in greater economy in respect of the compressor and the turbine. It is envisaged that this economy would more than compensate for the additional cost for this reservoir at moderately long storage times.

The skilled person will be aware that the TEES system, as illustrated in Figures 1 , 3 and 4, may be realized in several different ways. For example, the hot storage can consist of:

• A solid structure with embedded heat exchangers equipped with appropriate means of handling the expansion-contraction of the storage medium with changing temperatures.

• A two-tank molten salt storage system with heat exchangers between the tanks and flow of molten salt from the cold to the hot tank during charging and from the hot to the cold tank during discharging periods.

• A multiple-hot-tank multiple-cold-tank molten salt and liquid heat storage media cascaded at different temperatures between the evaporator operating temperature and the temperature of the heat pump working fluid at the exit of the compression processes.

• A phase change material with a suitable phase change temperature below the condensation temperature of the heat pump working fluid at the high operating pressure and above the boiling point of the turbine cycle working fluid at the high operating pressure. • Any combination of the above mentioned thermal storage options in series and in parallel.

• Two, three (as shown in Figures 3 and 4), four or more hot storage units in the hot storage.

The skilled person will be aware that the condenser and the evaporator in the TEES system may be replaced with a multi-purpose heat exchange device that can assume both roles, since the evaporation for the heat pump cycle and the condensation for the turbine cycle will be carried out in different periods. Similarly the turbine and the compressor roles can be carried out by the same machinery, referred to herein as a thermodynamic machine, capable of achieving both tasks.

The preferred working fluid for the instant invention is water; mainly due to the higher efficiencies of a water-based heat pump cycle and turbine cycle, and the amiable properties of water as a working fluid i.e. no global warming potential, no ozone depletion potential, no health hazards etc. For the operation of the present invention at ambient temperatures below the freezing point of water, a commercial refrigerant can be chosen as the heat pump working fluid, or a second bottoming heat pump cycle can be cascaded with the water-based cycle to provide the heat of evaporation.