Patent Publication Number: US-2021180457-A1

Title: Adiabatic salt energy storage

Description:
CROSS REFERENCE TO RELATED APPLICATIONS 
     This application is a continuation of U.S. patent application Ser. No. 16/111,151, filed Aug. 23, 2018, which is a Continuation of U.S. patent application Ser. No. 12/932,775, filed Mar. 4, 2011, now issued as U.S. Pat. No. 10,094,219, which claims the benefit of U.S. Provisional Application No. 61/339,577, filed Mar. 4, 2010, all of which are herein incorporated by reference in their entireties. 
    
    
     FIELD OF THE INVENTION 
     This invention relates to energy storage. 
     BACKGROUND 
     Large scale energy storage is of considerable interest for power generation and distribution systems, to assist with exploitation of capricious energy sources such as wind and solar. At the moment, the main technology in wide use for reversibly storing electric power is hydropumping—drawing electricity off the grid to pump water uphill and then letting the water back down through power turbines later on. Hydropumping is highly efficient (about 80%) but suffers from (1) the need to allocate land to build dams and lakes, (2) the need for mountains, which aren&#39;t always available nearby, and (3) the need for water. 
     Recent developments in solar energy have revealed the substantial cost effectiveness of storing heat in tanks of molten salt for later use in generating electricity, by means of steam turbines, when the sun isn&#39;t shining. However, these storage facilities are adapted to store solar thermal energy, and are therefore not directly applicable to the storage of wind energy, which is mechanical energy as opposed to thermal energy. Molten salt has also been used as a primary coolant in nuclear reactors. Another approach for energy storage is considered in US 2010/0251711, where hot and cold storage tanks are employed in connection with heat storage. 
     However, efficiency is critical for energy storage, and it is especially critical for large scale energy storage. Therefore, it would be an advance in the art to provide energy storage having improved efficiency, especially for capricious sources of mechanical energy (e.g. wind energy). 
     SUMMARY 
     Improved energy storage is provided by using a working fluid flowing in a closed cycle including a ganged compressor and turbine, and capable of efficient heat exchange with heat storage fluids on a hot side of the system and on a cold side of the system. This system can operate as a heat engine by transferring heat from the hot side to the cold side to mechanically drive the turbine. The system can also operate as a refrigerator by mechanically driving the compressor to transfer heat from the cold side to the hot side. Heat exchange between the working fluid of the system and the heat storage fluids occurs in counter-flow heat exchangers. 
     Preferably, the hot side and cold side heat storage fluids each have a corresponding pair of storage tanks, where heat transfer to/from a heat storage fluid entails flow of the heat storage liquid between its two corresponding storage tanks. In a preferred approach, molten salt is the hot-side heat storage fluid and water is the cold-side heat storage fluid. 
     This approach provides numerous significant advantages. The use of the same compressor and turbine for both storage and retrieval provides substantial cost savings relative to approaches where storage and retrieval are performed in separate machinery. This cost savings is expected to be extremely significant, because the cost of the compressor and turbine (or equivalent machinery) is expected to be the most significant capital expense for a large scale energy storage plant. Molten salt provides numerous advantages as a thermal energy storage medium, such as low vapor pressure, lack of toxicity, low chemical reactivity with typical steels, and low cost. The low vapor pressure of molten salt is a highly significant safety advantage, as can be appreciated by considering hypothetically the use of steam as an energy storage medium in a large scale (e.g., 1 GW) thermal energy storage facility. A steam explosion from such a facility could have an explosive force on the order of thousands of tons of TNT. Using a closed loop for the working fluid advantageously increases cold-side heat transfer rates, allows a broader selection of working fluids, allows for operation at elevated cold-side pressure, improves efficiency, and reduces the risk of turbine damage. 
    
    
     
       BRIEF DESCRIPTION OF THE DRAWINGS 
         FIG. 1  shows an exemplary embodiment of the invention. 
         FIG. 2  shows operation of the example of  FIG. 1  in a heat engine mode that uses heat energy to provide mechanical work. 
         FIG. 3  shows operation of the example of  FIG. 1  in a refrigerator mode that uses mechanical work to store heat energy. 
         FIG. 4  shows an idealized thermodynamic Brayton cycle that relates to operation of embodiments of the invention. 
         FIG. 5  show plots of compressor efficiency vs. number of compressor stages. 
     
    
    
     DETAILED DESCRIPTION 
       FIG. 1  shows an exemplary embodiment of the invention. In this example, a working fluid (schematically referenced as  108 ) flows in a closed cycle that includes, in sequence, a compressor  102 , a first heat storage unit  110 , a turbine  104 , and a second heat storage unit  120 . Compressor  102  and turbine  104  are ganged on a common mechanical shaft  106  such that they rotate together. Heat storage units  110  and  120  are both capable of exchanging heat with working fluid  108 . For ease of illustration, pipes for defining the flow paths of fluids (e.g., working fluid  108 ) are not shown on  FIGS. 1-3 . Suitable pipes for the working fluid and heat storage fluids described herein are known in the art. As described in greater detail below, this apparatus is capable of operating as a heat engine (to provide mechanical work from heat) or as a refrigerator (to use mechanical work to store heat). 
     The purpose of heat storage units  110  and  120  is to provide stored heat to working fluid  108  and to remove heat from working fluid  108  for storage. It can be helpful to refer to first heat storage unit  110  as the hot-side heat storage unit, and to refer to second heat storage unit  120  as the cold-side heat storage unit. This terminology can be understood by noting that hot-side heat storage unit  110  adds heat to working fluid  108  at the same point in the cycle that combustion of fuel adds heat to air in a conventional jet engine. Thus, it can be helpful to regard hot-side heat storage unit  110  as being analogous to the fuel in a jet engine, when the apparatus is operating as a heat engine. 
     Heat storage units  110  and  120  preferably have several features to improve efficiency, as shown on  FIGS. 1-3 . First heat storage unit  110  preferably includes a first hot heat storage tank  112 H, a first cold heat storage tank  112 C, a first heat storage fluid  118  capable of flowing between tanks  112 H and  112 C to store or release heat, and a first counter-flow heat exchanger  116 . In counter-flow heat exchanger  116 , it is important that working fluid  108  and first heat storage fluid  118  flow in opposite directions, as shown. First heat storage unit  110  also includes a valve  114  that can switch connections between heat exchanger  116  and tanks  112 H,  112 C as needed for the heat engine and refrigerator modes. 
     Second heat storage unit  120  preferably includes a second hot heat storage tank  122 H, a second cold heat storage tank  122 C, a second heat storage fluid  128  capable of flowing between tanks  122 H and  122 C to store or release heat, and a second counter-flow heat exchanger  126 . In counter-flow heat exchanger  126 , it is important that working fluid  108  and second heat storage fluid  128  flow in opposite directions, as shown. Second heat storage unit  120  also includes a valve  124  that can switch connections between heat exchanger  126  and tanks  122 H,  122 C as needed for the heat engine and refrigerator modes. 
     Counter-flow heat exchangers  116  and  126  can be designed according to known principles to reduce entropy generation in the heat exchangers to negligible levels compared to the compressor entropy generation. The basic idea is to have very small temperature differences between any two fluid elements that are exchanging heat, thereby reducing entropy production (and eliminating it entirely in the idealized case). 
     The heat storage tanks are thermally insulated tanks that can hold a suitable quantity of the relevant heat storage fluid. In other words, the heat storage fluids are the medium of heat storage. Liquids are preferred over solids or gases because of the need for extremely rapid exchange of large amounts of heat by convective counterflow. They also allow for relatively compact storage of large amounts of energy. For example, the size of each storage unit (i.e.  110  and  120  on  FIG. 1 ) for a 1 GW plant operating for 12 hours should be roughly 20 medium-size oil refinery tanks. Each heat exchanger (i.e.  116  and  126  on  FIG. 1 ) should be roughly the size of a large steam locomotive boiler. 
     On the hot side, it is preferred that the heat storage fluid (i.e., fluid  118 ) be a molten salt or mixture of molten salts. A preferred molten salt is a eutectic (i.e. lowest melting point) mixture of sodium nitrate and potassium nitrate. However, any salt or salt mixture that is liquid over the operating temperature range can be employed. Such molten salts can provide numerous advantages, including low vapor pressure (which is important for safety), melting point below the creep temperature of steels, low corrosiveness, low capacity to dissolve iron and nickel, chemical stability, lack of toxicity, and low cost. 
     On the cold side, it is preferred that the heat storage fluid (i.e., fluid  128 ) be liquid water. It is important to ensure that no steam is present on the cold side, because the presence of steam creates a significant explosion hazard. Thus, 100° C. is an upper limit for the temperature of heat storage fluid  128  if water is used. As will be seen below, efficiency is improved by increasing the temperature difference at which the system operates. Accordingly, in some preferred embodiments, a mixture of water and one or more antifreeze compounds (e.g., ethylene glycol, propylene glycol and glycerol) is employed to increase the cold side temperature range to greater than 100° C. (e.g., −30° C. to 100° C.). 
     The example of  FIG. 1  also preferably includes a radiator  130  for dissipating waste heat generated by operation of the apparatus. Preferably, the radiator is coupled to the second hot heat storage tank  122 H, as shown. However, practice of the invention does not depend critically on the location of the radiator, because waste heat can also be removed at other points in the cycle. 
     Before describing further preferred features of some embodiments of the invention, it will be helpful to consider the heat engine and refrigerator modes of this apparatus, in connection with  FIGS. 2-4 . A idealized thermodynamic Brayton cycle is shown on  FIG. 4  as a pressure-volume diagram. 
       FIG. 2  shows operation of the example of  FIG. 1  in a heat engine mode that uses heat energy to provide mechanical work. Here it is assumed that the hot-side storage tanks  112 H and  112 C have substantially different fluid temperatures (e.g., as would result from prior operation of the apparatus to store energy). Working fluid  108  at the input of compressor  102  is represented by point  408  on  FIG. 4 . Compression of working fluid  108  moves the system to point  402  on  FIG. 4 . Heat is added by heat storage unit  110  to move the system from  402  to  404  on  FIG. 4 . More specifically, valve  114  provides connections as shown such that heat storage fluid flows from tank  112 H to tank  112 C through heat exchanger  116 , thereby providing heat to working fluid  108 . Working fluid  108  expands in turbine  104  to move the system from  404  to  406  on  FIG. 4 . Mechanical energy is provided by the apparatus in this mode, because the work released by expanding from  404  to  406  on  FIG. 4  is greater than the work required to compress from  408  to  402  on  FIG. 4 . 
     Importantly, the thermodynamic cycle of  FIG. 4  is closed by connecting the exhaust of turbine  104  to the input of compressor  102  through the cold-side heat storage unit  120 . Heat is removed from working fluid  108  by heat storage unit  120  to move the system from  406  to  408  on  FIG. 4 . More specifically, valve  124  provides connections as shown such that heat storage fluid flows from tank  122 C to tank  122 H through heat exchanger  126 , thereby storing heat provided by working fluid  108 . This step can be understood as storing the heat energy present in the (hot) exhaust from turbine  104 . Hot-side heat storage unit  110  and cold-side heat storage unit  120  have comparable total heat capacity. The need for this can be appreciated in connection with the generation mode of  FIG. 2 , where it is apparent that cold-side heat storage unit  120  stores a fraction of the heat stored in hot-side heat storage unit  110  (i.e., the fraction of the stored hot-side heat that ends up in the exhaust from turbine  104 ). 
       FIG. 3  shows operation of the example of  FIG. 1  in a refrigerator mode that uses mechanical work to store heat energy. Working fluid  108  at the input of compressor  102  is represented by point  406  on  FIG. 4 . Compression of working fluid  108  moves the system to point  404  on  FIG. 4 . Heat is removed by heat storage unit  110  to move the system from  404  to  402  on  FIG. 4 . More specifically, valve  114  provides connections as shown such that heat storage fluid flows from tank  112 C to tank  112 H through heat exchanger  116 , thereby removing heat from working fluid  108 . Working fluid  108  expands in turbine  104  to move the system from  402  to  404  on  FIG. 4 . Mechanical energy must be provided to the apparatus in this mode, because the work released by expanding from  402  to  408  on  FIG. 4  is less than the work required to compress from  406  to  404  on  FIG. 4 . 
     Importantly, the thermodynamic cycle of  FIG. 4  is closed by connecting the exhaust of turbine  104  to the input of compressor  102  through the cold-side heat storage unit  120 . Heat is added to working fluid  108  by heat storage unit  120  to move the system from  408  to  406  on  FIG. 4 . More specifically, valve  124  provides connections as shown such that heat storage fluid flows from tank  122 H to tank  122 C through heat exchanger  126 , thereby providing heat to working fluid  108 . This step can be understood as warming up the (cold) exhaust from turbine  104 . 
     From the preceding description, it is apparent that in either mode of operation, two of the storage tanks  112 H,  112 C,  122 H, and  112 C will be feeding heat storage fluid to the system, and the other two tanks will be receiving heat storage fluid. The feed tanks set their own temperatures. The receiving tanks see fluid temperatures that depend on how the system is operating—i.e., its loads and/or power input. Ideally, the receiving tank fluid temperatures are set by the Brayton cycle conditions, but in practice there will be deviations from these conditions, and the pressure ratio varies in response to system demand. 
     A system controller (not shown) controls system parameters in order to approximately match the ideal temperature conditions. Suitable system parameters include but are not limited to: the flow rate of first heat storage fluid  118 , the flow rate of second heat storage fluid  128 , and operating parameters of compressor  102  and turbine  104  such as turbine stator blade positions. Because of entropy creation within the system, it will not be possible to match the ideal temperature conditions exactly, and at least one of the four tank temperatures will be too high. The purpose of radiator  130  is to reject this waste heat to the environment as part of system control. Suitable techniques for controlling systems as described herein are known in the art. 
     Some principles of the present invention can be better appreciated in connection with a specific example where hot-side heat storage fluid  118  is a molten salt and cold-side heat storage fluid  128  is water. In this example, there is a water side and a salt side, each characterized by two temperatures. However, these 4 temperatures are not independent of each other. Each salt temperature is the product of the corresponding water temperature and a factor that depends on the compressor pressure ratio (numerically, this factor is typically about 2). Thus, in steady state operation, there are only two independent temperatures. The water temperatures need to be in the liquid range for water (at 1 atmosphere) for safety, and the salt temperatures need to be in the liquid range for the relevant salt, and be at a temperature range that structural steels can handle. Fortunately, salts that are molten at temperatures on the order of 450-700 K are known, and such temperatures are well below typical steel melting or creep temperatures. 
     To better appreciate the present approach, it is helpful to note that it is possible to perform energy storage and retrieval without using a closed cycle for working fluid  108 . More specifically, the cold-side heat storage unit  120  could be removed from  FIG. 1 , thereby opening the cycle such that the compressor input is provided by the environment, and the turbine exhausts to the environment. 
     However, this open-cycle approach has numerous and severe disadvantages. The open-cycle approach entails employing atmospheric air as the cold-side heat reservoir. This automatically precludes the use of any working fluid other than air. It also precludes the use of counterflow heat exchange to minimize entropy production. It also exposes the system to environmental dangers, for example humidity fluctuations that could cause condensation or even freezing of water in the turbine operating in refrigerator mode, with total destruction of the turbine as the likely result. 
     A particularly important modification of the working fluid allowed by a closed cycle is pressurization. This enables the input pressure to compressor  102  to be higher than atmospheric pressure. It is helpful to consider the minimum pressure (P min ) of working fluid  108  in its closed cycle. The minimum pressure is typically found on the cold side of the apparatus (e.g., at the input to compressor  102 ). Although P min  can be as low as 1 atmosphere (atm), it is preferred for P min  to be about 10 atmospheres or greater. 
     This increase in power density provided by a high-pressure working fluid can be extremely significant. A storage turbine at 1 atm pressure generates about 1/10 the power of a combustion turbine of the same size. This can be seen by comparing the exhaust temperatures. For example, a large commercial power gas turbine has an output of 256 megawatts, a compression ratio of 15.3 and an exhaust temperature of 608° C. (i.e. 578° C. greater than the intake temperature). For a storage turbine based on air, which might have a compression ratio of 14 and an exhaust temperature rise (retrieval step) of 75° C., the same size as the above commercial power turbine and flowing the same amount of working fluid (643 kg/sec), the resulting power is (256 MW) (75° C.)/(578° C.)=33.2 MW. This is absurdly low for such a large machine. 
     To put this problem in perspective, the throat intake speed of industrial gas turbines is typically a significant fraction of the sound speed in air (e.g., Mach 0.5). Since sea level air has a mass density of 1.22 kg/m 3  and a sound speed of 343 m/s, the throat area required to accommodate the mass flow is about 3 m 2 . The power required merely to accelerate the air mass in question to Mach 0.5 is about 9.5 MW. Some of this power can be recovered from exhaust hydrodynamics, but not all, and the lost part is comparable to the energy one is trying to extract. 
     Thus it is important that the power output of the turbine of a given size be substantially raised. This can be done by raising the ambient pressure of the working fluid. If, for example, the pressure is raised to 10 atmospheres, something that steel can accommodate easily, the power output becomes 10 times what it was before, which is an amount comparable to that generated by a combustion gas turbine of the same size. The pressures and temperatures in question also feature in modern supercritical steam plants, so the steel is expected to be able to take the stress. The elevated working fluid density should also help raise the compressor efficiency, although the exact amount is difficult to estimate accurately. Water (i.e., a dense fluid) can be pumped uphill with 90% efficiency using Francis turbines. This high efficiency is what makes hydropumping the leading energy storage technology at the moment. 
     The closed loop also enables one to conserve momentum, as in a wind tunnel. This becomes increasingly important as the mass of the fluid rises, for then the total fluid kinetic energy passing by a point per second can become comparable to the power one is trying to store or extract. In a closed circuit this energy is automatically conserved (except for friction losses at the walls) so it doesn&#39;t reduce efficiency, but in an open circuit, where kinetic energy gets lost to the environment, it does reduce efficiency. 
     To better appreciate some further preferred embodiments, it is helpful to provide some results from an analysis of the Brayton cycle of  FIG. 4 . For adiabatic compression of a gas having temperature T 0  and pressure P 0  to a pressure P 1 , the resulting temperature T 1  is given by 
     
       
         
           
             
               
                 
                   
                     
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     where γ is the heat capacity ratio (i.e., C p /C v ) of the gas. The heat dumped to the environment per mole of working fluid compressed (Q dump ) is given by 
     
       
         
           
             
               
                 
                   
                     
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     where R is the ideal gas constant, η c  is the compressor efficiency, and T e  is the environment temperature. It is assumed that the compressor is the dominant source of entropy production in the cycle. This assumption is reasonable in view of the use of counter-flow heat exchangers and the high efficiencies provided by turbines in practice. The energy stored per mole of working fluid compressed (E store ) is given by 
     
       
         
           
             
               
                 
                   
                     
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     where ΔT is the temperature difference between the hot and cold storage tanks (e.g.,  112 H and  112 C). The thermodynamic efficiency of energy storage (η store ) is given by 
     
       
         
           
             
               
                 
                   
                     
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     For a numerical example, let P 1 /P 0 =14, η c =0.9, T e =300 K, ΔT=150 K and γ=1.4. The resulting storage efficiency is η store =0.857. The efficiency of retrieval is the same as for storage, so the total efficiency for storage+retrieval is η store   2 . 
     From these results, several further preferred features may be understood. Although air can be employed as a working fluid, a preferred working fluid is Argon. Argon is inexpensive, and has better properties than air. More specifically, γ for Argon is 1.66 and γ for air is 1.4, so Argon is seen to improve the efficiency given by Eqn. 4. Commonly employed working fluids in conventional refrigerators, such as ammonia and freon, are not preferred working fluids in this context, because drops of their liquid phase may form in operation and damage the turbine blades. 
     The effect of the use of Argon instead of air as the working fluid can be better appreciated in view of some compressor design considerations. The compressor is the dominant source of inefficiency in the present apparatus. Axial compressors, (e.g. those in jets and as shown on  FIGS. 1-3 ) tend to be the most efficient kind of compressor, particularly in applications requiring large volume flows. The fundamental limit of efficiency per stage in an axial compressor is about η s =0.9. The overall compressor efficiency degrades with stage number n according to: 
     
       
         
           
             
               
                 
                   
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       FIG. 5  shows plots of η c (n) for η s =0.9 (top curve) and η s =0.8 (bottom curve). The compression ratio per stage is taken to be r=1.4, and the specific heat ratio is taken to be γ=1.4. For P 1 /P 0 =14 as in the preceding example, the number of stages required is 7.8 (which rounds to 8). This number of stages degrades the overall compressor efficiency to 0.86 and reduces the storage-step efficiency to η store =0.80. The retrieval-step efficiency is the same, so the round-trip storage efficiency is the square of this number, or 0.64. 
     Thus there is a significant premium in increasing the stage efficiency even by a tiny amount and thereby reducing the number of stages. Substituting Ar for air as the working fluid, for example, increases the specific heat ratio to γ=1.66, reduces the overall compression ratio required from 14 to 6.7, and thus reduces the number of stages to 5.6 (which rounds to 6). The storage-step efficiency then rises to 0.84, which gives 0.71 when squared. 
     It is also clear from Eqn. 4 that there is a significant efficiency advantage in maximizing the temperature difference ΔT H  between the tanks on the hot side of the circuit. This is related by the Brayton cycle condition to the temperature difference ΔT c  between the tanks on the cold side of the circuit by 
       Δ T   H =( P   1   /P   0 ) (γ-1)/γ   ΔT   C .  (6)
 
     For P 1 /P 0 =14, γ=1.4 (i.e., air), and ΔT c =75 K, the resulting ΔT H  is about 150K (more specifically, it is 159 K). The value for ΔT c  in this example is a conservative liquid range for water. It is highly undesirable to pressurize the water to allow temperatures greater than 100° C., on account of the extreme explosion danger thereby created. Thus the only practical way to increase this range is extend the cold side to below room temperature. One can obtain a further 25 K by going down to the freezing point of water and a further 30 K if antifreeze is added as described above. Assuming ΔT c =130 K, the hot-side temperature difference then becomes ΔT H =276 K which gives a corresponding storage efficiency of η store =0.91. 
     In the preceding two examples, the effects of using Argon as the working fluid and of increasing the cold-side temperature difference were considered separately for ease of explanation. These approaches for improving efficiency can be practiced simultaneously, and can also be practiced in combination with any other ways of improving efficiency (e.g., operating at higher pressures). Preferably, efficiency is maximized by making use of all available methods of increasing efficiency. For example, the choice of working fluid can be considered and optimized in combination with compressor/turbine optimization.