Performance analysis of a novel medium temperature compressed air energy storage system based on inverter-driven compressor pressure regulation

Abstract

In compressed air energy storage systems, throttle valves that are used to stabilize the air storage equipment pressure can cause significant exergy losses, which can be effectively improved by adopting inverter-driven technology. In this paper, a novel scheme for a compressed air energy storage system is proposed to realize pressure regulation by adopting an inverter-driven compressor. The system proposed and a reference system are evaluated through exergy analysis, dynamic characteristics analysis, and various other assessments. A comprehensive performance analysis is conducted based on key parameters such as thermal storage temperature, component isentropic efficiency, and designated discharge pressure. The results show that the novel system achieves a relative improvement of 3.64% in round-trip efficiency, demonstrating its capability to enhance efficiency without significantly increasing system complexity. Therefore, the system proposed offers a viable solution for optimizing compressed air energy storage systems.

Graphical abstract

Keywords

adiabatic compressed air energy storage / throttle valve exergy loss / performance analysis / inverter-driven compressor

1 Introduction

In the current global context of fossil fuel scarcity and global warming, renewable energy is rapidly developing. In 2022, there was a notable increase of 295 GW in the global installed capacity of renewable energy, with wind and solar energy accounting for over 90% of this capacity expansion [1]. However, the intermittency of renewable energy sources and the limitations of grid integration have resulted in significant instances of curtailing wind and solar power, which leads to substantial economic losses [2]. To address this issue, energy storage technologies play a crucial role in balancing energy supply and demand, and these technologies are thus essential for maintaining grid stability.

Among all large-scale and long-term energy storage systems, compressed air energy storage (CAES) demonstrates significant advantages, making it a highly promising energy storage technology with significant energy and environmental benefits [3]. For instance, CAES is highly efficient, retaining approximately 70%–75% of input energy for future use; CAES is flexible, capturing energy during low demand periods and releasing it when demand rises, contributing to grid stability and minimizing energy waste; a CAES system seamlessly incorporates renewable energy sources, storing excess energy and reducing reliance on fossil fuels; CAES aids in decreasing greenhouse gas emissions by harnessing surplus renewable energy, thereby lessening the demand for fossil fuel-based power; CAES leverages either natural geological formations or man-made structures for storage, effectively minimizing land use; and adiabatic CAES systems have facilitated carbon-free operations, resulting in a significant increase in their scale of deployment.

The CAES technology has been continuously optimized and updated through the research and efforts of scientists and engineers in the field of thermodynamic systems. However, the current system efficiency of practical CAES implementations remains at approximately 70% [4,5]. Therefore, improving the cycle efficiency of CAES is of great significance for reducing generation costs and enhancing the market competitiveness of the technology. Many solutions have been proposed to enhance system efficiency, such as trigeneration [6,7] and coupling systems with wind power systems [8,9], solar power systems [10], and desalination systems [11].

Optimizing the throttle losses in the system is also an effective approach to improving efficiency. Some researchers have proposed methods to eliminate throttle devices, such as using isobaric air storage, switching valves pressure ratio control, and inverter-driven (ID) pressure regulation technologies.

In the field of isobaric air storage systems, Wang et al. [12] introduced a multilevel underwater CAES system, demonstrating the capability to attain exergy efficiencies ranging from 62% to 81% across diverse operational modes. Maisonnave et al. [13] developed a dynamic reversible model for a multiphysical conversion platform along with a control scheme and evaluated the dynamic performance. Jiang et al. [14] compared the performance of a 600 kW CAES air storage under isobaric and isochoric conditions. The results show that the exergy efficiency was 65.9% under isobaric conditions, which was 4.1% higher than that under isochoric conditions. Chen et al. [15] proposed an isobaric CAES system based on volatile fluid, which could achieve more than 4% exergy efficiency improvement compared with the reference adiabatic CAES system. Mazloum et al. [16] performed an exergy-economic analysis on an isobaric CAES system with a pressure stabilization system, which had an efficiency improvement of 2.7% and a capital investment reduction of 5.6%. The wide application of these methods is challenging due to the complex structures.

Using an injection device, pressure regulation is another method for reducing throttle losses. Zhou et al. [17] proved that the configuration of injectors could improve cycle efficiency by more than 2%. Cao et al. [18] achieved a relative reduction of 39.87% in system backpressure variation by placing injectors near the final-stage compressor. Zhang et al. [19] proposed a fully automatic injector that improved the system cycle efficiency by 0.93% compared to conventional injectors. However, these injection devices sacrificed some of the air storage pressure, and these prior studies have not considered the increase in along-path resistance losses caused by the complexity of the system pipelines.

Some efforts have also been made to propose a switching valve-based system topology dynamically adjusting method for throttle loss decrease. He et al. [20] proposed a novel variable pressure ratio compressed air system, whose efficiency was increased by 13.1%. Fu et al. [21] achieved a nearly 13% increase in system efficiency through dynamic topology adjustment. Despite achieving efficiency improvements, the use of variable pressure/expansion ratios in system design may lead to a decrease in the grid regulation capability of such CAES systems.

These abovementioned optimization schemes have the disadvantage of high implementation difficulty, making them difficult to be applied in practical engineering. To realize a simple and usable system throttling loss optimization, the authors of the present paper previously proposed a novel system optimization scheme which involves the addition of an extra inverter-driven air compressor (ID-AC) to control the outlet pressure of the isochoric air storage tank (AST). As a result, the system round-trip efficiency was improved by 1.8%–2.7% [22].

While ID technology has become mature in the power industry, its application in CAES systems is still in its infancy. Herein, a novel system (ID-CAES system) is proposed, which replaces the final stage compressor of the traditional CAES system with an ID-AC to regulate the charge and discharge process air pressure. As a result, system throttling losses are reduced without requiring any modifications to the primary equipment. A characterization analysis of the system is also conducted such as the consideration of the system thermal storage temperature, maximum air storage pressure, and discharging pressure ratio.

2 System description

To achieve a more cost-effective heat recovery, the medium temperature CAES technology that utilizes water as the thermal storage medium has been increasingly applied in various engineering projects. In this work, a typical medium temperature CAES system was chosen as the reference system for investigation. The system employs a four-stage compression and three-stage expansion process, whose schematic diagram is depicted in Fig.1.

The reference CAES system in the charging process employs the first three stages of ACs with equal compression ratios. The compression ratio of the first three stages of ACs is dependent on the thermal storage temperature and the temperature difference at the heat exchangers. To reduce compressor energy consumption and achieve a cascaded utilization of energy, the outlet temperature of the compressors is set to be 10 °C higher than the thermal storage temperature. To meet the pressure range of the AST, an additional compressor with a lower compression ratio (AC4) is added. Air coolers are installed to reduce the inlet temperature of the compressors and the temperature of the AST, thereby reducing compressor power consumption, and increasing energy storage density. In the discharging process, high-pressure air is heated by the thermal storage medium before entering the air turbine (AT) to generate axis work. The ATs are set with equal expansion ratios. Throttling valves are installed at the inlet and outlet of the AST to ensure the safety and stable power output of the compression and expansion devices.

The method proposed introduces a novel approach that includes utilization of an ID compressor as a substitute for throttle valves (TVs) to control the compression/expansion ratio during the charging and discharging processes. During the charging process, the ID-AC dynamically adjusts the pressure ratio in response to the pressure drop inside the AST, which ensures efficient compression of the air during storage. During the discharge process, as the pressure inside the AST decreases, ID-AC provides a pressure increase, enabling a consistent and stable pressure supply to the AT train. By replacing the traditional throttling process with the utilization of ID-AC, the pressure ratio and power output can be dynamically adjusted and optimized during the charging and discharging process. This flexibility allows for a better energy conservation, as the system can adapt to internal pressure variations, which contributes to achieving the goal of energy conservation in a more effective manner. A schematic diagram of the ID-CAES system proposed is depicted in Fig.2. Both the ID-CAES and the reference CAES systems exhibit a four-stage compression and three-stage expansion system configuration.

The ID-CAES system utilizes two directional valves (V1 and V2) to control the flow direction of the working fluid. During charging, V1 cuts off the pipeline entering the AST, while V2 cuts off the pipeline entering HEX4. As shown in Fig.3(a), air enters the ID-AC from HEX3, and the compression ratio of the ID-AC increases during the charging process to accommodate the rising internal pressure in the AST. During the discharging process, as illustrated in Fig.3(b), V1 cuts off the pipeline from HEX3 while allowing air from AST to enter ID-AC. V2 cuts off the pipeline from ID-AC to AST while guiding air into HEX4. The ID compressor (ID-AC) driven by an in-system generator pressurizes the air, and the outlet pressure of the ID-AC is maintained constant and set to the maximum operating pressure of the AST. Due to the adoption of different overall inlet pressures during the discharging process in the system proposed compared to the reference system, the expansion turbine and heat exchanger need to be reselected according to the new operating conditions.

3 Method

Modeling of the CAES system were conducted based on thermodynamic characteristics of its key components and an exergy balance. To analyze and optimize the two systems, thermodynamic analysis methods based on an evaluation indicator of the second law of thermodynamics were introduced. The EBSILON® Professional software was employed as the modeling tool, which is widely used in system analysis and optimization research [23] and considered as a reliable tool for thermodynamic modeling. The validation results are shown in Electronic Supplementary Material (ESM), and the error of all results show that the model meets the accuracy requirements of thermodynamic analysis.

During the modeling process of the CAES system, to simplify the analysis it was assumed that the air is an ideal gas, which consists of 75.5% N₂, 23.1% O₂, and 1.3% Ar by mass fraction [8,9]; the pressure drop and heat dissipation in the pipelines are neglected [8,9]; the power consumption of the pumps in the thermal energy storage system is neglected [24]; and there is no heat dissipation in the thermal storage systems, heat exchangers, and coolers [22].

3.1 Component model

1) Air compresser (AC)

The AC is classified as an energy-consuming component. The compression process is divided into adiabatic compression, isothermal compression, and polytropic processes. The CAES system employs staged compression and interstage cooling, often utilizing axial flow compressors. Axial-flow compressors offer advantages such as high individual capacity, high compression efficiency, and large unit area flow capacity. Moreover, their thermodynamic process closely approximates adiabatic conditions. It is assumed that the compression process is adiabatic. The compressor outlet temperature is described as [24]

(1)T_{\text{out}}=T_{\text{in}}\left[\right({\text{π}}_{\text{AC}}^{(\kappa -1)/\kappa }-1)/{\eta }_{\text{AC}}+1],

where T_{\text{out}}\left(\mathrm{K}\right) and T_{\text{in}} (K) are temperatures of the air leaving and entering the compressor, respectively; π_AC is compression ratio (π_AC = P_out/P_in); \kappa is adiabatic index of air; {\eta }_{\text{AC}} is isentropic efficiency of the compressor.

The power consumed by the compressor (W, MW) is expressed as

(2)W= m_{\text{air}}( h_{\text{out}}-h_{\text{in}})\times 10^{-3},

where m_air (kg/s) is the mass flow rate of air, and h_out (kJ/kg) and h_in (kJ/kg) are specific enthalpies at the compressor outlet and inlet, respectively.

2) Air turbine (AT)

The AT belongs to the work-producing component and is a key discharging component in CAES systems. The AT is regarded as a steady flow open system. Assuming the expansion process to be adiabatic, neglecting the influence of equipment structure and ignoring heat dissipation and losses during operation, the relationship between the air temperatures at the inlet and outlet is obtained by using Eq. (3) [24],

(3)T_{\text{out}}=T_{\text{in}} [1- {\eta }_{\text{AT}} (1- {\text{π}}_{\text{AT}}^{(1-\kappa )/\kappa }\left)\right],

where T_out (K) is the temperature of the air at the outlet of the compressor, T_in (K) is the temperature of the air entering the AT, π_AT is the expansion ratio (π_AT = P_out/P_in), and {\eta }_{\text{AT}} is the isentropic efficiency of the AT.

The power output of the AT W (MW) is given by

(4)W= m_{\text{air}}( h_{\text{out}}-h_{\text{in}})\times 10^{-3},

where m_air (kg/s) is the mass flow rate of air, and h_out (kJ/kg) and h_in (kJ/kg) are the specific enthalpies at the inlet and outlet of the AT, respectively.

3) Heat exchanger

The heat exchanger is an energy exchange device. The “heat exchanger effectiveness–—number of heat transfer units” method is adopted for the heat exchangers in CAES systems and newly added heat exchangers in coal-fired power plants [25]. The calculation of heat transfer through the heat exchanger is expressed using the heat exchanger effectiveness (EFF) as

(5)\mathrm{E}\mathrm{F}\mathrm{F}=\frac{(T_{\text{in}}-T_{\text{out}}{)}_{\text{max}}}{T_{\text{in}}^{\text{hot}}-T_{\text{in}}^{\text{cold}}},

where T_{\text{in}}^{\text{hot}}and T_{\text{in}}^{\text{cold}} are the inlet temperatures of the hot fluid and cold fluid, respectively.

For a counterflow heat exchanger, the EFF is expressed as the ratio of the larger actual temperature difference between the cold and hot fluids to the maximum possible temperature difference across the heat exchanger,

(6)\mathrm{E}\mathrm{F}\mathrm{F}=\frac{1-\text{exp}(-\mathrm{N}\mathrm{T}\mathrm{U}(1-\chi ))}{1-\chi \text{exp}(-\mathrm{N}\mathrm{T}\mathrm{U}(1-\chi ))},

where NTU is the number of transfer units, which is expressed as

(7)\mathrm{N}\mathrm{T}\mathrm{U}=\frac{kA}{{\dot{\tilde{C}}}_{\text{min}}},

where {\dot{\tilde{C}}}_{\text{min}} (kJ/(s∙°C)) is the minimum value of the thermal capacity ratio in the hot and cold fluids of the heat exchanger,

(8)\chi =\frac{{\dot{\tilde{C}}}_{\text{min}}}{{\dot{\tilde{C}}}_{\text{max}}},

where {\dot{\tilde{C}}}_{\text{max}} (kJ/(s∙°C)) is the maximum value of the thermal capacity ratio in the hot and cold fluids of the heat exchanger.

4) Air storage tank (AST)

Gas storage systems are classified into various types, including constant pressure and constant volume. It is assumed that the system utilizes a technologically mature constant volume gas storage device. Depending on the application scenario, constant volume compressed air storage devices can be further categorized into isothermal, adiabatic, or constant wall temperature types. To mitigate the impact of structural parameters on the calculation results, it is assumed that the temperature of the compressed air remains constant after entering the gas storage device. The outlet temperature and pressure of the storage tank are thus determined as

(9)T_{\text{out}}=T_{\text{in}} ,

(10)P_{\text{out}}\left(t\right)= P_{\text{max}}-(P_{\text{max}}-P_{\text{min}}){\int }_0^t\frac{m_{\text{dichar}}}{t_{\text{char}}{m}_{\text{char}}}\mathrm{d}t.

5) Throttle valve

During the charging and discharging processes, the pressure of the air in the storage chamber changes. It is thus necessary to ensure the safe and stable operation of the AC and AT units. This safe operation is achieved by installing TVs at the inlet and outlet of the storage chamber to maintain a stable back pressure for the compressor and inlet pressure for the AT, thereby stabilizing the output power during the discharge process. The process of air passing through the regulating valve is considered as an adiabatic throttling process where the enthalpy remains constant before and after throttling, which is described as

(11)h_{\text{in}}=h_{\text{out}} .

3.2 Energy analysis

Exergy analysis is employed to analyze CAES systems. The exergy analysis is based on the second law of thermodynamics and is used to quantify the exergy losses and exergy efficiencies of the system components. This technique can reveal the causes of losses and identify the primary equipment where losses occur, providing guidance for system optimization. Based on the established thermal balance model, an exergy analysis is conducted by establishing an exergy balance for the CAES system. To evaluate the performance of the system using exergy analysis, it is necessary to establish a reference state. Thus, reference states of 20 °C and 101.323 kPa are set. When neglecting the exergy associated with fluid kinetic energy and potential energy, the specific exergy rate of the working fluid \dot{E}x is expressed as the physical exergy rate of the working fluid, which is calculated as [26]

(12)\dot{E}x=m{e}_x= m\left[\right(h-h_0)-(T_0+273.15\left)\right(s-s_0\left)\right],

where m (kg/s) is the mass flow rate of the working fluid, e_x (kJ/kg) is the exergy flow rate, h (kJ/kg) is the enthalpy of the working fluid, s (kJ/(kg∙°C)) is the specific entropy of the working fluid, h_0 (kJ/(kg∙°C)) is the reference state enthalpy of the working fluid, s_0 (kJ/(kg∙°C)) is the reference state entropy of the working fluid, and T_0 (°C) is the reference state temperature.

The approach of exergy “fuel-product” analysis is applied to analyze the system components. This analysis method facilitates the quantification of individual equipment exergy losses, excluding losses caused by system factors. For the kth equipment, the exergy balance relationship is expressed as

(13)\dot{E}{x}_k^{\mathrm{F}}=\dot{E}{x}_k^{\mathrm{P}}+ \dot{E}{x}_k^{\mathrm{D}},

where \dot{E}{x}_k^{\mathrm{F}} (kW) is the fuel exergy rate released by equipment k,\dot{E}{x}_k^{\mathrm{P}} (kW) is the product exergy rate generated by equipment k, and \dot{E}{x}_k^{\mathrm{D}} (kW) is the exergy loss rate due to irreversibility in equipment k.

The main equipment and corresponding calculation formulas in exergy analysis are shown in Tab.1. For the compressor, the input fuel exergy is the total consumed power over one working cycle, while the product exergy is the increase in the exergy value of the working fluid (air) generated during the working cycle. For the throttle valve and storage tank, the fuel exergy is the total input exergy during the working time, while the product exergy is the total output exergy during the same time period. For the heat exchanger, the fuel exergy is the difference between the total exergy at the inlet and outlet of the hot fluid side, while the product exergy is the difference between the exergy at the outlet and inlet of the cold fluid side. For an AT, the fuel exergy is the decrease in the exergy of the working fluid during the operating time, while the product exergy is equal to the total electricity generation of the equipment.

3.3 Performance evaluation criteria

To comprehensively analyze CAES systems, some researchers proposed various evaluation criteria based on the first and second laws of thermodynamics [27]. Each one of these evaluation criteria has its own range of applicability. Herein, the widely accepted and commonly used concept of “round trip efficiency (RTE)” is implemented as a thermodynamic performance indicator for CAES systems. For an isolated CAES system, RTE is defined as the ratio of the electrical energy output during the discharge process to the electrical energy input from the external source during the charging process,

(14)\mathrm{R}\mathrm{T}\mathrm{E}=\frac{{\int }_0^t{W}_{\text{dischar}}\mathrm{d}t}{{\int }_0^t{W}_{\text{char}}\mathrm{d}t},

where W_{\text{dischar}} (MW) is the electrical power output during the discharge process and W_{\text{char}} (MW) is the electrical power output during the charge process.

4 Case study

To assess the performance of the novel ID-CAES system proposed, a comparative analysis of a typical CAES system is conducted. The equipment parameters for both systems are based on the actual engineering experience. The system utilizes an isochoric AST, with a discharge power of 300 MW and a charging time of 8 h for the reference system. In the ID-CAES system, the power generation of the AT train (ACs1−3) is fixed at 300 MW, and the total output power needs to account for the energy consumption of the ID-AC. The maximum pressure of the AST is 100 bar, while the minimum pressure is 80 bar. More detailed system parameters are provided in Tab.2.

4.1 Simulation results

The main simulation results for the reference system and the ID-CAES system proposed are provided in Tab.3. During the charging process, to ensure that the outlet temperatures of the first three-stage compressors are the same, AC1 needs to maintain higher pressure ratios and higher power, while the inlet temperature of AC1 is lower than that of AC2 and AC3. The ID-AC used in the ID-CAES system operates at variable power during both charging and discharging processes, with a power range of 3.08–14.92 MW during charging and 0–16.39 MW during discharging. Although the total output power of the ID-CAES system is slightly reduced by this operation, the RTE is improved by 0.026 compared to the reference system, representing a relative increase of 3.64%.

A “temperature−entropy” diagram of air during the operation of the reference system is illustrated in Fig.4. During the charging process, air is compressed and cooled from point 1 to point 12 and then stored in the AST. During the discharging process, air is throttled to point 13 from the AST and then heated and expanded in two stages. The temperatures at points 2, 5, and 8 during charging and points 14, 16, and 18 during discharging are influenced by the selected thermal storage temperature and the inlet/outlet temperature difference of the selected heat exchanger. The air cooler used in the system realizes lower temperatures at points 4, 7, and 10, such that the power consumption of the compressor is reduced. Due to the lower temperature rise of AC4, this component is not utilized during the discharging process. For more information regarding the working fluid data of the system, please refer to ESM.

To analyze the locations of irreversible losses occurring in the reference system and provide ways for further optimization, exergy analysis is conducted. The proportion of equipment exergy losses in the system is shown in Fig.5. The top three equipment with exergy losses are the ATs, ACs, and TVs. The results show that enhancing the efficiency of the “electric-to-power” conversion devices is crucial for improving system efficiency. The exergy loss of the TVs accounts for 16.1% of the total loss, which is also significant and presents another opportunity for optimization based on systemic improvement.

The exergy losses for each working cycle in the charging and discharging processes of the reference system and ID-CAES system are depicted in Fig.6. Compared to the reference system, the ID-CAES system exhibits increased losses in the discharging heat exchangers HEXs (HEXs4–6) and ATs (ATs1–3) due to the higher overall pressure ratio in the discharging process. However, the losses in Cooler4 are reduced due to the lower high-temperature working fluid parameters. Despite the additional losses introduced by ID-AC, the net losses in the system are significantly reduced due to the avoidance of exergy losses in TVs1 and 2.

The internal pressure and unit power of the AST in the charging and discharging processes of the ID-CAES system are shown in Fig.7. The discharging time of the ID-CAES system is 5.77 h, which is longer than that of the reference system. This increased discharge time is caused by the reduction in throttle losses in the ID-CAES system, which improves the compression capability of the compressed air. In the ID-CAES system, the power of the ID-AC changes with the variation of pressure in the AST, thereby altering the total input and output power of the system as the air storage pressure increases.

4.2 Parametric analysis

During optimization of the reference system with an ID compressor, it is crucial to select key indicators that have a significant impact on the thermodynamic characteristics of the system. Among various possible indicators, the efficiency and design outlet pressure of the ID compressors, as well as the thermal storage temperature and maximum air storage pressure parameters of the system, are particularly important.

4.2.1 ID-AC isentropic efficiency

It is necessary to conduct an analysis of the isentropic efficiency of ID-AC on the characteristics of the whole system. The impact of the isentropic efficiency of the ID compressor, ID-AC, on the RTE and discharging time of the system proposed is depicted in Fig.8. When the isentropic efficiency of ID-AC increases from 0.8 to 0.95, the RTE of the system rises from 0.733 to 0.741. From this observation, it can be concluded that the impact of the isentropic efficiency of the ID-AC on the RTE of the system is relatively small, and that, within a reasonable range of isentropic efficiency for ID-AC, the system proposed can maintain a significant efficiency advantage over the reference system. Observing the difference in RTE based on the variation in isentropic efficiency reveals that the RTEs of the reference system and the ID-CAES system proposed are the same when the isentropic efficiency is 0.42. Notably, the isentropic efficiency of ID-AC does not affect the charging and discharging time of the system.

The relationship between the improvement of the isentropic efficiency of ID-AC in the system proposed and the equipment power during the charging and discharging process is depicted in Fig.9. Because the power for ID-AC changes during the charging and discharging process, it is represented by a red-colored region in the graph. From Fig.9, it can be observed that the change in the isentropic efficiency of the ID-AC does not affect the power of ACs1−3 and ATs1−3 but only alters the power of ID-AC during the charging and discharging process. As the isentropic efficiency increases, the power consumption of ID-AC decreases, thereby improving the round-trip efficiency of the system.

4.2.2 ID-AC outlet pressure

The main function of the ID-AC in the novel system proposed is to adjust the compression and expansion ratio in coordination with the pressure of the AST while also stabilizing the inlet pressure of the AT train at the maximum pressure of the AST. Through system optimization, the ID-AC controls the inlet pressure of the AT to be greater than, equal to, or less than the maximum pressure of the AST.

As shown in Fig.10, when the set pressure is not lower than the maximum pressure of the AST, the air is directly pressurized through the ID-AC. When designing the inlet pressure of the AT1 to be lower than the maximum pressure of the AST, the air flow direction is controlled by V2 based on the internal pressure of the AST. When the pressure inside the AST is higher than the design inlet pressure of the AT, the air flows through pipeline 1, undergoes TV, and enters the ATs. When the pressure inside the AST is lower than the designed inlet pressure of AT1, the air flows through pipeline 2 and undergoes pressurization by ID-AC.

The influence of the ID-AC outlet pressure on the RTE and discharging time of the optimized system is illustrated in Fig.11(a). As the ID-AC outlet pressure increases, the RTE of the system initially rises and then declines, reaching its peak value of 0.741 at 94 bar. This outcome is attributed to the dual impact of ID-AC outlet pressure on the RTE of the system. First, the increase in outlet pressure leads to a rise in gas compression work, resulting in an extended discharging time. In addition, the higher outlet pressure also causes an increase in the power consumption of the ID-AC, as depicted in Fig.11(b), where the power range of the ID-AC gradually increases. In consideration of the overall complexity and economic performance enhancement of the system, this study selects 100 bar as the design parameter for the system.

4.2.3 AST maximum pressure

The impact of the AST maximum pressure on the RTE of the optimized system and discharging time is illustrated in Fig.12. As the pressure increases, the maximum and minimum pressure differentials of the system remain unchanged. Under the condition of a fixed AST volume, the properties of air in a high-pressure state determine that the charging time is shortened while the discharging time is increased. As a result, the RTE of the system decreases. The maximum pressure of the system rises from 100 to 140 bar, and the system RTE decreases by 0.004. However, the decrease in efficiency is relatively small.

The decrease in system efficiency is primarily caused by the increased power consumption of ID-AC during the charging process, as shown in Fig.13. Both the AC and AT power undergo changes during the charging and discharging processes. The power variation of ID-AC is attributed to two factors: the increase in the compression ratio which leads to a significant rise in its power range, and the high-pressure air which consumes less power than low-pressure air during compression. Therefore, the power of ID-AC during the discharging process decreases, albeit to a lesser extent. As a result, the overall power consumption of the system increases, leading to a decrease in the RTE of the system.

4.2.4 High-temperature storage temperature

The impact of high-temperature storage (HTS) temperature on the RTE and discharging time of the optimized system is illustrated in Fig.14. As the HTS temperature increases, the RTE of the system also increases. With the HTS temperature rising from 170 to 200 °C, the RTE of the system increases by 0.026. Compared to the previously studied parameters, the influence of temperature is more significant. Additionally, higher HTS temperatures enhance the air compression work, resulting in longer discharging times.

The influence of HTS temperature on equipment power is depicted in Fig.15. This temperature primarily affects the power of equipment during the charging process, while its impact on the power during the discharging process is relatively minimal. As the HTS temperature increases, ACs1–3 needs to increase the compression ratio to raise the exhaust temperature, resulting in an increase in its power consumption. However, due to the increased overall pressure ratio of the first three stages, the compression ratio of ID-AC decreases, leading to a decrease in its power consumption. The overall energy consumption during the charging process increases, but this increase is less than the energy released during the discharging process, resulting in an increase in RTE.

5 Conclusions

To reduce the throttling losses in the intake and exhaust of the air storage device, an ID-CAES system with an ID compressor is proposed in this paper based on the medium temperature CAES system. A thermodynamic model of the system is established, and a comparative analysis of the thermodynamic characteristics between the system proposed and the reference system, as well as an analysis of key parameters, are conducted. The main conclusions of this work are as follows:

1) The system proposed enhances the work capacity of the air by elevating the maximum air storage pressure, although it concurrently results in a reduction in the overall discharge power. The system has the capacity to diminish exergy losses associated with system throttling, leading to an enhancement in system round-trip efficiency by 3.64% under design parameters.

2) The outlet pressure of the ID compressor exerts contrasting effects on the work capacity of the air and the power consumption of the compressor itself, ultimately realizing an optimal pressure of 94 bar corresponding to the investigated conditions.

3) The thermal storage temperature has a significant impact on the round-trip efficiency. While increasing the thermal storage temperature may lead to higher compression power consumption, this increased temperature also enhances the capacity of the air for work. Overall, this results in an improvement in the cycle efficiency of the system as the thermal storage temperature rises.

4) The round-trip efficiency of the system is minimally influenced by both the isentropic efficiency of the ID compressor and the maximum air storage pressure. The system proposed manages to sustain its efficiency superiority over the reference system even when dealing with a comparatively low isentropic efficiency for the ID compressor.

To demonstrate the effectiveness of the system proposed, the design parameters employed in this study are set consistently with the reference system. In the future, the system design and will be further refined and optimized by considering the operational characteristics of an actual ID compressor. These advancements will make it possible to further analyze the practical operational features of the system proposed.

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