A novel solar -based compressed air energy storage system is develop ed and analyzed in this paper . The integrated system includes a multi -stage air compression unit, thermal oil loop, multi -stage gas turbine unit, high-temperature molten salt -based solar power tower unit, liquefied air power cycle , thermoelectric generator , and liquefied natural gas ( LNG ) regasification unit . A eutectic mixture of carbonate salt is used for thermal energy storage in the solar subsystem . Energy , exergy , and economic analyses are performed to evaluate the performance of the proposed system. A parametric study investigates the effects of important parameters on the performance of the integrated system. The proposed system stores 356 MWh of grid electricity during the charging mode . Considering all the available energy sources, t he energy output of the integrated system is 797.6 MWh. Energy and exergy efficiencies of the integrated system are 5 5.3% and 4 6.4%, respectively. The results show that the highest contributor to the overall exergy destruction rate of the integrated system is the combustion chamber unit. Finally, levelized cost of electricity is evaluated as 12.4 ¢/kWh. Keywords: Compressed air e nergy storage ; Solar energy ; Economic analysis ; Energy and exergy analys es; LNG regasification ; Thermal energy storage Nomenclature A Area (m2) 𝐸𝑥̇ Exergy (W) ex Specific exergy (kJ/kg) h Specific enthalpy (kJ/kg) I Irradiance (W/m2) 𝑚̇ Mass flow rate (kg/s) P Pressure (kPa) 𝑄̇ Heat transfer rate (W) t Time T Temperature (K) 𝑊̇ Power (W) yx Coefficient of specie x Subscripts ch Chemical exergy D Destruction h Heliostats ke Kinetic exergy pe Potential exergy ph Physical exergy Greek symbols ηI Energy efficiency ηII Exergy efficiency 𝜆̅ Molar fuel to air ratio ∆ Difference 𝜀 Normalized carbon dioxide emissions (kg CO 2/MWh) Abbreviations CAES Compressed air energy storage CCHP Combined cooling, heating, and power GTIT Gas turbine inlet temperature PEC Purchase equipment cost TRR Total revenue requirement 1. Introduction With the exhaustion of fossil fuel resources and aggravation of global energy demand, the utilization of renewable energies to achieve sustainability has become an important subject of research and development in th e modern era [1]. One of the main challenges associated with renewable energies is the ir intermittent nature [2]. To deal with this obstacle and to manage the outgrowing problems of high energy demand, it is necessary to use energy storage systems that store energy when it is not needed an d use it later when the demand is high [ 3]. Energy may be stored in the form of mechanical, chemical , or thermal energy [4]. Compressed air energy storage (CAES) , as a mechanical energy storage method, is one of the most promising methods to resolve the intermittent issues of renewable energies [ 5]. This method posses ses various advantages including technolog ical maturity, availability of equipment, low investment cost, and high energy storage capacity [ 6,7]. CAES systems store energy through a series of air compress ors to pressurize and store air in underground reservoirs such as salt caverns or high-pressure tanks above ground . When energy is required, the pressurized air flow s through a series of expansion units in order to generate power [8]. To increase the system efficiency, it is convenient to use the heat generated in the compression process to heat up the air before enteri ng the expanders [8]. Liu et al. [9] investigated the CAES system through an advanced exergetic analysis. Their results indicated that the round -trip and exergy efficiencies of the cycle were 47.4% and 41.2%, respectively. Advanced exergetic analysis showe d that 34.6 MW of the total exergy destruction of the system is not related to the irreversibilities, and therefore, it is avoidable. More specifically, intercooler, aftercooler, and combustion chamber account for around 58% of the avoidable exergy destruc tion in the system [9]. In another study, Krawczyk et al. [10] investigated two compressed air energy storage systems including a conventional diabatic CAES and a liquefied air energy storage (LAES). The round - trip efficiency of the CAES and LAES systems were reported as 39.77 % and 55.2%, respectively [10]. Peng et al. [1 1] proposed and analyzed a n adiabatic CAES system coupled with a packed bed thermal energy storage system and reported an overall efficiency of 56. 74% [ 11]. In another comparative study, Liu and Wang [12] investigated two different adiabatic CAES system s for energy storage. One of which was a conventional CAES system and the other was a modified version of CAES that used a pneumatic motor, heat accum ulator, and preheater. The results indicated that the modified version had a round trip efficiency of 66.9% which is 13% higher than that of the conventional CAES system [1 2]. Coupling of the CAES system with the Kalina cycle was proposed by Zhao et al. [ 13]. Kalina cycle is used to recover the waste heat of the CAES system and a r ound trip efficiency of 47.64% was reported [ 13]. In another study, Szablowski et al. [14] investigated an adiabatic CAES system in terms of energetic and exergetic perspectives. The round -trip efficiency of their proposed cycle was 50% [ 14]. A transient thermodynamic model for an underwater CAES system was developed and investigated by Carriveau et al. [1 5]. The results indicated that the heat exchangers had the highest rates of exergy destruction among other system components considering the steady -state condition . Under transient condition , the highest exergy destruction contributors are air compressors with a value of 40% of the total exergy destruction of the system [15]. For increasing the performance of gas turbines (GT) or CAES systems, one convenient way is to use natural gas to achieve higher temperature s for the inlet of the gas turbines [16]. As the outlet temperature of the gas turbine is high, they can be coupled with a bottoming cycle to increase the performance of the system [16]. A commonly used combined power cycle utilizes a steam Rankine cycle coupled with the gas turbine units for this mean [17]. Natural gas is mostly liquefied in order to be transported more efficiently in large distances. When it is needed again, it is regasi fied using cryogenic pumps, heat exchangers, and expanders. The cold waste heat of this process can be used as a heat sink of many power generation processes to yield additional power and increase the cycle performance [ 18]. Several working fluids are suit able for this purpose [ 19]. As an example of the compressed gas energy storage, t he feasibility of compressed carbon dioxide energy storage (CCES) integrated with low -temperature thermal energy storage has been studied in the literature [20] in which the r esults show the feasibility of the cycle to be used in cogeneration systems . For the assessment of compressed air energy storage systems, the heat transfer mechanism and heat losses from the storage tank or cavern should be considered. An experimental and theoretical study has been conducted for a high -pressure compressed air tank storage system in which the experimental data was used to assess the wall and air temperatures in the pressurized storage tank of 100 bar. The results showed that the tank wall sh ould be taken into account in the thermal analysis of CAES systems [21]. In another study, a CAES system integrated with a biomass -driven power plant with thermal energy storage was studied and the effect s of heat loss in the cavern and the temperature of the soil around the cavern on the performance of the system we re investigated [22]. Air, as a free and non-toxic working fluid , is a great choice for this mean. Air with a composition of 75.57% nitrogen, 23.16% oxygen, and 1.27%, has a critical temperature of about -140 ℃ while the temperature of LNG at atmospheric pressure is about -160 ℃. Such temperature difference is enough to liquefy air in the cryogenic condenser of a liquefied air power cycle . The technology of such power systems is similar to those used in the LAES systems [ 23]. To improve the thermal efficiency and reliability of a CAES system various thermodynamic cycle can be used and integrated with renewable energy -based systems . Wang et al. [24] proposed and analyzed an underwater CAES system driven by marine renewable energies for an island using energy and exergy thermodynamic analyses. Cold and hot thermal oil loops were utilized in their studies to improve the system performance. Both energy and exergy round -trip efficienc ies are evaluated as 58% for their proposed system [24]. In another study, Wang et al. [25] investigated the integration of O RC with an adiabatic CAES system to produce power, heating, and cooling. Four different configurations of CAES systems were proposed and analyzed in their study. These configurations differed from the number of expansion stages, thermal oil loops, and heat exchanger configurations. Their results indicated that the energy efficiency of these systems was between 52% to 59.6% [25]. Hao et al. [26] proposed a novel energy storage system similar to a conventional CAES system that utilize s carbon dioxide as its working fluid, and also, used a heat pump to eliminate the use of natural gas in the discharging mode . Energ etic and exerg etic efficiencies of the system were obtained as 58.4% and 67.9%, respectively [26]. An integrated wind/solar -based CAES system was investigated by Ji et al. [27]. The round -trip and exergy efficiencies of the proposed system were 61.2% and 65.4%, respectively [27]. Mohammadi and Mehrpooya [28] investigated the coupling of a solar dish with the CAES system. The sys tem consumed 152 kW during charging mode for compression of the air and by using the solar heat to increase the inlet temperature of the gas turbine produced 228 kW during discharging mode. Exergy efficiency of 53.4% and energy efficiency of 76.5% were re ported [28]. Yang et al. [29] proposed an integration of a solar CAES system with combined heating, cooling, and power system for a hotel building in south China. The integrated system considered in their study included solar collector assembly, gas turbin e cycle, absorption chiller, heat recovery steam generator unit, and CAES system. Solar collectors were utilized to enhance the temperature of the air before entering the gas turbine to improve the system efficiency. It was concluded that their proposed system enhanced the efficiency of previous systems by 1.01%. Under different design conditions , the energy efficiency of their integrated system varied between 6 -39% [29]. A novel integration of CAES system with turbocharger and fuel cell subsystems for the production of hot water and electricity is proposed and analyzed thermodynamically by Roushenas et al. [30]. Energ etic and exerg etic efficiencies were 58% and 78%, respectively [30]. The thermodynamic performance of an integrated wind -powered CAES system coupled with an absorption refrigeration cycle and an ORC for the production of cooling, electricity, and heating is investigated by Mohammadi et al. [31]. The energy efficiency of their sy stem was 53.9% [31]. A comparison of selected past s tudies is presented in Table 1. Table 1 A comparison of past studies on the CAES systems. Study Year Renewable energy source System outputs Energy efficiency (%) Exergy efficiency (%) Reference Wang et al. 2019 Marine energy Power 58 58 [24] Wang et al. 2019 - Power, heating, and cooling 52-59.6 - [25] Hao et al. 2020 - Power 58.4 67.9 [26] Ji et al. 2017 Solar and wind Power and hot water 61.2 65.4 [27] Mohammadi et al. 2016 Solar Power 76.5 53.4 [28] Yang et al. 2017 Solar Power, heating, and cooling 6-39 - [29] Roushenas et al. 2020 - Power and hot water 58 78 [30] Mohammadi et al. 2017 Wind Power and hot water 53.9 - [31] As mentioned earlier, there have been afforts to improve the performance of the conventional CAES systems by utilizing modified configurations or integrat ing with other thermodynamic cycles and renewable energ y-based systems . However , coupling of CAES with liquefied air power cycle, utilization of LNG regasification waste heat, utilization of thermoelectric generators (TEG) , and utilization of solar power tower s has not been studided in the past . The main objective of th e current study is to conduct a ther modynamic analysis using energy and exergy methods for a novel integrated CAES system that uses solar energy, LNG regasification waste heat, air power cycle, and TEG for the cogeneration of cooling and electricity. A high - temperature carbonate molten salt is used as the working fluid of the solar power tower unit to heat the air before entering the expanders in the discharging phase of the system. Natural gas is also used in this hybrid system to increase the system performance and power yield. The waste he at of the exhaust gas is used to drive an air power cycle and a TEG unit. A parametric study is carried out to investigate the effects of varying major design parameters on the system performance. The results presented in this paper enable types, locations , and magnitudes of losses to be determined to reduce sources of inefficiency in the proposed system. The main features of the paper are as follows : • A cogeneration CAES system is proposed for the production of electricity and cooling. • Solar and natural gas are used in the discharging phase of the system to increase the system performance and sustainability. • High -temperature molten salt is utilized as the heat transfer fluid and thermal energy storage medium of the solar power tower unit. • The waste heat of the exhaust gas is used to drive an air power cycle and TEG unit. • The waste heat of the LNG regasification unit is used to improve the performance of the integrated system. • A thermal oil loop is utilized to capture the heat of the compression process and u se it in the discharge phase of the plant. 2. Proposed System Description Schematic of the diabatic solar -assisted CAES system is presented in Fig. 1. The integrated system comprises an air compression unit, thermal oil loop, solar power tower with thermal energy storage , gas turbine with a combustor , thermoelectric generator , liquefied air power cycle , and LNG regasification unit. During charging mode , the air is pressurized by a double -stage air compressor to store the energy . Thermal oil leaving the cold tank is used in the intercooler and aftercooler units to recover the heat produced during the compression process . The recovered heat is then used during the discharging mode to increase the cycle performance . Compressed air is stored in the salt cavern at near ambient temperature. The heated thermal oil is stored in the hot tank. Base on past studies, a pressure regulating valve at the outlet of the salt c avern was not considered in this paper [28,29,31]. During the discharging mode , the air flowing out of the cavern is firstly heated up in heat er 1 by the thermal oil, and then in heater 2 by the high -temperature molten salt before entering the first stage of the gas turbine unit to generate power. High -temperature molten sa lt, which is a eutectic mixture of Li2CO 3(32.1%) -Na2CO 3(33.4%) -K2CO 3(34.5%) , is used for thermal energy storage and als o, as the heat transfer fluid in the solar power tower [32]. After the first expansion process, the air enters the combustion chamber unit where natural gas is used to increase the temperature of the air before the next expansion process. For the base case, exhaust gas es with a temperature of 1100 ℃ enter the second stage of the gas turbine unit to produce power. Exhaust gases of the second stage pass through the heater 3 to provide the required heat of the liquefied air power cycle . As the temperature of the flue gas leaving the heater 3 is still high, a thermo -electric generator is utilized to recover the remaining thermal energy of the flue gas before discharging it to the environment at 150 ℃. In the liquefied air power cycle , air is pressurized by the pump and enters two sequential heat exchangers where it is super heated before entering the turbine to generate power . The expanded air exits the turbine at a lower temperature ( 455 ℃ for the base case) and enters the recuperator where rejects heat to the pressurized cold air. The air at the exit of the recuperator , state 37 is a very low temperature of -109.6 ℃. At the same time, LNG exiting the tank is pressurized by the LNG pump for the regasification process . The temperature of the liquefied air power cycle condenser is a function of LNG temperature at the outlet of the pump. For the LNG regasification process, the natural gas must be heated up before entering the expander to increase the power yield of the regasification process . Therefore, LNG then goes through heater 4 to be heated up by the thermal oil. It then enters heater 5 which produce s cooling before entering the natural gas expander. The natural gas leave s the system to be transported by pipeline. Figure 1 Schematic of the solar -assisted CAES system. 2. Methodology and Analysis As mentioned earlier, energy and exergy approaches are used for the thermodynamic analysis in this study [3 3]. MATLAB software integrated with REFPROP 9 library is utilized to develop a thermodynamic model used for the integrated system. The conservations of mass and energy for the st eady -state condition are presented respectively as follows: ∑ṁi−∑ṁe=0 (1) Q̇−Ẇ=∑(ṁh)e−∑(ṁh)i (2) Exergy rate balance at steady -state is expressed by: EẋW−EẋQ=∑(ṁexf)i−∑(ṁexf)e−EẋD (3) where EẋW and EẋQ are rates of exergy transfer accompanying heat transfer and work, respectively, and exf is the total flow exergy: EẋQ=(1−T0 Ti)×Qi̇ (4) EẋW=Ẇ (5) exf=exPH+exCH+exKE+exPT (6) where exPH, exCH, exKE, and exPT denote physical, chemical, kinetic, and potential exergies. Apart from the common mass, energy, and exergy balance equations a series of other equations are solved to complete the thermodynamic modeling of this study. The emitted heat from the sun is eval uated as follows [ 34]: Q̇s=A×n×I (7) Here, A, n, and I represent the aperture area, number of mirrors, and irradiance, respectively. A portion of the heat from the sun is reflected by the mirrors while the remaining heat is wasted due to the optical efficiency of the mirrors [ 34]: Q̇h=Q̇s×ηh (8) where ηh is the optical efficiency of the heliostat mirrors. The rate of absorbed heat by the high -temperature molten salt is evaluated by calculating the heat losses associated with the receiver towe r: Q̇a=Q̇h−Q̇l,total =Q̇h−Q̇l,cond −Q̇l,conv −Q̇l,e−Q̇l,r (9) where Q̇l,total, Q̇l,cond, Q̇l,conv, Q̇l,e, and Q̇l,r denote total heat loss of the receiver, conduction heat loss, convection heat loss, emission heat loss, and radiation heat loss, respectively. Considering charging and discharging modes, e ach system component is studied using the first and second law s of thermodynamics. The reaction that takes place in the combustion chamber is given by [17]: λ̅Cx1Hy1+(xN2N2+xO2O2+xArAr) →yCO 2CO 2+yN2N2+yO2O2+yH2OH2O+yArAr (10) Energy balance on the combustion chamber is expressed as follows : ṅAirh̅Air+ṅFuelLHV FuelMFuel =ṅProduct h̅Product +(1−ηCC)LHV FuelMFuel (11) where LHV Fuel represent s a lower heating value of the fuel . The TEG power output is calculated using the following equations [35]: ẆTEG =ηTEG/Q̇TEG (11) ηTEG =ηCarnot ×√1+ZM−1 √1+ZM+TL TH (12) TM=1 2×(TL+TH) (13) Q̇TEG =ṁ15(h16−h15) (14) where ηCarnot , ZM, TL, and TH are the Carnot efficiency, figure of merit, temperature of cold plate, and temperature of hot plate of the TEG unit. The power consumed during charging , and also, the power produced during discharging mode are evaluated as follows: Ẇcharge =ẆAC,1+ẆAC,2 (15) Ẇdischarge =ẆGT,1+ẆGT,2+ẆTEG +ẆLAPC ,t+ẆNG,t−ẆLAPC ,p−ẆLNG ,p (16) where subscript AC, GT, TEG, NG, LAPC, t, and p denote air compressor, gas turbine, thermoelectric generator, natural gas, liquefied air power cycle, turbine, and pump, respectively. Cooling load produced during discharge is given as follows : QC=ṁ45×(h45−h44) (17) Energy and exergy efficiencies of the integrated system are presented as follows : ηI=(Ẇdischarge ×td)+(QC×td) (Ẇcharge ×tc)+(ṁ38×(h43−h38))×td+(Q̇s×td)+(ṁ10×LHV ×td) (18) ηII =1−(Eẋd,total ,charge ×tc)+(Eẋd,total ,discharge ×td) (Ẇcharge ×tc)+(ṁ38×(exf,38−exf,43))×td+Eẋs×td+ṁ10×exf,10×td (19) Normalized carbon dioxide emission is determined by: ε=ṁCO 2 Ẇnet (20) Here, ṁCO 2 is the mass flow rate of carbon dioxide determined from the following equation: ṁCO 2=λ̅×ṁ9×MCO 2 Mair (21) where MCO 2 and Mair denote the molar mass of carbon dioxide and air, respectively. Also, λ̅ is the ratio of fuel to air on a molar basis [17]: λ̅=nFuel nAir (22) Major i nput parameters are presented in Table 2 [23,3 4-36]. Table 2 Input parameters for the base case of this study. Parameter Value Mass flow rate of air entering the system (kg/s) 100 Charging time (h) 8 Discharg ing time (h) 8 Pressure ratio of each compression/expansion stage 6 Pinch point ( ℃) 10 Area of a single mirror (m2) 100 Number of mirrors 400 Optical efficiency (%) 75 Wind velocity (m/s) 5 Molten salt type (LiNaK) 2CO 3 Inlet temperature of molten salt ( ℃) 400 Outlet temperature of molten salt ( ℃) 700 Outlet temperature of combustor ( ℃) 1100 TEG figure of merit 0.8 Liquefied air power cycle turbine inlet pressure (kPa) 15000 Pressure of LNG tank (kPa) 130 Pressure of LNG pump outlet (kPa) 10000 Temperature of liquefied air power cycle condenser ( ℃) T39 +pinch point NG turbine outlet pressure (kPa) 7000 Isentropic efficiency of air compressors (%) 0.8 Isentropic efficiency of gas turbines (%) 0.86 Isentropic efficiency of liquefied air power cycle and NG turbines (%) 0.8 Isentropic efficiency of liquefied air power cycle and LNG pumps (%) 0.75 Energy and exergy balance equations for each component present in the proposed energy storage system are presented in Table 3. Table 3 Energy and exergy balance equations for each component of the integrated system Component Energy balance equation Exergy balance equation Air compressor I WAC,1=(ṁ2h2−ṁ1h1)×tc ExD,AC,1=(ṁ1ex1−ṁ2ex2+ẆAC,1)×tc Air compressor II WAC,2=(ṁ4h4−ṁ3h3)×tc ExD,AC,2=(ṁ3ex3−ṁ4ex4+ẆAC,2)×tc Intercooler (ṁ2h2+ṁ18h18)×tc =(ṁ3h3 +ṁ20h20)×tc ExD,IC=(ṁ2ex2+ṁ18ex18−ṁ3ex3 −ṁ20ex20)×tc Aftercooler (ṁ4h4+ṁ19h19)×tc =(ṁ5h5 +ṁ21h21)×tc ExD,AC=(ṁ4ex4+ṁ19ex19−ṁ5ex5 −ṁ21ex21)×tc Heater I (ṁ6h6+ṁ23h23)×td =(ṁ7h7 +ṁ24h24)×td ExD,H,1=(ṁ6ex6+ṁ23ex23−ṁ7ex7 −ṁ24ex24)×td Heater II (ṁ7h7+ṁ30h30)×td =(ṁ8h8 +ṁ31h31)×td ExD,H,2=(ṁ7ex7+ṁ30ex30−ṁ8ex8 −ṁ31ex31)×td Gas turbine I WGT,1=(ṁ8h8−ṁ9h9)×td ExD,GT,1=(ṁ8ex8−ṁ9ex9−ẆGT,1)×td Gas turbine II WGT,2=(ṁ11h11−ṁ12h12)×td ExD,GT,2=(ṁ1ex1−ṁ2ex2−ẆGT,2)×td Combustion chamber (ṁ9h9+ṁ10LHV Fuel)×td =(ṁ11h11 +(1−ηCC)LHV Fuel) ×td ExD,CC=(ṁ9ex9+ṁ10ex10 −ṁ11ex11)×td Heater III (ṁ12h12+ṁ34h34)×td =(ṁ13h13 +ṁ35h35)×td ExD,H,3=(ṁ12ex12+ṁ34ex34−ṁ13ex13 −ṁ35ex35)×td LAPC turbine WLAPC ,t=(ṁ35h35−ṁ36h36)×td ExD,LAPC ,t=(ṁ35ex35−ṁ36ex36 −ẆLAPC ,t)×td Recuperator (ṁ33h33+ṁ36h36)×td =(ṁ34h34 +ṁ37h37)×td ExD,REC =(ṁ33ex33+ṁ36ex36−ṁ34ex34 −ṁ37ex37)×td Condenser (ṁ37h37+ṁ39h39)×td =(ṁ32h32 +ṁ40h40)×td ExD,C=(ṁ37ex37+ṁ39ex39−ṁ32ex32 −ṁ40ex40)×td LAPC pump WLAPC ,p=(ṁ33h33−ṁ32h32)×td ExD,LAPC ,p=(ṁ32ex32−ṁ33ex33 +ẆLAPC ,p)×td LNG pump WLNG ,p=(ṁ39h39−ṁ38h38)×td ExD,LNG ,p=(ṁ38ex38−ṁ39ex39 +ẆLNG ,p)×td Heater IV (ṁ26h26+ṁ40h40)×td =(ṁ27h27 +ṁ41h41)×td ExD,H,4=(ṁ26ex26+ṁ40ex40−ṁ27ex27 −ṁ41ex41)×td Heater V (ṁ41h41+ṁ44h44)×td =(ṁ42h42 +ṁ45h45)×td ExD,H,5=(ṁ41ex41+ṁ44ex44−ṁ42ex42 −ṁ45ex45)×td Natural gas turbine WNG,t=(ṁ42h42−ṁ43h43)×td ExD,NG,t=(ṁ42ex42−ṁ43ex43 −ẆNG,t)×td Heliostat mirrors Qs=(Q̇h+Q̇l,opt)×td ExD,HM=(Eẋs−Eẋh)×td Solar tower Qh=(Q̇a+Q̇l,cond +Q̇l,conv +Q̇l,e +Q̇l,r)×td ExD,ST=(Eẋh+ṁ28ex28−ṁ29ex29)×td TEG WTEG =(ṁ13h13+ṁ15h15−ṁ14h14 −ṁ16h16)×td ExD,TEG =(ṁ13ex13+ṁ15ex15−ṁ14ex14 −ṁ16ex16−ẆTEG)×td To perform the economic analysis, total revenue requirement method (TRR) is used. TRR for each year include s a return on investments, recovery of the capital, income tax, fuel costs, insurance and other taxes, and operation and maintenance costs [37]: TRR j=ROI j,d+ROI j,ce+ROI j,ps+TCR j+ITX j+FCj+OTXI j+OMC j (23) where subscripts d, ce, and ps denote debt, common equity, and preffered stock, respectively. Levelized TRR is then evaluated by [37]: TRR L=CRF ∑TRR j (1+i)jn 1 (24) where CRF is the capital recovery factor, n is the lifetime of plant, and i is the interest rate [37]. Levelized cost of el ectricity is obtained as follows [37]: LCOE =TRR L−BPV MPQ (25) Here, MPQ and BPV are main product quantity and byproduct value, respectively [37]. The ssumptions used for the economic analysis are listed in Table 4 [37]. Table 4 Economic inputs and assumptions used for economic analysis. Parameter Value Lifetime of plant (years) 30 Interest rate (%) 6 ROR of debt (%) 15 ROR common equity (%) 11.7 ROR preferred stock (%) 10 Debt fraction of total investment cost (%) 35 Common equity fraction of total investment cost (%) 15 Preferred stock of total investment cost (%) 50 O&M factor (%) 2 Tax life of plant (years) 15 Rate of income tax (%) 30 Other taxes and insurance factor (%) 2 Cost of LNG ($/GJ) 5.66 Cost of natural gas ($/GJ) 3.15 Reference economic year 2020 For evaluation of total capital investment, it is necessary to calculate the purchase equipment costs (PEC). Cost functions listed in Table 5 are used to evaluate the total PEC of the plant [37- 42]. Total PEC of the plant is around 23% of the fixed capital investment, while installation of equipments and instrumentation account for 12% of the fixed capital investment. Piping, electrical, building, yard improvement, service facilities, land, engineeri ng and supervision, expenses of construction, contractor’s fee, contingencies, and working capital investment account for 7, 4, 8, 2, 13, 1, 9, 10, 2, 9, and 17.5% of the fixed capital investment, respectively [37]. Table 5 Cost fu nctions used for the estimation of total capital invesetment. Component Cost function Heliostat mirrors Z=145 Nhelio Ahelio Solar receiver Z=1.03×108(Arec/1571 )0.7 Solar tower Z=300 ×exp (0.0113 (Htower −(Hrec−Hhelio )/2) Storage tank s Z=2.9232exp (11.662 −0.6104 (ln(V)+0.04536 (ln(V))2)) Molten salt Z=1.2m Air compressor Z=71.1ṁ 0.92−η×Rp×ln (Rp) Gas turbine Z=479 .34ṁ 0.92−η×ln (Rp)×(1+exp(0.036Ti−54.4)) Combustion chamber Z=46.08ṁair 0.995 −(Poutlet Pinlet⁄ )(1+exp(0.018 Toutlet −26.4)) Pump Z=705 .48(Ẇ)0.7(1+0.2 1−η) Heat exchangers Z=2143 (A)0.514 Condenser Z=1397 (A0.89) TEG Z=1500 (Ẇ) 4. Results and Discussion The r esults of the energy and exergy analyses of the integrated system proposed in this study are presented in Table 6. Using Eqs. 15 and 16, t he energy and exergy efficiencies are found as 55.3% and 4 6.4%, respectively. Additionally, it is observed that for producing 797.6 MWh of electricity during discharge, 63.4 tons/day of fuel is required. The charging power required for the compressors is 356 MWh. The terms of Eqs. 12 and 13 are calculated by using the energy balance equations listed in Table 3. By applying the balance equations on the two air compressor, power required for the c harging mode of the plant is calculated. Also, the cooling load produced during discharging mode is calculated by Eq. 14. Moreover, by using the exergy balance equations listed in Table 3, the overall exergy destruction of the plant is found to be 1,035.2 MWh. In the next section of results and discussion, detailed results of the exergy analysis will be presented. After determining the unknown species coefficients of the combustion process ( see Eq. 10) and the molar based fuel to air ratio ( see Eq. 19), the mass of produced carbon dioxide is calculated using Eq. 18. Finally, by knowing the net power generation of the plant, t he normalized carbon dioxide emission s are calculated as 210.1 kg CO 2/MWh (see Eq. 17). Thermodynamic properties including temperature, pressure, enthalpy, and entropy for the various streams of the system are presented in Table 7. As mentioned earlier, for evaluating the thermophysical properties, REFPROP 9 library is utilized i n this study. Also, energy balance equations listed in Table 3 are used to find unknown properties such as mass flow rate, enthalpy, temperature, or pressure. For instance, the enthalpy and temperature of state 8, are calculated from the energy balance equ ation for heater 2. Table 6 Thermodynamic modelling results for the proposed system . Parameter Value Energy efficiency (%) 55.3 Exergy efficiency (%) 46.4 Power consumption during charging (MW h) 356 Power generation during discharging (MW h) 797.6 Cooling power produced during discharging (MW h) 423.2 Overall exergy destruction (MW h) 1,035.2 Consumed fuel ( tons/day ) 63.4 Rate of consumed LNG ( tons/day) 5,539.5 Required volume of the salt cavern (m3) 100,188.3 Total capital investment (Million dollars) 248.6 Levelized cost of electricity (¢/kWh) 12.4 Normalized carbon dioxide emissions (kg CO 2/MWh) 210.1 Table 7 Thermodynamic properties of various streams in the proposed solar -assisted CAES system. Stream number Fluid State of matter Pressure (kPa) Temperature (℃) Enthalpy (kJ/kg) Entropy (kJ/kg.K ) 1 Air Gas 101.3 25.2 298.4 6.9 2 Air Gas 506.5 240.3 517 6.9 3 Air Gas 506.5 35.2 307.7 6.4 4 Air Gas 2532.5 257.6 534.3 6.5 5 Air Gas 2532.5 35.2 303.4 6 6 Air Gas 2532.5 35.2 303.4 6 7 Air Gas 2532.5 227 502.4 6.5 8 Air Gas 2532.5 448.4 737.4 6.8 9 Air Gas 506.5 227.8 504 6.9 10 Methane Gas 101.3 25.2 910 6.7 11 Flue gas Gas 506.5 1100 1484.1 8 12 Flue gas Gas 101.3 714.6 1032.6 8.1 13 Flue gas Gas 101.3 368.3 651.1 7.6 14 Flue gas Gas 101.3 150 424.7 7.2 15 Water Liquid 101.3 25.2 104.9 0.4 16 Water Liquid 101.3 30.2 125.8 0.4 17 Therminol 66 Liquid 101.3 25.2 52.8 0.2 18 Therminol 66 Liquid 101.3 25.2 52.8 0.2 19 Therminol 66 Liquid 101.3 25.2 52.8 0.2 20 Therminol 66 Liquid 101.3 228.3 479.5 1.3 21 Therminol 66 Liquid 101.3 245.6 515.8 1.3 22 Therminol 66 Liquid 101.3 237 497.8 1.3 23 Therminol 66 Liquid 101.3 237 497.8 1.3 24 Therminol 66 Liquid 101.3 47.2 99 0.3 25 Therminol 66 Liquid 101.3 237 497.8 1.3 26 Therminol 66 Liquid 101.3 141.3 296.7 0.9 27 Therminol 66 Liquid 101.3 26.2 54.9 0.2 28 Carbonate salt Liquid 101.3 400 644 1.5 29 Carbonate salt Liquid 101.3 700 1127 2 30 Carbonate salt Liquid 101.3 700 1127 2 31 Carbonate salt Liquid 101.3 400 644 1.5 32 Air Liquid 3401 -143.2 -0.6 4.1 33 Air Gas 15000 -119.6 30.7 4.2 34 Air Gas 15000 356.3 639.6 6.2 35 Air Gas 15000 704.6 1030.2 6.7 36 Air Gas 3401 455.5 745.4 6.8 37 Air Gas 3401 -109.6 136.5 5.1 38 Natural gas Liquid 130 -158.2 10.9 0.1 39 Natural gas Gas 10000 -153.2 42.4 0.2 40 Natural gas Gas 10000 -121.6 153.7 1 41 Natural gas Gas 10000 -75 348.3 2.1 42 Natural gas Gas 10000 15.2 778.8 3.9 43 Natural gas Gas 7000 -7.9 745.5 4 44 R134a Gas 101.3 25.2 424.5 1.9 45 R134a Liquid 101.3 -60 123.2 0.7 The power consumption/ generation of each system component is presented in Fig. 2. The first and second stages of the air compressor unit consume 174.9 and 181.3 MWh of electricity daily , respectively . The power produced by the first and second stages of the gas turbine are 186.7 and 369 MWh, respectively. The difference is due to the high difference in the inlet temperatures of gas entering the turbines . Inlet temperatures of the first and second stages are 448.4 and 1 ,100 ℃, respectively. Therefore, when the mass flow rate of gas remains constant, the high temperature of gas means higher enthalpy difference, and therefore, higher values of power is produced. The turbine of the liquefied air power cycle also produces 227.3 MWh of electricity which is quite significant , mainly due to the high heat addition an d low heat rejection temperatures of this cycle. The power generated by the TEG for the waste heat recovery process is calculated as 52.8 MWh daily . Moreover, the electric power consumed by the LNG and liquefied air power cycle pump s are calculated as 30.9 and 3 2.7 MWh, respectively. Also, the power produced in the natural gas expander is evaluated as 32.75 MWh per day. 050100150200250300350400 Air Compressor 1 Air Compressor 2 Gas Turbine 1 Gas Turbine 2 Thermoelectric Generator Liquefied Air Power Cycle Turbine Liquefied Air Power Cycle Pump Natural Gas Expander LNG PumpPower Consumption /Generation (MWh) Figure 2 Power consumption/generation of various components in the integrated system. Dynamic performance of the solar power tower in the integrated CAES system along the DNI of the selected location for a typical day is presented in Fig. 3. The solar irradiation starts from 6 A.M. as the sun ri ses and reach a maximum of 976.9 W/m2 at 11 A.M. The maximum solar heat input to the heliostat mirrors is evaluated as 39.1 MW. Consider the heat losses associated with the heliostats and solar receiver, the net abosrbed heat by the molten salt is 23.74 MW . Therefore, the efficiency of the solar subsystem is 60.7% at 11 A.M. Figure 3 Dynamic performance of the solar power tower subsystem along the DNI of the selected location for a typical day in year. Grid power storage and power generation of the energy storage system proposed i n this study are shown in Fig. 4. From about 2 to 9 A .M., the demand profile of the selected location, which is the Bandar Abbas city located in southern Iran, is below the hourly average power demand. Therefore, t he charging time of the plant for the grid power storage is set as 8 hours from 2 to 9 A.M. The power demand starts to rise steadily at around 7 A.M. and exceeds the average hourly demand from around 1 P.M. , and reaches its peak at around 3 P.M. The electricity demand decreas es significantly from around 9 P.M. Therefore, the discharging mode of the CAES plant is set from 1 P.M. to 9 P.M. to include the time period that the grid demand is high. Figure 4 Grid power storage and power generation of the plant . Exergy destruction of various components in the integrated system are presented in Fig. 5. As can be seen, the highest contributors to the overall exergy destruction of the system are the combustion chamber, heater 4, solar tower, recuperator , and heliostats with a value of 370.9 , 164.9 , 89.8, 81.5, and 69.7 MWh, respectively. The high exergy destruction in the combustion chamber unit is due to the significant loss of available chemical exergy of the fuel during the combustion process . The significant exergy destruction rate of heater 4 is because of the high-temperature difference between the hot and cold streams of thermal oil and natural gas flowing through this heat exchanger. Similarly, h igh temperature difference between the hot and cold streams in the recuperator is the reason for its high exergy destruction. The exergy destruction of the solar tower unit is higher compared to heliostats because of the utilization of high -temperature salt. If a lower temperature salt is utilized, the exergy destruction of the solar tower unit will reduce. 020406080100120 0 4 8 12 16 20 24Power (MW) Time (h)Grid Power Storage Net Power Generation Figure 5 Exergy destruction of various components in the integrated system. The effects of the gas turbine s/compressors pressure ratio s on the performance of the integrated system are presented in Fig. 6. By enhanc ing the pressure ratio of each stage of the compression/expansion process from 3 to 8, the energy efficiency reduces from 56.5% to 53.6%, while the daily exergy destruction increases from 1,013.1 to 1,094.2 MWh. As the pressure ratio increases, more heat is stored by the thermal oil and hence, the e xergy destructi on in heater 1 is increased . However, due to the higher temperature of the air at state 7, the exergy destruction of heater 2 reduces. On the other hand, the outlet temperature of the air at state 9 reduces as the pressure ratio increases. I t will result in more fuel consumption to achieve the fixed gas turbine inlet temperature at state 11. An increase in fuel consumption causes an aggravation of the exergy destruction rate of the integrated system. 0 50 100 150 200 250 300 350 400OthersThermoelectric GeneratorSolar TowerHeliostatsNatural Gas ExpanderHeater 5Heater 4Heater 3Heater 2Heater 1LNG PumpPower Cycle PumpCondenserRecuperatorPower Cycle TurbineCombustion ChamberGas Turbine 2Gas Turbine 1AftercoolerIntercoolerAir Compressor 2Air Compressor 1 Exergy Destruction (MWh) Figure 6 Effects of pressure ratio of each compression/expansion stage on the performance of the integrated system. The influence s of the heliostat numbers on the fuel consumption and power generation of the system are shown in Fig. 7. When the heliostat numbers a re increased from 300 to 850, the daily generated power increases from 798.1 to 876 MWh, while the daily fuel consumption is reduce d from 65.1 to 48.9 tons , respectively. This is because when the number of heliostats increases, the temperature of the air leaving heater 2 is increase d drastically (from 391 to 695) , and hence, the outlet temperature of the first stage gas turbine and its generated power are increase d. On the other hand, whe n the inlet temperature of air entering the first stage of the gas turbine increases, by considering a constant value for the pressure ratio, its outlet temperature is also increased. This means that lower values of fuel will be required to heat up the air the chosen temperature. Therefore, the fuel consumption rate in the combustion chamber is decreased . 100010201040106010801100 5353.55454.55555.55656.557 2.5 3.5 4.5 5.5 6.5 7.5 8.5 Exergy Destruction Rate (MWh)Energy Efficiency (%) Pressure RatioEnergy Efficiency Exergy Destruction Rate Figure 7 Influence of heliostat numbers on the integrated system performance. The effect of the combustor outlet temperature on the performance of the integrated system is presented in Fig. 8. When the combustion chamber outlet temperature increases from 1 ,000 to 1,400, the daily power generated by the integrated system increases from 748 to 1,016.1 MWh, and energy efficiency increases from 55.1% to 55.6% . As the combustor outlet temperature increases, the power output of the GT cycle increases. This causes an increase in the outlet temperature of the gas turbine which in turn result s in an increase in power generation of the liquefied air power cycle and TEG unit. To be more specific, with the same amount of increase in the GTIT, the power generated by the TEG, liquefied air power cycle, and GT cycle are increased from 36, 183.2 , and 527.2 MWh to 112, 260, and 642.4 MWh, respectively. 780795810825840855870885 4550556065 250 350 450 550 650 750 850 Power Generation (MWh)Daily Fuel Consumption (tons) Number of Heliostat MirrorsMass Flow Rate of Fuel Power Generation Figure 8 Influence of combustor outlet temperature on integrated system performance. The influence s of liquefied air power cycle turbine inlet pressure on the exergy destruction rate and generated power of the proposed system are presented in Fig. 9. As the inlet pressure of the liquefied air power cycle turbine increases from 5 ,000 to 18 ,000, the daily generated power increases fr om 735.9 to 821.2 MWh, while the exergy destruction of the system is reduce d from 1,348.4 to 1,015.9 MWh, respectively . As the turbine inlet pressure increases , more heat is required in the liquefied air power cycle , and therefore, the temperature of flue gas leaving heater 3 is decreased . This translates into a more efficient consumption of the exergy content of the flue gas before discharging it to the atmosphere . 70075080085090095010001050 5555.155.255.355.455.555.6 950 1050 1150 1250 1350 1450 Power Generation (MWh)Exergy Efficiency (%) Combustor Outlet Temperature ( ℃)Energy Efficiency Power Generation Figure 9 Influence of liquefied air power cycle turbine inlet pressure on integrated system performance. The effects of LNG pump outlet pressure on the power generation and normalized CO 2 emission s of the cycle are presented in Fig. 10. The system is optimized at the LNG pump outlet pressure of 12,000 kPa. The optimum results of daily power generation and normalized CO 2 emission s are 815.8 MW and 209 kg CO 2/MWh , respectively. The liquefied air power cycle condenser temperature is dependent on the temperature of natural gas at the outlet of the LNG pump. The best performance of the liquefied air power cycle, and hence the integrated system is achieved a t the LNG pump outlet pressure of 12000 kPa. 9501050115012501350 720740760780800820840 4000 6000 8000 10000 12000 14000 16000 18000 Exergy Destruction (MWh)Power Generation (MWh) Liquefied Air Power Cycle Turbine Inlet Pressure (kPa)Power Generation Exergy Destruction Rate Figure 10 Effects of LNG pump outlet pressure on the integrated system performance. 5. Comparison with past studies The energ etic and exerg etic efficienc ies of the integrated solar -assisted CAES system are compared with those presented previously in Table 8. Past studies selected for comparison includ e a trigeneration CAES plant with absorption chiller and parabolic trough collectors [ 43], an advanced CAES system integrated with organic Rankine cycle/Kalina cycle for the waste heat recovery purposes [ 44], a hybrid CAES system which uses both wind and solar energ ies [45], another hybrid CAES integrated with solar PV panels, batteries, and fuel cell [46], and an integrated CAES system with waste heat recovery and wind farm [47] . It is shown that the integrated system proposed in this study has a better performance in terms of energy and exergy. Table 8 A comparison between the present CAES system and past studies. Study Subsystems Compressors and turbine effici encies (%) Energy efficiency (%) Exergy Efficiency (%) LCOE (¢/kWh) Present study CAES, LNG regasification, Solar tower, and liquid air power cycle 80 and 86 54.9 46.9 12.4 Wang et al. [43] CAES, solar collector, gas turbine, and absorbtion chiller 86 and 88 53.1 45.4 - 208209210211212213 799.5802.5805.5808.5811.5814.5817.5 7500 8500 9500 10500 11500 12500 13500 Normalized CO2 Emissions (kg CO2/MWh)Power Generation (MWh) LNG Pump Outlet Pressure (kPa)Power Generation Normalized Carbon Dioxide Emissions Soltani et al. [44] CAES, low and high temperature energy storage, Kalina cycle, and ORC 88 and 85 50.7-51.2 44.2-44.6 - Chen et al. [45] CAES, wind farm, and solar collecto r ~90 46.5 - - Sadeghi and Askari [46] CAES, TES, fuel cell, solar PV pannels, battery, and water reservoir 82 and 86 55 - 15.6 Meng et al. [47] CAES, waste heat recovery unit, and wind farm 84 and 90 53.3 - 5-8 6. Conclusion s A novel solar -assisted integrated compressed air energy storage system is designed and analyzed in this paper . The proposed system comprise s an air compression unit, thermal oil loop, gas turbine, solar power tower with thermal energy storage , liquefied natural gas regasification unit, thermoelectric generator, and liquefied air power cycle . The performance of the energy storage system is studied through exergy, energy, and economic analyses. A parametric study is performed to examine the effect s of varying major parameters on the performance of the proposed integrated system . Major findings are presented as follows: • The energy and exergy efficiencies of the integrated system are 5 5.3% and 4 6.4%, respectively. • 356 MWh of power is consumed during charging mode and 797.6 MWh of power is generated during the discharging mode . This indicates that the energy storage system proposed here uses less than half of the power it produces during the discharging mode with the aid of solar energy and cold waste heat of liquefied natural gas . • 5,539.5 tons of LNG is required per day for the integrated plant to be operational for 8 hours of discharging. • Excess air is used in the combustion process to reduce fuel consumption and improve the performance of the energy storage system. It is observed that a daily value of 63.4 tons of fuel is requ ired during the discharging mode of the system. • Dynamic modelling of solar subsystem showed that the solar heat input to the heliostat mirrors is maximized at 11 A.M. with a value of 39.1 MW. The abosrbed heat by the molten salt at this time is 23.74 MW. T herefore, the solar subsystem efficiency is evaluated as 60.7% at 11 A.M. • The normalized carbon dioxide emission s of the system are 210.1 kg CO 2/MWh. The current system is three times less harmful to the environment compared to the carbon impact of Iranian national grid. • The highest contributors to the overall exergy destruction of the system are the combustion chamber, heater 4, solar tower, recuperator, and heliostats with values of 370.9, 164.9, 89.8, 81.5, and 69 .7 MWh, respectively . Therefore, to improve environmental sustainability , these components require further investigation in future research . • Overall exergy destruction of the system during both charging and discharging phases are calculated as 1,035.2 MW h. • Enhancing the pressure ratio of the compressors will reduce the performance of the plant, while increasing the combustion chamber outlet temperature will improve it. Also, the normalized carbon dioxide emissions of the plant are minimized at LNG pump ou tlet pressure of 12 MPa. Increasing the number of heliostat mirrors from 300 to 850 will increase the daily power generation of the plant 798.1 to 876 MWh, while the daily fuel consumption is reduced from 65.1 to 48.9 tons . • Economic analysis indicated that the levelized cost of electricity is 12.4 ¢/kWh.