Exergoeconomic and Exergoenvironmental Analysis of a Novel Power and Cooling Cogeneration System Based on Organic Rankine Cycle and Ejector Refrigeration Cycle

Abstract

A novel combined power and refrigeration system is proposed based on organic Rankine and jet refrigeration cycles. The system has a wider application range and can be adjusted to different cooling and evaporation temperatures. To meet the needs of diverse populations, the cooling and evaporation temperature can be as low as −60 degrees Celsius. The genetic algorithm is used to optimize the system, and the proposed system’s energy, exergy, economy, and environment are analyzed under optimal conditions. The results desmonstrate that the exergy damage, environmental impact rate, and exergy economic coefficient of steam turbine are the largest. The system’s exergy damage and the turbine’s investment cost are reduced, and the system’s performance is improved. The condenser has the greatest potential for improvement and should be considered a priority component for system improvement. In addition, the system parameters are analyzed. Higher low-pressure steam generation temperature, dryness of low-pressure steam generator outlet, turbine steam extraction ratio, refrigeration evaporation temperature, and compressor compression ratio are advantageous to system cooling capacity output but not the system net power.High-pressure evaporation temperature is unfavorable to the system’s output of net power and cooling capacity. Still, it is beneficial to improve the thermal and energy efficiency of the system. Under the same operating conditions, compared with the system proposed by predecessors, the system’s net power is increased by 12.52 kW, the thermal efficiency is increased by 4.27%, and the energy efficiency is increased by 2.57%. The system was optimized by taking low-pressure evaporation temperature, high-pressure evaporation temperature, outlet dryness of low-pressure steam generator, suction ratio of steam turbine and compression ratio of compressor as decision variables, and thermal efficiency, exergy efficiency, SUCP and SUEP as objective functions. The low-pressure evaporation temperature, high-pressure evaporation temperature, outlet dryness of low-pressure steam generator, suction ratio of steam turbine, and compression ratio of compressor are 357.99 K, 385.72 K, 0.1, and 0, respectively. The system thermal efficiency is 15.01%, exergy efficiency is 43.18%, SUCP is 45.525USD/MWh, and SUEP is 5122.6 MPTS/MWh.

1. Introduction

With today’s rapid economic expansion, energy reserves are depleting, and environmental pollution is getting worse, forcing people to find effective ways to use energy and renewable energy sources. The world’s primary energy is inappropriately converted to waste heat at about 72%, which has increased interest in waste heat recovery studies [1]. Combined cooling and power supply is a mature technology for the utilization of low-grade energy [2]. The refrigeration and generation cycle are coupled by the combined cooling and power supply system. Different cycle and coupling modes will result in variations in the system’s thermodynamic performance [3]. To recover low-temperature waste heat, ORC has reached a reasonably advanced stage. The benefits of ERC include its straightforward structure, inexpensive equipment investment, and low operating and maintenance expenses [4]. As a result, domestically and internationally, researchers have examined and built various cooling and power supply systems based on ORC and ERC. Zhang et al. evaluated different common alternative refrigerants before proposing an ORC-ERC combined cycle system that uses low-grade solar energy or other waste heat (below 150 °C) in hot and dry areas: In hot regions, they discovered that R290 (propane) and R601 (pentane) have intrinsic advantages for cooling or power generation [5]. Wang et al. proposed a new combined cooling and power supply system, which combined ORC and ERC by adding an extraction turbine between the heat recovery steam generator (HRVG) and ejector. They carried out parameter analysis and exergy loss of the system and found that the biggest exergy loss occurred in the heat recovery steam generator, followed by the ejector and turbine [6]. Zheng et al. proposed a combined power-cooling system driven by a low-temperature heat source. The system uses the exhaust of the ejector turbine to generate power and refrigeration at the same time. They found that when R245fa was used as a working medium, the steam generation temperature was 395 K, condensation temperature was 298 K and refrigeration and evaporation temperature was 280 K, and the energy efficiency, exergy efficiency, and exergy efficiency of the system could reach 34.1%, 18.7% and 56.8%.The majority of exergy loss occurred in the ejector [7] Wang et al. proposed a new combined refrigeration cycle of power and ejection absorption, which used ammonia as a working fluid. It was found through parameter analysis that the heat source, condenser, and evaporator temperature had significant effects on the net power output, cooling capacity, and exergy efficiency of the combined cycle. Yu et al. designed a new type of CCP based on dual-pressure evaporation ORC (DORC) and ERC, carried out thermodynamic and exergoeconomic analysis of the system and used a genetic algorithm to optimize the system. The results desmonstrated that the steam generator and condenser accounted for the largest proportion of overall exergy loss of the system under an optimized condition [8] Yu et al. proposed a new combined cooling and power generation system with dual pressure evaporation ORC (DORC) and ERC coupling by changing the coupling mode of the system, carried out thermodynamic and exergoeconomic analysis on the system, and evaluated the influence of critical parameters on the system performance. It is found that, compared with the parallel steam generator system, the unit cost of the system is obviously reduced, and the net output power is obviously increased [3].

Despite the fact that many researchers have studied CCP system that couples ORC and ERC, the evaporation temperature of the refrigeration cycle is higher than −5 °C, which can not provide a lower temperature cooling capacity for freezing, and the application range is small. Therefore, by changing the coupling mode of ORC and ERC, this paper proposes a CCP system that can provide both electric quantity and low-temperature cooling capacity. In addition, the researchers’ integrated cooling and power supply system based on ORC and ERC is only limited to thermodynamic analysis and exergy economic analysis. Therefore, this paper will build a thermodynamic, exergy economic, and exergy environment model of the system and analyze the system from multiple perspectives.

2. System Description

The cycle diagram of the system is shown in Figure 1. From the figure, we can observed the components of the system. Abbreviations represent the names of the components of the system, and the meanings of the abbreviations have been listed in the glossary. The system works as follows: the working fluid flows through the condenser to cool down and becomes saturated. After that, the working liquid is divided into two parts. One part is throttled by THV and enters EVA for refrigeration. After absorbing heat, the working liquid becomes saturated gas and enters the ejector as an ejector fluid after pressurizing by COM. LP first pressurized the other part, then preheated by IHE, changed into a wet steam state by endothermic in LPVG, and finally entered SEP. The saturated gas working fluid in the upper part of SEP is confluent with the exhaust of TUR and acts as the working fluid of EJ. The saturated liquid working fluid in the lower part of SEP is pressurized by HP and then enters HPVG for heat absorption. The endothermic working fluid has a certain superheat and enters TUR for work output. The working fluid after work is merged with the fluid at the EJ outlet, and the merged fluid enters IHE to preheat the fluid at the LP outlet and finally enters the condenser for heat dissipation. Thus, it is a complete cycle.

3. Mathematical Models and Assumptions

3.1. Assumptions

To simplify the simulation process, it is necessary to make some assumptions about the ejector and system before system simulation [9,10,11]:

(1)

The system is steady, regardless of heat loss and pressure drop in heat exchangers and connecting pipes.

(2)

The working medium at the outlet of the separator and refrigeration evaporator is saturated.

(3)

The compression process in the compressor and expansion process in the turbine are non-isentropic processes.

(4)

The throttle valve’s throttling process is equal enthalpy.

(5)

The refrigerant in the ejector is in a one-dimensional steady flow state, and the ejector’s working process meets the law of conservation of energy, momentum, and energy.

(6)

The ejector inlet and outlet velocities are ignored, and the primary and secondary flow process in the mixing chamber is isobaric.

(7)

Three efficiencies represent the irreversible loss of the injector: the isentropic efficiency of the nozzle and the diffusion section and the mixing section efficiency.

(8)

The pinch temperature difference in the heat absorption process is (ΔTPP) 10 K, the pinch temperature difference in the heat transfer process of IHE, the condensation process of condenser and the heat absorption process of the evaporator is 5 K, and the superheat (Tsup) of the steam entering the turbine is 5 K [11,12].

3.2. Energy Model of the System

The ejector is the core connecting part of the system, which combines the power and refrigeration cycles of the system. Ejectors have no moving parts and no energy loss, making them safer, environmentally friendly and energy efficient than compressors. This paper’s ejector is modeled based on the one-dimensional isobaric mixture model theory. Many existing studies demonstrate that the model accurately predicts ejector performance [13,14]. The calculation process is as follows:

$${u_{11}}^{\prime }=\sqrt{2{\eta }_n(h_{11}-h_{11s})/1000}$$

(1)

According to the conservation of energy and momentum, the velocity and enthalpy at the exit of the mixing chamber can be calculated using the mixing efficiency of the ejector mixing section.

$$u_m=\frac{{u_{11}}^{\prime }}{1+\mu }\sqrt{{\eta }_m}$$

(2)

$$h_m=(h_{11}+\mu \ast h_{14})/(1+\mu )$$

(3)

Ignoring the fluid velocity at the outlet of the diffusion section, the fluid enthalpy value at the outlet of the diffuser is:

$$h_{15}=h_m+({u_m}^2/2000)$$

(4)

By applying the isentropic efficiency η_d of the diffuser, the ideal specific enthalpy at the diffuser outlet can be obtained as follows:

$$h_{15s}=h_{15}+(h_{15}-h_m){\eta }_d$$

(5)

When the system is in a steady flow, all components in the cycle follow the conservation of mass, working medium components, and energy. When the system is in a constant flow, kinetic and potential energy changes can be ignored. According to the conservation law of mass, energy, and working medium components, the mathematical model of each element of the system are as follows [15]:

$${\displaystyle \sum m_{in}={\displaystyle \sum m_{out}}}$$

(6)

$${\displaystyle \sum \stackrel{·}{Q}}-{\displaystyle \sum \stackrel{·}{W}}+{\displaystyle \sum {\stackrel{·}{m}}_{in}{h}_{in}-{\displaystyle \sum {\stackrel{·}{m}}_{out}{h}_{out}=0}}$$

(7)

The equation of exergic balance, exergic efficiency and exergic loss improvement potential of each component of the system is as follows [3]:

$${\stackrel{·}{E}}_D={\stackrel{·}{E}}_{in}-{\stackrel{·}{E}}_{out}$$

(8)

$$\stackrel{·}{E}=\stackrel{·}{m}\left[h-h_o-T_o(s-s_o)\right]$$

(9)

$${\eta }_e=\frac{\stackrel{·}{E_{out}}}{\stackrel{·}{E_{in}}}=1-\frac{{\stackrel{·}{E}}_D}{{\stackrel{·}{E}}_{in}}$$

(10)

$$ELIP=(1-{\eta }_e){\stackrel{·}{E}}_D$$

(11)

The net output work of turbine can be expressed as:

$$Wt={\stackrel{·}{m}}_{{}_{10}}(h8-h10)+{\stackrel{·}{m}}_9(h8-h9)$$

(12)

Calculation formula of working medium pump power consumption:

$$Wp={\stackrel{·}{m}}_2(h2-h1)+{\stackrel{·}{m}}_6(h7-h6)$$

(13)

The power consumption of the compressor is described as:

$$Wc=\stackrel{·}{m_{13}}(h_{14}-h_{13})$$

(14)

The net work of the system is calculated as follows:

$$Wnet=Wt-Wp-Wc$$

(15)

Exergy of cooling capacity and cooling capacity can be calculated as follows:

$$Qc={\stackrel{·}{m}}_{12}(h_{13}-h_{12})$$

(16)

$$Ec={\stackrel{·}{m}}_{12}(h_{13}-h_{12})(\frac{To}{Tevap}-1)$$

(17)

System external input system heat:

$$Qg={\stackrel{·}{m}}_{22}(h_{22}-h_{24})$$

(18)

$$Eg={\stackrel{·}{m}}_{22}(E_{22}-E_{24})$$

(19)

The thermal efficiency of the system is:

$${\eta }_t{}_h=\frac{Wnet+Qc}{Qg}$$

(20)

The exergy efficiency of system is as follows:

$${\eta }_{ex}=\frac{Wnet+Ec}{Eg}$$

(21)

3.3. Exergoeconomic Model

For a thermodynamic system, it should be determined that the proposed system is cost-effective before its application in industry. We often need to know the cost of inefficiency in the system. Knowing this information about costs can be very beneficial in improving the system’s cost-effectiveness, that is, in reducing the cost of the final product that the system produces. Exergy economic assessment is used to evaluate this system from both technical and economic perspectives. Exergy economic evaluation is a good way to evaluate the cost and efficiency of a thermal system [16]. Exergy balance and exergy financial balance equation combined with Equations (22)–(24) [17,18]. can be used to calculate the exergy carried by each working medium flow of the system, fuel exergy rate of each equipment, exergy rate of products and capital investment cost rate.

$$\stackrel{·}{E{x}_{f,k}}=\stackrel{·}{E{x}_{p,k}}\stackrel{·}{+E{x}_{L,k}}\stackrel{·}{+E{x}_{D,k}}$$

(22)

$$C_{p,k}=C_{F,k}-C_{L,k}+{\stackrel{·}{Z}}_k$$

(23)

$$c_{p,k}{\stackrel{·}{Ex}}_{p,k}=c_{F,k}{\stackrel{·}{Ex}}_{p,k}+(c_{F,k}{\stackrel{·}{Ex}}_{L,k}-C_{L,k})+\stackrel{·}{Z_k}+c_{F,k}{\stackrel{·}{Ex}}_{D,k}$$

(24)

In the above expression, ${\dot{C}}_{p,k}$ ($/h) and ${\dot{C}}_{F,k}$ ($/h) refers to the product cost rate and fuel cost rate of each equipment, $\dot{Z_k}$ ($/h) refers to the equipment cost rate, while $c_{p,k}$ ($/MWh) and $c_{F,k}$ ($/MWh) refer to the average unit cost of product and fuel, respectively.

The investment cost equation of each piece of equipment in the system is shown in

Table 1. The investment cost equation of equipment.

Equipment	PEC Functions ($)
VG	$Z_{VG}=\mathrm{17,500}(\frac{A_{VG}}{100}{)}^{0.6}$
Condenser	$Zcon=8000(\frac{A_{condenser}}{100}{)}^{0.6}$
IHE	$Z_{IHE}=\mathrm{12,000}(\frac{A_{IHE}}{100}{)}^{0.6}$
Evaporator	$Z_{EVA}=\mathrm{16,000}(\frac{A_{EVA}}{100}{)}^{0.6}$
Turibe	$Z_{turibe}=4405{\stackrel{·}{W}}_{turibe}^{0.7}$
Pump	$Z_{Pump}=2100{\left(\frac{\stackrel{·}{W_{pump}}}{10}\right)}^{0.26}{\left(\frac{1-{\eta }_{is,pump}}{{\eta }_{is,pump}}\right)}^{0.5}$
THV	$Z_{THV}=114.5\stackrel{·}{m_{thv}}$
Compressor	$Z_{COM}=71.7{\stackrel{·}{m}}_{com}\left(\frac{1}{0.92-{\eta }_{is,com}}\right)(\frac{p_o}{p_i})\ln (\frac{p_o}{p_i})$

. The purchase cost (PEC) of the system equipment listed in Table 1 was proposed before this study. To calculate the equipment purchase cost in this study year, it is required to multiply the proportion of a chemical factory cost index. The PEC of the reference year can be converted by using the inflation index—Chemical Plant Cost Index (CEPCI) [19,20]:

$$PE{C}_{reference}=PE{C}_{orign}\times (\frac{CEPC{I}_{reference}}{CEPC{I}_{orign}})$$

(25)

$$CRF=\frac{i{(1+i)}^n}{{(1+i)}^n-1}$$

(26)

The above equation represents the capital recovery coefficient since the initial investment that can be recovered at the end of the useful life of the equipment is calculated, where i represents the interest rate (assumed to be 15%) and n represents the system life (supposed to be 20 years) [21].

$${\stackrel{·}{Z}}_k=\left(\frac{Z_k\times CRF\times \Phi }{N\times 3600}\right)$$

(27)

The investment cost rate of each piece of equipment is calculated by the above equation, where Z_k, CRF, N and $\Phi$ respectively represent the purchase cost, capital recovery coefficient (CRF), annual working hours (7000 h in this study) and maintenance coefficient (assumed to be 1.06) of the k_th component [19]. To make a more comprehensive evaluation of the system from the perspective of thermal economy, three parameters, exergy loss cost rate ${\dot{C}}_{D,k}$, relative cost difference and exergy economic factor $f_k$, are introduced, and their calculation formula is [22,23,24]:

$${\stackrel{·}{C}}_{D,k}=c_{Fk}{\stackrel{·}{E}}_{D,k}$$

(28)

$$r_k=\frac{c_{p,k}-c_{F,k}}{c_{p,k}}$$

(29)

$$f_k=\frac{\stackrel{·}{Z_k}}{{\stackrel{·}{Z}}_k+{\stackrel{·}{C}}_{D,k}}$$

(30)

For the whole system, the sum unit cost of products can be expressed as [3]:

$$SUCP=\frac{{\displaystyle \sum \stackrel{·}{Z_K}}+{\stackrel{·}{C}}_{22}}{Wnet+Ec}$$

(31)

In order to solve the exergy cost of each flow, Equation (22) is deformed to obtain the balance equation of the exergy cost of parts in a general form. By applying Equation (30) and the auxiliary equation of each component, a set of linear equations [25]. can be written out. Exergy cost balance equations of each component in the system and supplementary equations for solving exergy cost balance equations are listed in

Table 2. Exergy balance equation and supplementary equation of each component in the system.

Component	Cost Balance Equation	Supplementary Equation
Condenser	${\stackrel{·}{C}}_1+{\stackrel{·}{C}}_{19}={\stackrel{·}{C}}_{17}+{\stackrel{·}{C}}_{18}+{\stackrel{·}{Z}}_{condenser}$	$c_{18}=0,c_1=c_{17}$
Lppump	${\stackrel{·}{C}}_2={\stackrel{·}{C}}_{1^{\prime }}+{\stackrel{·}{C}}_{w,lppump}+{\stackrel{·}{Z}}_{lppump}$	$c_{w,lppump}=c_{w,turibe},c_1=c_{1^{\prime }}$
IHE	${\stackrel{·}{C}}_3+{\stackrel{·}{C}}_{17}={\stackrel{·}{C}}_2+{\stackrel{·}{C}}_{16}+{\stackrel{·}{Z}}_{IHE}$	$c_{16}=c_{17}$
LPVG	${\stackrel{·}{C}}_4+{\stackrel{·}{C}}_{24}={\stackrel{·}{C}}_3+{\stackrel{·}{C}}_{23}+{\stackrel{·}{Z}}_{LPVG}$	$c_{23}=c_{24}=4.68 \text{\$}/\mathrm{MWh}$
Separator	${\stackrel{·}{C}}_5+{\stackrel{·}{C}}_6={\stackrel{·}{C}}_4+{\stackrel{·}{Z}}_{separator}$	$c_4=x_4{c}_5+(1-x_4)c_6,{\stackrel{·}{Z}}_{seprator}=0$
Hppump	${\stackrel{·}{C}}_7={\stackrel{·}{C}}_6+{\stackrel{·}{C}}_{w,hppump}+{\stackrel{·}{Z}}_{hppump}$	$c_{w,hppump}=c_{w,turibe}$
HPVG	${\stackrel{·}{C}}_8+{\stackrel{·}{C}}_{23}={\stackrel{·}{C}}_7+{\stackrel{·}{C}}_{22}+{\stackrel{·}{Z}}_{HPVG}$	$c_{23}=c_{24}=4.68 \text{\$}/\mathrm{MWh}$
Turibe	${\stackrel{·}{C}}_9+{\stackrel{·}{C}}_{10}+{\stackrel{·}{C}}_{w,turibe}={\stackrel{·}{C}}_8+{\stackrel{·}{Z}}_{turibe}$	$c_9=c_8=c_{10}$
THV	${\stackrel{·}{C}}_{12}={\stackrel{·}{C}}_{1^{″}}+{\stackrel{·}{Z}}_{THV}$	$c_1=c_{1^{″}}$
Evaporator	${\stackrel{·}{C}}_{13}+{\stackrel{·}{C}}_{21}={\stackrel{·}{C}}_{12}+{\stackrel{·}{C}}_{20}+{\stackrel{·}{Z}}_{EVA}$	$c_{12}=c_{13},c_{20}=c_{21}$
Compressor	${\stackrel{·}{C}}_{14}={\stackrel{·}{C}}_{13}+{\stackrel{·}{C}}_{w,compressor}+{\stackrel{·}{Z}}_{COM}$	$c_{w,compressor}=c_{w,turibe}$
Ejector	${\stackrel{·}{C}}_{15}={\stackrel{·}{C}}_{11}+{\stackrel{·}{C}}_{14}+{\stackrel{·}{Z}}_{Ejector}$	$c_{11}=(x_4{c}_5+x_{10}{c}_{10})/(x_4+x_{10}),{\stackrel{·}{Z}}_{ejector}=0$

, and the expression of the linear equation is as follows:

$${\displaystyle \sum _{out}{\stackrel{·}{C}}_{out,k}}+{\stackrel{·}{C}}_{w,k}={\stackrel{·}{C}}_{q,k}+{\displaystyle \sum _{in}{\stackrel{·}{C}}_{in,k}}+\stackrel{·}{Z_k}$$

(32)

$$\left[{\stackrel{·}{E}}_{x,k}\right]\times \left[c_k\right]=\left[\stackrel{·}{Z_k}\right]$$

(33)

3.4. Exergoenvironmental Model

Exergy environment analysis of the system requires knowing the environmental impact of all components and products in the system in its life cycle. Life cycle assessment is a good method. The life cycle of each element in the system includes the acquisition, production, operation, and final disposal and recovery of materials. There are a variety of guidelines to follow when determining the environmental impact of a product. Many evaluation approaches have been developed to assess the environmental impact and performance of energy conversion and production systems [26]. In this paper, Eco-Indicator 99 method is used for the environmental analysis life cycle assessment. The life cycle impact assessment method can be found in [27]. This methodology assesses five life cycle assessment stages of the components: (i) materials used in manufacturing (per kilogram of materials), and (ii) the manufacturing process-the handling and processing of various materials. (iii) the transportation process of materials, components, fuel (per unit ton-km), (iv) the energy required for a production such as electrical and thermal loads, and (v) the disposal of materials and waste (per kilogram of materials).

The environmental impact of each device is represented by “ecopoints” (1000 mpts = 1 pts), and the ecopoints of each device are obtained by weighting the above results. One of these points is equivalent to 1/1000 of a European resident’s average annual environmental impact [28].

Table 3. Eco-indicator for materials for system equipments.

Equipment	Material	Proportion of Materials	Eco-Indicator mpts/kg	Points mpts/kg
Condenser	Steel	100%	86	86
Pump	Steel Cast iron	35% 65%	86 240	186
Generator	Steel Cast iron Copper Aluminum	20% 60% 15% 5%	86 240 1400 780	410
Turbine	Steel Steel high alloy	25% 75%	86 240	202
Evaporator	Steel	100%	86	86
Compressor	Steel Steel low alloy Cast iron	33% 45% 22%	86 110 240	131

presents the ecological indicators of the materials of the system components, and

Table 4. Environmental impact data related to equipment in life cycle assessment.

Component	Material mpts/kg	Process mpts/kg	Disposal mpts/kg	Weight Function
Condenser	86	12.1	−70	$Wcon=0.073\times {\stackrel{·}{Q}}^{0.99}$
Pump	410	12.1	−70	$Wp=0.175\times \ln (W)-1.06$
Generator	202	16.9	−70	$W_G=2.340\times {\stackrel{·}{Q}}^{1.15}$
Turbine	86	11.7	−70	$W_{Tur}=4.90\times W^{0.73}$
Evaporator	131	12.1	−70	$Wcon=0.073\times {\stackrel{·}{\mathrm{Q}}}^{0.7}$
Compressor	131	11.7	−70	$\mathrm{Equations} (33) \mathrm{and} (34)$

lists the data related to the life cycle of the system components.

Due to the lack of available data on compressor weight, the following equation was used to assess the compressor weight from the mass flow rate of fluid in the component and the rupture pressure in the cylindrical steel shell [29].

$$\stackrel{·}{m}=\rho vA$$

(34)

$$t=\frac{P·D·FS}{2\sigma }$$

(35)

In the above formula, $\rho$(kg/m³), v (m/s), and A (m²) are the density, the velocity of the working medium and the cross-sectional area of the compressor inlet, respectively. FS is the safety factor, whose value is 2.0, and σ is the fracture stress, whose value is 21 Mpa.

The balance equation of the environmental impact of the system components is as follows:

$${\stackrel{·}{B}}_{p,k}={\stackrel{·}{B}}_{F,k}+\stackrel{·}{Y_k}$$

(36)

$$b_{p,k}{\stackrel{·}{E}}_{p,k}=b_{F,k}{\stackrel{·}{E}}_{F,k}+\stackrel{·}{Y_k}$$

(37)

where ${\dot{B}}_{p,k }$ and ${\dot{ B}}_{F,k}$ are the environmental impact rate associated with product and fuel, respectively, while $b_{p,k}$ and $b_{F,k}$ are the environmental impact corresponding to the unit exergy of product and fuel.

The environmental impact of each system component $\dot{ Y}$ considers the entire life cycle of the component, and the sum of the following parameters calculates the environmental impact of each component:

$$\stackrel{·}{Y_k}=\stackrel{·}{Y_k{}^{CO}}+\stackrel{·}{Y_k{}^{OM}}+\stackrel{·}{Y_k{}^{DI}}$$

(38)

where ${\dot{Y}}_k^{CO}$ is the construction environmental impact related to manufacturing, transportation, installation and other processes; ${\dot{Y}}_k^{OM}$ is the environmental impact associated with operation and maintenance; ${\dot{Y}}_k^{DI}$ is the environmental impact associated with the disposal and recovery of equipment, and the above environmental impact data for each system component are listed in Table 4.

Environmental impacts related to equipment damage are as follows:

$${\stackrel{·}{B}}_{D,k}=b_{F,k}{\stackrel{·}{E}}_{D,k}$$

(39)

In order to obtain the improvement potential of environmental impact of each equipment, the difference of relative environmental impact of parameter $r_b$ is introduced; to determine which environmental impact is most important for each component, an environmental impact factor $f_b$ is introduced. The two indicators are calculated as follows:

$$r_{b,k}=\frac{b_{p,k}-b_{F,k}}{b_{F,k}}$$

(40)

$$f_{b,k}=\frac{\stackrel{·}{Y}}{\stackrel{·}{Y}+{\stackrel{·}{B}}_{D,K}}$$

(41)

The environmental impact of the total unit product of the system is represented by the following equation:

$$SUEP=\frac{{\displaystyle \sum {\stackrel{·}{Y}}_k+b_w}\ast (W_P+W_{com})}{Wnet+Ec}$$

(42)

Similar to the exergoeconomic analysis, a set of linear equations [30]. can be written by applying Equation (35) and auxiliary equation of each component to solve the environmental impact of each exergy flow. Overall exergy environment impact balance equation of each component in the system and supplementary equations for solving exergy environment balance equations are listed in

Table 5. Exergy environmental impact balance equation and supplementary equation of each component.

Component	Environmental Impact Balance Equation	Supplementary Equation
Condenser	${\stackrel{·}{B}}_1+{\stackrel{·}{B}}_{19}={\stackrel{·}{B}}_{17}+{\stackrel{·}{B}}_{18}+{\stackrel{·}{Y}}_{condenser}$	$b_{18}=0,b_1=b_{17}$
Lppump	${\stackrel{·}{B}}_2={\stackrel{·}{B}}_{1^{\prime }}+{\stackrel{·}{B}}_{w,lppump}+{\stackrel{·}{Y}}_{lppump}$	$b_{w,lppump}=6206\frac{mpts}{GJ}$
IHE	${\stackrel{·}{B}}_3+{\stackrel{·}{B}}_{17}={\stackrel{·}{B}}_2+{\stackrel{·}{B}}_{16}+{\stackrel{·}{Y}}_{IHE}$	$b_{16}=b_{17},b_2=b_3$
LPVG	${\stackrel{·}{B}}_4+{\stackrel{·}{B}}_{24}={\stackrel{·}{B}}_3+{\stackrel{·}{B}}_{23}+{\stackrel{·}{Y}}_{LPVG}$	$b_{23}=b_{24},b_3=b_4$
Separator	${\stackrel{·}{B}}_5+{\stackrel{·}{B}}_6={\stackrel{·}{B}}_4$	$None$
Hppump	${\stackrel{·}{B}}_7={\stackrel{·}{B}}_6+{\stackrel{·}{B}}_{w,hppump}+{\stackrel{·}{Y}}_{hppump}$	$b_{w,hppump}=b_{w,lppump}$
HPVG	${\stackrel{·}{B}}_8+{\stackrel{·}{B}}_{23}={\stackrel{·}{B}}_7+{\stackrel{·}{B}}_{22}+{\stackrel{·}{Y}}_{HPVG}$	$b_{23}=b_{24},b_7=b_8$
Turbine	${\stackrel{·}{B}}_9+{\stackrel{·}{B}}_{10}+{\stackrel{·}{B}}_{w,turibe}={\stackrel{·}{B}}_8+{\stackrel{·}{Y}}_{turibe}$	$b_9=b_8=b_{10,}{b}_{w,turibe}=b_{w,hppump}$
THV	${\stackrel{·}{B}}_{12}={\stackrel{·}{B}}_{1^{″}}$	$None$
Evaporator	${\stackrel{·}{B}}_{13}+{\stackrel{·}{B}}_{21}={\stackrel{·}{B}}_{12}+{\stackrel{·}{B}}_{20}+{\stackrel{·}{Y}}_{EVA}$	$b_{12}=b_{13},b_{20}=b_{21}$
Compressor	${\stackrel{·}{B}}_{14}={\stackrel{·}{B}}_{13}+{\stackrel{·}{B}}_{w,compressor}+{\stackrel{·}{Y}}_{COM}$	$b_{w,compressor}=b_{w,turibe}$
Ejector	${\stackrel{·}{B}}_{15}={\stackrel{·}{B}}_{11}+{\stackrel{·}{B}}_{14}+{\stackrel{·}{Y}}_{Ejector}$	$b_{11}=(x_4{b}_5+x_{10}{b}_{10})/(x_4+x_{10}),{\stackrel{·}{Y}}_{ejector}=0$

, and the expressions of the linear equations are as follows:

$$\left[{\stackrel{·}{E}}_{x,k}\right]\times \left[b_k\right]=\left[\stackrel{·}{Y_k}\right]$$

(43)

4. Model Validation

To verify whether the system energy model established in this paper is accurate, a verification program is written in MATLAB to compare the calculation results with the results of references. The accuracy of the energy model of the dual pressure ORC system in this paper is verified by comparing the calculation results of this paper with those of Manente et al. [31]. under the same internal and external conditions.

Table 6. Validation of dual pressure ORC systems model.

Title 1	Present Work	Reference	Error (%)
PHP (bar)	6.65	6.66	0.15
PLP (bar)	4.14	4.13	−0.24
mhp (kg/s)	43.71	43.7	−0.02
mlp (kg/s)	36.0	36.0	0
Tsat_HP (°C)	73.3	73.3	0
ΔTsh_hp (°C)	2.0	2.0	0
Tsat_LP (°C)	56.2	56.2	0
ΔTsh_Lp (°C)	0.01	0.01	0
Tout (°C)	60.7	60.2	−0.82
Wnet (kW)	969.3	973.1	0.39
Φ (%)	49.23	49.9	1.3
ηth (%)	5.88	5.82	−1.0
ηsys (%	2.89	2.905	0.52

Note: Working fluid is R245fa, inlet temperature of heat source is 100 °C.

lists the validation results.

In addition, the calculation results of the ejector mathematical model established in this paper are compared with the experimental data of Huang et al., which verifies the accuracy of the ejector model [32], and it was found that the error was small and had little influence on the simulation results.

Table 7. Validation of ejector model.

Pg, Mpa (Tg, °C)	Te (°C)	Tc (°C)	μ = ms/mp		Error (%)
			Present Work	Experiment
0.604(95)	8	31.3 33.6 41	0.4175 0.3570 0.2097	0.4377 0.3457 0.2043	−4.62 3.27 2.64
0.538(90)	8	38.9	0.2198	0.2246	2.13
0.604(95)	12	34.2 39.3	0.4271 0.2970	0.4048 0.3040	5.5 −2.3

Note: The experimental fluid is R141b.

lists the injector validation results.

5. System Optimization

In this paper, four output parameters are defined to evaluate the system’s performance: thermal efficiency, exergy efficiency, SUCP, and SUEP. For a perfect system, the higher the thermal and exergy efficiency, the better it would be, while the smaller the SUCP and SUEP, the better. The genetic algorithm has been proved to be the most comprehensive and robust method [33], thus, a genetic algorithm is selected to optimize the system, and through MATLAB there is a genetic algorithm optimization toolbox to achieve.

Because of the conflict of output parameters, it is necessary to make the optimal decision under the condition of tradeoffs among multiple objectives. Exergy efficiency and thermal efficiency should be maximized, while system cost and environmental impact should be minimized. In this case, the Pareto method is only one of many methods to achieve this goal. Multiply each objective function by a weight coefficient and then add it up. Changing the weight coefficient of each objective function can also obtain different Pareto frontier optimal solutions. For this system, the objective function MOF can be defined as the following equation:

$$Max(MOF=w_1\times {\eta }_{th}+w_2\times {\eta }_{ex}+w_3\times (1-SUCP/100)+w_4\times (1-SUEP/b_w))$$

(44)

$$0\le w_1,w_2,w_3,w_4\le 1;w_1+w_2+w_3+w_4=1$$

(45)

where w₁, w₂, w₃, and w₄ are the weight coefficients of thermal efficiency, exergy efficiency, SUCP, and SUEP, respectively, and b_w is the unit environmental impact of electricity supply in the proposed system.

$$\{\begin{array}{l}356.15\le T_{lpvg}\le 368.15\\ T_{lpvg,E}\le T_{hpvg}\le \min (T_{22}-\Delta T_{pp}-T_{\sup },T_{cr}-10)\\ 0.1\le x_{lpvg,out}\le 0.3\\ 0\le r_{suction}\le 0.3\\ 1.2\le p_{cond}/(r_{compress}\times p_{evap})\le 2.5\end{array}$$

(46)

In the above formula, T_lpvg is the low-pressure evaporation temperature; T_hpvg is the high-pressure evaporation temperature; T_cr is the critical temperature of the working medium used by the system; X_{LPVG, OUT} is the dryness of the outlet working fluid of the low-pressure evaporator generator; r_suction is the ratio of suction from the steam turbine, P_cond is the condensing pressure, P_evap is the evaporation pressure of refrigeration, and r_Compress is the compression ratio of the compressor.

6. Results and Discussion

Based on the above description, a program was developed to calculate the energy, exergy economy and exergy environment models of the proposed new power/refrigeration cycle. The program is designed in the numerical analysis software MATLAB, using iterative methods to evaluate the system from the energy, economic, and environmental points of view.

Table 8. Initial condition of simulation.

Parameter	Value
Working fluid	R245fa
Ambient temperature (K)	293.15
Ambient pressure (kpa)	101.325
Inlet temperature of geothermal water (K)	423.15
Pressure of geothermal water (kpa)	500
Mass flow rate of geothermal water (kg/s)	20
Evaporating temperature of refrigeration (K)	243.15
Isentropic efficiency of turbine/η	0.75
Isentropic efficiency of pump/ηp	0.75
Isentropic efficiency of compressor/ηc [3]	0.75
Isentropic efficiency of compressor/ηn [3]	0.9
Mixing efficiency of mix chamber/ηm	0.85
Isentropic efficiency of diffuser/ηd	0.9

lists the initial parameters used to simulate the proposed cycle. Under the optimal working condition, the parameters of each state point of the system are shown in

Table 9. Parameters of each state point under optimized working conditions.

Point	Fluid	T (K)	P (kpa)	m (kg/s)	h (kJ/kg)	s (kJ/kg.K)	E (kW)
1	R245fa	303.15	177.78	21.840	239.10	1.135	77.9
1′	R245fa	303.15	177.78	20.315	239.10	1.135	72.5
1″	R245fa	303.15	177.78	1.525	239.10	1.135	5.4
2	R245fa	303.67	889.32	20.315	240.00	1.137	83.6
3	R245fa	318.88	889.32	20.315	260.53	1.203	107.9
4	R245fa	357.99	889.32	20.315	313.31	1.409	314.4
5	R245fa	357.99	889.32	2.031	464.91	1.783	80.6
6	R245fa	357.99	889.32	18.284	316.47	1.368	233.8
7	R245fa	358.81	1656.5	18.284	317.57	1.369	247.4
8	R245fa	390.72	1656.5	18.284	488.35	1.816	976.2
9	R245fa	334.21	177.78	18.284	456.55	1.848	221.9
10	R245fa	370.34	889.32	0	479.79	1.824	0
11	R245fa	357.99	889.32	2.031	464.91	1.783	80.6
12	R245fa	243.15	10.85	1.525	239.10	1.170	−9.9
13	R245fa	243.15	10.85	1.525	382.76	1.761	−55
14	R245fa	308.47	114.82	1.525	433.70	1.803	3.8
15	R245fa	329.05	177.78	3.556	451.53	1.833	41.1
16	R245fa	333.37	177.78	21.84	455.73	1.846	262.9
17	R245fa	313.67	177.78	21.84	436.63	1.787	223.9
18	water	293.15	101.325	195.634	84.01	0.296	0
19	water	298.42	101.325	195.634	106.06	0.371	38.3
20	air	253.15	101.325	43.570	379.18	3.716	131.7
21	air	248.15	101.325	43.570	374.15	3.696	168.8

, and the system’s performance is shown in

Table 10. The optimal condition calculated based on genetic algorithm.

Parameter	Value
Tlpvg (K)	357.99
Thpvg (K)	385.72
Xlpvg,out	0.1
rsuction	0
rcompress	10.58
$\mu$	0.7506
Wnet (kW)	465.36
Qc (kW)	219.06
ηth (%)	15.01
ηex (%)	43.18
SUCP ($/MWh)	45.525
SUEP (mpts/MWh)	5122.6

. It can be odserved from the Table 10 that the net power output of the system is 465.36 kW, and the cooling output is 219.06 kW. The system’s thermal, exergy efficiency, SUCP and SUEP were 15.01%, 43.18%, 45.525 USD/MWh and 5122.6 mpts/MWh, respectively.

The calculated parameters of each state point in the system under optimal conditions are shown in Table 9. The performance parameters of the system can be easily calculated using the results in Table 9. Table 10 shows that a net power output of 465.36 kW and a cooling output of 219.06 kW can be obtained under optimal conditions, and the corresponding system energy efficiency, exergy efficiency, and SUCP and SUEP are 15.01%, 43.18%, and 45.525 $/MWh, 5122.6 mpts/MWh respectively.

6.1. Exergy Analysis

The exergy destruction of each component of the system can be calculated based on the above optimal conditions. As can be observed from

Table 11. Exergy analysis results of the system.

Component	${\dot{\mathit{E}}}_F$ (kW)	${\dot{\mathit{E}}}_P$ (kW)	${\dot{\mathit{E}}}_D$ (kW)	ηe (%)	ELIP (kW)
Condenser	146.0	38.4	107.6	26.30	79.30
Lppump	18.2	11.2	7	61.54	2.69
IHE	39.1	24.3	14.8	62.15	5.60
LPVG	322.7	206.5	116.2	63.99	41.84
Separator	31.44	31.44	0	100	0
Hppump	20.2	13.6	6.6	67.33	2.16
HPVG	860.3	728.9	131.4	84.73	20.06
Turibe	754.3	581.5	172.8	77.09	39.59
THV	66.3	51.0	15.3	76.92	3.53
Evaporator	45	37.1	7.9	82.44	1.39
Compressor	77.7	58.8	18.9	75.68	4.60
Ejector	57.1	13.8	43.3	24.17	32.83

, LPVG, and HPVG, the steam turbine and condenser were the four parts with the biggest exergy loss in the system, and their exergy loss was 116.2 kW, 131.4 kW, 172.8 kW and 107.6 kW respectively. Exergy efficiency was 63.99%, 84.73%, 77.09% and 26.30%, respectively. The improvement potentials are 41.84 kW, 20.06 kW, 39.59 kW, and 79.30 kW, respectively. Figure 2 shows the ratios of exergy destruction of the four parts in the total exergy destruction of the system, which are 18.1%, 20.5%, 26.9%, and 16.8%, respectively. From the perspective of exergy efficiency and improvement potential, the condenser has the greatest improvement potential, followed by LPVG.

6.2. Exergoeconomic Analysis

Table 12. Exergoeconomic analysis results of system components.

Component	cF (USD/MWh)	cp (USD/MWh)	$\dot{\mathit{Z}}$ (USD/h)	${\dot{\mathit{C}}}_D$ (USD/h)	r	f
Condenser	23.4302	127.6784	1.4742	2.5211	4.4493	0.3689
Lppump	37.3694	66.2622	0.0599	0.2616	0.7732	0.1858
IHE	23.4302	47.4552	0.2361	0.3468	1.0254	0.4052
LPVG	4.6800	9.1450	0.3780	0.5438	0.9541	0.4100
Separator	17.4205	17.4205	0	0	0	0
Hppump	37.3694	60.0417	0.0616	0.2466	0.6067	0.1994
HPVG	4.6800	6.7493	0.8930	0.6150	0.4421	0.5921
Turibe	10.0485	37.3694	14.1490	1.7364	2.7189	0.8907
THV	23.4302	24.6272	0.0065	0.3585	0.0511	0.0178
Evaporator	13.5240	24.7664	0.3109	0.1068	0.8313	0.7444
Compressor	13.5240	55.6031	0.3667	0.2556	3.1114	0.5896
Ejector	46.5427	95.6343	0	2.0153	1.0548	0

lists the thermal economic evaluation parameters of the system, the average unit cost of fuel and product, the exergy consumption cost rate, the total cost rate, the relative cost difference and the economic energy coefficient of each component. The turbine had the highest equipment investment, operation and maintenance cost ratio at USD14.149 per hour. The ejector and condenser fuel costs are very high, at USD46.5427/MWh and USD23.4302/MWh, respectively. Therefore, the ejector and condenser are the parts with the most increased destruction cost rates of exergy in the system, which are USD2.0153/h and USD2.5211/h, respectively. The relative cost rate of the condenser is the largest, which is 4.4493, followed by the compressor and turbine, which are 3.1114 and 2.7189, respectively. Exergy economic coefficients of the turbine are the largest, respectively, at 0.8907, indicating that reducing equipment investment costs can improve the performance of this system. Exergy economic coefficient of two pumps and throttle valve is lower than 0.3. For these components, reducing exergy destruction can improve the system exergy efficiency and system performance.

6.3. Exergoenvironmental Analysis

The thermal environment variables of this system, average unit environmental impact of fuel and products, environmental impact rate of waste, total environmental impact rate, the relative environmental impact difference, and environmental impact coefficient of each component are given in

Table 13. Exergoenvironmental analysis results of system components.

Component	bF (mpts/MWh)	bp (mpts/MWh)	$\dot{\mathit{Y}}$ (mpts/h)	${\dot{\mathit{B}}}_D$ (mpts/h)	rb	fb (%)
Condenser	18,401	70,071	0.0498	1979.9476	2.8079	0.0025
Lppump	22,342	36,411	0.0690	156.394	0.6289	0.044
IHE	18,401	29,722	2.3663	272.3348	0.6152	0.86
LPVG	14,577	22,809	5.6278	1693.8474	0.5648	0.33
Separator	22,809	22,809	0	0	0	0
Hppump	22,342	33,200	0.0765	147.4572	0.4860	0.0517
HPVG	14,577	17,219	9.6843	1915.4178	0.1812	0.50
Turibe	17,219	22,342	2.7084	2975.4432	0.2975	0.091
THV	18,401	33,509	0	281.5353	0.8210	0
Evaporator	10,105	12,252	0.0026	79.8295	0.2125	0.0033
Compressor	22,342	29,519	0.3169	422.2638	0.3213	0.075
Ejector	13,438	55,733	0	581.8654	3.1474	0

. Among the components of the system, HPVG has the highest environmental impact rate in the whole life cycle, which is 9.6843 mpts/h, followed by LPVG, which is 5.6278 mpts/h. The main reason is that these two components need to use 15% copper in the manufacturing process, and the environmental impact rate of copper in its life cycle is 1400 mpts/kg. The exergy loss of the steam turbine, condenser, HPVG and LPVG accounted for more than 16%, thus, the environmental impact rate of the four parts was relatively high. They are 2975.4432 mpts/h, 1979.9476 mpts /h, 1915.4178 mpts/h and 1693.8474 mpts/h, respectively. The relative environmental impact rate of ejector is the largest, which is 3.1474, followed by the condenser, which is 2.8079. The main reason is that the average unit environmental impact of these two parts is the largest, which is 55,733 mpts/MWh and 70,071 mpts/MWh, respectively. The internal heat exchanger (IHE) has the largest environmental impact coefficient, which is 0.86%, followed by HPVG and LPVG, which are 0.50% and 0.33%, respectively. The main reason is that the production materials of these three components contain copper, and their total environmental impact rate is relatively large.

6.4. Parameter Analysis

The inlet pressure of the ejector mainstream had a decisive effect on the ejector suction ratio. The inlet pressure of the ejector mainstream was determined by the low-pressure evaporation temperature (T_lpvg), and the outlet wet vapour dryness (x_lpvg,out) of the low-pressure evaporator and the steam turbine suction ratio (r_suction) affected the mass flow rate of the secondary fluid that the ejector could induce. The system’s cooling capacity is affected by the ejection ratio and the secondary flow rate of the ejector. The ejector ejection ratio is also affected by the secondary fluid pressure, controlled by the refrigeration and evaporation temperature (T_evap) and the compressor compression ratio (r_compress). The system’s power output is determined by the thermal properties of the steam at the inlet of the turbine, and the high-pressure evaporation temperature (T_hpvg) controls the thermal properties of the steam at the inlet of the turbine. At the same time, it also affects the matching of the heat transfer temperature difference of the organic working medium in heat absorption. All the parameters affect the output of the net work and cooling capacity of the system, thus, this paper selects the above six parameters to analyze the system’s performance. The working condition of the analysis process is the optimal working condition calculated by the genetic algorithm.

Figure 3 shows the impact of T_lpvg on the system performance. As shown in Figure 3, with the increase in T_lpvg, the net power output of the system gradually decreased from 467.39 kW to 455.19 kW. The cooling capacity is slowly increasing, from 213.6 KW to 246.96 KW. Because the increase in T_lpvg will lead to a rise in the outlet temperature of the heat source, the mass flow rate through the two vapor generators will be reduced, and the system output work will be reduced under the condition that T_hpvg is unchanged. Secondly, with the increase in the ejector primary fluid pressure and enthalpy value, the ejection coefficient will be improved, and the working medium flow through the evaporator will be increased accordingly. Under the condition that the evaporation temperature is unchanged, the cooling capacity will be increased. As output work decreased and cooling capacity increased, the system’s exergy efficiency decreased, while the system’s thermal efficiency inflated. The exergy efficiency decreased from 43.26% to 42.77%, and the thermal efficiency increased from 14.93% to 15.385%. SUCP increased from $45.4 /MWh to $46.15 /MWh because the cost of cool exergy was higher than the cost of work exergy. When the compression ratio is unchanged, increasing the working medium flow through the evaporator will increase compression power consumption and have a greater environmental impact on the system. SUEP increases from 5025 mpts/MWh to 5617.1 mpts/MWh.

Figure 4 shows the impact of T_hpvg on system performance. As observed in Figure 4, with the increase in T_hpvg, the net power output of the system gradually decreased from 533.93 kW to 389.45 kW. The cooling capacity is also reducing progressively, from 289.51 KW to 171.55 KW. This is because when T_hpvg increases, the working medium’s enthalpy value at the high-pressure steam generator outlet will increase. When the mass flow rate of the heat source is certain, the mass flow rate of the working medium flowing through the whole system will decrease.

Although the enthalpy value at the outlet of the high-pressure evaporator will increase, the output work of the system still demonstrate a decreasing trend, and the decreasing speed is increasing. In the case of T_lpvg and X_{LPVG, OUT} is unchanged, the flow of working medium through the evaporator will be reduced, the evaporation temperature is entire, and the cooling capacity will be reduced when the output work and cooling capacity of the system decrease at the same time, and the increase in T_hpvg leads to the decline of flow through the two steam generators. According to Equations (20) and (21), the decline rate of denominator Qg and Eg is greater than that of the numerator, thus, the system’s thermal and exergy efficiency will show an increasing trend, with the thermal efficiency rising from 13.85% to 15.6%. Exergy efficiency increased from 41.15% to 43.71%. As can be odserved from the diagram, SUCP first reduces after an increasing trend, and this is due to the rise of T_hpvg flows through two flow decreases, the steam generator system of heat absorption capacity decreased, and the cost of fuel consumption is reduced, according to the Formula (30),with a declining rate of fuel costs less than initially lower than that of the exergy of output work and the exergy of cooling capacity., However, when T_hpvg increases to a certain degree, the decreasing rate of fuel cost is greater than that of exergy of output work and exergy of cooling capacity. SUEP shows a downward trend and then an upward trend because the downward trend of output work is slow at first and then fast. When the downward trend of output work is greater than that of the cooling capacity, according to Formula (41), the downward trend of the numerator is smaller than that of the denominator, thus, SUEP shows a downward direction first and then an upward trend.

Figure 5 shows the impact of X_Lpvg,out on system performance. With the increase in X_Lpvg,out, the net power output of the system gradually decreases from 465.36 kW to 238.24 kW. The cooling capacity is increasing from 219.06 kW to 844.93 kW. This is because the increase in X_Lpvg will increase the flow of the working fluid flowing through the low-pressure evaporator and the primary working fluid flow of the ejector. Under the condition that the evaporator temperature remains unchanged, the cooling capacity will gradually increase. In addition, the working flow of the low-pressure evaporator increases, and the power consumption of the low-pressure pump and compressor will also increase. However, the increase in X_Lpvg,out will not change the flow of the working fluid flowing through the high-pressure steam generator; that is, the output work of the turbine will not change, so the net output work of the whole system gradually decreases. As the output work of the system decreased and the cooling capacity increased, the exergy efficiency of the system decreased, while the system’s thermal efficiency increased. Exergy efficiency decreased from 43.18% to 29.37%, and thermal efficiency increased from 15.01% to 18.95%. SUCP increased from $45.525/MWh to $64.17/MWh. When the compression ratio is unchanged, the increased working fluid flow through the evaporator will increase compression power consumption and the greater environmental impact on the system. The SUEP increases from 5122.6 mpts/MWh to 18,673 mpts/MWh.

Figure 6 shows the effect of r_suction on system performance. As shown in Figure 6, when the system’s r_suction increases, the system’s net output energy decreases from 465.36 kW to 82.66 kW. The cooling capacity has been increased from 219.06 KW to 898.79 KW. This is because the increase of r_suction could not only reduce the flow of working fluid flowing through the turbine but also increase the primary working fluid flow of the ejector, the work conducted by the turbine is reduced, the work of the input compressor is increased, and the net power output of the whole system is reduced. In addition, under the condition that the evaporator temperature remained unchanged, the cooling capacity would also gradually increase. The increase in r_suction did not change the system’s heat absorption, and the exergy efficiency of the system decreased. In contrast, the thermal efficiency of the system increased instead, with exergy efficiency declining from 43.18% to 23.15% and thermal efficiency rising from 15.01% to 22.14%. As the cost of the cooling capacity exergy was higher than that of electric exergy, SUCP increased from 45.525 USD/MWh to 83.53USD/MWh. When the compression ratio is unchanged, the increase in the working medium flow through the evaporator will increase compression power consumption and the greater environmental impact on the system. SUEP increases from 5122.6 mpts/MWh to 29,908 mpts/MWh.

Figure 7 shows the impact of T_evap on system performance. As shown in Figure 7, with the increase in T_evap, the net power output of the system gradually decreased from 540.16 kW to 347.06 kW. The cooling capacity is slowly increasing, from 7.69 KW to 558.77 KW. This is because the increase in T_evap will increase the pressure of secondary fluid entering the ejector, and the ejector coefficient will also increase. Under the condition that r_compress is unchanged, the compressor’s power consumption will gradually increase, so the net work output of the system will decrease. Secondly, the ejector’s ejection coefficient increases, and the working fluid’s flow rate through the evaporator increases accordingly. When the T_evap gradually increases, the cooling capacity will also increase. As output work decreased and cooling capacity increased, the system’s exergy efficiency decreased, while the system’s thermal efficiency increased. Exergy efficiency decreased from 45.96% to 37.99%, and thermal efficiency increased from 12.02% to 19.9%. Because the cooling capacity cost is greater than the work capacity, SUCP increases from 45.525 USD/MWh to 83.53 USD/MWh. When the r_compress is unchanged, increasing the working fluid flow through the evaporator will increase the compression power consumption and the greater environmental impact on the system. The SUEP increases from 1736.7 mpts/MWh to 11,726 mpts/MWh.

Figure 8 shows the impact of r_compress on system performance. As shown in Figure 8, with the increase in r_compress, the net output power of the system decreases from 511.26 kW to 342.84 kW. The cooling capacity is gradually increasing, from 110.06 KW to 500.48 KW. When the compressor compression ratio increases, the pressure of the injected fluid in the ejector will increase, and the injection coefficient of the ejector will also increase. As a result, the mass flow rate of the working fluid flowing through the compressor and the compressor’s power consumption will increase, so the net work output of the system will decrease, but the cooling capacity will increase. As the output work decreased and the cooling capacity increased, the system’s exergy efficiency decreased, while the system’s thermal efficiency increased. Exergy efficiency decreased from 45.12% to 37.82%, and thermal efficiency increased from 13.61% to 18.56%. As the cost of unit cooling capacity is greater than that of unit work capacity, SUCP rose from 42.77 USD/MWh to 54.68 USD/MWh. Due to the increased flow of working fluid through the compressor, the compressor’s power consumption will increase, and the system’s environmental impact will be larger. The SUEP increased from 2976.1 MPTS /MWh to 12,007 mpts/MWh.

6.5. System Comparison

The new system proposed in this paper is based on DORC and ERC. However, the new system can adjust the evaporation temperature according to different needs, up to −60 degrees Celsius. To illustrate the advantages of the proposed system over the system of Yu et al. [3], a comprehensive comparative study is conducted between the proposed system and the system of Yu et al. under constant input parameters. The simulation was carried out based on the constant input parameters and assumptions of the research of Yu et al., and the results are shown in

Table 14. Comparison between the thermodynamic performance of the system in this study and that of the reference system.

System Performance	New System	System in Literature [3]
Fluid	R245fa	R245fa
Thpvg (K)	383.15	383.15
Tlpvg (K)	360.15	360.15
Tevap (K)	275.15	275.15
XLPVG,OUT	0.118	0.4
THS,OUT (K)	365.36	351.71
Qin	4910.2	6057.61
WT	549.08	529.82
Wp	38.21	46.91
Wc	68.45	-
μ	1.48	0.2925
Wnet	499.47	486.95
Qc	555.37	555.37
Hth,sys (%)	21.48	17.21
ηex,sys (%)	28.15	25.58
SUCP (USD/MWh)	45.06	53.37

. It can be observed from the table that with the same cooling capacity of 555.37 kW, more work was obtained in this study, and the new system absorbed less heat into the system than Yu et al.’ s system, thus, so the thermal efficiency and exergy efficiency were also higher than the system of Yu et al. As the cost of fuel exergy decreased, the SUCP of the present system was much lower than that of Yu et al. In summary, the proposed new system not only has a broader application range but also has a certain improvement in performance compared with the system of Yu et al.

7. Conclusions

This paper presents a new type of electric cooling co-generation system. The system parameters are optimized, and the system performance is analyzed from the perspectives of thermodynamics, exergy economy and environment combined with the system parameters. The new system’s performance is compared with that of the previous system. The main conclusions are as follows:

(1)Taking thermal efficiency, exergy efficiency, SUCP, and SUEP as the objective functions, GA was used to optimize the system. After optimization, the low-pressure evaporation temperature, high-pressure evaporation temperature, outlet dryness of low-pressure steam generator, suction ratio of steam turbine, and the compression ratio of compressor are 357.99 K, 385.72 K, 0.1, and 0, respectively. the thermal efficiency was 15.01%, exergy rate was 43.18%,SUCP was 45.525 USD/MWh, and SUEP was 5122.6 MPTS/MWh.

(2) The exergy destruction of steam turbine is the greatest. However, from the perspective of exergy efficiency and improvement potential of each component, condenser should be the priority component for system improvement, followed by LPVG.

(3) Among the components of the system, the equipment investment, operation and maintenance cost rate of turbine is the highest, which is 14.149 USD/h; the exergy economic coefficient of steam turbine is the largest, which is 0.8907. Reducing the investment cost can improve the system performance. The exergy economic coefficient of the two pumps and throttle valve is lower than 0.3, so reducing the exergy destruction of these components can also improve the system performance.

(4) Among all components of the system, turbine had the highest impact rate of exergy destruction, which is 2975.4423 mpts/h. The relative environmental impact rate of ejector is the largest, followed by condenser. The exergy environmental influence coefficient of internal heat exchanger (IHE) is the largest.

(5) Higher Tl_pvg, X_{lpvg, out}, r_suction, T_eva and r_compress are beneficial to the output of cooling energy of the system but unfavorable to the output of net work of the system, and the SUCP and SUEP of the system are also larger. However, higher T_hpvg is not conducive to the output of net power and cooling capacity of the system, but is conducive to the improvement of thermal efficiency and exergy efficiency of the system. As for the larger SUCP of the system, the opposite is true for SUEP.

(6) The proposed new system is more widely used through a comprehensive comparative study with the system proposed by Yu et al. Moreover, under the same operating conditions, the net power of the new system increases by 12.52 kW, the thermal efficiency increases by 4.27% and the energy efficiency increases by 2.57%. The results demonstrate that the system’s performance can be significantly improved by using the compressor to increase the pressure of the secondary flow into the ejector.

(7) In this paper, geothermal energy is used to realize the combined supply of cold and electricity, which cannot meet the needs of different users for different products. Distributed energy supply networks can promote the further utilization of renewable energy technologies, reduce costs and improve energy supply efficiency. For example, Sadi [34]. et al. used parabolic trough solar collectors to jointly supply the main heating demand of the chiller and district heating network, realizing the combined heating and cold supply in Denmark. Therefore, solar energy and geothermal energy can supply energy for the system in the future, improving energy utilization efficiency and realizing the triple production of cold, heat, and electricity. The triple production of cold, electricity and distilled water for island areas can also be realized.