Energy and Exergy Analysis of Solar Organic Rankine Cycle Coupled with Vapor Compression Refrigeration Cycle

Abstract

This study investigates a combined power generation and cooling system comprising an organic Rankine cycle (ORC) and a vapor compression cycle (VCC). Thermodynamic analyses of the system were conducted for two operation modes, i.e., the basic ORC mode and the ORC–VCC mode, using six different working fluids: R245fa, R114, R600, R142b, R152a, and R1234yf. The results showed that the thermal efficiency of the combined ORC–VCC system was almost twice that of the basic ORC system. The effects of thermodynamic parameters, such as the turbine inlet temperature and pressure and the condensing temperature, on the system performance were discussed. The second-law efficiency, cooling capacity, and coefficient of performance were addressed by varying the condensing temperature.

1. Introduction

Global environmental and energy problems are emerging as significant issues as the demand for various energy-consuming facilities, including refrigeration and air-conditioning equipment, increases owing to industrial advancement and life standard improvement. The energy efficiency in industry must be improved to overcome global energy and environmental problems [1,2]. Energy recovery technologies employing standalone and combined cycle configurations have garnered increasing attention. All energy-related challenges encompassing resources, demand, and supply, as well as their applications, have been a global concern. Implementing thermally efficient practices and exploiting industrial waste heat and renewable energy sources are possible measures for improving industrial energy efficiency. Owing to technical advancements, renewable energy and low-grade waste heat sources are prime candidates for various applications, such as electricity generation, heating and cooling, hydrogen production, and water desalination [3].

Combined heat and power (CHP) systems are widely used for electricity production associated with simultaneous water and space heating or cooling; these systems involve the use of renewable energies and waste heat recovery. This technology exhibits high potential in improving the overall system performance because it allows waste heat and renewable energy sources to be transformed into useful energy. These characteristics of CHP systems have become particularly attractive in recent years owing to the expected increase in global energy demand of up to 80% by 2050 [4], along with regulations for reducing greenhouse gas emissions to mitigate global warming. Applying small-scale, renewable-energy-based cogeneration technology to decentralized, off-grid areas can further improve the socioeconomic environment [5], providing fuel independence and reducing energy costs. This implementation has become a driving force for economic development and sustainability. Combined cooling, heat, and power systems equipped with different types of primary mover engines, such as organic Rankine cycle (ORC), Stirling, and reciprocating engines, have been investigated extensively [6,7,8].

Recently, Khatoon et al. [3] conducted a thermal performance analysis on an ORC-powered vapor compression refrigeration cycle for eight working fluids, also known as a dual-fluid system. They performed a cycle analysis to satisfy a setting temperature of −16 °C for different condenser temperatures of 34, 36, 38, and 40 °C. Bao et al. [9] compared the performances of single- and dual-fluid system configurations of ORC–vapor compression cycle (VCC) systems. Using a geothermal energy source that provided water at 140 °C, they investigated the vapor injection cycle of a flask tank coupled with an ORC for several different refrigerants. Javanshir et al. [10] proposed and investigated a cooling–power cogeneration cycle comprising vapor-compression refrigeration and ORCs. They used geothermal water as a low-temperature heat source and conducted a simulation to investigate the effects of various working fluids on the thermal efficiencies and size parameters using EES software. Kim and Blanco [11] investigated the performance of an ORC–VCC system based on eight working fluids. Mudasar et al. [12] performed a thermodynamic analysis on an ORC using flue gas from biogas accumulation as a heat source and subsequently calculated the cycle performance using a MATLAB code developed with REFPROP real-gas properties. Aphornratana and Sriveerakal [13] investigated the performance of an ORC–VCC system that shared a single working fluid through a common condenser coupled with a single-piston expander–compressor unit. They demonstrated that R22 exhibited the best coefficient of performance (COP) under the design parameters considered. Khatoon and Kim [14,15] investigated the performance of a power-generating Brayton cycle using supercritical CO₂ as the working fluid.

Molés et al. [16] performed a theoretical evaluation of the low-global-warming-potential (GWP) refrigerants R1234yf and R1234ze as alternatives to R134a for use as low-temperature heat sources in ORC systems. The results revealed that the net cycle efficiency of R1234ze was 13.8% higher than that of R134a, whereas the cycle efficiency of R1234yf was 13.9% lower than that of R134a. Le et al. [17] optimized the system efficiencies of basic and regenerative supercritical ORCs using low-GWP working fluids driven by a low-grade heat source. They investigated two cases: basic and regenerative ORC systems. The results revealed that R32 and R152a were the best working fluids for optimizing the system efficiency of the basic and regenerative cycles, respectively, whereas R1234ze was the best working fluid for optimizing the net electrical power of the basic cycle. Rayegan and Tao [18] investigated a method for selecting working fluids in solar ORCs and discovered that 11 working fluids were suitable for solar ORC plants used in low- or medium-temperature solar collectors. Quoilin et al. [19] presented an overview of several ORC applications, along with working fluid selection and expansion machine issues. They addressed technological constraints and optimization methods. Wang et al. [20] conducted a detailed off-design analysis of solar-driven ORCs. Song et al. [21] performed a thermo-economic optimization study on several different types of ORC configurations using six working fluids, namely R134a, R1234yf, R600a, R1233zd, R245fa, and isopentane, based on a geothermal heat-source temperature of 180 °C and a mass flow rate of 40 kg/s. Bertrand et al. [22] investigated 20 working fluids for a low-temperature solar ORC system and discovered that R134a was the most suitable refrigerant for small-scale solar ORC applications. Hydrofluorocarbon and hydrocarbon working fluids, such as R152a, R600, R600a, and R290, exhibit excellent performances but require safety precautions because they are flammable refrigerants. Dragomir–Stanciu et al. [23] investigated the effect of condensing temperatures in the range of 10–25 °C on a solar-driven ORC power system using two working fluids: R134a and R600a. The results revealed that the system efficiency and amount of electricity produced by the solar ORC system increased when the condensing temperature was decreased. Freeman et al. [24] conducted a systematic analysis to evaluate a range of sensible and latent thermal energy storage options in a solar CHP system based on an ORC engine. Aziz et al. [25] conducted a thermodynamic performance analysis on a high-temperature cascade ORC system coupled with a water-heating system based on local weather data. Nasir and Kim [26] presented the thermal performances of several different combinations of seven working fluids, namely R134a, R1234yf, R1234ze (E), R123, R245fa, R600a, and butane, in an ORC-powered VCC system for residential air-conditioning applications.

This study investigates the performance of a solar ORC–VCC system for electricity generation and space cooling. The system operates in either the basic ORC mode (in the winter) or in ORC–VCC mode (in the summer). A system configuration description, thermodynamic modelling, and an analysis for different operating conditions are presented herein. A series of system simulations is conducted for six different working fluids: R245fa, R114, R600, R142b, R152a, and R1234yf. The system performance is presented in terms of the net power output, thermal efficiency, refrigerant mass flow rate, second-law efficiency, cooling capacity, and COP under several different operating conditions.

2. System Configurations

Figure 1 shows a schematic of a solar ORC coupled with a VCC refrigeration cycle. The two cycles of the ORC and ORC–VCC were designed such that a common condenser could be used between them, as shown in the figure. The solar collector loop was designed to use solar energy as a heat source for the ORC, and the collector type was a parabolic trough solar collector with Terminol-VP1 as the heat transfer fluid. The cooling water entered the condenser at state 14 and exited at state 15. The saturated working fluid exiting the condenser was directed to the ORC and VCC. In the combined ORC–VCC mode, the refrigerant mass flow rate of the VCC was 20% of that of the ORC, whereas in the basic cycle mode when the VCC was not in operation, the entire refrigerant flowed to the ORC. The working fluid at state 1 from the condenser in the ORC was pressurized through the ORC pump and heated through the heat exchangers (recuperator, economizer, evaporator, and superheater), which received heat from the solar energy source and entered the turbine inlet as superheated vapor. Subsequently, the working fluid expanded as an isentropic process in the turbine expander to generate electric energy and entered the condenser in saturation state 8 after undergoing heat exchange in the recuperator. The recuperator preheated the low-temperature refrigerant before entering the economizer by recovering heat from the refrigerant expanded in the turbine and cooling the refrigerant entering the condenser to saturate it, thus reducing the heat input and improving the system efficiency.

In the VCC, the saturated working fluid at state 18 from the condenser entered the expansion valve, where the working fluid underwent isenthalpic expansion. Subsequently, the refrigerant entered the evaporator in saturation state 19, absorbed the heat from the conditioned space, and evaporated. The low-temperature saturation vapor at state 16 from the evaporator entered the compressor, where it was compressed in high-pressure superheated state 17. The compressed high-pressure working fluid (state 17) merged with the working fluid (state 8) from the ORC recuperator and entered the condenser.

3. Modelling and Analysis

A thermodynamic analysis of the system was conducted based on mass and energy conservation laws. Cycle simulations were performed using EES software [27]. The mass-and-energy-balance relationships for each component of the system considered are presented in this section. To avoid an overly complex simulation, the following constraints and assumptions were introduced.

(1)

The system operation is at steady-state conditions;

(2)

The heat and frictional losses in the cycle are negligible;

(3)

The variations in kinetic and potential energy are negligible;

(4)

The pumping, compression, and expansion processes are isentropic;

(5)

The throttling process in the expansion valve is isenthalpic;

(6)

The operation of the solar collector pump is negligible;

(7)

The economizer and evaporator of the ORC exchange heat such that the evaporator inlet and outlet qualities are 0 and 1, respectively.

A pinch-point temperature of 5 °C was considered for all the heat exchangers. The overall efficiency of the power generation was assumed to be 90%, considering the generator and mechanical efficiency. The energy-balance relationships for each component of the system were as follows.

-

ORC turbine

The power generated from the ORC expander can be calculated as:

$${\mathcal{W}}_{tb}={\dot{m}}_{ORC}\left(h6-h7\right)={\dot{m}}_{ORC}\left(h6-h7s\right){\eta }_{tb}.$$

(1)

-

ORC condenser

The heat rejected from the ORC condenser can be calculated as:

$${\mathcal{Q}}_{cond}={\dot{m}}_{ORC}\left(h8-h1\right).$$

(2)

-

ORC pump

The power consumed by the pump can be calculated as:

$${\mathcal{W}}_{pump}={\dot{m}}_{ORC}\left(h2-h1\right)={\dot{m}}_{ORC}\left(h2s-h1\right)/{\eta }_{pump}.$$

(3)

-

ORC recuperator

The heat exchange in the recuperator, whose effectiveness is 0.8, can be expressed as:

$$h7-h8=h3-h2.$$

(4)

-

Heat input

The heat to be received by the ORC heat exchangers (economizer, evaporator, and superheater) can be calculated as:

$${\mathcal{Q}}_{in}={\dot{m}}_{ORC}\left(h6-h3\right)={\dot{m}}_{sc}\left(h10-h13\right),$$

(5)

where $m_{sc}$ is the mass flow rate of the heat transfer fluid in the solar energy collector loop.

-

VCC compressor

The power consumed by the compressor can be calculated as:

$${\mathcal{W}}_{comp}={\dot{m}}_{VCC}\left(h17-h16\right)={\dot{m}}_{VCC}\left(h17s-h16\right)/{\eta }_{comp}.$$

(6)

-

VCC condenser

The heat rejected from the VCC condenser can be calculated as:

$${\mathcal{Q}}_{condVCC}={\dot{m}}_{VCC}\left(h17-h18\right).$$

(7)

-

Expansion valve

The working fluid flowing through the expansion valve can be considered an isenthalpic throttling process, as follows:

$$h18=h19.$$

(8)

-

Evaporator

The heat transferred in the evaporator VCC can be calculated as:

$${\mathcal{Q}}_{evapVCC}={\dot{m}}_{VCC}\left(h16-h19\right).$$

(9)

The net power outputs of the ORC and ORC–VCC systems are expressed as:

$$\begin{array}{c}{\mathcal{W}}_{net}=\left({\mathcal{W}}_{tb}-{\mathcal{W}}_{pump}\right)\hspace{1em}\hspace{1em}\hspace{1em}forORC,\\ W_{net}=\left({\mathcal{W}}_{tb}-{\mathcal{W}}_{pump}-{\mathcal{W}}_{comp}\right)\hspace{1em}\hspace{1em}\hspace{1em}forORC-VCC.\end{array}$$

(10)

The thermal efficiency (${\mathsf{\eta }}_{th}$) of the ORC system depends on the net power output (${\mathcal{W}}_{net}$) and heat input to the system, whereas in case of the ORC–VCC system, the evaporator capacity should be considered as an output. Therefore, they are expressed as:

$$\begin{array}{c}{\mathsf{\eta }}_{th}=W_{net}/{\mathcal{Q}}_{in}\hspace{1em}\hspace{1em}\hspace{1em}forORC,\\ {\mathsf{\eta }}_{th}=({\mathcal{W}}_{net}+{\mathcal{Q}}_{evapVCC})/{\mathcal{Q}}_{in}\hspace{1em}\hspace{1em}\hspace{1em}forORC-VCC.\end{array}$$

(11)

The COP of the VCC system is a function of the evaporator capacity and compressor power input to the system, as follows:

$$COP=\frac{Q_{evapVCC}}{W_{comp}}.$$

(12)

The exergy efficiency (${\mathsf{\eta }}_{ex}$), or the second-law efficiency, can be written as:

$${\mathsf{\eta }}_{ex}=\frac{{\mathcal{W}}_{net}+E{x}_{evapVCC}}{E{x}_{in}},$$

(13)

where the exergy input for the system and the exergy in the VCC evaporator are expressed as follows:

$$\begin{gathered} E{x}_{in}=Q_{in}-{\dot{m}}_{sc}{T}_o\left(s_{10}-s_{13}\right), \\ E{x}_{evapVCC}=Q_{evapVCC}-{\dot{m}}_{VCC}{T}_o\left(s_{16}-s_{19}\right). \end{gathered}$$

(14)

The working fluids used for the system were R245fa, R114, R600a, R142b, R152a, and R1234yf. Some refrigerants have ODP and high GWP values, and R600, R142b, and R152a are classified as high-flammability groups A2 or A3, as presented in

Table 1. Thermophysical properties of working fluids used [28,31,32].

Working Fluid	Molar Mass (kg/kmol)	Normal Boiling Point (°C)	Critical Temperature (°C)	Critical Pressure (kPa)	ODP/ GWP	Safety Group
R245fa	134.05	15.0	153.9	3651	0/858	A1
R114	170.92	3.6	145.7	3257	1/8590	A1
R600	58.12	−0.5	152.0	3796	0/1	A3
R142b	100.50	−9.1	137.1	4055	0.065/1980	A2
R152a	114.04	−29.5	94.7	3382	0/138	A2
R1234yf	66.05	−24.0	113.6	4517	0/1	A2L

. Refrigerants with properties suitable for the operating conditions of the system were selected. The relationship between saturation vapor pressure and temperature is shown in Figure 2. All the properties were obtained using REFPROP 10 [28]. A thermodynamic analysis of the system was conducted based on the mass and energy conservation laws using EES software. The operating conditions are listed in

Table 2. Operating conditions.

Parameter	Value
Atmospheric temperature, To [°C]	25
Atmospheric pressure, Po [kPa]	101.3
Solar intensity, G [W/m2]	700
$\mathrm{Solar}\mathrm{collector}\mathrm{efficiency},{\mathsf{\eta }}_{\mathrm{sc}}$ [-]	0.81
Maximum thermal oil temperature, T10 [°C]	150
Collector working pressure, Psc [kPa]	2000
$\mathrm{Mass}\mathrm{flow}\mathrm{rate}\mathrm{of}\mathrm{thermal}\mathrm{oil},{\dot{m}}_{sc}$ [kg/s]	2
$\mathrm{Heat}\mathrm{exchanger}\mathrm{pinch}\mathrm{point},\mathsf{\eta }{T}_{PP}$ [°C]	5
Turbine inlet pressure, P6 [kPa]	1500–2500
Turbine inlet temperature, T6 [°C]	130–140
$\mathrm{Isentropic}\mathrm{turbine}\mathrm{efficiency},{\mathsf{\eta }}_{\exp }$ [-]	0.8
$\mathrm{Isentropic}\mathrm{compressor}\mathrm{efficiency},{\mathsf{\eta }}_{\mathrm{comp}}$ [-]	0.75
$\mathrm{Isentropic}\mathrm{pump}\mathrm{efficiency},{\mathsf{\eta }}_{\mathrm{pump}}$ [-]	0.85
$\mathrm{Recuperator}\mathrm{efficiency},{\mathsf{\eta }}_{\mathrm{rec}}$[-]	0.8
Condensing temperature, T1 [°C]	30–40
Condenser water inlet temperature, T14 [°C]	15
$\mathrm{Cooling}\mathrm{water}\mathrm{flow}\mathrm{rate},{\dot{m}}_{water}$ [kg/s]	2

. The evaporation temperature of the system was set at 7 °C to reflect the operating conditions of the air-conditioning system. The heat transfer fluid for the solar collector was Therminol VP-1, whose properties are included in EES software. Therminol VP-1 can generally be used up to a maximum temperature of 400 °C in a closed system [29]. The solar intensity and collector efficiency were 700 kW/m² and 81%, respectively [30].

4. Results and Discussion

Based on the computer code developed with EES software, the considered systems were analyzed for different operating conditions. Two operation modes were considered: the basic ORC mode and the ORC–VCC mode. In the basic mode, the VCC was turned off, and the ORC system operated only with the solar collector. In the ORC–VCC mode, the VCC system was turned on to operate, and the refrigerant mass flow rate of the VCC was divided by 20% of the ORC flow rate from the condenser. The performances of the ORC and ORC–VCC modes were evaluated by calculating the net work output, thermal efficiency, cooling capacity, coefficient of performance, and exergy efficiency for various operating conditions. The calculated results are presented in the following section.

4.1. Performance Analysis of ORC

Using the energy balance equations described in Section 3, the basic ORC system was first analyzed for a set of working fluids, as listed in Table 1, where the turbine inlet temperature and pressure were varied within the ranges of 130–140 °C and 1500–2500 kPa, respectively, whereas the other variables were fixed. For the basic solar ORC system, the performance parameters, such as the net power output, thermal efficiency, and refrigerant mass flow rate, were calculated using EES code. Figure 3 depicts the system characteristics as a function of the turbine inlet temperature computed for each working fluid. As expected, the thermal efficiency of the system increased with the turbine inlet temperature owing to an increase in the enthalpy difference. Consequently, the temperature difference between the inlet and outlet temperatures of the heat transfer fluid of the solar collector increased. Therefore, a larger collector area was required, which increased the collector cost. The net work output decreased with the turbine inlet temperature, thereby reflecting the thermophysical properties of the working fluids and the refrigerant mass flow rates of the ORC system. Figure 4 shows the system characteristics as the turbine inlet pressure varied. The net power outputs for refrigerants R245fa, R114, and R600 increased with the turbine inlet pressure, whereas they decreased for R152a and R1234yf, reflecting their thermodynamic properties. For R142b, the net work increased with the turbine inlet pressure and then decreased, with the maximum value at a turbine inlet pressure of 2200 kPa. As shown in the figures, the refrigerant mass flow rate of R600 was the lowest value, whereas that of R1234yf was the highest, which reflected the specific volume of the compressor inlet. As shown in the figures obtained from the parametric study, the effect of the turbine inlet pressure for each working fluid on the system performance was greater than that of the turbine inlet temperature in the pressure and temperature ranges considered. Thus, the optimum turbine inlet pressure control for each refrigerant must be ensured to improve the system efficiency.

4.2. Performance Analysis of ORC–VCC

Because the working fluid of the ORC and VCC is a single fluid, the refrigerant mass flow rate is segregated in the condenser such that the VCC mass flow rate is 20% of the ORC mass flow rate, as explained earlier. The overall system characteristics depend on the performances of the ORC and VCC systems; hence, based on the compressor power input in the VCC calculated to satisfy the required cooling temperature (9 °C), system simulations were conducted for different turbine inlet temperatures and pressures, as well as condensing temperatures of 30–45 °C, for the six working fluids considered. The results calculated for the ORC–VCC system included the net power output, thermal efficiency, second-law efficiency, cooling capacity, compressor power input, and COP for different operating conditions, as presented in

Table 3. Performance characteristics of basic ORC and ORC–VCC systems at a condensing temperature of 40 °C, a turbine inlet temperature of 135 °C, and a pressure of 2300 kPa.

Working Fluid	Operation Mode	Net Work (kW)	Thermal Efficiency * (%)	Second-Law Efficiency * (%)	Evaporator Capacity (kW)	Mass Flow Rate (kg/s)	Compressor Power (kW)	COP
R245fa	ORC	15.82	13.1 (11.8)	48.0 (43.7)
	ORC–VCC	13.00	25.0 (22.6)	42.4 (38.6)	17.29	0.55	3.12	5.54
R114	ORC	17.37	12.3 (10.7)	46.1 (40.4)
	ORC–VCC	14.17	23.6 (20.5)	40.5 (35.5)	19.08	0.94	3.55	5.37
R600	ORC	18.86	12.9 (11.3)	48.5 (43.1)
	ORC–VCC	15.47	24.7 (21.7)	42.8 (38.0)	20.69	0.35	3.76	5.51
R142b	ORC	23.62	11.5 (10.1)	46.2 (41.5)
	ORC–VCC	18.66	23.9 (21.0)	39.9 (35.8)	30.44	0.89	5.50	5.54
R152a	ORC	22.17	8.6 (7.2)	37.3 (33.1)
	ORC–VCC	15.56	21.7 (18.2)	30.0 (26.7)	40.22	0.84	7.34	5.48
R1234yf	ORC	19.77	8.1 (5.9)	34.3 (28.0)
	ORC–VCC	13.75	19.4 (14.2)	27.2 (22.2)	33.81	1.50	6.69	5.06

* The value in parentheses indicates the case in which the recuperator was not applied.

and Figure 5, Figure 6, Figure 7, Figure 8 and Figure 9.

Table 3 lists the simulated results for the ORC–VCC system compared with those for the basic ORC mode under a condenser temperature of 40 °C, a turbine inlet temperature of 135 °C, and a pressure of 2300 kPa. The net work output of the ORC–VCC mode for each refrigerant was less than that of the basic ORC mode because it considered the compressor power input. In contrast, the thermal efficiency of the ORC–VCC mode was approximately twice that of the basic ORC mode because the cooling capacity of the VCC was included in the net work output. The thermal efficiencies of high-pressure working fluids R245fa, R114, and R600a (see Figure 2) were higher than those of low-pressure refrigerants R152a and R1234yf. This was attributed partly to the effect of the inlet pressure of the turbine on the net work output, as shown in Figure 6. The highest thermal efficiency was 25% for R245fa, which was 29% higher than the 19.4% achieved by R1234yf under this specific operating condition. When the recuperator was not applied in the cycles, the thermal efficiency and second-law efficiency increased from 10% and 11% (based on R245fa) to 27% and 18% (based on R1234yf), respectively. The recuperator in both cycles considerably affected the thermal and second-law efficiencies, and applying the recuperator could reduce the total system cost (as compared with using the basic simple cycle) by reducing the solar collector and condenser loads.

Figure 5 and Figure 6 depict the net power output and thermal efficiency as the turbine inlet temperature and pressure were varied, respectively. The overall performance trends for the ORC–VCC mode were similar to those of the basic ORC mode. The effect of the turbine inlet temperature on the net work output and thermal efficiency was less prominent than that of the turbine inlet pressure. Refrigerant R142b yielded the best net work output, and R245fa indicated the least net work in the turbine inlet temperature range considered. In terms of thermal efficiency, R245fa exhibited the highest value, whereas R1234yf exhibited the lowest value in the turbine inlet temperature range investigated. The overall effect of the turbine inlet pressure on the net work output and thermal efficiency for the ORC–VCC mode was similar to that of the basic ORC mode. As previously explained, in both the ORC–VCC and basic ORC modes, the turbine inlet pressure considerably affected the system performance; hence, the turbine inlet pressure should be controlled appropriately for the selected refrigerants.

Figure 7, Figure 8 and Figure 9 show the system performances for the ORC–VCC mode vs. the condensing temperature. In particular, the net work, thermal efficiency, cooling (evaporator) capacity, compressor work, COP, and second-law efficiency were calculated for a range of condensing temperatures. As shown in Figure 7, the net work and thermal efficiency decreased monotonically with the condensing temperature owing to an increase in the heat transfer rate to the environment (condenser heat transfer rate). Figure 8 shows the cooling capacity and compressor power input of the VCC vs. condensing temperatures of 130–140 °C. The cooling capacity decreased with the condensing temperature, whereas the compressor work increased, as shown in Figure 8. The decrease in the cooling capacity was attributed to the reduction in the compressor volumetric efficiency, as increasing the condensing temperature while maintaining the evaporating temperature increased the compression ratio. The higher the condensing temperature, the higher the compressor discharge temperature during the isentropic process, and thus, more compressor work was required. The COP and second-law efficiency of the ORC–VCC mode decreased with the condensing temperature, as expected. R600a and R1234yf indicated the highest and lowest values of second-law efficiency, respectively. As presented in Table 3 and Figure 9, the COP values dis not differ considerably, barring the lowest value for R1234yf. The refrigerant mass flow rate of R1234yf presented in Table 3 was 1.5 kg/s at the operating condition, which was considerably higher than those of the other working fluids; this was attributable to its extremely small specific volume at the compressor inlet at the evaporating temperature of 7 °C, i.e., 0.045 m³/kg.

5. Conclusions

A thermodynamic analysis of an ORC system combined with a parabolic trough solar collector applicable for small-scale power generation was presented herein. The goal was to increase the thermal efficiency of the power generation system using the heat of the ORC and incorporating a VCC into the system. Six refrigerants, namely R245fa, R114, R600, R142b, R152a, and R1234yf, were considered as working fluids. The VCC system utilized some of the dissipative heat in the condenser and was used to cool the air in a closed space; the cooling air discharge temperature was set to 9 °C. Using the VCC cycle simultaneously with the ORC cycle considerably increased the thermal efficiency of the system. The effects of turbine input temperature and pressure on the net power output, thermal efficiency, refrigerant mass flow rate, secondary efficiency, cooling capacity, and COP were investigated under various operating conditions. The performance analysis results indicated that applying a recuperator improved the thermal efficiency and second-law efficiency by 10–27% and 11–18%, respectively, via heat recovery from the outlet of the turbine and reduction of the solar collector and condenser loads. Furthermore, applying the recuperator to the system reduced the total system cost. Results from the parametric study revealed that the inlet pressure of the turbine considerably affected the thermal and exergy efficiency of the system. R245fa exhibited the highest thermal efficiency of 25%, which was 29% higher than the 19.4% achieved by R1234yf under the operating conditions used in this study.