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# Energy and exergy analysis of solar organic Rankine cycle coupled with vapor compression refrigeration cycle using different working fluids

Merouane Saka<fnref n="1" />, Zakaria Triki<fnref n="1" />, Zineb Fergani<fnref n="1" />, Kamilia Madi<fnref n="2" />, Jie Zhang<fnref n="3" />, Abdeltif Amrane<fnref n="4" /> and Hichem Tahraoui<fnref n="4" />,<fnref n="5" />*

## Abstract

This study investigates the performance of a solar-powered organic Rankine cycle (ORC) coupled with a vapor compression cycle (VCC) for simultaneous power generation and cooling applications in the region of Ghardaïa (Southern Algeria). The analysis focuses on the energy and exergy performance of five low-GWP refrigerants: R134a, R1234ze(E), R600a (isobutane), R152a, and R1234yf, under realistic operating conditions. In addition, key parameters such as turbine inlet temperature and pressure, VCC condensing and evaporation temperatures, and the VCC/ORC mass flow rate ratio were analyzed to evaluate their impact on net power output, cooling capacity, coefficient of performance (COP), and exergy efficiency. Results indicate that R600a provides the highest net power output (7.08 kW) and exergy efficiency (9.87%), making it ideal for power-dominant applications. R152a, on the other hand, excels in cooling performance, achieving a COP of 6.10. R1234ze(E) emerges as the most environmentally sustainable option, offering a balanced performance with 8.02% exergy efficiency and a negligible GWP of 7. Exergy analysis reveals that the solar collector is the primary source of irreversibility, with 58 kW of exergy destruction, followed by the ORC and VCC subsystems. These findings offer valuable insights for designing sustainable cogeneration systems in arid regions, highlighting the trade-off between thermodynamic performance and environmental impact when selecting the optimal working fluid for ORC-VCC applications.

**Keywords** Solar ORC-VCC system, Power generation, Cooling, low-GWP refrigerants, Thermal performance, Exergy analysis

## 1 Introduction

With the continual growth in worldwide energy consumption, adopting sustainable approaches is becoming ever more essential. In response, the development of sustainable and efficient energy systems has become a critical research area. Hybrid energy systems, which combine multiple energy conversion technologies, present a viable solution for optimizing energy utilization and minimizing environmental impact [1]. Among

<footnote>*Correspondence:</footnote>

<footnote>Hichem Tahraoui
hichemm.tahraouii@gmail.com</footnote>

<footnote n="1">¹Renewable Energies and Materials
Laboratory, University of Medea,
26000 Medea, Algeria</footnote>

<footnote n="2">²Department of Process
Engineering, Faculty of Technology,
University of Sétif 1, 19000 Sétif,
Algeria</footnote>

<footnote n="3">³School of Engineering, Merz Court,
Newcastle University,
Newcastle Upon Tyne NE1 7RU, UK</footnote>

<footnote n="4">⁴Univ Rennes, Ecole Nationale
Supérieure de Chimie de Rennes,
CNRS, ISCR – UMR6226,
F-35000 Rennes, France</footnote>

<footnote n="5">⁵Laboratory of Biomaterials and
Transport Phenomena, University of
Medea, 26000 Medea, Algeria</footnote>
these hybrid systems, the integration of an organic Rankine cycle (ORC) with a vapor compression cycle (VCC) has emerged as a promising approach for simultaneous power generation and cooling, particularly when driven by low-to-medium temperature heat sources such as solar energy [2–4].

Solar energy, an abundant and environmentally benign resource, presents a significant potential for powering such hybrid systems [5]. The integration of solar energy into combined ORC-VCC systems not only reduces reliance on fossil fuels but also enhances the overall sustainability of the energy generation process. The ORC, which utilizes organic working fluids with lower boiling points than water, enables the efficient conversion of heat from solar thermal sources, even at relatively low-to-medium temperatures [6, 7]. The VCC, a well-established and widely implemented cooling technology, can be effectively coupled with the ORC to enhance overall system efficiency and provide valuable cooling services, creating a synergistic energy system [4]. This combined approach allows for a more comprehensive utilization of the solar energy input, improving the overall performance of the hybrid system and potentially increasing its economic viability.

Numerous studies have explored the potential of integrating ORC and VCC systems, particularly when powered by solar energy. Research has focused on optimizing system performance through working fluid selection, component design, and operational parameter adjustments. Wang et al. [8] proposed a novel thermally activated cooling system combining an ORC and a VCC, capable of utilizing solar, geothermal, or waste heat. The cycle achieved a coefficient of performance (COP) of 0.66 under extreme conditions (48.9 °C) when both subcooling and cooling recuperation were applied. A parametric study identified five key design parameters, and two advanced configurations, adding a secondary ORC recuperator and using dual fluids, were suggested for further improvement.

Wang et al. [9] developed a heat-activated cooling system combining an ORC with a vapor compression cycle, designed to utilize waste heat from engines for portable cooling applications. A 5 kW prototype was built and tested, achieving 4.4 kW cooling and a COP of 0.48 under off-design conditions. High performance was supported by micro-channel heat exchangers and scroll-based devices, with the scroll expander reaching 84% isentropic efficiency.

Sherwani [10] integrated an ORC-VCC system with an N-series photovoltaic thermal compound parabolic collector to simultaneously generate power and cooling. System performance was evaluated using COP, exergy efficiency, and CO₂ mitigation per collector area. Maximum energy and exergy performance (COP = 1.81, exergy efficiency = 5.8%) occurred at an expander inlet temperature of 423 K and a packing factor of 0.25. The highest environmental benefit (1999 g/m² CO₂ reduction) was achieved at 373 K and 0.89 packing factor. Increasing the number of collectors from 1 to 5 further improved exergy efficiency to 17.4% and CO₂ mitigation to 8546 g/m².

In another work, Sherwani [11] proposed a multi-evaporator vapor-compression refrigeration system driven by an ORC for applications like supermarkets and restaurants. Using a zeotropic mixture of hexane and R245fa, the study analyzed thermodynamic performance and optimized the system via a genetic algorithm. Maximum COP (0.425) and exergetic efficiency (31.1%) were achieved at a 0.5 mass fraction. Optimization yielded a peak exergetic efficiency of 58% with optimal operating temperatures of 388 K (ORC evaporator), 268 K (both VCC evaporators), and 303 K (condenser).
Khan and Mishra [12] conducted a parametric analysis of a supercritical ORC-VCC system powered by solar parabolic trough collectors (PTCs) for simultaneous power and cooling generation. Key performance metrics included thermal efficiency, exergy efficiency, COP, and exergy destruction. Maximum exergy efficiency (92.9%) and thermal efficiency (51.13%) were achieved with R227ea, while the highest COP (2.278) was obtained with R134a. PTCs contributed to 76.32% of total exergy destruction, highlighting the need for careful PTC design.

Malwe et al. [13] analyzed an integrated ORC-VCC system using a dynamic MAT-LAB-Simulink model across 150 refrigerant combinations. The study evaluated COP, exergy efficiency, and exergy destruction under varying load and temperature conditions. The R141b/R1234ze(Z) pair yielded the best results, with an exergy efficiency of 33.05%, a net COP of 4.593, and a minimum exergy destruction of 2.591 kW. The COP improved with higher evaporator temperatures, and the integrated system achieved a 10.62% higher COP than a standalone VCC system.

Further investigations have examined the economic performance of integrated ORC-VCC systems. Sherwani and Tiwari [14] conducted a comprehensive thermodynamic, economic, and environmental assessment of a low-grade heat source-driven ORC-VCC system for multifunctional applications. Using the TOPSIS method for multi-criteria decision-making, the optimal exergetic efficiency, total cost, and environmental cost were determined to be 28.4%, $69,949, and $3,491, respectively.

Braimakis et al. [15] examined two microscale trigeneration systems integrating an ORC and a VCC with biomass boilers and PTCs, comparing them economically to a photovoltaic (PV) powered heat pump. The ORC efficiencies ranged from 3.7% to 10.05%, depending on the working fluid and operating conditions, with cogeneration efficiencies reaching up to 73.5%. The PV heat pump demonstrated an average efficiency of 15%. Economic analysis revealed similar payback periods for both systems, 12.3 years for the ORC-VCC system and 12.1 years for the PV heat pump, indicating competitive viability for renewable energy-based multigeneration technologies in domestic heating and cooling applications. Working fluid selection has been a critical aspect of research in this field.

Ngangué et al. [16] evaluated a cogeneration system combining ORC, vapor absorption cycle (VAC), and VCC, using $NH_3-NaSCN$ in the VAC and seven refrigerants for the ORC-VCC. R717 was identified as the most efficient fluid, achieving 1205 kW cooling capacity, 158.3 kW net power, 64.32% thermal efficiency, and 39.39% exergy efficiency. Major exergy destruction occurred in the steam generator (57.29%) and condenser (22.58%).

Aryanfar et al. [17] investigated a solar-driven ORC-VCC system for combined power and cooling suitable for rural applications. The ORC was powered by a solar PTC, and four working fluids (R245fa, R114, R600, R142b) were evaluated through thermodynamic and economic simulations. The integration of ORC and VCC improved thermal efficiency. R245fa exhibited the lowest exergy destruction, while R114 and R142b minimized the cost function (PCEU) at a turbine inlet temperature of 137 °C and a pressure of 2500 kPa, respectively. The turbine inlet pressure was found to critically influence both systems' cost and thermal efficiency.

Kumar et al. [18] performed a thermodynamic analysis of a low-grade heat-driven ORC-VCC system using Engineering Equation Solver (EES). They compared R290,
R152a, and R1234yf with R22 based on cooling capacity and COP. R152a showed strong performance with only 2.8% lower COP than R22, while R290, with zero GWP, had the highest CC but a COP 8.5% lower than R22, making both environmentally friendly alternatives.

Nasir and Kim [19] investigated ORC-powered VCC systems for domestic air conditioning using seven working fluids, evaluating 49 combinations. Isobutane performed best in standalone VCC, while R134a was optimal for the ORC. The R134a/isobutane pair yielded the best combined performance, achieving a COP of 0.219 with dry air and 0.281 with 50% relative humidity at the VCC evaporator inlet.

González et al. [20] studied an ORC-driven VCC system using 16 combinations of 8 working fluids, including traditional and low-GWP options. Each cycle uses a separate fluid, with operating temperature ranges of 60–100 °C for the ORC and – 15 to 0 °C for the VCC. Using PC-SAFT EOS for property estimation, the analysis focused on compressor outlet temperature and COP values. The combination of n-pentane (ORC) and R1233zd (VCC) showed superior performance among the tested pairs.

Kim [21] investigated a combined power generation and cooling system integrating an ORC and a VCC. Thermodynamic analyses were performed for two operating modes: the basic ORC and the combined ORC-VCC, using six working fluids (R245fa, R114, R600, R142b, R152a, and R1234yf). Results indicated that the thermal efficiency of the ORC-VCC system was nearly twice that of the basic ORC. The influence of key thermodynamic parameters, including turbine inlet temperature and pressure, and condensing temperature, on system performance was examined. Additionally, second-law efficiency, cooling capacity, and coefficient of performance were evaluated with varying condensing temperatures.

Choosing the right working fluids is essential for maximizing the efficiency and sustainability of these hybrid systems. Recent studies have examined how different working fluids perform, focusing on their thermodynamic properties and environmental impacts, especially GWP. Karthick et al. [22] looked at the energy efficiency of compression refrigeration cycles using eco-friendly refrigerants, offering insights into low-GWP fluids that are crucial for making cooling and power systems more sustainable. Similarly, Hacıpaşaoğlu [23] studied various fluids for ejector organic Rankine cycles that use low-temperature sources, highlighting how important fluid choice is for boosting the efficiency and environmental friendliness of thermodynamic cycles.

While previous studies have examined the integration of ORC and VCC systems, few have compared the performance of multiple environmentally friendly working fluids in such a hybrid system. This study offers a unique contribution by analyzing the energy and exergy performance of five working fluids, R134a, R1234ze(E), R600a (isobutane), R152a, and R1234yf, highlighting the trade-offs between thermodynamic efficiency and environmental impact. The results offer valuable insights into the advantages of integrating ORC and VCC technologies within a solar-driven system, with a focus on identifying the most suitable working fluid for optimizing performance while reducing environmental impact.
<figure><img src="image_2.png" /><figcaption>**Fig. 1** Schematic of the solar-powered ORC-VCC system</figcaption></figure>

<figure><img src="image_3.png" /><figcaption>**Fig. 2** Temperature-Entropy diagram of the ORC-VCC system</figcaption></figure>

## 2 System description

The schematic diagram of the solar ORC-VCC is shown in Fig. 1, and its T-S diagram is shown in Fig. 2. The primary energy source is solar radiation, captured by a PTC field that heats a thermal oil, which then circulates between the hot and cold thermal energy storage tanks via a solar pump. The hot tank stores high-temperature heat transfer fluid (stream 7), which is directed to a heat exchanger for transferring energy to the ORC working fluid.

In the ORC subsystem, the working fluid is preheated and vaporized in the heat exchanger (stream 1) and expanded in a turbine (stream 2), producing mechanical work that drives an electrical generator. After expansion, the partially expanded vapor enters a recuperator (stream 3), where it preheats the liquid returning from the condenser, enhancing system efficiency. The flow exiting the recuperator (stream 6), now at a lower temperature and pressure, is condensed to a saturated liquid (stream 4), pressurized
by a pump (stream 5), and finally preheated again in the recuperator before being rein-
troduced to the heat exchanger. The VCC subsystem is thermally driven by the ORC
through a coupling between streams 9 and 10.

Superheated vapor (stream 9) enters the evaporator, where it absorbs heat from the environment, enabling the refrigeration function. The vapor is then compressed (stream 10) and passes through the condenser (stream 11), rejecting heat to a cooling water stream (streams 13–14). The resulting saturated liquid (stream 11) expands via an expansion valve, producing a two-phase mixture (stream 12) that reenters the evaporator, thereby closing the refrigeration cycle.

## 2.1 Working fluid selection

In this study, five environmentally friendly low-GWP working fluids are considered for both ORC and VCC cycles: R600a (isobutane), R134a, R152a, R1234yf, and R1234ze(E). These fluids are selected due to their favorable thermophysical properties [2, 3], including moderate critical temperatures, acceptable pressure ranges. All fluids exhibit zero ozone depletion potential (ODP), aligning with the Montreal Protocol [24].

Figure 3 illustrates pressure-temperature curves for all selected fluids, highlighting their comparative thermodynamic behavior. Table 1 summarizes the key properties of each fluid, including molecular weight, critical temperature and pressure, boiling point, GWP, and ODP. These characteristics are critical in assessing the performance, environmental impact, and suitability of each fluid for both power and cooling generation.

<figure><img src="image_4.png" /><figcaption>**Fig. 3** Liquid vapor pressure for the used working fluids</figcaption></figure>
<table><caption>**Table 1** Thermophysical properties of refrigerants [24]</caption><thead><tr><th>Refrigerant</th><th>Molar mass (kg/kmol)</th><th>Normal boiling Point [°C]</th><th>Critical temperature [°C]</th><th>Critical pressure [kPa]</th><th>ODP / GWP</th><th>Safety group</th></tr></thead><tbody><tr><td>R600a</td><td>58.12</td><td>-11.7</td><td>134.7</td><td>3.63</td><td>0 / 20</td><td>A3</td></tr><tr><td>R134a</td><td>102.03</td><td>-26.30</td><td>122.00</td><td>4059</td><td>0 / 1430</td><td>A1</td></tr><tr><td>R152a</td><td>114.04</td><td>-29.5</td><td>94.7</td><td>3382</td><td>0 / 138</td><td>A2</td></tr><tr><td>R1234yf</td><td>66.05</td><td>-24.00</td><td>113.60</td><td>4517</td><td>0 / 1</td><td>A2L</td></tr><tr><td>R1234ze(E)</td><td>114.00</td><td>-19.26</td><td>109.40</td><td>3634</td><td>0 / 7</td><td>A2L</td></tr></tbody></table>

## 3 Mathematical model

Mathematical modeling of the integrated solar ORC-VCC system was performed by applying the first and second laws of thermodynamics, incorporating mass, energy, and exergy balance equations. The simulation and performance evaluation of the system were implemented using MATLAB. All thermophysical properties of the working fluids were obtained from the REFPROP software [25].

To simplify the theoretical model and make the simulation process more efficient, the following assumptions are adopted:

* The system operation is at steady-state conditions;

* The heat and frictional losses in the cycle are negligible;

* The changes in kinetic and potential energies are neglected.

* The pumping, compression, and expansion processes are isentropic;

* The throttling process in the expansion valve is isenthalpic;

* A pinch-point temperature of 5 °C is considered for all the heat exchangers.

### 3.1 Thermodynamic analysis

The available energy from the sun can be calculated as follows:

$$Q_s = A_c \times G_b \qquad (1)$$

where $A_c$ is the aperture area of the collector (m²) and $G_b$ is the beam irradiation (W/m²).

The solar collector efficiency is given by:

$$\eta_c = \frac{Q_c}{Q_s} \qquad (2)$$

where $Q_c$ is the energy absorbed by the HTF in the collector.

Given the temperatures of the HTF at the inlet and outlet of the collector and the mass flow rate through the collector, the heat absorbed in each time interval by the HTF in the solar collector can be calculated as:

$$Q_c (i) = \dot{m}_c (i) \cdot C_p^{HTF} (T_{hot} - T_{cold}) \qquad (3)$$

where:

$T_{hot}$ and $T_{cold}$ are the temperatures of the hot and cold tanks (°C).

$m_c (i)$ is the hourly mass flowrate of the thermal oil (kg/s).

The specific heat capacity of the heat transfer fluid $C_p^{HTF}$ is given by [26]:

$$C_p^{HTF} = 3.3811 \cdot T_m + 1509.7 \qquad (4)$$
The collector efficiency can be determined from the following equation [27]:

$$ \eta_c (i) = 0.73116 - 1.2402 \times 10^{-6} \frac{(\Delta T (i))^3}{G (i)} - 5.4012 \times 10^{-5} \Delta T (i) \quad (5) $$

$\Delta T (i)$ is the difference between ambient temperature, $T_a (i)$, and mean temperature, $T_m$, in the PTC:

$$ \Delta T (i) = T_m - T_a (i) \quad (6) $$

where:

$$ T_m = \frac{(T_{hot} + T_{cold})}{2} \quad (7) $$

The mass flow rate of the thermal oil to the ORC sub-system can be calculated by:

$$ \dot{m}_{HTF} = \frac{\sum_{i=1}^{24} \dot{m}_c (i)}{24} \quad (8) $$

where i denotes the time intervals assumed to be 24 in this study.

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

$$ \dot{Q}_{in} = \dot{m}_{ORC} \cdot (h_6 - h_1) = \dot{m}_{HTF} \cdot C_p^{HTF} (T_{hot} - T_{cold}) \quad (9) $$

where:

$\dot{m}_{ORC}$ denotes the mass flow rate of the organic working fluid in the ORC system.

$h_6$ and $h_1$ are the specific enthalpies of the working fluid at state points 6 and 1, respectively.

$T_1$ and $P_1$ are respectively the temperature and pressure of the turbine inlet stream.

$h_6$ is the enthalpy of the saturated liquid.

The heat recovered in the recuperator is given by the heat balance equation:

$$ \dot{m}_{ORC} \cdot (h_3 - h_2) = \dot{m}_{ORC} \cdot (h_6 - h_5) \quad (10) $$

The effectiveness of the recuperator can be expressed as:

$$ \epsilon = \frac{T_2 - T_3}{T_2 - T_5} \quad (11) $$

The turbine's isentropic efficiency is calculated as:

$$ \eta_t = \frac{h_2 - h_1}{h_{2is} - h_1} \quad (12) $$

$h_{2is}$ is the isentropic outlet enthalpy of the turbine.

The pump's isentropic efficiency is given by:

$$ \eta_p = \frac{h_{5is} - h_4}{h_5 - h_4} \quad (13) $$

$h_{5is}$ is the isentropic outlet enthalpy of the pump.

The net power output of the ORC-VCC system is calculated by:
$$ \dot{W}_{net} = \dot{W}_{turb} - \dot{W}_{pump} - \dot{W}_{comp} \qquad (14) $$

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

$$ \dot{W}_{turb} = \dot{m}_{OCR} (h_1 - h_2) \qquad (15) $$

The power consumed by the pump can be calculated as:

$$ \dot{W}_{pump} = \dot{m}_{OCR} (h_5 - h_4) \qquad (16) $$

The power consumed by the compressor can be calculated as:

$$ \dot{W}_{comp} = \dot{m}_{VCC} (h_{10} - h_9) \qquad (17) $$

$\dot{m}_{VCC}$ denotes the mass flow rate of the refrigerant in the VCC system.

$h_9$ is the inlet enthalpy of the compressor.

$h_{10}$ is the outlet enthalpy of the compressor.

In the present study, the VCC's refrigerant mass flow rate is divided by 20% of the ORC flow rate from the condenser.

$$ \dot{m}_{VCC} = 0.2 \times \dot{m}_{ORC} \qquad (18) $$

The compressor isentropic efficiency is defined as:

$$ \eta_c = \frac{h_{10is} - h_9}{h_{10} - h_9} \qquad (19) $$

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

$$ \dot{Q}_{cond} = (\dot{m}_{ORC} + \dot{m}_{VCC}) (h_{mix} - h_{11}) \qquad (20) $$

$h_{mix}$ is the specific enthalpy at the mixing point of states 3 and 10.

$h_{11}$ is the inlet enthalpy of the compressor.

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

$$ \dot{Q}_{evap} = \dot{m}_{VCC} (h_9 - h_{12}) \qquad (21) $$

$h_{12}$ is the inlet enthalpy of the expansion valve (isenthalpic).

The energy balance of the adiabatic expansion process in the throttling valve gives:

$$ h_{12} = h_{11} \qquad (22) $$

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

$$ \text{COP} = \frac{\dot{Q}_{evap}}{\dot{W}_{comp}} \qquad (23) $$

The thermal efficiency of the VCC system is defined as:

$$ \eta_{th,VCC} = \frac{\text{COP}}{\text{COP}_{max}} \qquad (24) $$

The theoretical maximum COP can be expressed as:
$$COP_{max} = \frac{T_{evap}}{T_{cond} - T_{evap}} \qquad (25)$$

Where $T_{evap}$ and $T_{cond}$ refer to the evaporation and condensation temperatures, respectively.

The thermal efficiency of the ORC system is expressed as:

$$\eta_{th,ORC} = \frac{\dot{W}_{turb} - \dot{W}_{pump}}{\dot{Q}_{in}} \qquad (26)$$

The thermal efficiency of the solar ORC-VCC system is defined by:

$$\eta_{th,sys} = \frac{\dot{W}_{net} + \dot{Q}_{evap}}{Q_s} \qquad (27)$$

## 3.2 Exergetic analysis

In the steady-state condition, mass, energy, and exergy balance equations can be given as:

$$\sum \dot{m}_{in} - \sum \dot{m}_{out} = 0 \qquad (28)$$

$$\dot{Q}_{in} + \dot{W}_{in} + \sum_{in} \dot{m}_{i}h_{i} = \dot{Q}_{out} + \dot{W}_{out} + \sum_{out} \dot{m}_{i}h_{i} \qquad (29)$$

$$\sum_{in} \left(1 - \frac{T_0}{T_i}\right) \dot{Q}_i + \dot{W}_{in} + \sum_{in} \dot{m}_i ex_i = \dot{Q}_i + \dot{W}_{out} + \sum_{out} \dot{m}_i ex_i \qquad (30)$$ $$= \sum_{out} \left(1 - \frac{T_0}{T_i}\right) \dot{Q}_i + \dot{W}_{out} + \sum_{in} \dot{m}_i ex_i + \dot{E}_D$$

where: $\dot{Q}$ is the rate of heat transfer between the control volume and its surroundings, $\dot{W}$ is the rate of work, $h$ is the specific enthalpy.

The term $(1 - \frac{T_0}{T_i}) \dot{Q}_i$ represents the exergy destruction due to heat transfer between the system and the surroundings. This term is non-zero only when there is a temperature difference between the system's inlet temperature ($T_i$) and the surrounding environment temperature ($T_0$). When $T_0$ equals $T_i$, there is no temperature difference, indicating no exergy destruction due to heat transfer.

The rate of exergy loss, $\dot{E}_D$, is proportional to the generated entropy, $\dot{S}_g$, according to the Gouy-Stodola theorem [28]:

$$\dot{E}_D = T_0 \dot{S}_g \qquad (31)$$

The specific flow exergy is evaluated using the following expression:

$$ex = h - h_0 - T_0 (s - s_0) \qquad (32)$$

where $h_0$ and $s_0$ are the specific enthalpy and entropy of the environment, respectively, and.

$h$ and $s$ are the specific enthalpy and entropy at the system state.

The energy and exergy balances in Eqs. (29) and (30) were applied to each component of the system, and the obtained results are summarized in Table 2.
<table><caption>**Table 2** Exergy and energy balance equations</caption><thead><tr><th>Component</th><th>Energy balance</th><th>Exergy balance</th></tr></thead><tbody><tr><td>ORC subsystem</td><td></td><td></td></tr><tr><td>Turbine</td><td>W&#x0307;<sub>turb</sub> = ṁ<sub>OCR</sub> (h<sub>1</sub> - h<sub>2</sub>)</td><td>Ex<sub>d,turb</sub> = ṁ<sub>OCR</sub> (ex<sub>1</sub> - ex<sub>2</sub>) - Ṽ<sub>turb</sub></td></tr><tr><td>Pump</td><td>W&#x0307;<sup>·</sup><sub>pump</sub> = ṁ<sub>OCR</sub> (h<sub>5</sub> - h<sub>4</sub>)</td><td>Ex<sub>d,pump</sub> = ṁ<sub>OCR</sub> (ex<sub>4</sub> - ex<sub>5</sub>) + Ṽ<sup>·</sup><sub>pump</sub></td></tr><tr><td>Recuperator</td><td>ṁ<sub>ORC</sub>·(h<sub>3</sub> - h<sub>2</sub>) = ṁ<sub>ORC</sub>·(h<sub>6</sub> - h<sub>5</sub>)</td><td>Ex<sub>d,recup</sub> = ṁ<sub>OCR</sub> [(ex<sub>2</sub> - ex<sub>3</sub>) - (ex<sub>6</sub> - ex<sub>5</sub>)]</td></tr><tr><td>Heat exchanger</td><td>Q&#x0307;<sub>in</sub> = { ṁ<sub>ORC</sub>·(h<sub>6</sub> - h<sub>1</sub>) ṁ<sub>HTF</sub>·C<sup>HTF</sup><sub>p</sub> (T<sub>hot</sub> - T<sub>cold</sub>)</td><td>Ex<sub>d,exch</sub> = Ḃ<sub>hx</sub> (1 - T<sub>0</sub> / T<sub>hm</sub>) - ṁ<sub>OCR</sub> (ex<sub>1</sub> - ex<sub>6</sub>)</td></tr><tr><td>Total system</td><td>η<sub>ORC</sub> = W&#x0307;<sub>turb</sub> - Ḃ<sub>pump</sub> / Q&#x0307;<sub>in</sub></td><td>Δ ḃ<sub>x</sub> = ∑<sub>i</sub> Δ ḃ<sup>·</sup><sub>i</sub></td></tr><tr><td>VCC Subsystem</td><td>W<sub>comp</sub> = mV<sub>CC</sub> (h<sub>10</sub> - h<sub>9</sub>)</td><td>Ex<sub>d,comp</sub> = ṁ<sub>VCC</sub> (ex<sub>9</sub> - ex<sub>10</sub>) + Ḃ<sub>comp</sub></td></tr><tr><td>Compressor</td><td>Q<sub>cond</sub> = mT<sub>OT</sub> (h<sub>mix</sub> - h<sub>11</sub>)</td><td>Ex<sub>d,cond</sub> = ṁ<sub>TOT</sub> (ex<sub>mix</sub> - ex<sub>11</sub>) - Ḃ<sub>cond</sub> (1 - T<sub>0</sub> / T<sub>mw</sub>)</td></tr><tr><td>Condenser</td><td>Q<sub>evap</sub> = ṁ<sub>VCC</sub> (h<sub>9</sub> - h<sub>12</sub>)</td><td>Ex<sub>d,evap</sub> = ṁ<sub>VCC</sub> (ex<sub>12</sub> - ex<sub>9</sub>) - Ḃ<sub>evap</sub> (1 - T<sub>0</sub> / T<sub>evap</sub>)</td></tr><tr><td>Evaporator</td><td>h<sub>12</sub> = h<sub>11</sub></td><td>Ex<sub>d,valve</sub> = ṁ<sub>VCC</sub> (ex<sub>11</sub> - ex<sub>12</sub>)</td></tr><tr><td>Expansion valve</td><td>COP<sub>VCC</sub> = Ḃ<sub>evap</sub> / Ḃ<sub>comp</sub></td><td>Δ ḃ<sub>x</sub> = ∑<sub>i</sub> Δ ḃ<sup>·</sup><sub>i</sub></td></tr><tr><td>Total system</td><td></td><td></td></tr></tbody></table>
The exergy efficiency of the ORC subsystem can be given by:

$$ \eta_{ex,ORC} = \frac{\dot{W}_{net} + \dot{Q}_{evap}}{\dot{Q}_{in} \left(1 - \frac{T_0}{T_{mh}}\right)} \qquad (33) $$

where $T_{mh}$ is the mean logarithmic temperature of the main flow (solar HTF).

$$ T_{mh} = \frac{T_{hot} - T_{cold}}{\ln\left(\frac{T_{hot}}{T_{cold}}\right)} \qquad (34) $$

The exergy efficiency of the VCC subsystem is given by:

$$ \eta_{ex,VCC} = \frac{\dot{Q}_{evap} \left( \frac{T_0}{T_{evap}} - 1 \right)}{\dot{W}_{comp}} \qquad (35) $$

For the whole system, exergy efficiency takes into account the fuel, which corresponds to the exergy of solar radiation absorbed by the collector, and the product, which is the sum of the effective mechanical power and the amount of exergy in the VCC evaporator.

$$ \eta_{ex,sys} = \frac{\dot{W}_{net} + \dot{E}x_{evap}}{\dot{E}x_{in}} \qquad (36) $$

The inlet exergy to the system can be calculated by Petela's formula for undiluted direct solar radiation [29]:

$$ \dot{E}x_{in} = A_c G_b \left( 1 + \frac{1}{3} \left( \frac{T_0}{T_s} \right)^4 - \frac{4}{3} \left( \frac{T_0}{T_s} \right) \right) \qquad (37) $$

where $T_s$ is the equivalent temperature of the sun as a black body (~5800 K).

# 4 Results and discussion

This section presents the simulation results and provides a comprehensive analysis of the thermodynamic behavior of the solar ORC-VCC system using various working fluids. The influence of key parameters is examined in terms of net power output, cooling capacity, COP, and exergy efficiency. An exergy analysis is also carried out to identify the primary sources of irreversibility and assess the overall thermodynamic performance of the system. A comparative evaluation of the selected working fluids is conducted to highlight differences in system response and efficiency. The primary input parameters and assumptions used in the simulation are outlined in Table 3.

## 4.1 Daily solar thermal simulation

Figure 4 illustrates the hourly variation of both solar radiation and ambient temperature for a typical day in June at the Ghardaïa site. The solar radiation exhibits a characteristic bell-shaped curve, peaking at approximately 2500 W/m² around midday. The ambient temperature follows a similar trend, increasing from a minimum in the early morning to a peak of about 31.7 °C around midday, and then gradually decreasing.

shows the hourly changes in solar collector efficiency and the HTF mass flow rate. Both parameters begin to increase between 05:00 and 06:00 h, closely following the onset
<table><caption>**Table 3** System operating parameters for the solar ORC-VCC configuration</caption><thead><tr><td>Parameter</td><td>Value</td></tr></thead><tbody><tr><td>Solar collector area, m²</td><td>100</td></tr><tr><td>Cold tank temperature, °C</td><td>290</td></tr><tr><td>Hot tank temperature, °C</td><td>390</td></tr><tr><td>Turbine inlet temperature, °C</td><td>135</td></tr><tr><td>Turbine inlet pressure, kPa</td><td>2300</td></tr><tr><td>Isentropic turbine efficiency</td><td>0.8</td></tr><tr><td>Isentropic compressor efficiency</td><td>0.75</td></tr><tr><td>Isentropic pump efficiency</td><td>0.85</td></tr><tr><td>Recuperator efficiency</td><td>0.8</td></tr><tr><td>Condensing temperature, °C</td><td>35</td></tr><tr><td>Evaporation temperature, °C</td><td>5</td></tr><tr><td>Condenser water inlet temperature, °C</td><td>25</td></tr></tbody></table>

<figure><img src="image_5.png" /><figcaption>**Fig. 4** Hourly variation of solar radiation and ambient temperature for a typical day in June (Ghardaïa region)</figcaption></figure>

of solar radiation, and decline to zero by approximately 19:00–20:00 h. The collector efficiency rapidly rises and stabilizes within the range of 68–70% between 08:00 and 17:00 h, indicating optimal operating conditions during the core daylight period. Meanwhile, the HTF mass flow rate increases more gradually, reaching a peak of approximately 0.7 kg/s between 12:00 and 13:00 h, coinciding with the maximum solar irradiance.

## 4.2 Effect of the turbine inlet temperature

Figure 6 illustrates the influence of turbine inlet temperature on the net power output of the solar ORC-VCC system using different working fluids: R600a, R134a, R152a, R1234yf, and R1234ze(E). It can be observed that increasing the turbine inlet temperature from 120 °C to 150 °C leads to a corresponding rise in net power output for all working fluids. This is due to the greater enthalpy difference across the turbine at higher temperatures, which allows more work to be extracted. Among the tested fluids, R600a
<figure><img src="image_6.png" /><figcaption>**Fig. 5** Variation of HTF flow rate and solar collector efficiency with time of day (Ghardaïa region)</figcaption></figure>

<figure><img src="image_7.png" /><figcaption>**Fig. 6** Effect of turbine inlet temperature on net power output</figcaption></figure>

offers the highest net power output across the entire temperature range, increasing from 6.4 kW to 7.4 kW. On the other hand, R134a and R1234yf consistently show the lowest power outputs. This can be explained by their thermodynamic properties, which make them less efficient in converting heat into work.
Figure 7 shows the effect of turbine inlet temperature on the exergy efficiency of the solar ORC-VCC system. As with net power output, exergy efficiency increases steadily with rising turbine inlet temperature for all working fluids. This trend is explained by the fact that higher inlet temperatures provide a greater potential for work recovery, as more of the supplied heat can be effectively converted into mechanical energy. R600a also achieves the highest exergy efficiency, increasing from around 9.0% at 120 °C to nearly 10.0% at 150 °C. R1234ze(E) shows moderate performance, while R134a and R1234yf consistently record the lowest exergy efficiencies. These lower values reflect the same trends observed in net power output and are mainly attributed to the thermophysical limitations of these fluids, which reduce their effectiveness in exergy conversion under the given ORC operating conditions.

## 4.3 Effect of the turbine inlet pressure

Figure 8 illustrates the impact of turbine inlet pressure on the net power output of the solar ORC-VCC system for the five refrigerants considered previously. A consistent increase in power output is observed with rising inlet pressure across all working fluids, primarily due to the enhanced enthalpy drop across the turbine. Among the refrigerants, R600a shows the highest net power output, ranging from approximately 5.4 kW to 7.3 kW. R1234ze(E) demonstrates moderate performance. In contrast, R134a and R1234yf yield the lowest values.

Figure 9 illustrates the influence of turbine inlet pressure on the exergy efficiency of the solar ORC-VCC system for the five refrigerants analyzed earlier. Increasing the turbine inlet pressure leads to a notable improvement in exergy efficiency for all working fluids. This is attributed to the higher-pressure ratio across the turbine, which elevates the quality of the energy input and enhances its ability to be converted into useful work.

<figure><img src="image_8.png" /><figcaption>**Fig. 7** Effect of turbine inlet temperature on exergy efficiency</figcaption></figure>
<figure><img src="image_9.png" /><figcaption>**Fig. 8** Effect of turbine inlet pressure on net power output</figcaption></figure>

<figure><img src="image_10.png" /><figcaption>**Fig. 9** Effect of turbine inlet pressure on exergy efficiency</figcaption></figure>
Among the refrigerants, R600a consistently achieves the highest exergy efficiency, rising from approximately 7.8% to 10.0%. R1234ze(E) follows with efficiencies between 5.5% and 8.9%, while R152a shows moderate performance. R134a and R1234yf exhibit the lowest efficiencies.

## 4.4 Effect of the VCC condensing temperature

Figure 10 illustrates the effect of the VCC condensing temperature on its COP for all refrigerants studied. It can be seen that increasing the condensing temperature leads to a consistent decrease in COP for all refrigerants. This decline is primarily due to the higher compressor pressure ratio at elevated condensing temperatures, which increases the work input required to maintain the same cooling capacity, thus reducing efficiency. Among the refrigerants, R600a consistently demonstrates the highest COP, starting above 7.0 at 30 °C and decreasing to just above 2.5 at 60 °C. R1234ze(E) and R152a show intermediate COP values, whereas R134a and R1234yf exhibit the lowest performance throughout the temperature range.

Figure 11 illustrates the influence of the VCC condensing temperature on the cooling capacity of the system for the five refrigerants under consideration. As the condensing temperature increases, the cooling capacity declines. This trend is primarily attributed to the increased compression ratio required, which leads to a lower mass flow rate through the evaporator. Among the refrigerants, R600a consistently achieves the highest cooling capacity, starting above 8.4 kW at 30 °C and decreasing to approximately 7.0 kW at 60 °C. R134a ranks second, followed by R1234ze(E) and R152a, while R1234yf exhibits the lowest cooling performance throughout the temperature range.

Figure 12 presents the effect of VCC condensing temperature on the exergy efficiency of the integrated ORC-VCC system for the same refrigerants. As the condensing

<figure><img src="image_11.png" /><figcaption>**Fig. 10** Effect of VCC condensing temperature on the coefficient of performance</figcaption></figure>
<figure><img src="image_12.png" /><figcaption>**Fig. 11** Effect of VCC condensing temperature on cooling capacity</figcaption></figure>

<figure><img src="image_13.png" /><figcaption>**Fig. 12** Effect of VCC condensing temperature on system-level exergy efficiency</figcaption></figure>
temperature increases, the exergy efficiency consistently decreases for all working fluids. This decline is primarily due to the reduced temperature gradient between the heat source and sink, which limits the system's potential to convert thermal energy into useful work, following the second law of thermodynamics. R600a demonstrates the highest exergy efficiency across the examined range, decreasing from approximately 8.7% at 30 °C to slightly above 5.0% at 60 °C. R1234ze(E) maintains intermediate values, while R134a, R152a, and R1234yf yield comparatively lower exergy efficiencies.

## 4.5 Effect of the VCC evaporation temperature

Figure 13 illustrates the effect of the VCC evaporation temperature on the COP for the five selected refrigerants. As the evaporation temperature increases from –15 °C to 15 °C, the COP increases for all refrigerants. R600a and R152a yield the highest COPs, reaching values close to 9.8 at 15 °C. R1234ze(E) performs moderately well, while R134a and R1234yf consistently record the lowest COP throughout the examined range.

Figure 14 presents the influence of the VCC evaporation temperature on the cooling capacity of the system. Increasing the evaporation temperature from –15 °C to 15 °C results in a continuous rise in cooling capacity across all working fluids. R152a achieves the highest cooling capacity, increasing from approximately 8.0 kW at –15 °C to nearly 8.7 kW at 15 °C. R134a follows closely, while R600a and R1234ze(E) exhibit moderate values. R1234yf records the lowest cooling capacity.

Figure 15 presents the relationship between VCC evaporation temperature (–15 °C to 15 °C) and exergy efficiency in the ORC-VCC system. A consistent positive correlation is observed across all refrigerant candidates, with efficiency increasing monotonically with temperature elevation. This improvement may be attributed to the higher quality of heat absorbed at elevated temperatures, which reduces exergy destruction within the

<figure><img src="image_14.png" /><figcaption>**Fig. 13** Effect of VCC evaporation temperature on the coefficient of performance</figcaption></figure>
<figure><img src="image_15.png" /><figcaption>**Fig. 14** Effect of VCC evaporation temperature on cooling capacity</figcaption></figure>

<figure><img src="image_16.png" /><figcaption>**Fig. 15** Effect of VCC evaporation temperature on system-level exergy efficiency</figcaption></figure>
evaporator and compressor. R600a consistently delivers the highest exergy efficiency, increasing from approximately 8.0% at – 15 °C to nearly 8.9% at 15 °C. R1234ze(E) performs second best, with efficiencies ranging from about 6.5% to 7.3%, while R152a, R134a, and R1234yf yield lower values.

## 4.6 Effect of the VCC/ORC mass flow rate

Figure 16 illustrates the influence of the VCC/ORC mass flow ratio on the COP of the integrated ORC-VCC system for the five selected refrigerants. As the mass flow ratio increases from 0.05 to 0.5, the COP generally rises for all refrigerants. This trend can be attributed to the improved utilization of the ORC-generated power in driving the VCC, resulting in more effective energy exchange between the two subsystems. Among the working fluids, R152a consistently achieves the highest COP values, increasing from approximately 2.9 to over 5.5 at the upper end of the ratio range. R134a and R600a show comparable intermediate performance, while R1234yf and R1234ze(E) exhibit relatively lower COP.

Figure 17 presents the effect of the VCC/ORC mass flow ratio on the overall exergy efficiency of the system. As the ratio increases from 0.05 to 0.5, a consistent decline in exergy efficiency is observed for all working fluids, ranging from approximately 10.5% to 5.5%. This is primarily attributed to the increased power demand of the VCC at higher flow ratios, which intensifies irreversibilities in compression and heat exchange processes. Among the fluids, R600a exhibits the highest exergy efficiency across the entire range, followed by R1234ze(E) and R152a, while R134a and R1234yf perform less favorably.

Figure 18 depicts the influence of the VCC/ORC mass flow rate ratio on both the cooling capacity and net power output. Overall, an increase in the mass flow rate ratio

<figure><img src="image_17.png" /><figcaption>**Fig. 16** Effect of VCC/ORC mass flow rate on the coefficient of performance</figcaption></figure>
<figure><img src="image_18.png" /><figcaption>**Fig. 17** Effect of VCC/ORC mass flow rate on exergy efficiency</figcaption></figure>

<figure><img src="image_19.png" /><figcaption>**Fig. 18** Effect of VCC/ORC mass flow rate on net power output and cooling capacity</figcaption></figure>

induces an approximately linear increase in cooling capacity, while the net power output decreases progressively across all refrigerants analyzed. This opposite behavior is attributed to the increasing fraction of power generated by the ORC cycle being consumed to operate the VCC, which enhances the cooling effect while concurrently reducing the
net power available at the output. Among the refrigerants considered, R600a demonstrates the highest net power output, making it particularly suitable for applications prioritizing electricity generation. In contrast, R152a and R134a exhibit superior cooling capacities, exceeding 24 kW at a mass flow rate ratio of 0.5, thus being more appropriate for cooling-centric systems. R1234ze(E) presents intermediate performance, providing a balanced compromise between power generation and cooling capacity, whereas R1234yf generally exhibits lower overall efficiency.

Furthermore, for each working fluid, there is a specific intersection point where the cooling capacity and net power output curves intersect. This point corresponds to an optimal VCC/ORC mass flow rate ratio that achieves a thermodynamic compromise between maximizing electrical power generation and cooling performance. Selecting a mass flow rate ratio near this optimum enables the hybrid ORC-VCC system to balance its dual objectives effectively.

## 4.7 Exergy analysis

A detailed exergy analysis was conducted to assess the thermodynamic irreversibilities across the hybrid solar ORC-VCC system and to identify areas where performance improvements are most critical. The exergy destruction for the three main subsystems: the solar collector, the ORC unit, and the VCC system, is presented in Fig. 19. Among these, the solar collector exhibits the highest exergy destruction, reaching approximately 58 kW. This is mainly due to the inherent irreversibilities associated with the solar energy conversion process. The ORC system follows with values ranging between 22 and 25 kW, while the VCC subsystem exhibits the lowest exergy destruction, limited to around 4–5 kW. Moreover, the choice of working fluid has a negligible impact on

<figure><img src="image_20.png" /><figcaption>**Fig. 19** Specific exergy destruction in the solar ORC-VCC system</figcaption></figure>
the collector's performance, underlining that optimizing the solar collector is critical for enhancing the overall efficiency of the system.

In addition, the exergy destruction for each component within the solar ORC-VCC system is illustrated in Fig. 20 for the working fluids considered. For all fluids, the heat exchanger exhibits the highest exergy destruction, with values ranging from 19.35 kW for R600a to 22.11 kW for R134a, primarily due to the pronounced irreversibilities associated with heat transfer across finite temperature differences. The evaporator and condenser also contribute significantly to the overall exergy destruction, particularly for R152a and R134a, where the values exceed 2.1 kW and 1.7 kW, respectively.

The turbine losses vary from 1.33 kW to 1.99 kW, while the compressor shows relatively consistent destruction levels around 0.36–0.38 kW, reflecting mechanical losses and work input demands. The pump and valve remain the least impactful components, with exergy destruction values below 0.2 kW for all fluids. Although R1234ze(E) exhibits marginally lower destruction in several components, the results indicate that optimizing the heat exchanger, evaporator, and condenser with careful selection of the working fluid is essential for improving the system’s thermodynamic performance.

## 4.8 Performance comparison

A comprehensive comparison of the system's thermodynamic and exergetic performance across the evaluated working fluids is presented in Tables 4 and 5. R600a demonstrates the highest net power output (7.08 kW) and exhibits favorable thermal and exergy efficiencies within the ORC subsystem (13.92% and 28.29%, respectively). Additionally, its overall system exergy efficiency (9.87%) is the highest among all fluids, making it a suitable candidate for applications prioritizing power generation and exergy recovery.

<figure><img src="image_21.png" /><figcaption>**Fig. 20** Exergy destruction by component and working fluid</figcaption></figure>
<table><caption>**Table 4** Performance metrics of the ORC-VCC system</caption><thead><tr><th>Working fluid</th><th>ṁ<sub>ORC</sub> (kg/s)</th><th>ṁ<sub>VCC</sub> (kg/s)</th><th>Ẇ<sub>net</sub> (kW)</th><th>Ẇ<sub>comp</sub>(kW)</th><th>Q̇<sub>evap</sub> (kW)</th><th>COP</th></tr></thead><tbody><tr><td>R600a</td><td>0.15526</td><td>0.031052</td><td>7.0822</td><td>1.4291</td><td>8.6132</td><td>6.0271</td></tr><tr><td>R134a</td><td>0.30076</td><td>0.060151</td><td>4.4166</td><td>1.5489</td><td>9.1722</td><td>5.9216</td></tr><tr><td>R152a</td><td>0.19338</td><td>0.038676</td><td>4.9279</td><td>1.5756</td><td>9.6102</td><td>6.0996</td></tr><tr><td>R1234yf</td><td>0.35786</td><td>0.071571</td><td>4.627</td><td>1.493</td><td>8.5081</td><td>5.6986</td></tr><tr><td>R1234ze(E)</td><td>0.31345</td><td>0.062691</td><td>5.5458</td><td>1.4844</td><td>8.7637</td><td>5.9039</td></tr></tbody></table>

<table><caption>**Table 5** Thermal and exergy efficiencies for the solar ORC-VCC system</caption><thead><tr><th>Working fluid</th><th>η<sub>th,ORC</sub><br>(%)</th><th>η<sub>ex,ORC</sub><br>(%)</th><th>η<sub>th,VCC</sub><br>(%)</th><th>η<sub>ex,VCC</sub><br>(%)</th><th>η<sub>th,sys</sub><br>(%)</th><th>η<sub>ex,sys</sub><br>(%)</th></tr></thead><tbody><tr><td>R600a</td><td>13.923</td><td>28.292</td><td>65.005</td><td>70.567</td><td>17.783</td><td>9.8721</td></tr><tr><td>R134a</td><td>9.7584</td><td>19.829</td><td>63.868</td><td>69.332</td><td>15.396</td><td>6.6994</td></tr><tr><td>R152a</td><td>10.638</td><td>21.617</td><td>65.787</td><td>71.416</td><td>16.472</td><td>7.3859</td></tr><tr><td>R1234yf</td><td>10.011</td><td>20.343</td><td>61.463</td><td>66.721</td><td>14.883</td><td>6.8613</td></tr><tr><td>R1234ze(E)</td><td>11.5</td><td>23.369</td><td>63.677</td><td>69.125</td><td>16.213</td><td>8.019</td></tr></tbody></table>

On the other hand, R152a achieves the highest COP (6.10) due to its favorable compressor work and heat input characteristics, reflecting its superior performance in cooling scenarios. However, its net power output remains moderate (4.93 kW), and its system-level exergy efficiency (7.39%) is lower than that of R600a. R1234ze(E) provides a balanced performance with moderate values across all indicators, including a net power output of 5.55 kW and a system exergy efficiency of 8.02%. R134a and R1234yf yield the lowest net power outputs (4.42 kW and 4.63 kW, respectively) and also exhibit the lowest overall exergy efficiencies, at 6.99% and 6.86%, respectively.

In terms of the VCC subsystem, R152a again shows the highest thermal (71.42%) and exergy (17.45%) efficiencies, reinforcing its suitability for cooling-dominated configurations. The results confirm that no single fluid optimizes all performance criteria, emphasizing the trade-offs between power output, cooling capacity, and system-wide efficiency. The working fluid selection should therefore be aligned with the dominant operational objective, whether power recovery, refrigeration, or combined energy performance.

# 5 Environmental analysis

The environmental impact of the solar ORC-VCC system is evaluated through three key metrics: (1) direct emissions (GWP), (2) indirect emissions (energy-related CO₂ savings), and (3) comparative sustainability of the working fluids.

## 5.1 Direct emissions

The selected refrigerants' environmental footprints are quantified using their 100-year GWP values (Table 1). Key observations are:

* R134a has the highest GWP (1430), making it environmentally unfavorable despite its thermodynamic performance.

* R1234ze(E) (GWP = 7) and R1234yf (GWP = 1) offer negligible climate impact, complying with international regulations (e.g., EU F-Gas Regulation).

* R600a (GWP = 20) and R152a (GWP = 138) are low-GWP alternatives but require careful handling due to flammability (safety group A2/A3).
<table><caption>**Table 6** Environmental and performance trade-offs of working fluids</caption><thead><tr><td>Fluid</td><td>GWP</td><td>Safety</td><td>Net power (kW)</td><td>COP</td><td>Exergy efficiency (%)</td><td>Recommendation</td></tr></thead><tbody><tr><td>R600a</td><td>20</td><td>A3</td><td>7.08</td><td>6.03</td><td>9.87</td><td>Best for power generation</td></tr><tr><td>R152a</td><td>138</td><td>A2</td><td>4.93</td><td>6.10</td><td>7.39</td><td>Best for cooling</td></tr><tr><td>R1234ze(E)</td><td>7</td><td>A2L</td><td>5.55</td><td>5.90</td><td>8.02</td><td>Best eco-friendly compromise</td></tr><tr><td>R1234yf</td><td>1</td><td>A2L</td><td>4.63</td><td>5.70</td><td>6.86</td><td>Low GWP but moderate performance</td></tr><tr><td>R134a</td><td>1430</td><td>A1</td><td>4.42</td><td>5.92</td><td>6.70</td><td>Least sustainable</td></tr></tbody></table>

## 5.2 Indirect emission

The system's solar-driven design reduces reliance on fossil fuels. Assuming a 100 m² PTC field:

* The system offsets ~3.2–4.1 tons of CO₂/year compared to grid-powered VCC systems (based on Algeria's grid emission factor: 0.48 kg CO₂/kWh).

* Exergy destruction in the solar collector (58 kW) implies a trade-off; optimizing collector efficiency could further reduce embodied emissions.

## 5.3 Comparative sustainability assessment

A multi-criterion ranking (Table 6) balances thermodynamic performance and environmental impact:

## 5.4 Lifecycle considerations

* While R600a and R152a have higher GWPs than HFOs, their thermodynamic performance may justify use in controlled environments.

* Solar ORC-VCC systems typically operate for 20–25 years, with refrigerant leakage rates <5%/year. Over this period, R1234ze(E) reduces cumulative emissions by 92% compared to R134a.

* R1234ze(E) and R1234yf are future-proof under phasedown schedules (e.g., U.S. AIM Act, EU F-Gas).

## 5.5 Recommendations for sustainable deployment

* For power-dominant applications, prioritize R600a but implement leak detection (flammability risk: A3).

* For cooling-dominant systems, R152a is optimal if safety measures (A2) are enforced.

* For eco-design, R1234ze(E) offers the best balance of efficiency, safety (A2L), and compliance.

## 6 Conclusion

In the present study, the thermodynamic performance of a solar-assisted ORC-VCC system was investigated using five working fluids: R600a (isobutane), R134a, R152a, R1234yf, and R1234ze(E). The analysis examined the effects of turbine inlet conditions, VCC condensing and evaporation temperatures, and the VCC/ORC mass flow rate ratio on net power output, cooling capacity, COP, and exergy efficiency. The main findings that can be drawn from this study are as follows:
* Increasing the turbine inlet temperature and pressure improved both net power output and exergy efficiency across all fluids, with R600a consistently showing the best performance. In contrast, higher condensing temperatures reduced COP, cooling capacity, and efficiency, emphasizing the need to maintain low condensing conditions.

* Increasing the evaporation temperature enhanced COP and efficiency, particularly for R600a and R152a.

* The VCC/ORC mass flow rate ratio had a dual effect: increasing the ratio improved cooling capacity but reduced power output and exergy efficiency. For each fluid, a best compromise point was found where cooling capacity and power output intersected, indicating an optimal balance between thermal and electrical objectives.

* Exergy analysis showed the solar collector as the main source of irreversibility, followed by the ORC and VCC subsystems.

* The environmental analysis revealed that R600a achieves superior power generation (7.08 kW net output, 9.87% exergy efficiency), while R152a excels in cooling performance (COP 6.10). Meanwhile, R1234ze(E) emerges as the most balanced eco-friendly alternative with a GWP of 7 and an exergy efficiency of 8.02%.

These results highlight the importance of effective solar energy use and careful fluid selection to improve the performance of integrated ORC-VCC systems. Some fluids proved more suitable under the studied conditions, making them promising candidates for combined solar cooling and power generation. However, the study's steady-state assumptions and idealized modeling conditions limit the generalization of the results to real-world transient conditions. Further research should focus on conducting comprehensive economic assessments and analyzing the impact of real-world transient operating conditions to facilitate the large-scale implementation of solar-assisted ORC-VCC systems. Future work could also include dynamic simulations and environmental comparisons under varying operational conditions.

**Author contributions**

Merouane Saka: Conceptualization, Methodology, Software, Validation, Formal analysis, Investigation, Resources, Data curation, Writing—original draft, Visualization. Zakaria Triki: Conceptualization, Methodology, Software, Validation, Formal analysis, Investigation, Resources, Data curation, Writing—original draft, Visualization. Zineb Fergani: Conceptualization, Methodology, Software, Validation, Formal analysis, Investigation, Resources, Data curation, Writing—Review & Editing, Visualization, Supervision, Project administration. Kamilia Madi: Conceptualization, Methodology, Supervision, Writing—review & editing. Jie Zhang: Conceptualization, Methodology, Supervision, Writing—review & editing, Project administration. Abdeltif Amrane: Conceptualization, Methodology, Supervision, Writing—review & editing, Project administration. Hichem Tahraoui: Conceptualization, Methodology, Software, Validation, Formal analysis, Investigation, Writing—original draft, Supervision, Project administration. All authors approved the final version.

**Funding**

This research received no external funding.

**Data availability**

The data and materials for the current study are included in this published article.

**Declarations**

**Ethics approval and consent to participate**

Not applicable. This manuscript does not involve research with human participants or animals.

**Consent to publish**

Not applicable.

**Competing interests**

The authors declare no competing interests.

**Clinical trial number**

Clinical trial is not applicable.
Received: 13 June 2025 / Accepted: 22 September 2025

Published online: 17 October 2025

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