Introduction

Air conditioning is a widely adopted technology that delivers cooling and comfort in various settings, encompassing residential, commercial, and industrial buildings. Pérez-García et al. [1] highlights the staggering figure of 1.5 billion refrigerators and freezers currently in use, underscoring the profound impact of refrigeration systems on human life. Besides providing personal comfort during hot days, air conditioning plays a vital role in preserving food and medicine [2]. Freezing has long been recognized as the most effective technique for long-term food storage, as evidenced by the consistent increase in consumption of frozen foods across industrialized nations [3][4]. However, frozen foods inevitably experience a decline in quality compared to their fresh counterparts due to food tissue damage and drip loss during thawing, primarily caused by the formation of ice [5]. To mitigate the quality disparity between fresh and frozen foods, rapid freezing and storage at low temperatures are imperative [6]. To ensure food preservation, maintaining temperatures of at least -18°C or below is essential, necessitating continuous operation typically spanning 24 hours a day [1]. The conventional approach involves employing a refrigeration system to extract heat from a given area and cool it under controlled conditions [7].

However, traditional cooling systems such as vapor compression (VC) refrigeration systems consume substantial amounts of energy, resulting in high electricity demand and greenhouse gas emissions [8]. The International Energy Agency (IEA) [9] reported that approximately one-fifth of the total electricity consumed in buildings globally can be attributed to the use of air conditioners and electric fans for maintaining comfortable temperatures. Addressing the issue of greenhouse gas emissions and tackling climate change while sustaining economic development presents a critical challenge, as nearly 40% of the potential for reducing such emissions lies in enhancing energy efficiency [10].

Growing concerns about global warming and ozone depletion led to the adoption of the Montreal Protocol in 1987, aiming to regulate the use of refrigerants [11]. Factors such as flammability, ozone depletion potential (ODP), and global warming potential (GWP) have been consistently monitored in the selection and utilization of refrigerants to address environmental concerns [12][13][14][15]. In the past, refrigeration systems commonly employed ozone-depleting substances like chlorofluorocarbons (CFCs) and hydrofluorocarbons (HFCs) [16]. While CFCs were completely phased out by 2010, certain HFCs, such as Methylene Fluoride (R-32), are still accepted by the U.S. Environmental Protection Agency (EPA) [17]. R-32 also boasts a reduced Global Warming Potential (GWP) of 675 compared to typical refrigerants. Moreover, the refrigerant system requires a smaller charge due to R32’s lower density compared to conventional options [18].

Although HFCs exhibit low or zero ODP, their GWP remains relatively high. Consequently, the Kigali Amendment, signed by over 150 countries in 2016, was established to reduce the production and consumption of HFCs [19]. Subsequently, numerous refrigerants have been developed as replacements. However, the introduction of new refrigerants often raises concerns about flammability or necessitates the redesign of components such as compressors, heat exchangers, or expansion devices [20]. Consequently, certain HFCs continue to be used as refrigerants in various parts of the world.

In a study conducted by Mishra et al. [22], it was found that the coefficient of performance (COP) of R-32 is higher when compared to R-410A. Additionally, R-32 was found to have a lower global warming potential (GWP) compared to R-410A. On the other hand, Gaurav and Kumar [23] noted that R-134a remains widely favored for mobile air conditioning systems due to its exceptional thermal stability, non-corrosive properties, and low toxicity [24]. Hadi et al. [25] also reported that R-32 exhibits lower COP for heat pump applications but is more environmentally friendly than R-134a, while R-134a is safer in terms of its impact on safety. Table 1 provides an overview of the physical and environmental properties associated with both refrigerants.

Several studies have also demonstrated the potential of R-32 as an alternative working fluid to R-134a in energy conversion and refrigeration-related systems. Al Azhari et al. [26] studies on low- to medium-temperature energy systems, including Organic Rankine Cycle applications, have shown that R-32 outperforms R-134a in terms of net power output under comparable operating conditions. In addition to its thermodynamic advantages, Life Cycle Climate Performance (LCCP) analyses indicate that R-32 has lower direct emissions and improved environmental performance compared to R-134a. Prabha and Rambabu [27] evaluated the suitability of R32 as a replacement for R-134a in a 1.5-ton room air-conditioning system using both experimental testing and simulation. Their results show that R32 provides higher performance efficiency and lower compressor power demand than R-134a, and, as an ASHRAE A2L-classified refrigerant with a shorter atmospheric lifetime, represents an effective option for retrofitting air-conditioning systems originally designed for R-134a.

Moreover, Rayhan et al. [28] compared R-134a, R-744, and R-32 for fishing boat refrigeration systems using thermodynamic simulations across varying pressures (800–1000 kPa) and operating temperatures. The results show that R-32 achieves higher COP and lower energy consumption than R-134a, while R-744 exhibits the poorest performance and practical challenges for Indonesian fishing vessels. Bolaji [29] experimentally evaluated R-32 and R-152a as alternatives to R-134a in a domestic refrigerator originally designed for R-134a. Results showed that, compared to R-134a, R-32 achieved the target ISO design temperature and pull-down time more slowly and exhibited an average COP about 8.5% lower, indicating reduced efficiency under domestic refrigeration conditions. Although R-32 is environmentally preferable to R-134a, the findings suggest that its performance in small domestic refrigerators may be inferior, highlighting that its suitability depends strongly on system design and operating conditions.

In terms of heat transfer performance, an experimental investigation of condensation heat transfer on a horizontal tube conducted by Mohammed and Kumar [30] demonstrated that R-32 exhibits significantly superior heat-transfer performance compared to R-134a under saturation temperatures of 40–50 °C and varying wall sub-cooling conditions. At identical operating conditions, R-32 achieved 50–55% higher condensing-side heat-transfer coefficients than R-134a, primarily due to its favourable thermophysical properties. An experimental study by Shete et al. [31] demonstrated that R-32 exhibits boiling heat transfer characteristics comparable to or superior to those of R-134a, particularly on enhanced tube surfaces, where higher nucleate boiling heat transfer coefficients were observed across a range of heat fluxes and saturation conditions; these results indicate that R-32 can offer improved evaporator heat transfer performance relative to R-134a, supporting its suitability as a low-GWP alternative refrigerant in refrigeration and air-conditioning applications. Experimental investigations on flow condensation comparing pure R-32 and R-134a conducted by Shao and Granyd [32] have shown that R-134a generally experiences higher pressure drops due to its lower vapor density, whereas R-32 exhibits more favourable pressure drop characteristics under similar operating conditions, while differences in heat transfer performance are influenced by mass flux, heat flux, and flow regime; overall heat transfer degradation results from the combined effects of temperature and concentration gradients, slip effects, and non-ideal thermophysical properties rather than any single dominant factor.

This paper presents the thermodynamic design of a cold storage facility for fish, aiming to develop an optimized and efficient storage system that fulfils the specific requirements for preserving fish quality and extending shelf life. The study utilizes refrigeration systems operating with R-32 and R-134a. The vapor compression system, a well-established and widely used system in thermodynamic engineering, is comprehensible and consists of components such as the evaporator, condenser, compressor, and expansion valve [33]. A higher coefficient of performance (COP) in VC systems indicates reduced energy consumption per unit of refrigeration capacity [34][35]. The thermodynamic properties of refrigerants, including enthalpy, entropy, temperature, and pressure, significantly influence the COP.

The research aims to investigate various design parameters, encompassing refrigerant characteristics, refrigeration systems, airflow patterns, and performance evaluation. Its objective is to determine the most suitable combination of these factors to ensure optimal conditions for fish storage. Furthermore, the research addresses challenges related to energy efficiency. Through comprehensive analyses and simulations, the study seeks to generate data and insights that inform the development of a cold storage facility specifically tailored to the unique requirements of fish storage. The ultimate goal is to provide practical design recommendations and guidelines that enhance the performance and functionality of the cold storage facility, leading to improved fish quality, reduced waste, and enhanced economic viability for the fisheries industry.

Methodology

System Description

Figure 1 illustrates a cold storage system that utilizes a vapor compression mechanism. The system operates based on the fundamental principle of utilizing fluid characteristics, such as the boiling temperature associated with pressure and the latent heat of phase change, to enable heat transfer from a lower temperature region to a higher temperature region within the cycle [36].

In cold storage, the refrigerants absorb the heat and undergo a phase change from a liquid state to a vapor in the evaporator. The vapor is then compressed by the compressor, resulting in an increase in pressure and temperature of the refrigerant. This high-pressure and high-temperature refrigerant vapor is directed to the condenser, where it releases heat to the surrounding environment and condenses into a high-pressure liquid. Typically, this heat transfer occurs through a heat exchanger with the assistance of a fan or cooling medium. The high-pressure liquid refrigerant then passes through the expansion valve, where its pressure and temperature are rapidly reduced. This expansion causes a pressure drop, transforming the refrigerant into a low-pressure mixture of liquid and vapor. The refrigerant undergoes this cycle of compression, condensation, expansion, and evaporation repeatedly, facilitating the transfer of heat from the lower temperature region (the evaporator) to the higher temperature region (the condenser). This cyclic process enables the achievement of cooling or refrigeration.

The objective of the designed cold storage system was to maintain a temperature of -20 ᵒC while accommodating a daily storage capacity of 2 tons of fish. This goal had to be achieved even in an ambient temperature of 30 ᵒC. In the studied refrigeration system, several conditions were set; R-32 and R-134a were chosen as the refrigerants; air humidity was disregarded; the evaporator temperature was set at -30 ᵒC; the condenser temperature was set at 35 ᵒC; and the compressor isentropic efficiency was assumed to be 90% [37].

Numerical Equations

The following assumptions are used in the cycle calculations:

• Steady-State Operation

  The refrigeration cycle calculations assume steady-state operation, where the system has reached a stable operating condition without significant temperature, pressure, or parameter variations. Start-up and shut-down transient effects are neglected for simplicity.

• Negligible Pressure Drops

  Pressure drops at the condenser and evaporator in the refrigeration system are assumed to be negligible.

• Negligible Heat Transfer Resistance

  Heat transfer between the refrigerant and surrounding components, such as evaporator and condenser walls, is assumed to occur without significant thermal resistance. This assumption simplifies the analysis by considering perfect heat transfer across these boundaries.

• No Superheat Evaporation and Subcooling Condensation

  In some simplified calculations, the presence of superheating (vapor above its saturation temperature) and subcooling (liquid below its saturation temperature) may be neglected. Although these phenomena are common in actual refrigeration systems and can impact performance, they are often disregarded for simplicity.

Cooling Load

In this scenario, the amount of cooling required in the cold storage is dependent on the fish being stored. With an average specific heat capacity of the fish of 3.6 kJ/kg K [38][39], the cooling load can be computed using Eq. (1).

Heat Transfer

The process of a refrigerant’s work cycle involves absorbing heat from a cold storage area and then releasing that heat into the environment. This heat transfer rate at each component can be calculated using the energy method formula as written in Eq. (2).

In addition, by using the enthalpy change during the work cycle of the refrigerant, the heat transfer can be calculated as described in Eq. (3).

Mass Flow Rate of Refrigerant

In designing this cold storage, the mass flow rate of the refrigerant must first be determined by using Eq. (4) as the product of density and volumetric flow rate.

Compressor Work

To ensure that the refrigeration system can work at its maximum potential, a suitable compressor power is required. Therefore, it is necessary to calculate the isentropic efficiency that can be produced by the compressor. This efficiency can be defined as the ratio of ideal (isentropic) work input to the actual work input for the same inlet state and outlet pressure and can be calculated by comparing the ideal (isentropic) and actual enthalpy increases of the refrigerant between the inlet and outlet of the compression process. Eq. (5) can be used to calculate efficiency.

Coefficient of Performance (COP)

After determining the cooling load required for this cold storage and knowing the output power of the compressor, the coefficient of performance (COP) of this cold storage system can be obtained by comparing the cooling load with the compressor’s power consumption. This will provide us with a COP value that can be used as a reference in designing this cold storage. The COP calculation can be done using Eq. (6).

Cooling Air Mass Flow Rate

The airflow rate at the evaporator and condenser is needed to determine how fast cooling air can exchange at the evaporator with the air inside the system, as well as at the condenser with the ambient air. To calculate both, Eq. (7) can be used.

However, there is a slight difference in the calculation of the airflow rate at the evaporator and the condenser. To calculate the airflow at the evaporator, the heat within the system was used. On the other hand, at the condenser, the air flow was calculated using the heat that exits through the condenser and the temperature difference between the ambient air temperature and the temperature within the system.

Results and Discussions

Numerical analysis was conducted to compare the performance of R-32 and R-134a under the same conditions, specifically at low target temperatures. The temperature of the refrigerant’s saturated liquid in the condenser was set at 35 ᵒC, while the temperature of the refrigerant’s saturated vapor in the evaporator was set at -30 ᵒC for both refrigerants. These temperature settings were chosen to ensure that the refrigerant temperature is higher than the air temperature in the condenser for efficient heat transfer, and lower than the air temperature in the cold storage for effective heat transfer. Various aspects, such as the work cycle, temperature, pressure, coefficient of performance (COP), and performance in the condenser and evaporator, were evaluated to understand the performance of R-32 and R-134a.

Cycle Performance

Figure 2 illustrates the actual refrigeration cycle of each refrigerant on temperature-entropy diagrams. Point 1 represents the saturated vapor condition of the refrigerant leaving the evaporator. Subsequently, the refrigerants undergo compression during the process (1–2). During this process, the refrigerant enters the compressor as a low-pressure vapor and is compressed to a high-pressure state, resulting in increased temperature and pressure. In the condenser, heat from the refrigerants is transferred to the ambient air at the outside temperature using blowers. This heat rejection causes the refrigerant to condense into a high-pressure saturated liquid (2-3).

Following that process, the high-pressure saturated liquid refrigerant undergoes an isenthalpic process through an expansion valve or throttle, resulting in a significant pressure drop along with a change in temperature and pressure (3-4). The low-pressure, low-temperature refrigerant then enters the evaporator, where it absorbs heat from the cold storage, such as the air or the product being cooled. This heat absorption leads to the evaporation of the refrigerant into a low-pressure vapor state (4-1).

To determine the properties of the refrigerant at each state point, certain parameter values and conditions must be known. State 1 represents the initial state of the refrigerant entering the compressor, typically in a saturated vapor state. The temperature and pressure of the refrigerant can be determined based on the desired cooling requirements and the properties of the specific refrigerant being utilized. Other thermal properties, such as specific enthalpy and specific entropy, can also be determined from the refrigerant’s saturation properties. State 2 corresponds to the refrigerant’s state after the compression process. The pressure at this point is typically equal to the pressure at State 3, which is usually high due to compression.

To determine the properties at State 2, an assumption is made that the isentropic efficiency of the refrigerant compression process is 90%. This assumption is commonly adopted in practical refrigeration cycle analyses to simplify the modelling, and it is consistent with the reported efficiency range of low- to high-performance compressors, which typically spans from 65% to 100% [37]. Recent studies on compressor performance and system-level thermodynamic analyses report isentropic efficiencies approaching 90% for modern, well-designed compressors operating near their rated conditions, supporting the validity of this assumption [40][41]. State 3 represents the refrigerant’s state after the heat rejection process in the condenser, assumed to be in a saturated liquid state. The pressure at this point is typically high, similar to State 2. The temperature and other properties of the refrigerant at State 3 can be determined based on the heat rejection process occurring in the condenser. The specific enthalpy and specific entropy values can be obtained from the refrigerant’s saturation properties. Lastly, State 4 refers to the refrigerant’s state after the expansion process in the expansion valve or throttle. The pressure at this point is the same as the pressure at State 1, which is generally low. In the ideal refrigeration cycle, the enthalpy at State 4 is assumed to be the same as the enthalpy at State 3, disregarding any energy losses or non-idealities.

Peak Temperature

During refrigerant operation, the condenser experiences elevated temperatures due to the high thermal energy of the compressed refrigerant. As the refrigerant flows through the condenser, the heat exchanger removes heat from the system and rejects it to the surroundings. Figure 2 illustrates the actual vapor compression cycle of the refrigeration system, in which the compressor efficiency is taken into account. As shown, R-32 reaches a peak compressor outlet temperature of approximately 96 °C to achieve a saturation liquid temperature of 35 °C.

This elevated outlet temperature does not directly indicate superior heat-release capability but rather signifies that a relatively larger amount of heat must be rejected in the condenser.

The heat-transfer performance in the condenser is strongly influenced by the thermophysical properties of the refrigerant. Compared with R-134a, R-32 exhibits higher thermal conductivity and lower dynamic viscosity, which enhances convective heat-transfer coefficients on the refrigerant side [30]. In addition, R-32 has a higher vapor density, contributing to improved heat-transfer characteristics at comparable mass flow rates. Although the latent heat of vaporization of R-32 is lower than that of R-134a, its favorable transport properties compensate for this limitation, enabling effective heat rejection during the condensation process.

Nevertheless, the elevated condenser temperatures associated with R-32 introduce certain practical challenges, particularly under high ambient temperature conditions. The reduced temperature margin within the condenser may increase the risk of approaching supercritical operating conditions, which can degrade system efficiency and stability.

Moreover, compressor outlet temperatures approaching 96 °C impose additional safety and reliability considerations. Such temperatures can accelerate thermal aging of compressor components, increase mechanical and thermal stresses, and potentially shorten equipment lifespan. From an environmental perspective, the requirement to reject a larger amount of heat in the condenser is generally associated with higher energy consumption and increased indirect greenhouse gas emissions.

Therefore, while R-32 offers favorable heat-transfer characteristics due to its thermophysical properties, careful thermal management and system design are essential to mitigate the associated drawbacks. Strategies such as optimized condenser design, appropriate thermal insulation, and operation within safe temperature limits are necessary to ensure efficient, reliable, and safe system performance.

Pressure Increment

In a vapor compression system, both compression and expansion play crucial roles in the cycle for heat release from the cold storage. Compression is necessary to increase the refrigerant’s pressure, which allows it to raise the temperature while minimizing changes in stored energy, and then release heat in the condenser. Figure 3 provides a comparison between different refrigerants.

One notable observation is that the pressure of R-32 can be up to three times higher than that of R-134a. This high-pressure characteristic of R-32 brings certain challenges and considerations. Over-pressurization can lead to potential issues such as leakage or even pipe blow-up in the capillary pipe, which regulates the flow of refrigerant. To mitigate these risks, it becomes necessary to use thicker pipes in the vapor compression system of R-32, ensuring that the system can handle the higher pressures without compromising its integrity.

However, the use of thicker pipes in the system comes at a cost. It leads to higher expenses during the construction and maintenance of the system. Thicker pipes require more materials and may necessitate adjustments to the system design and installation processes. Additionally, the increased complexity of the system due to higher pressures can result in higher maintenance costs, as specialized equipment and expertise might be required for servicing and repairs.

Therefore, while R-32 operates at higher pressures and temperatures than R-134a, these conditions provide several performance advantages, including higher refrigerating capacity, improved convective heat-transfer coefficients in the heat exchangers, and the potential for more compact system design, which can contribute to improved system efficiency when the system is properly designed and controlled. However, the associated challenges and costs should not be overlooked.

Overall, the use of R-32 in a vapor compression system necessitates careful planning, considering factors such as pressure requirements, pipe thickness, and associated costs. This approach ensures the system’s efficiency, safety, and longevity while effectively managing the potential challenges posed by the higher pressures of R-32.

System Efficiency

The coefficient of performance (COP) is a metric used to assess the efficiency of a refrigeration system. It is determined by calculating the ratio of the heat absorbed from the cold storage to the energy input required from the compressor, as defined by Eq. (6). The rated heat and power of each component, refrigerant mass flow rate needed to reach the cooling load temperature, and the COP are presented in Table 2.

Figure 4 illustrates that when considering the same mass flow rate, R-32 exhibits superior heat transfer capabilities compared to R-134a. This indicates that R-32 can transfer a larger amount of energy per unit mass of refrigerant. This observation aligns with the information presented in Figure 3, which highlights that R-32 possesses a higher enthalpy than R-134a.

It is important to note that although R-32 demonstrates better heat transfer characteristics, this does not necessarily result in a higher COP. The COP takes into account both the heat transfer capabilities and the energy input required from the compressor. In the case of R-134a, despite its slightly lower heat transfer capabilities compared to R-32, it compensates by requiring less energy input from the compressor to achieve the desired cooling effect. Therefore, while R-32 may have advantageous heat transfer properties and a higher enthalpy, the overall system efficiency, as measured by the COP, is slightly better for R-134a. This underscores the need to strike a balance between heat transfer capabilities and energy consumption to achieve optimal efficiency in a refrigeration system.

The examination of Figure 4 uncovers several significant findings. Firstly, it demonstrates that increasing the flow rate of the refrigerant results in a greater workload for the compressor. This connection is logical since a higher workload necessitates a larger amount of refrigerant circulating through the system to meet the heat transfer requirements. Conversely, a lower compressor workload would result in a lower flow rate of the refrigerant. Moreover, the analysis indicates that as the compressor workload increases, the time needed to heat the cold-water supply decreases. This relationship suggests that a more powerful compressor can transfer heat more efficiently, leading to faster water heating. Conversely, a lower compressor workload would lead to a longer heating time.

As mentioned earlier, Figure 4 a) reveals that a 7-kW increase in compressor workload can significantly reduce the heating time, as evidenced by the initial downward trend of the cooling time curve. However, after the blue and yellow curves intersect, the slope of the cooling time graph begins to level off, indicating that further increases in compressor workload do not cause significant changes in heating time. This point is referred to as the optimum point. Additionally, it is evident that when the compressor workload reaches 15 kW and continues to increase to 42 kW, the heating time remains nearly constant at 1-3 hours.

A similar trend is observed in Figure 4 b), which uses a different refrigerant, R-134a. When the blue and yellow curves intersect, the compressor workload is noted as 7.5 kW, and the cooling time is 5 hours. The cooling time then starts to level off, resulting in minimal or no significant changes in heating time as the compressor workload increases from 14 kW to 42 kW, with the cooling time remaining between 1 and 3 hours.

The analysis also highlights a comparison between refrigerants R-32 and R-134a. It reveals that, at the same flow rate, R-32 requires more compressor workload than R-134a. This implies that R-32 demands a higher electrical power input to operate the compressor and generate the necessary workload. However, despite requiring more compressor workload, R-32 achieves a shorter time to heat the cold-water supply compared to R-134a. This finding suggests that R-32 exhibits higher performance, enabling it to rapidly heat the water despite the greater compressor workload requirement.

Evaporator and Condenser Performances

Based on the calculations and Figure 5, it can be observed that the air flow rate in the evaporator for both R-32 and R-134a is the same, measuring 272 m./h. However, a significant difference can be seen in the air flow rate of the condenser for these two working fluids.

Although the condenser heat duty of the R-32 system is slightly higher than that of R-134a, a lower condenser airflow rate is required. This behavior can be attributed to the superior thermophysical properties of R-32, including higher thermal conductivity, higher vapor density, and lower dynamic viscosity, which enhance the refrigerant-side heat-transfer coefficient. In addition, the lower refrigerant mass flow rate of R-32 (Table 2) and its higher condensing temperature increase the temperature driving force for heat transfer. As a result, the required heat rejection can be achieved with a reduced airflow rate on the air side of the condenser.

For R-32, the air flow rate in the condenser is determined to be 867 m./h, whereas for R-134a, it is recorded as 1580 m./h. This indicates that the R-134a working fluid requires a larger air flow rate in the condenser compared to R-32. The air flow rate represents the volume of air passing through a specific point per unit of time.

The larger air flow rate required by R-134a in the condenser has implications for the cross-sectional area needed to facilitate efficient fluid cooling. In practical terms, it means that if R-134a is used as the working fluid, a larger physical space is required to accommodate the necessary air flow and ensure effective cooling in the condenser.

This disparity in air flow rates and the subsequent need for a larger space have implications for system design and installation. When opting for R-134a as the working fluid, engineers and designers must take into account the increased space requirements to accommodate the higher air flow rate. This may involve using larger fans or modifying the system layout to allow for the proper circulation of air and ensure optimal cooling performance efficiency.

Conclusions

Thermodynamic and energetic performance were evaluated for R-32 and R-134a refrigerants for cold storage applications. The main results were summarized as follows:

1. R-134a has a higher efficiency system, which reduces the energy input by compressor work. Conversely, R-32 has the advantage of being able to carry more energy per unit mass, resulting in a reduced space requirement for the condenser to dissipate heat.

2. The temperature and pressure of R-32 after compressing show challenging features, including the risk of leakage and the possibility of experiencing supercritical heating.

3. When assessing the environmental risks, it is observed that despite having zero ozone depletion potential (ODP), R-134a has a greater contribution to global warming than R-32. On the other hand, R-32 poses a low combustion risk as a refrigerant, but presents hazards to the surrounding environment and equipment due to its high-temperature operating cycle.

To avoid harm to the operation of cold storage, it is advisable to use R134 rather than R-32 as a refrigerant due to its potential to cause damage to the surrounding environment.

References

Notes

^* Research paper

Author notes

^a Corresponding author. E-mail: hifni.m.a@ugm.ac.id

Additional information

How to cite this article: H M Ariyadi, I Muhammad, W Kartika, M A Saputro, “R-32 and R-134a as Refrigerants for Fisheries Cold Storage: An Energy Performance Analysis” Ing. Univ. vol. 30, 2026. https://doi.org/10.11144/Javeriana.iued30.rfcs

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