Energy and Exergy Analysis on Zeotropic Refrigerants R-455A and R-463A as Alternatives for R-744 in Automotive Air-Conditioning System (AACs)

Abstract

The popularity of vehicles and the increased time spent in cars with air conditioning systems has led to regulations in many countries that require the use of environmentally friendly refrigerants with minimal global warming and zero ozone depletion potential (GWP and ODP). Cars need high-performance, eco-friendly air conditioning systems to reduce their impact on the environment, lower fuel consumption, and decrease carbon emissions. The aim of the current work was to propose CO₂-based blend zeotropic refrigerants, R-455A (R-744/32/1234yf) and R-463A (R-744/32/125/1234yf/134a), to improve the thermodynamic performance of pure CO₂ refrigerants. The thermodynamic energy and exergy analysis and system optimization of an AAC system for the new zeotropic refrigerant blends compared to carbon dioxide (R-744), using Aspen HYSYS software, were investigated. The influence of cooler/condenser pressure, average evaporator temperature, cooler/condenser outlet temperature, and refrigerant flow rate on the cycles’ COP and exergy efficiency were conducted and are presented. The results showed that, at the same operating condition parameters, the cycle COP improved by 57.6 and 76.5% when using R455A and R463A instated of R744, respectively, with the advantage of reducing leakage problems due to the higher operating pressure of R744 (5–7 times higher than those of R455A and R463A), as well as requiring heavy equipment, but at optimal operating condition parameters, R744 and R-463A had a maximum COP of 14.58 and 14.19, respectively. The maximum COPs of R744, R455A, and R463A based on the optimal pressure of the cooler/condenser were 3.1, 4.25, and 5.4, respectively. Additionally, regarding the need for environmentally friendly air conditioning systems with acceptable performance in cars due to their impact on the environment and their contribution to global warming, the blend R455A is recommended for use as a refrigerant in AAC systems.

1. Introduction

Most automobiles now come equipped with automotive air conditioning (AAC) systems due to their prevalence and the amount of time people spend in their vehicles. These systems typically use refrigerant fluids, which can be harmful to the environment and contribute to global warming. To mitigate this, many countries have mandated the use of eco-friendly refrigerants with minimal global warming potential (GWP) and zero ozone depletion potential (ODP). The popularity of cars and the increasing time that people spend in them have led to the installation of air conditioning systems in most vehicles. However, the use of refrigerant fluids in these systems can be damaging to the environment, prompting many countries to enforce laws requiring the use of environmentally friendly refrigerants with zero ozone depletion potential and low global warming potential.

Due to scientific advancements and the detrimental effects of non-natural refrigerants on the ozone layer and greenhouse gas emissions, it is inevitable that current refrigerants will be phased out and replaced with new, ecologically friendly alternatives. As a result, R134a, with its 100-year GWP of 1430 [1], has been banned from air conditioning systems in the European Union since 2017 [2], being substituted with refrigerants that have a GWP of under 150 [3]. The Kyoto Protocol’s focus on the impact of certain greenhouse gases on global warming has led to the gradual replacement of traditional hydrofluorocarbons. Natural CO₂ (R744) has gained attention as a safe refrigerant due to its non-flammability and non-toxic properties [4].

Research on refrigerant replacements is now divided into two key approaches that are driven by environmental concerns. The first approach involves seeking innovative refrigerants with low GWP and zero ODP, such as United Signal Company’s azeotropic combination R410A and DuPont Company’s R1234yf and R1234ze. The second approach explores the use of natural refrigerants [5,6]. CO₂ has long been used as a refrigerant by humans [7,8], offering several advantages, such as non-toxicity, environmental friendliness, safety, affordability, and excellent thermal properties. Studies by Robinson and Groll [9] from Purdue University have demonstrated that the CO₂ transcritical cycle is a cutting-edge and reliable technique for automobile air conditioning, with great potential for efficiency enhancement. Research by Yu Binbin, Wang Dandong, Xiang Wei et al. [10] has shown that CO₂ electric car air conditioning systems perform comparably to R134a.

Further studies have investigated the impact of R32 refrigerant on CO₂ water heat pump systems, revealing varying reductions in exhaust pressure when R32 was applied [11]. Massuchetto LHP et al. conducted tests on three different refrigerant combinations (R744/R1270, R744/R717, and R744/RE170), determining that R744/RE170 yielded the optimal refrigeration outcome [12]. Mancini et al. explored the use of CO₂ and dimethyl ether as an azeotropic combination refrigerant, finding that the addition of dimethyl ether benefited the CO₂ system [13]. Adding propane to CO₂ refrigerant was also found to enhance the system’s coefficient of performance by 7–9% according to the study by W. I. Mazyan et al. [14].

While CO₂ refrigerant is a popular research area for refrigerant replacement, it does have some disadvantages, including its high working pressure and relatively lower system efficiency. These factors somewhat hinder its rapid promotion and implementation [15,16]. Sun et al. conducted an experimental investigation on CO₂/R32 blends in a water-to-water heat pump system [17]. Yu et al. studied the temperature difference in heat transfer using theoretical and experimental methods based on the nonlinear temperature enthalpy of R236fa/R32 mixtures, identifying distinct characteristics of the glide temperature depending on the composition [18]. Liu et al. examined the detonation of R32/R1234ze(E) mixtures, analyzing the flame structure, explosion pressure, and lowest flammability boundaries of the refrigerant explosion [19].

The development of synthetic refrigerants such as R-32, R-125, R-1234yf, and R-134a with zero ODP and low GWP is hindered by safety concerns, processing complexity, and high cost [20,21]. R744, an alternative to R134a, belongs to the Hydrofluoroolefin family [22]. Abbood et al. [23] studied and presented environmentally friendly alternative refrigerants to R134a and compared the thermodynamic performance of an AAC system using R134a with blends of hydrocarbons (HCs) such as R290/R600a. Their theoretical analysis demonstrated that the blend (R600a/R290/134a) with a mass ratio of (43/35/22) exhibited a low global warming potential (GWP), while maintaining comparable refrigerant performance. This blend can be directly utilized in the system without requiring any modifications. A reviewed, compared, and comprehensively described outline of the main enhancements of recent advances and sustainable solutions in automobile air conditioning systems category was presented by Sagar and Rakshit [24]. The thermodynamic and thermophysical properties for new mixed refrigerants, R13I1/R152a, R1234yf/R290, and R1234yf/R600a, as alternatives to R134a in automotive air conditioning systems, to solve the problem of the high global warming potential of R134a, were presented by [25,26]. Savitha et al. [27] presented a review of the thermodynamic and flammability properties of the low GWP refrigerants and the compatibility of these refrigerants with construction material and lubricant. A novel design that integrated evaporative cooling with an automotive CO₂ air conditioning system was introduced and experimentally investigated by [28] to address the significant decrease in energy efficiency experienced in hot climates when using carbon dioxide (CO₂) refrigerant for vehicles. Lei et al. [29] presented thermotical and experimental studies to increase the heat transfer area of the evaporator and to improve the system performance of air condition systems using CO₂ in hot climate conditions. Vaccaro et al. [30] investigated and presented the effects of the addition of the second element in the CO₂ transcritical cycles to overcome the heavy expansion loss, requiring specific means for its mitigation. Luo et al. [31] theoretically studied the vapor–liquid equilibrium (VLE) properties of two eco-friendly zeotropics, CO₂/R1234ze(Z) and CO₂/R1336mzz(E), by using Peng-Robinson (PR) and Soave-Redlich-Kwong (SRK) models. Chen et al. [32] proposed new correlations, along with the application of deep learning-based modeling techniques, to analyze and predict saturated flow boiling heat transfer and two-phase pressure drops in the evaporating flow. Hussain et al. [33] predicted and optimized the two-phase pressure drop of refrigerant R1234yf across a diverse range of testing conditions. To achieve this, feature engineering techniques were employed to identify and select the most influential features that could accurately estimate the desired output.

However, pure R744 systems can experience leakage problems due to their higher operating pressures (7–10 times higher than conventional R134a systems) and require heavy equipment, which reduces their coefficient of performance. These issues can be mitigated by using new blends, although the properties of such blends can differ from their original constituents. Therefore, the aim of this study was to explore the potential of two R744-based blends, R-455A (R-744/32/1234yf) and R-463A (R-744/32/125/1234yf/134a), to enhance the thermodynamic performance of pure R744 refrigerants, leading to lower engine fuel consumption and hence lower carbon emissions. The need for environmentally friendly air conditioning systems with high performance in cars, due to their impact on the environment and contribution to global warming, is essentially required. To achieve this, the thermodynamic energy and exergy characteristics of an AAC system were evaluated and compared for the proposed zeotropic refrigerants and for carbon dioxide (R-744) using Aspen HYSYS software. Additionally, the effects of cooler/condenser pressure, average evaporator temperature, cooler/condenser outlet temperature, and refrigerant flow rate on the cycle coefficient of performance (COP) and exergy efficiency were studied, analyzed, and presented. The anticipated outcomes of this investigation aim to demonstrate the most advantageous refrigerant performance in comparison to R-744 and to provide suitable operating conditions for the proposed refrigerants. Furthermore, these findings will serve as a foundation for constructing an AAC system that can be mass-produced economically and in an environmentally conscious manner.

2. Thermodynamic Properties of Refrigerant Blends

2.1. Physical and Environmental Properties

Table 1. Physical and environmental properties of individual refrigerants [34].

Refrigerant Name	Type	Safety Group (ASHRAE Classification)	ODP	GWP	Molar Mass [kg/kmol]	Normal Boiling Point [°C]	Critical Temperature [°C]	Critical Pressure [MPa]
R134a	HFC	A1	0	1370	102.03	−26.1	101.1	4.06
R32	HFC	A2L	0	677	52.02	−51.7	78.2	5.78
R125	HFC	A1	0	3170	120	−48.5	66.18	3.629
R1234yf	HFO	A2L	0	<1	114.04	−29.5	94.7	3.38
R744	Natural	A1	0	1	44.01	−78.51	31.05	7.38

presents the key physical and environmental characteristics of individual refrigerants: R134a, R32, R125, R1234yf, and R744. As shown in Table 1, all refrigerants have an ozone depletion potential (ODP) of 0. However, R125 and R134a have higher global warming potential (GWP) values compared to the other refrigerants. Notably, R744 and R1234yf have the lowest GWP values. Additionally, R744 has the highest critical pressure and the lowest normal boiling point, with values of 7.3 MPa and −78 °C, respectively.

Table 2. Physical and environmental properties of studied zeotropic refrigerants [34].

Refrigerant Designation	Composition (Mass %)	Molecular Weight (kg/kmol)	Normal Boiling Point (Bubble/Dew) (°C)	Critical Temperature (°C)	Critical Pressure (MPa)	Safety Group (ASHRAE Standard 34)	ODP	GWP
R-744	R-744 (100)	44.01	−78.51	31.05	7.38	A1	0	1
R-455A	R-744/32/1234yf (3.0/21.5/75.5)	87.45	−51.6/−39.1	85.48	4.32	A2L	0	~146
R-463A	R-744/32/125/1234yf/134a (6.0/36.0/30.0/14.0/14.0)	74.72	−58.4/−46.9	75.38	5.04	A1	0	~1386

displays the physical and environmental properties of the studied zeotropic refrigerants. The composition of the two blends, R-455A (R-744/32/1234yf) and R-463A (R-744/32/125/1234yf/134a), by mass are (3.0/21.5/75.5) and (6.0/36.0/30.0/14.0/14.0), respectively. Furthermore, the blend R-463A has a high GWP of approximately 1386, while the GWP of the blend R-455A is less than one.

2.2. Temperature Glide, P-T Envelope

During the evaporation process, the refrigerant undergoes a transition from a saturated liquid state, known as the bubble point, to a saturated vapor state, known as the dew point. The temperature at which this transition occurs remains constant under constant pressure. However, when multiple components are present in a refrigerant mixture, a temperature glide can occur. The temperature glide refers to the difference between the dew point and bubble point temperatures. In this study, the temperature glide properties of R455A and R463A were investigated by analyzing the vapor–liquid equilibrium (VLE). The Peng-Robinson equation of state (EOS) was employed to accurately calculate the saturated vapor pressure and saturated liquid density. However, it is important to use appropriate mixing rules when evaluating the thermodynamic properties of the mixture using the EOS. Figure 1 illustrates the pressure-temperature (P-T) envelopes of the studied zeotropic refrigerants (R455A and R463A) compared to pure R744. Pure R744 does not exhibit a temperature difference between the bubble and dew point temperatures since it is a single component refrigerant (see dashed lines in Figure 1a). Azeotropic blends behave similarly to pure refrigerants, where the boiling and condensation points coincide at the same composition for both vapor and liquid phases. However, for zeotropic blends, the more volatile component starts boiling first in the evaporator, followed by the less volatile component. This sequential boiling can lead to a change in concentration as the temperature changes, resulting in a temperature glide (t_glide > 2 K). If the refrigerant begins boiling at t_i and ends at t_f, then the temperature glide can be expressed as t_glide = t_f − t_i (see dashed lines in Figure 1b,c). The changing composition of the evaporating liquid, caused by the varying boiling point, contributes to the temperature glide. Additionally, leakage from the system can introduce changes in the composition and properties of the refrigerant, further contributing to t_glide.

3. System Modeling

A simulation system was constructed using Aspen HYSYS V12.1^® Software (AspenTech, Bedford, MA, USA) [35] to assess the thermodynamic performance of zeotropic alternative refrigerants in AACs, as depicted in Figure 2. Aspen HYSYS has gained recognition among academics and engineers for its reliability and capability to evaluate complex industrial processes. The user-friendly nature of the Aspen HYSYS platform enables the optimization of conceptual design and operational parameters. This robust process simulator offers a vast library of pre-built component models and property packages. Aspen HYSYS allows for the seamless integration of multiple modules, facilitating the connection of material and energy streams. This capability enables the static and dynamic modeling of various complex chemical and hydrocarbon fluid-based processes. The simulation model developed for AACs in Aspen HYSYS offers easy integration with a range of energy systems, including compressors, condensers, gas coolers, heat exchangers, evaporators, expansion valves, and more.

By leveraging Aspen HYSYS, the thermodynamic performance of zeotropic alternative refrigerants in AACs can be thoroughly examined. The software’s extensive functionality and versatility contribute to the advancement of energy-efficient and environmentally conscious air conditioning technologies.

3.1. System Description and Assumptions

It is crucial to investigate and evaluate the performance of AACs cycles considering the physical and environmental characteristics of the newly studied zeotropic alternative refrigerants. Figure 2a illustrates the schematic diagram of the traditional vapor compression refrigeration cycle in AACs, comprising an evaporator, compressor, cooler/condenser, expansion valve, and internal heat exchanger. The following is a description of the system cycle: (i) 1′–2: non-isentropic compression; (ii) 2–3: constant pressure heat rejection from the cooler/condenser; (iii) 3–3′: subcooling in the internal heat exchanger (IHE); (iv) 3′–4: isenthalpic throttling in the expansion valve (EXV); (v) 4–1: constant pressure heat absorption in the evaporator; (vi) 1–1′: superheating in the internal heat exchanger (IHE). By adjusting the average evaporation temperature, cooler/condenser refrigerant output temperature, cooler/condenser pressure, and refrigerant mass flow rate, the operating conditions of the cycle can be modified. Because the critical point pressure of the refrigerant can be comparatively low compared to the cooler pressure, a transcritical model for the cycle was developed based on the operating conditions. In transcritical cycles, a gas cooler is utilized instead of a condenser, as in subcritical cycles, since there is no phase change above the critical point. As the pressure and temperature of the refrigerant are independent of each other, both pressure and temperature need to be defined at pressures above the critical point. The coefficient of performance of the cycle varies at the same gas cooler outlet temperature but different operating pressures. Therefore, the simulation model’s gas cooler was optimized by running it at the optimum gas cooler pressure [36]. Figure 2b presents the pressure-enthalpy (p-h) diagram of subcritical/transcritical cycles using R-744 in AACs.

In order to ensure accuracy and simplicity in the current simulation, the following assumptions were considered to evaluate the performance of zeotropic alternative refrigerants in AACs:

• The entire system operates under steady-state conditions.
• The effects of kinetic energy and gravity are neglected.
• A reference analysis is conducted with an ambient temperature of 25 °C and atmospheric pressure of 101.325 kPa.
• Heat loss and pressure drop during the heat transfer process are disregarded.
• The refrigerant at the evaporator outlet is assumed to be in a saturated state.
• The operating parameter settings and modeling assumption values are presented in

  Table 3. Operating parameters conditions.

  Parameter	Value/Range
  Cooler/condenser pressure, P2	5–15 MPa
  Average evaporator temperature, tevap	5–15 °C
  Cooler/condenser outlet temperature, t3	20–40 °C
  Refrigerant flow rate, ${\dot{m}}_r$	0.05–0.15 kg/s

  and

  Table 4. Energy and exergy equations and modelling assumptions.

  Component	Symbol	Characteristic Equations [37]	Modelling Assumptions
  Compressor		$\begin{array}{ll}{\dot{W}}_{comp}& ={\dot{m}}_r(h_{comp,out,act}-h_{comp,in})\\ & ={\dot{m}}_r(h_{comp,out}{}_{,s}-h_{comp,in})/{\eta }_{is}\end{array}$ ${\eta }_{is}=\frac{h_{comp,out}{}_{,s}-h_{comp,in}}{h_{comp,out,act}-h_{comp,in}}$ ${\dot{I}}_{comp}={\dot{E}}_{comp,in}-{\dot{E}}_{comp,out}+{\dot{W}}_{comp}$	ηis = 80%, Polytropic method: ShultzOperation mode: centrifugal
  Cooler/ condenser		${\dot{Q}}_{cooler/cond.}={\dot{m}}_r(h_{cooler/cond,out}-h_{cooler/cond,in})$ ${\dot{I}}_{Cooler/cond.}={\dot{E}}_{cooler/cond,in}-{\dot{E}}_{cooler/cond,out}-{\dot{E}}_{cooler/cond.}$ ${\dot{E}}_{cooler/cond}=\left(1-\frac{T_o}{T_{r\hspace{0.17em}(cooler/cond),out}}\right)\dot{Q}{}_{cooler/cond}$	Heat exchanger model: Simple end point ΔP = 0 kPa
  Expansion valve		${\dot{m}}_r{h}_{EXV,in}={\dot{m}}_r{h}_{EXV,out}$ ${\dot{I}}_{EXV}={\dot{E}}_{EXV,in}-{\dot{E}}_{EXV,out}$
  Evaporator		${\dot{Q}}_{evap}={\dot{m}}_r(h_{evap,in}-h_{evap,out})$ ${\dot{I}}_{evap}={\dot{E}}_{evap,in}-{\dot{E}}_{evap,out}-{\dot{E}}_{evap}$ ${\dot{E}}_{evap}=\left(1-\frac{T_o}{T_{r\hspace{0.17em}evap,out}}\right)\dot{Q}{}_{evap}$	x1 = 1 ΔP = 0 kPa
  IHE		${\dot{Q}}_{IHE}={\dot{m}}_r{(h_{IHE,in}-h_{IHE,out})}_{Hot}={\dot{m}}_r{(h_{IHE,out}-h_{IHE,in})}_{cold}$ ${\dot{I}}_{IHE}={\displaystyle \sum _{}^{}{\dot{E}}_{IHE,in}}-{\displaystyle \sum _{}^{}{\dot{E}}_{IHE,out}}$	Heat exchanger model: Simple end point ΔP Hot stream = 0 kPa ΔP Cold stream = 0 kPa

  .

These assumptions are made to simplify the simulation and focus on the primary aspects of the performance evaluation of zeotropic alternative refrigerants in AACs.

3.2. Thermodynamic Model Analysis

When evaluating the system, it is important to consider several key parameters, including evaporator capacity, compressor power, pressure ratio, compressor discharge temperature, coefficient of performance (COP), and exergy efficiency. These parameters are used to analyze the energy and exergy performance of the AAC system when using different zeotropic alternative refrigerants. Furthermore, model validation is necessary to verify the accuracy of the AAC system simulation. The validation process involves comparing the model predictions with experimental or reference data to ensure the reliability of the results. The details of these parameters and the importance of model validation are discussed in the following sections.

3.2.1. Energy Analysis

The energy analysis of the modeled AAC system involves applying the principle of conservation of energy to each component of the system. This analysis assumes that the system is in a steady-state condition, with negligible changes in kinetic and potential energies. To calculate the evaporator capacity, the principle of conservation of energy is specifically applied to the evaporator unit, as follows:

$${\dot{Q}}_{evap}={\dot{m}}_r(h_1-h_4)$$

(1)

where ṁ_r is the refrigerant mass flow rate, and h is the refrigerant’s specific enthalpy. The same approach can also be used to determine the cooler/condenser capacity of the cooler/condenser unit, as follows:

$${\dot{Q}}_{cooler/cond}={\dot{m}}_r(h_2-h_3)$$

(2)

The compressor power absorbed by the refrigerant during the compression process can be calculated using the following equation, assuming the compressor operates adiabatically:

$${\dot{W}}_{comp}={\dot{m}}_r(h_2-h_{1\prime })={\dot{m}}_r(h_{2,s}-h_{1\prime })/{\eta }_{is}$$

(3)

where η_is is the compressor isentropic efficiency, which is given by:

$${\eta }_{is}=\frac{h_{2,s}-h_{1\prime }}{h_2-h_{1\prime }}$$

(4)

where h_2,s represents the specific enthalpy of the refrigerant at the outlet of the compressor during isentropic compression, and h₂ represents the specific enthalpy of the refrigerant at the outlet of the compressor during the actual compression operation.

The compressor pressure ratio is defined as:

$$\gamma =\frac{P_2}{P_{1\prime }}$$

(5)

where γ is the pressure ratio of the compressor in the system, P₂ is the compressor’s exhaust pressure in MPa, and P_1′ is the compressor’s suction pressure in MPa.

The cycle’s coefficient of performance:

$$COP=\frac{{\dot{Q}}_{evap}}{{\dot{W}}_{comp}}$$

(6)

3.2.2. Exergy Analysis

The study of exergy, unlike energy, reveals that exergy is not conserved and is subject to destruction in processes that involve both useful work and waste. Conducting a thermal exergy analysis is valuable in identifying the sources and magnitudes of thermodynamic inefficiencies. To achieve this objective, the following general form of the steady-state exergy rate balance equation for control volumes [38] can be applied to the components of the AAC system:

$${\dot{I}}_{}={\displaystyle \sum _{}^{}\left(1-\frac{T_0}{T_j}\right)}{\dot{Q}}_j-{\dot{W}}_{cv}+{\displaystyle \sum _{}^{}{\dot{E}}_{in}}-{\displaystyle \sum _{}^{}{\dot{E}}_{out}}$$

(7)

where T₀ represents the environmental temperature dead state, ${\dot{Q}}_j$ denotes the time rate of heat transfer at the boundary location, T_j is the boundary temperature, ${\dot{W}}_{cv}$ is the amount of work conducted in the control volume, $\dot{E}$ is the flow exergy rate, whereas in and out denote the inlet and outlet, respectively, and $\dot{I}$ is the rate of exergy destruction in the control volume.

The exergy rate in the system at state point i is defined as follows:

$${\dot{E}}_i={\dot{m}}_i[(h_i-h_0)-T_0(s_i-s_0)]$$

(8)

where s is the refrigerant’s specific entropy, and subscript “0” indicates the reference (dead) condition.

The exergy destruction in the compressor is influenced by the internal heat transfer, gas friction, and mechanical friction of the moving elements. Assuming adiabatic compression, the exergy rate balance equation leads to the following equation for the rate of exergy destruction in the compressor:

$${\dot{I}}_{comp}={\dot{E}}_{1\prime }-{\dot{E}}_2+{\dot{W}}_{comp}$$

(9)

The rate of exergy destruction in the condenser, resulting from heat transfer between the refrigerant and air streams, can be calculated using the following equation [22]:

$${\dot{I}}_{Cooler/cond.}={\dot{E}}_2-{\dot{E}}_3-\left(1-\frac{T_o}{T_r{{}_{(}}_{cooler/cond.),out}}\right)\dot{Q}{}_{cooler/cond}$$

(10)

where T_r is the refrigerant temperature.

Assuming there is no heat transfer to or from the environment, the exergy destruction in the internal heat exchanger, which is caused by the internal heat transfer between the refrigerant streams, can be calculated using the following equation:

$${\dot{I}}_{IHE}={\dot{E}}_3-{\dot{E}}_{3\prime }+{\dot{E}}_1-{\dot{E}}_{1\prime }$$

(11)

The exergy destruction in the expansion valve, which is caused by internal friction and a rapid pressure drop, can be calculated by disregarding the heat transfer with the environment using the following equation:

$${\dot{I}}_{EXV}={\dot{E}}_{3\prime }-{\dot{E}}_4$$

(12)

To calculate the rate of exergy destruction in the evaporator, which is caused by the heat transfer between the refrigerant and air streams, the following equation can be used [22]:

$${\dot{I}}_{evap}={\dot{E}}_4-{\dot{E}}_1+\left(1-\frac{T_o}{T_r{\hspace{0.17em}}_{evap,out}}\right)\dot{Q}{}_{evap}$$

(13)

The rate of overall exergy destruction in the AAC system’s cycle can be calculated by summing up the individual exergy destruction rates in the system’s components, as follows:

$${\dot{I}}_{total}={\dot{I}}_{comp}+{\dot{I}}_{cooler/cond.}+{\dot{I}}_{IHE}+{\dot{I}}_{EXV}+{\dot{I}}_{evap}$$

(14)

Finally, the following expression for exergetic efficiency can be used to evaluate the overall exergetic performance of the AAC system:

$${\eta }_{ex}=1-\frac{{\dot{I}}_{total}}{{\dot{W}}_{comp}}$$

(15)

4. Model Validation

The accuracy of the AAC system simulation model, generated with Aspen HYSYS, was validated through model validation. The transcritical CO₂ refrigeration cycle with a regenerator model was validated using the same boundary and operating parameters as reported by Rigola et al. [39] and Elattar and Nada [37]. The experimental [39] and numerical [37] results based on COP were compared with the results obtained from the current model, and the maximum relative errors were found to be 11.24% and 2.79% for experimental and numerical results, respectively, as presented in

Table 5. Model validation with reported results [37,39].

State Parameters						Performance Parameter
Experimental Data, [39]						COP
${\dot{Q}}_{evap}$ (W)	Tevap (°C)	Tcomp,in (°C)	Pcomp,out (bar)	Tcooler,out (°C)	ηcomp,overall (%)	Exp. [39]	Present model	Error (%)
351.65	−11.30	35.38	89.71	35.24	46.30	1.08	0.98	9.26
525.11	−2.03	35.62	90.37	35.23	54.80	1.56	1.45	7.05
832.70	9.48	35.42	90.14	36.45	65.70	2.49	2.21	11.24
469.09	−11.58	31.06	84.95	31.94	56.00	1.42	1.34	5.63
649.51	−1.77	31.56	84.98	31.94	60.60	1.85	1.75	5.41
842.02	5.32	31.83	85.78	32.11	64.50	2.29	2.24	2.18
Numerical data, [37]						COP
${\dot{Q}}_{evap}$ (W)	Tevap (°C)	Tcomp,in (°C)	Pcomp,out (bar)	Tcooler,out (°C)	ηcomp,overall (%)	Elattar & Nada [37]	Present model	Error (%)
367.25	−9.56	34.39	89.71	35.24	46.8	1.05	1.06	0.95
544.63	−1.62	34.52	90.37	35.23	52.6	1.43	1.47	2.79
816.03	9.53	34.44	90.14	36.45	62.8	2.25	2.28	1.33
488.58	−10.27	31.19	84.95	31.94	55.3	1.36	1.34	2.0
695.93	−0.99	31.17	84.98	31.94	61.7	1.80	1.78	2.0
851.32	5.09	31.03	85.78	32.11	65.6	2.21	2.22	0.45

. These results demonstrate that the simulation model technique implemented in this study using Aspen HYSYS software is sufficiently accurate.

5. Results and Discussion

5.1. Zeotropic Refrigerants Environmental Impacts

Considering their potential for advancing refrigeration cycles, it is important to note that R134a, R32, R125, R1234yf, and R744 have negative impacts on the depletion of the ozone layer. However, R134a and R32 also have positive impacts on global warming, whereas R1234yf and R744 do not contribute to global warming. To assess the environmental impact of the proposed refrigerant mixtures, Table 1 provides information on the ozone depletion potential (ODP) and global warming potential (GWP) of the individual pure components of refrigerants [14,40]. To evaluate the environmental impact of the proposed zeotropic refrigerants, Equations (16) and (17) were used, and the results are presented in Table 2.

$$OD{P}_{Zeotrope}={\displaystyle \sum _{i=1}^n{w}_{pure,i}}OD{P}_{pure,i}$$

(16)

$$GW{P}_{Zeotrope}={\displaystyle \sum _{i=1}^n{w}_{pure,i}}GW{P}_{pure,i}$$

(17)

In the equations, w_pure,i represents the mass fraction of component i in a zoetrope. The evaluated ozone depletion potential (ODP) and global warming potential (GWP) of the proposed zeotropic refrigerants are presented in Table 2. It is worth noting that all the studied zeotropic blends have an ODP of zero, as their individual components have an ODP of zero. Furthermore, based on the classification by UNIP, 2019 [41], for 100-year global warming potential levels, R-455A is classified as having a negligible GWP level, while R-463A is classified as having a high GWP level.

5.2. Parametric Studies

A transcritical/subcritical cycle was developed for R744, R455A, and R463A due to the critical temperatures of these refrigerants. In transcritical cycles, a gas cooler is used instead of a condenser since there is no phase change above the critical point. At pressures above the critical point, the pressure and temperature of the refrigerant become independent variables and both need to be specified. The coefficient of performance (COP) of the cycle is influenced by the cooler/condenser outlet temperature and the operating pressures. Figure 3 illustrates the optimum cooler/condenser pressure, P₂, for the studied zeotropic refrigerants, R455A and R463A, compared to R744, based on the maximum COP of the cycle. The common operating conditions for the comparison are t_evap = 7.5 °C, t₃ = 35 °C, and ṁ_r = 0.075 kg/s. According to the figure, the maximum COP values achieved with R744, R455A, and R463A were 3.1, 4.25, and 5.4, respectively. These values were obtained at the optimum pressures of 8.9, 1.65, and 2.3 MPa for R744, R455A, and R463A, respectively.

5.2.1. Impact of Cooler/Condenser Pressure, P₂

Figure 4 displays the variations of evaporator capacity (${\dot{Q}}_{evap}$), compressor power (${\dot{W}}_{comp}$), exergy efficiency (η_ex), and coefficient of performance (COP) with the cooler/condenser pressure (P₂) at t_evap = 7.5 °C, t₃ = 35 °C, and ṁ_r = 0.075 kg/s. In Figure 4a, it is evident that as P₂ increased, the evaporator capacity (${\dot{Q}}_{evap}$) sharply increased for R744, while it remained nearly constant for R455A and R463A. This behavior is due to the difference in refrigerant enthalpy through the evaporator. With R744, the refrigerant enthalpy decreased at the inlet of the evaporator, while it remained approximately constant for R455A and R463A as P₂ increased. This is because the cycle is transcritical for R744 using a gas cooler, while it is subcritical for R455A and R463A using a condenser. Furthermore, R463A exhibited the highest evaporator capacity (${\dot{Q}}_{evap}$) among the three refrigerants, while R455A had a higher ${\dot{Q}}_{evap}$ than R744 until P₂ ≤ 8.9 MPa, after which R744 surpassed R455A. Figure 4b illustrates the compressor power (${\dot{W}}_{comp}$) of R744, R455A, and R463A. It was observed that ${\dot{W}}_{comp}$ increased with an increase in the cooler/condenser pressure (P₂). This behavior can be attributed to the thermodynamic relationship between pressure and temperature and the nature of the thermodynamic cycle.

Figure 4c,d illustrates the impact of cooler/condenser pressure (P₂) on exergy efficiency (η_ex) and coefficient of performance (COP). For R744, it can be observed that as the pressure of the cooler/condenser increased, both η_ex and COP initially rose, reached a peak around P₂ = 9 MPa, and then started to decline. This behavior is primarily attributed to the increase in evaporator capacity (${\dot{Q}}_{evap}$), which dominated over the increase in compressor power (${\dot{W}}_{comp}$) until P₂ ≤ 9 MPa. However, after P₂ > 9 MPa, the situation reversed, and the COP became inversely correlated with ${\dot{Q}}_{evap}$ and ${\dot{W}}_{comp}$. On the other hand, for R455A and R463A, the exergy efficiency (η_ex) and coefficient of performance (COP) decreased with increasing P₂. This decrease is mainly due to the increase in compressor power (${\dot{W}}_{comp}$) and the nearly constant values of evaporator capacity (${\dot{Q}}_{evap}$) as P₂ rose. At P₂ = 15 MPa, the maximum ${\dot{Q}}_{evap}$ values for R744, R455A, and R463A were 12.2 kW, 9 kW, and 12.5 kW, respectively. Furthermore, within the studied range of P₂, the increase in evaporator capacity (${\dot{Q}}_{evap}$) was 614% for R463A compared to R744 at P₂ = 7 MPa and 2.5% at P₂ = 15 MPa. Additionally, within the studied range of P₂, the COP of R455A and R463A decreased by 27% and 37%, respectively.

5.2.2. Impact of Average Evaporator Temperature, t_evap

Figure 5 illustrates the impact of the average evaporator temperature (t_evap) on the following key performance parameters: evaporator capacity (${\dot{Q}}_{evap}$), compressor power (${\dot{W}}_{comp}$), exergy efficiency (η_ex), and coefficient of performance (COP). The analysis was performed at a condenser temperature of t₃ = 35 °C, mass flow rate of ṁ_r = 0.075 kg/s, and condenser pressure of P₂ = P_opt. The optimum pressures for R744, R455A, and R463A were 8.9, 1.65, and 2.3 MPa, respectively, as shown in Table 2. Table 2 also provides the critical pressures and temperatures for R744, R455A, and R463A, indicating that the R744 cycle operates in the transcritical region, while the cycles for R455A and R463A are in the subcritical region. The ${\dot{Q}}_{evap}$ of R455A and R463A increased with rising t_evap, while it decreased for R744. Additionally, the ${\dot{Q}}_{evap}$ of R744 was higher than that of R455A until t_evap reached 9 °C.

In Figure 5b,c, it can be observed that for all refrigerants, ${\dot{W}}_{comp}$ decreased and COP increased with increasing average evaporator temperature (t_evap). This trend occurs due to the decrease in compression work with higher t_evap while maintaining a constant cooler/condenser pressure (P₂). Furthermore, Figure 5d demonstrates the effect of t_evap on exergy efficiency (η_ex), clearly showing a decrease as t_evap increased. At t_evap = 15 °C, the maximum ${\dot{Q}}_{evap}$ values were 10 kW and 12.5 kW for R455A and R463A, respectively, while R744 achieved a maximum ${\dot{Q}}_{evap}$ of 9.48 kW at t_evap = 5 °C. Within the studied range of t_evap, using R455A and R463A instead of R744 resulted in ${\dot{Q}}_{evap}$ enhancements of 16.3% and 49%, respectively, at t_evap = 15 °C. The maximum COP values were 4.25, 6.7, and 7.5 for R744, R455A, and R463A, respectively, at t_evap = 15 °C. Additionally, within the studied range of t_evap, the COP for R744, R455A, and R463A increased by 52%, 76%, and 56%, respectively. The improvements in COP were 57.6% and 76.5% when using R455A and R463A instead of R744, respectively, at t_evap = 15 °C.

5.2.3. Impact of Cooler/Condenser Outlet Temperature, t₃

Figure 6 demonstrates the influence of cooler/condenser outlet temperature (t₃) on the following performance parameters: evaporator capacity (${\dot{Q}}_{evap}$), compressor power (${\dot{W}}_{comp}$), exergy efficiency (η_ex), and coefficient of performance (COP). The analysis was conducted at a fixed average evaporator temperature of t_evap = 7.5 °C, mass flow rate of ṁ_r = 0.075 kg/s, and condenser pressure of P₂ = P_opt. The optimum pressures for R744, R455A, and R463A were 8.9, 1.65, and 2.3 MPa, respectively. As shown in Figure 6a,c, the ${\dot{Q}}_{evap}$ and COP decreased as the cooler/condenser outlet temperature (t₃) increased for all refrigerants. Furthermore, the decline in ${\dot{Q}}_{evap}$ and COP became more pronounced when t₃ exceeded 35 °C. This phenomenon occurs due to the reduction in refrigerant enthalpy difference across the evaporator with increasing t₃, while the compression work remained constant. Figure 6b indicates that t₃ had no significant effect on compressor power (${\dot{W}}_{comp}$) since the cooler/condenser pressure (P₂) and average evaporator temperature (t_evap) were fixed at their optimum values of 7.5 °C.

Exergy efficiency (η_ex) serves as a fundamental measure of the system’s thermodynamic performance. Figure 6d illustrates that for R744, R455A, and R463A, as t₃ increased, ηex initially rose, reached a peak, and then started to decline. Typically, this peak in η_ex occurs around t₃ = 35 °C. The decrease in η_ex after t₃ > 35 °C can be attributed to the increase in total irreversibility of the cycle while keeping the compressor power (${\dot{W}}_{comp}$) constant. At t₃ = 20 °C, the maximum ${\dot{Q}}_{evap}$ values were 14.1 kW, 12.9 kW, and 15 kW for R744, R455A, and R463A, respectively. Within the studied range of t₃, the ${\dot{Q}}_{evap}$ for R744, R455A, and R463A decreased by 57%, 77%, and 67%, respectively. The maximum COP values were 4.8, 6.5, and 6.5 for R744, R455A, and R463A, respectively, at t₃ = 20 °C. Moreover, within the studied range of t₃, the COP for R744, R455A, and R463A decreased by 58.3%, 77%, and 66%, respectively. When compared to R744 at t₃ = 20 °C, the improvements in COP were 35.4% and 76.5% when using R455A and R463A, respectively.

5.2.4. Impact of Refrigerant Flow Rate, ṁ_r

Figure 7 illustrates the effects of the refrigerant flow rate (ṁ_r) on the following performance parameters: evaporator capacity (${\dot{Q}}_{evap}$), compressor power (${\dot{W}}_{comp}$), exergy efficiency (η_ex), and coefficient of performance (COP). The analysis was conducted at a fixed average evaporator temperature of t_evap = 7.5 °C, cooler/condenser outlet temperature of t₃ = 35 °C, and condenser pressure of P₂ = P_opt. The optimum pressures for R744, R455A, and R463A were 8.9, 1.65, and 2.3 MPa, respectively. As depicted in Figure 7, the ${\dot{Q}}_{evap}$ and ${\dot{W}}_{comp}$ increased with an increase in the refrigerant flow rate (ṁ_r). However, ṁ_r had no significant impact on the exergy efficiency (η_ex) and coefficient of performance (COP). The maximum ${\dot{Q}}_{evap}$ values were 18.8 kW, 17.7 kW, and 24.9 kW for R744, R455A, and R463A, respectively, at ṁ_r = 0.15 kg/s. Within the studied range of ṁ_r, the ${\dot{Q}}_{evap}$ for R744, R455A, and R463A increased by 208%, 200%, and 207%, respectively. Similarly, the maximum ${\dot{W}}_{comp}$ values were 6 kW, 4 kW, and 4.6 kW for R744, R455A, and R463A, respectively, at ṁ_r = 0.15 kg/s. Within the studied range of ṁ_r, the ${\dot{W}}_{comp}$ for R744, R455A, and R463A increased by 200%, 700%, and 206%, respectively. The maximum COP values were 3.15, 4.4, and 5.3 for R744, R455A, and R463A, respectively, at t₃ = 20 °C. Within the studied range of ṁ_r, the improvements in COP were 40% and 68% when using R455A and R463A instead of R744, respectively. It is worth noting that while ${\dot{Q}}_{evap}$ and ${\dot{W}}_{comp}$ were significantly influenced by the refrigerant flow rate (ṁ_r), the exergy efficiency (η_ex) and coefficient of performance (COP) remained unaffected by changes in ṁ_r.

5.3. Comparisons and Assessments of Studied Refrigerant Blends

Figure 8 provides a cycle analysis and the state properties at the inlet and exit of each component for the R744 refrigerant using Aspen HYSYS software. The purpose of this analysis was to compare and assess the performance indicators of the Advanced Adiabatic Compressed Air Energy Storage (AACs) system when charged with pure R744 and with R455A and R463A blend refrigerants as working fluids. The selected performance indicators include evaporator capacity (${\dot{Q}}_{evap}$), compressor power (${\dot{W}}_{comp}$), exergy efficiency (η_ex), and coefficient of performance (COP).

Figure 9 presents the system enhancements achieved with the different working fluids. It is evident that the blend R463A exhibited the highest values for ${\dot{Q}}_{evap}$, COP, and ηex, with values of 12.35 kW, 5.39, and 47.8%, respectively, surpassing both R744 and R455A. To further understand the system behavior, the percentage of exergy destruction for each component in the studied refrigerant blends compared to R744 is illustrated in Figure 10. This provides insight into the influence of each component on the overall exergy destruction. It can be observed that the expansion valve (EXV) contributed the highest irreversibility, accounting for 42% in the R455A cycle, followed by the compressor with 35% in the R463A cycle. The evaporator and condenser also contributed to exergy destruction in all configurations, albeit to a lesser extent. Overall, the findings from these analyses demonstrated the performance advantages and trade-offs associated with different refrigerant blends in the AACs system, highlighting the potential of R463A as the most promising option based on the evaluated performance indicators.

5.4. System Optimization

After conducting a parametric analysis, the performance of two refrigerant blend mixtures (R455A and R463A) and pure CO₂ (R744) was optimized to maximize the coefficient of performance (COP) within the studied ranges of operating condition parameters. The optimal operating conditions, based on maximum COP, are presented in

Table 6. Optimal operating parameters based on maximum COP.

Parameter	Unit	R-744	R-455A	R-463A
P1, P1′, P4	Mpa	5.09	0.793	1.182
P2, P3, P3′	Mpa	6.3	1.243	1.76
t1	°C	15	17.89	18.47
t1′	°C	20	22.89	23.47
t2	°C	38.63	33.57	41.03
t3	°C	24	24.00	24
t3′	°C	22.43	22.38	21.61
t4	°C	15	10.29	10.26
${\dot{m}}_r$	Kg/s	9.73 × 10−2	8.97 × 10−2	9.91 × 10−2
${\dot{W}}_{comp}$	kW	0.975	1.17	1.35
${\dot{Q}}_{evap}$	kW	14.21	15.05	19.15
${\dot{Q}}_{cooler/cond}$	kW	15.19	16.22	20.505
COPmax	---	14.58	12.86	14.19
ηex	%	45.4	26.8	26.7

. Furthermore,

Table 7. Exergy destruction and efficiency for each component for the optimal COP case.

Term	Unit	R-744	R-455A	R-463A
${\dot{I}}_{comp}$	kW	0.23	0.23	0.26
${\dot{I}}_{cooler/cond}$	kW	0.078	0.289	0.356
${\dot{I}}_{IHE}$	kW	0.0211	0.0127	0.0045
${\dot{I}}_{EXV}$	kW	0.20	0.15	0.11
${\dot{I}}_{evap}$	kW	5.27 × 10−8	1.82 × 10−1	2.60 × 10−1
${\dot{E}}_{in}$	kW	0.975	1.17	1.35
ηex	%	45.4	26.8	26.7

lists the exergy destruction and efficiency for each component, calculated based on the optimal COP case. The table provides the maximum system performance and refrigeration capacity achievable for R744, R455A, and R463A, with values of 14.21 kW, 15.05 kW, and 19.15 kW, respectively. The corresponding maximum COP values were 14.58, 12.86, and 14.19, and the exergy efficiencies (η_ex) were 45.4%, 26.8%, and 26.7%, respectively.

The outcomes of the optimization study were further analyzed in terms of COP, cycle operating pressures, and refrigerant environmental impact. In terms of COP, R744 and R463A yielded the best results, with their cycle performance indicators being most similar. Regarding cycle operating pressures, R463A and R455A exhibited lower pressures compared to R744. This aspect may have implications for the materials used in cycle components, cycle leakages, longevity, and compressor lubrication. In terms of environmental impact, R455A showed the least negative impact on the environment with a global warming potential (GWP) less than 1. However, it had a slightly lower COP compared to R744 and R463A. Additionally, the average difference in performance (COP) between R455A and R744 and R463A was approximately 11%.

Considering the need for environmentally friendly air conditioning systems with acceptable performance in cars, given their impact on the environment and contribution to global warming, it is recommended to use the R455A blend as the refrigerant in the AAC system.

6. Conclusions

The growing popularity of vehicles has led to increased time spent in cars equipped with air conditioning systems. However, the refrigerant fluids used in these systems often have significant negative environmental effects. To address this issue, many countries have implemented regulations mandating the use of refrigerants with minimal global warming potential (GWP) and zero ozone depletion potential (ODP) in cars. Therefore, there is a pressing need for environmentally friendly and high-performance air conditioning systems in vehicles, given their detrimental impact on the environment and contribution to global warming in addition to lower engine fuel consumption and hence lower carbon emissions.

The objective of this study was to propose the use of CO₂-based blend zeotropic refrigerants, specifically R-455A (R-744/32/1234yf) and R-463A (R-744/32/125/1234yf/134a), to improve the thermodynamic performance of a pure CO₂ refrigerant in automotive air conditioning systems. The study involved analyzing the thermodynamic energy and exergy of an AAC system and optimizing it using Aspen HYSYS software. The investigation focused on comparing the performance of the new zeotropic refrigerant blends with that of carbon dioxide (R-744). Additionally, the study examined the impact of cooler/condenser pressure, average evaporator temperature, cooler/condenser outlet temperature, and refrigerant flow rate on the cycles’ coefficient of performance (COP) and exergy efficiency.

The results of the analyses indicate that R-463A provided the best system performance among the investigated refrigerants, resulting in the longest possible driving range when used in automotive air conditioning systems compared to R744 and R-455A at the same operating conditions. The study identified optimal cooler/condenser pressures for R744, R455A, and R463A as 8.9, 1.65, and 2.3 MPa, respectively. The maximum COP values based on these optimal cooler/condenser pressures were 3.1, 4.25, and 5.4 for R744, R455A, and R463A, respectively. Furthermore, the study revealed the maximum refrigeration capacities (${\dot{Q}}_{evap}$) for R744, R455A, and R463A as 12.2, 9, and 12.5 kW, respectively, at a specific pressure (P₂) of 15 MPa. The improvement in ${\dot{Q}}_{evap}$ when using R463A instead of R744 was 614% at P₂ = 7 MPa and 2.5% at P₂ = 15 MPa. Similarly, the study examined the influence of average evaporator temperature (t_evap) and cooler/condenser outlet temperature (t₃) on ${\dot{Q}}_{evap}$ and COP, demonstrating the advantages of R455A and R463A over R744 in terms of performance enhancement. The improvement in COP was 35.4% and 76.5% when using R455A and R463A instead of R744 when t₃ = 20 °C, respectively. The maximum ${\dot{W}}_{comp}$ values of R744, R455A, and R463A were 6, 4, and 4.6 kW, respectively, at ${\dot{m}}_r$ $=0.15 \mathrm{k}\mathrm{g}/\mathrm{s}$. The maximum system performance and refrigeration capacities that can be achieved at the optimal operating parameters for R744, R455A, and R463A are =14.21, 15.05, 19.15 kW, COP_max = 14.58, 12.86, 14.19, and η_ex =45.4, 26.8, 26.7% respectively.

In conclusion, the study recommends using R455A as the refrigerant for AAC systems in cars, considering the need for environmentally friendly air conditioning systems with acceptable performance due to their impact on the environment and contribution to global warming. The optimal operating conditions for R455A and R463A result in improved COP and refrigeration capacity compared to R744, while also reducing leakage problems associated with higher operating pressures.

The future recommendations to improve and expand on the current study are (1) to continuously invest in research and development to explore and understand the behavior and performance of zeotropic refrigerant blends under various operating conditions (this will help in improving their efficiency and reliability); (2) to evaluate the compatibility of zeotropic blends with different materials used in the refrigeration system, including leakage, seals, gaskets, and lubricants; (3) to select the materials that are compatible with the specific blend to ensure long-term system reliability.