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The present work emphasis on theoretical computation of thermodynamic performance of window air conditioner using various sustainable R290/RE170 and R1270/RE170 refrigerant mixtures as substitutes to R22. In this work, apart from R407C, twelve new binary mixtures comprising of R290, R1270 and RE170 at various compositions were developed. And also in this investigation, a MATLAB code was developed to compute the thermodynamic performance characteristics of various considered R22 alternatives at T_{k}=54.4 ℃ and T_{e}=7.2 ℃. The various performance characteristics computed are mass flow rate, refrigeration effect, compressor work, coefficient of performance (COP), pressure ratio, compressor discharge temperature, power consumed per ton of refrigeration, condenser heat rejection and volumetric cooling capacity. Results showed that the COP of refrigerant mixture RM7 (R1270/RE170 95/5 by mass %) was the highest among twelve refrigerants studied and it was 0.23 % higher than R22. Pressure ratio of RM7 (3.174) was 7.49 % lower than that of R22 (3.431). Compressor discharge temperature of all the twelve investigated refrigerants was lower in the range of 9.35 ℃ to 17.15℃ when compared with R22. Power consumed per ton of refrigeration of RM7 was 0.27 % lower than that of R22. Volumetric cooling capacity of RM7 (3833 kJ/m^{3}) was very close to that of R22 volumetric capacity (3863 kJ/m^{3}). Heat transfer through condenser of RM7 (6.372 kW) was similar to that of R22 (6.377 kW). Overall, thermodynamic performance of RM7 matches well with the performance of base line refrigerant R22 and hence, refrigerant RM7 can be considered as sustainable alternative to R22 used in air conditioners.
COP, discharge temperature, R22 alternatives, sustainable refrigerants, volumetric cooling capacity
The ongoing phase out schedule of ozone depleting substance like R22 needs the development of ecofriendly refrigerants, since Hydrochlorofluorocarbon (HCFC) refrigerant R22 causing significant ozone layer depletion and high global warming in the atmospheric environment due to the presence of chlorine atoms in R22 [1]. Therefore, Montreal protocol have decided to phaseout R22 by the year 2030 in all the nations [2, 3]. In this context, many nations are keenly focusing on the development of their own R22 alternatives. Earlier, various experimental and theoretical performance studies were conducted in order to find viable alternatives to R22.
Experimental investigation on airtowater heat pumps was carried out with different refrigerants including R22, R134a and R404A in order to find the optimal refrigerant charge [4]. Results showed that the overcharging and undercharging of refrigerant tends to decrease the performance of heat pump. Theoretical performance investigation was done with various refrigerants like R134a, R430A, R440A and R450A used in domestic refrigerator [5]. Results revealed that the coefficient of performance (COP) of R440A was the highest among R134a, R430A and R450A. Experimental performance assessment of air conditioner was done with R410A under various indoor working conditions [6]. Results showed that the performance of air conditioner with R410A was better under wide range of operating conditions.
An extensive study reported that refrigerant R32 was recommended as a viable candidate to replace R410A used in air conditioners [7]. SegmenttoSegment modeling of microchannel heat exchangers (condenser and evaporator) was developed, in order to predict the performance of air conditioning system using various alternative refrigerants [8]. The benefit of microchannel heat exchangers was it reduces the charge of refrigerant required to the system. From this modeling of microchannel heat exchangers, it was observed that R32 shows the better performance compared to R410A whereas R22 and R290 had similar performance, however R290 required a larger displacement volume of compressor.
Theoretical performance investigation was conducted on simple vapour compression refrigeration (VCR) system working with R432A and R433A as alternatives to R22 [9]. Results exhibited that the performance of R432A and R433A was closer to refrigerant R22. Experimental studies revealed that R407C was considered as an appropriate retrofit refrigerant to R22 [10]. Theoretical analysis of exergy revealed that R507A was an appropriate refrigerant to substitute R502 rather than R404A [11]. Theoretical performance investigation was done on simple VCR system working with R134a, RE170 and R510A, under various operating conditions [12]. Results showed that the COP of RE170 and R510A was better than R134a.
Experimental performance tests were conducted in a vapour compression system of 5TR capacity using both variable speed compressor and electronic expansion valve, operating at different evaporator temperatures. In this investigation, refrigerants tested were R22, R32, R438A, R404A, R410A, R290 and R1270 [13]. Test results shown that COP of R290 and R1270 was better than R22 and other investigated refrigerants. Experimental and theoretical investigations were done in air conditioner of 3.5kW capacity working with R22 and R161 under various operating conditions [14]. Results revealed that the performance of R161 was better than refrigerant R22. Experimental investigation on ice cream refrigerator was done with R404A and its alternative refrigerant mixture R290/RE170 [15]. Test results shown that the time taken to form ice cream for the mixture R290/RE170 (65/35 by mass %) was less compared to R404A and also its COP was better than R404A. Experimental studies were conducted in a heat pump bench tester with R22 and R432A (R1270/RE170 80/20 by mass %) under both the cooling mode (air conditioning) and heating mode (heat pumping) conditions [16]. Test results exhibited that the COP of R432A was higher in the range of 8.5 to 8.7%, compared to R22 in both the conditions.
A comprehensive analysis of an automotive condenser under maldistribution of intake air was carried out by both experimental and numerical simulations [17]. Simulation studies were done using Coil Designer software, in order to simulate the heat transfer and fluid dynamics of cross flow heat exchanger (condenser), whereas experiments were carried out in a condenser of a stationary test facility for an automotive air conditioning system over a wide range of blockage conditions. In this study, the experimental results exhibited good agreement with the simulation results, which assures the applicability of simulation results for situations where experimental results were not available. From this investigation it was also observed that, in case of the blocked condition, the heat exchange zone was extended to a longer length depending upon the category and magnitude of the blockage. Experimental performance study and exergy analysis was carried out with R1234yf used in automotive air conditioning system [18]. In this study, the performance of R1234yf systems with and without internal heat exchanger (IHX) was compared with an R134a system. Results showed that COP of R1234yf system without IHX was 3.6 to 4.5 % lower compared to the R134a system, whereas COP of R1234yf system with IHX was 0.9 % higher compared to the R134a system at the compressor speed of 2500 rpm. The exergy destruction ratio (EDR) of R1234yf system with IHX was 1.2 % lower compared to the R134a system at the compressor speed of 2500 rpm.
Experimental performance tests were conducted with low global warming potential (GWP) refrigerant blend R1234yf/R134a (89/11 by mass %) as substitute to R134a under both the cooling and heating modes respectively [19]. Results revealed that, capacities of R1234yf/R134a (89/11 by mass %) and R134a were similar in both cooling and heating modes. The COP of R1234yf/R134a (89/11 by mass %) was 4 to 16% lower than R134a in heating mode, whereas it was 4 to 9% lower than R134a in cooling mode. Experimental test results revealed that the refrigerant mixture R290/R600/R600a (50/40/10 by mass %) would be considered as a suitable replacement to R134a used in automotive air conditioners, since it has higher COP than R134a in all the tested conditions and also it has similar working temperature and pressure as that of R134a [20]. A literature review revealed that hydrocarbons and their mixtures would be considered as suitable alternatives to halogenated refrigerants [21]. A literature review reported that R717 and R744 were considered as feasible refrigerants for industrial applications whereas R290, R600a, R1150 and R1270 were considered as viable refrigerants for air conditioning applications [22].
From the literature, it is noticed that the refrigerant Dimethylether (RE170) is considered as a viable ecofriendly refrigerant and it can be used as a blending component with different refrigerants in order to reduce the global warming potential (GWP) as well as to enhance the COP of refrigerant blend. Therefore, the present study focuses on theoretical thermodynamic performance investigation of standard vapour compression refrigeration system working with various sustainable refrigerant blends like R290/RE170 and R1270/RE170 as substitutes to R22 used in air conditioners. And also in this study, thermodynamic performance results found from the MATLAB code have been validated with the results published in the literature.
In this work apart from R407C, twelve new binary refrigerant blends comprising of R290, R1270 and RE170 at various compositions were developed. Environmental properties like ozone depletion potential (ODP) and global warming potential (GWP) of various considered refrigerants were taken from the ASHRAE data hand book [23]. The designation of refrigerant blends that were developed in this study is given in Table 1. Similarly basic properties of all the considered R22 alternative refrigerants are given in Table 2.
Table 1. Designation of various developed alternative refrigerants
Designation of refrigerants 
Composition (mass %) 
R22 
Pure refrigerant 
RM1 (R290/RE170) 
95/5 
RM2 (R290/RE170) 
90/10 
RM3 (R290/RE170) 
85/15 
RM4 (R290/RE170) 
80/20 
RM5 (R290/RE170) 
75/25 
RM6 (R290/RE170) 
70/30 
RM7 (R1270/RE170) 
95/5 
RM8 (R1270/RE170) 
90/10 
RM9 (R1270/RE170) 
85/15 
RM10 (R1270/RE170) 
80/20 
RM11 (R1270/RE170) 
75/25 
RM12 (R1270/RE170) 
70/30 
R407C (R32/R125/R134a) 
23/25/52 
From Table 2, it is noticed that all the twelve developed refrigerant mixtures (RM1 to RM12) possess zero ODP and very low GWP compared to R22. Therefore refrigerants (RM1 to RM12) can be considered as ecofriendly refrigerants. From Table 2, it is also observed that the refrigerant mixtures (RM1 to RM4) are considered as near azeotropic due to their lower temperature glide (less than 0.5 ℃), whereas refrigerants (RM5 to RM12 and R407C) are categorized into non azeotropic mixtures, since these refrigerants possess higher temperature glide.
Table 2. Basic properties of developed alternative refrigerants
Refrigerants 
MW (kg/kmol) 
BP (^{0}C) 
T_{bub} (^{0}C) 
T_{dew} (^{0}C) 
T*_{glide} (^{0}C) 
T_{c} (K) 
P_{c} (MPa) 
ODP 
GWP (100 years) 
R22 
86.5 
40.81 
0 
0 
0 
369.3 
4.99 
0.055 
1760 
RM1 
44.19 
 
41.99 
41.97 
0.02 
370.13 
4.287 
0 
<3 
RM2 
44.285 
 
41.76 
41.67 
0.09 
370.48 
4.325 
0 
<3 
RM3 
44.381 
 
41.42 
41.20 
0.22 
370.97 
4.365 
0 
<3 
RM4 
44.477 
 
40.93 
40.47 
0.46 
371.59 
4.407 
0 
<3 
RM5 
44.573 
 
41.11 
40.30 
0.81 
372.36 
4.453 
0 
<3 
RM6 
44.669 
 
39.51 
38.23 
1.28 
373.29 
4.502 
0 
<3 
RM7 
42.263 
 
46.36 
45.68 
0.68 
365.69 
4.604 
0 
<3 
RM8 
42.447 
 
45.07 
43.69 
1.38 
367.22 
4.654 
0 
<3 
RM9 
42.633 
 
43.76 
41.69 
2.07 
368.79 
4.705 
0 
<3 
RM10 
42.821 
 
42.45 
39.71 
2.74 
370.41 
4.756 
0 
<3 
RM11 
43.011 
 
44.53 
41.15 
3.38 
372.06 
4.807 
0 
<3 
RM12 
43.202 
 
39.86 
35.88 
3.98 
373.76 
4.858 
0 
<3 
R407C 
86.204 
 
43.62 
36.62 
7.0 
359.18 
4.631 
0 
1774 
T*_{glide} = (T_{dew}T_{bub}) at 0.101325MPa
2.1 Development of properties of alternative refrigerants
Generally, thermodynamic properties of refrigerants are essential to compute performance characteristics of R22 and its various considered alternatives. Therefore, in this work, thermodynamic properties of all the considered R22 alternative refrigerants were developed with the help of a MartinHou equation of state (MHEOS). The significance of MHEOS is that, it gives better accuracy in results while computing thermodynamic properties of refrigerants [2426]. Literatures on applicability of MHEOS were discussed below.
Experimental studies were conducted using a constant volume apparatus, in order to measure the vapor phase PvT Properties of refrigerant R1225ye(Z) over temperature ranges from (263 to 368) K and pressure ranges from (135 to 777) kPa [27]. In this study, the experimental vapour phase PvT Properties of R1225ye(Z) were compared with a MartinHou equation of state (MHEOS). Results showed that the deviations between computed values of PvT using MHEOS and experimental values were (0.54 to 0.52) % with an average absolute deviation (AAD) of 0.170. Experimental tests were performed using a constant volume apparatus, in order to measure the vapor phase PvT Properties of refrigerant R1243zf over temperature ranges from (268 to 368) K and pressure ranges from (220 to 910) kPa [28]. In this work, the experimental vapour phase PvT Properties of R1243zf were compared with a MartinHou equation of state (MHEOS). Results revealed that the deviations between computed values of PvT using MHEOS and experimental values were (0.65 to 1.27) % with an average absolute deviation (AAD) of 0.305.
Experimental investigations were carried out by using a constant volume apparatus, in order to measure the vapor phase PvT Properties of refrigerant R1234ze(E) over temperature ranges from (243 to 373) K and pressure ranges from (57 to 1024) kPa [29]. In this study, the experimental vapour phase PvT Properties of R1234ze(E) were compared with a MartinHou equation of state (MHEOS). Results shown that the deviations between computed values of PvT using MHEOS and experimental values were (0.35 to 0.38) % with an average absolute deviation (AAD) of 0.198. Experimental tests were done by using a constant volume apparatus, in order to measure the vapor phase PvT Properties of refrigerant HFO1234yf over temperature ranges from (243 to 373) K and pressure ranges from (84 to 3716) kPa [30]. In this study, the experimental vapour phase PvT Properties of HFO1234yf were compared with a MartinHou equation of state (MHEOS). Results exhibited that the deviations between computed values of PvT using MHEOS and experimental values were (1.18 to 1.27) % with an average absolute deviation (AAD) of 0.44.
Dong et al., proposed an empirical correlation, to conveniently compute parameter B4 of the MartinHou equation of state (MHEOS) at various temperatures by relating B4(T) with B4(0.7) (B4 at reduced temperature=0.7) after analyzing B4(T) of more than 200 compounds [31]. In this study, the accuracy of MHEOS to compute volume of various refrigerants by using proposed correlation was verified by comparing the calculated data with literature data. An empirical correlation used to compute B4 was given below.
Results showed that the proposed equation improves the accuracy in results of properties obtained from the MHEOS and also significantly decreases the computational time to determine parameter B_{4}, which would enhance the application of the MHEOS in engineering calculations.
From the available literature, it is evident that the MartinHou equation of state (MHEOS) can be conveniently used in chemical engineering, mechanical engineering, and refrigeration technology design. Therefore, MHEOS was used in the present study to compute thermodynamic properties of various considered R22 alternative refrigerants. The methodology followed to establish thermodynamic properties of refrigerants was explained below.
2.2 Methodology to establish thermodynamic properties of pure and mixture refrigerants
Procedure followed to establish thermodynamic properties of pure and mixture refrigerants was taken from literature and it was discussed below [32, 33].
2.2.1 Methodology to develop properties of pure refrigerants
Correlations used to develop thermodynamic properties of refrigerants are given in this section. PressureEnthalpy (Ph) chart used while developing properties of pure refrigerants is shown in Figure 1. Step by step procedure followed to develop the properties of refrigerants is given below.
Figure 1. Ph chart for computing properties of pure refrigerant
where x=1T/T_{c}; A, B, C and D are constants for a particular refrigerant and these constants for several refrigerants were available in the literature [34]. For example, saturation vapour pressure constants for refrigerant R22 are given in Table 3.
Table 3. Constants of R22 for equation (4)
A 
B 
C 
D 
7.0682 
1.52369 
1.8545 
2.8439 
where ω is acentric factor; T_{r}=T/Tc, ω, ρ_{c} and T_{c }are constants for a particular refrigerant and these constants for several refrigerants were available in the literature [33]. For example, density correlation constants for refrigerant R22 are given in Table 4.
Table 4. Constants of R22 for equation (5)
ω 
ρ_{c} (kg/m^{3}) 
T_{c} (K) 
0.221 
523.8 
369.15 
$P=\frac{R T}{vb}+\frac{A_{2}+B_{2} T+C_{2} e^{\frac{5.475 T}{T_{C}}}}{(vb)^{2}}+\frac{A_{3}+B_{3} T+C_{3} e^{\frac{5.475 T}{T_{C}}}}{(vb)^{3}}+\frac{A_{4}}{(vb)^{4}}+\frac{B_{5} T}{(vb)^{5}}$ (6)
where
$b=V_c\frac{\beta V_c}{15Z_c}$ (7)
$\beta=31.883 Z_{C}^{2}+20.533Z_C$ (8)
$Z_C=\frac{P_cV_c}{RT_c}$ (9)
where A_{2}, A_{3}, A_{4}, B_{2}, B_{3}, B_{5}, C_{2}, C_{3} and b are dimensionless coefficients of MHEOS for various refrigerants. Procedure followed to compute above coefficients was described in the literature [24, 31]. By solving the above equation (6), dimensionless coefficients of MHEOS for R22 are listed in Table 5.
Table 5. Dimensionless coefficients of R22 for equation (6)
Dimensionless coefficients of R22 
Values 
A_{2} 
139.154038231457 
A_{3} 
0.295289024195263 
A_{4} 
0.000104165697806786 
B_{2} 
0.128645931301646 
B_{3} 
0.000446322328392750 
B_{5} 
8.14900447033360 x10^{11} 
C_{2} 
2292.28498497122 
C_{3} 
3.44337587584321 
b 
0.000407841281333333 
(v) Compute enthalpy and entropy properties (both liquid and vapour phase) of given refrigerant by using departure method. Generally, the reference state of enthalpy and entropy was fixed while computing properties. İn case of refrigerants, the reference state chosen is that of saturated liquid at 0 ℃. Enthalpy and entropy values assigned to the reference state of saturated liquid at 0 ℃ are usually h_{1}=h_{f1}=200 kJ/kg and S_{1}=S_{f1}=1.0 kJ/kg K respectively [32, 33].
The significance of departure function is, to compute the enthalpy and entropy at various points as shown in Figure 1. For example, to compute enthalpy at point 3, enthalpy departure method is applied and corresponding departure term (h_{3}h_{2}) is given as follows.
${{h}_{4}}{{h}_{3}}=\int\limits_{3}^{4}{C_{P0}dT}$(13)
In the present work, ideal gas heat capacity $(C_{P0})$ correlation was taken from the literature and it is given below [33].
$C_P0=H_0+H_1 T+H_2 T^2+H_3 T^3+H_4 T^4$ (14)
where H_{0}, H_{1}, H_{2}, H_{3} and H_{4} are constants for a particluar refrigerant and these constants for several refrigerants were avialable in the literature [33]. For example, constants of ideal gas heat capacity correlation for R22 are given in Table 6.
Table 6. Constants of R22 for equation (14)
C_{P0} constants 
Values 
H_{0} 
3.164 
H_{1} 
10.422 x10^{3} 
H_{2} 
1.179 x10^{5} 
H_{3} 
2.650 x10^{8} 
H_{4} 
1.222 x10^{11} 
$h_5h_4=(U_5U_4 )+(P_5 V_5P_4 V_4 )$ (15)
${{U}_{5}}{{U}_{4}}=\int\limits_{4}^{5}{\left[ T{{\left[ \frac{\partial P}{\partial T} \right]}_{V}}P \right]}dV$(16)
By solving above equations (15) and (16), the value of h_{5} can be found. On the other hand, saturated liquid enthalpy at state point 6 can be found by using the following relation.
$h_5h_6=h_fg$ (17)
$h_6=h_5h_fg$ (18)
where h_{fg }can be found by using ClasiusClayperon equation at a given temperature.
(vi) Find the liquid entropy of given refrigerant. In order to compute the thermodynamic properties (enthalpy and entropy) of any given refrigerant at any given pressure and temperature, the departure method was used, and the corresponding saturated liquid enthalpy and saturated liquid entropy was computed by using the ClausiusClapeyron equation. Entropy of liquid for any given refrigerant can be calculated as follows.
$S_fg=S_gS_f$ (19)
$S_f=S_gS_fg$ (20)
(vii) Entropy of vapour for any given refrigerant can be computed as follows.
$S_{fg}=\frac{h_{fg}}{T_{sat}}$ (21)
$S_{fg}=\frac{h_{g}}{T_{sat}}$ (22)
By following the above methodology, thermodynamic properties of various pure refrigerants can be found.
2.2.2 Methodology to develop properties of mixture refrigerants
Correlations used to develop thermodynamic properties of refrigerant mixtures were given in this section. PressureEnthalpy (Ph) chart used while developing properties of refrigerant mixtures is shown in Figure 2.
Figure 2. Ph diagram for computing properties of refrigerant mixtures
(i) Generally, thermodynamic properties data of pure refrigerants was taken into account, while computing properties of refrigerant mixtures.
(ii) In order to compute the bubble point temperature and dew point temperature of refrigerant mixtures, use the interpolation method by taking saturation temperature and saturation pressure data of pure refrigerants.
(iii) For example, correlation used to compute the bubble point temperature of refrigerant mixture is given below.
$p=x_1 p_1^{sat}+x_2 p_2^{sat}$ (23)
where x_{1} and x_{2} are mole fractions of pure components in the liquid phase.
(iv) Similarly, correlation used to compute the dew point temperature of refrigerant mixture is given below.
$p=\frac{p_{1}^{sat}p_{2}^{sat}}{p_{1}^{sat}y_1(p_{1}^{sat}p_{2}^{sat})}$ (24)
where y_{1} and y_{2} are mole fractions of pure components in the vapour phase.
(v) Mixing rules and binary interaction parameter used while developing and establishing the thermodynamic properties of refrigerant mixtures were taken from the literature and these rules were used to find critical temperature and critical pressure of the refrigerant mixture [33].
$T_{cm}= y_i^2 T_i+y_j^2 T_j+2y_i y_j T_{cij}$ (25)
$T_{cij}=(1k_{ij} ) (T_{ci} T_{cj} )^{1/2}$ (26)
$1k_{ij}=\frac{8(V_{ci}V_{cj})^{1/2}}{(V_{ci}^{1/3}+V_{cj}^{1/3})^{3}}$ (27)
${{P}_{cm}}=({{Z}_{cm}}R{{T}_{cm}})/({{V}_{cm}})={\left( \left( \sum\limits_{i=1}^{n}{{{y}_{i}}{{Z}_{ci}}} \right)(R)\left( \sum\limits_{i=1}^{n}{{{y}_{i}}{{T}_{ci}}} \right) \right)}/{\left( \sum\limits_{i=1}^{n}{{{y}_{i}}{{V}_{ci}}} \right)}\;$ (28)
The binary interaction parameter k_{ij} was given by
$k_{ij}=1\frac{8(V_{ci}V_{cj})^{1/2}}{(V_{ci}^{1/3}+V_{cj}^{1/3})^{3}}$ (29)
(vi) Compute the specific volume of vapour for a given refrigerant mixture. In the present investigation, specific volume of vapour for all the considered refrigerant mixtures was computed by using MartinHou equation of state (MHEOS) [24, 31].
(vii) In order to compute the liquid density of refrigerant mixtures for a given pressure, use the interpolation method by taking liquid density properties data of pure refrigerants.
(viii) Similarly departure method was used to find the enthalpy and entropy of various refrigerant mixtures and the procedure used for the departure method was described in the literature [32, 33].
By following the above methodology, thermodynamic properties of various considered R22 alternative refrigerant mixtures were computed. For example, the computed thermodynamic properties of various refrigerants like R22, R407C and RM7 (R1270/RE170 95/5 by mass %) were compared with the properties obtained from NIST REFPROP 9.1 [36]. Results showed that the computed properties show good agreement with NIST REFPROP 9.1. The deviation between computed properties of R22, R407C and RM7 using MHEOS and REPROP properties was within 2% for the given operating conditions. Therefore, the methodology followed to establish properties of R22, R407C and RM7 can be considered as reliable. Hence, the same methodology was followed to develop thermodynamic properties of various new refrigerants considered for the study, since properties of various new refrigerants were not available in the literature.
Investigation on flammability of refrigerants is important for researchers while developing the alternative refrigerants. ASHRAE safety standard 34 exhibits that, flammability of refrigerants were classified into various safety groups like nonflammable (ASHRAE A1), weakly flammable (ASHRAE A2) and flammable (ASHRAE A3) groups respectively [37]. From this safety standard, it was found that, refrigerants R22, R134a, R125 and R407C were classified into nonflammable category (A1) whereas R32 was classified into weakly flammable (A2). Similarly refrigerants R290, R1270 and RE170 were classified into flammable group (A3).
Table 7. RF number and flammability group of various investigated refrigerants
Refrigerants 
RF Number (kJ/g) 
ASHRAE Flammability Group 
RM1 
52.01 
A3* 
RM2 
51.83 
A3* 
RM3 
51.68 
A3* 
RM4 
51.56 
A3* 
RM5 
51.46 
A3* 
RM6 
51.40 
A3* 
RM7 
61.02 
A3* 
RM8 
60.50 
A3* 
RM9 
60.01 
A3* 
RM10 
59.53 
A3* 
RM11 
59.08 
A3* 
RM12 
58.66 
A3* 
* Estimated values of RF number
Table 8. ASHRAE safety group of various investigated R22 alternative refrigerants
Refrigerants 
ASHRAE Safety Group 
ASHRAE Flammability 
R22 
A1 
A1 Nonflammable 
R134a 
A1 
A1 Nonflammable 
R125 
A1 
A1 Nonflammable 
R32 
A2 
A2 Weakly flammable 
R407C 
A1 
A1 Nonflammable 
R290 
A3 
A3 Flammable 
R1270 
A3 
A3 Flammable 
RE170 
A3 
A3 Flammable 
RM1 
A3* 
A3* Flammable 
RM2 
A3* 
A3* Flammable 
RM3 
A3* 
A3* Flammable 
RM4 
A3* 
A3* Flammable 
RM5 
A3* 
A3* Flammable 
RM6 
A3* 
A3* Flammable 
RM7 
A3* 
A3* Flammable 
RM8 
A3* 
A3* Flammable 
RM9 
A3* 
A3* Flammable 
RM10 
A3* 
A3* Flammable 
RM11 
A3* 
A3* Flammable 
RM12 
A3* 
A3* Flammable 
*Estimated
However flammability category of various new refrigerant mixtures (RM1 to RM12) considered in this study were not available in the ASHRAE safety standard 34 and hence, refrigerant flammability number (RF number) was used in this investigation for assessing the flammability of new refrigerant blends. RF number shows good agreement with that of ASHRAE safety standard 34 which was used for classifying the refrigerants into various flammability categories [38]. It is reliable to express the hazards of combustion with respect to flammability limits of each refrigerant by using RF number. Based on the values of RF number, refrigerants were categorized into various safety groups [38]. If RF number of refrigerants is less than 30 kJ/g, then they are classified into weakly flammable group (ASHRAE A2) and if it is in between 30 to 150 kJ/g, then they are classified into flammable group (ASHRAE A3). An empirical correlation used for computing the RF number of different refrigerants studied is given below.
$RF=\left\{ {{\left( \frac{U}{L} \right)}^{0.5}}1 \right\}\times \frac{HOC}{MW}$(30)
By using above correlation, values of RF number of various new refrigerant mixtures (RM1 to RM12) were computed and they are given in Table 7. Similarly, the summary of flammability groups of all the R22 alternative refrigerants studied in this investigation are given in Table 8.
From Table 7, it was found that the flammability category of all the twelve investigated refrigerants (RM1 to RM12) are classified into ASHRAE A3 flammability group, since RF number of these refrigerants was in between 30 to 150 kJ/g.
Basically, window air conditioners work on the principle of vapour compression refrigeration (VCR) system. The basic representation of VCR system is shown in Figure 3. Normally, VCR cycle consists of four basic processes like isentropic compression, constant pressure condensation, isenthalpic expansion and constant pressure evaporation. In majority of literature, thermodynamic performance analysis of air conditioners was done based on either simple saturation vapour compression refrigeration cycle or standard vapour compression cycle [3942].
Table 9. Description of simple saturation vapour compression refrigeration cycle
State points of the cycle 
Description 
12 
Isentropic compression 
23 
Constant pressure condensation 
34 
Isenthalpic expansion 
41 
Constant pressure evaporation 
Table 10. Description of standard vapour compression refrigeration cycle
State points of the cycle 
Description 
1”2” 
Isentropic compression 
2”3” 
Constant pressure condensation 
3”4” 
Isenthalpic expansion 
4”1” 
Constant pressure evaporation 
11” 
Degree of superheating 
33” 
Degree of subcooling 
PressureEnthalpy (Ph) diagrams of simple and standard vapour compression cycle are shown in Figure 4a and 4b respectively. Either in simple saturation cycle or standard cycle, pressure losses and heat losses to the surroundings from condensers and evaporators were neglected. Similarly suction line pressure drop, discharge line pressure drop and heat gain or heat losses occur at various devices of the system were neglected for the ease of theoretical computations. Description of various state points of simple saturation cycle is given in Table 9. Similarly, description of various thermodynamic stages of standard vapour compression refrigeration cycle is given in Table 10.
Figure 3. Schematic diagram of vapour compression refrigeration system
Figure 4a. Ph diagram of simple saturation vapour compression refrigeration cycle without superheating and subcooling
Figure 4b. Ph diagram of standard vapour compression refrigeration cycle with superheating and subcooling
In case of simple saturation cycle, effect of superheating and subcooling on performance of the system was not considered, whereas in standard vapour compression cycle, effect of superheating and subcooling on thermodynamic performance of the system was considered. Therefore, the present study focuses on theoretical performance evaluation of air conditioner using various alternative refrigerants based on standard vapour compression refrigeration cycle. Assumptions made, while doing thermodynamic analysis of standard vapour compression system were taken from literature and they are given below [3942].
(i) Pressure drops in the condenser and evaporator are negligible.
(ii) Heat losses to the environment from the various devices like evaporator and condenser are negligible
(iii) Flow across the expansion valve is isenthalpic.
(iv) Flow across the compressor is isentropic.
Commonly, thermodynamic performance analysis of air conditioners is carried out, in order to find the suitable alternative to R22. Thermodynamic performance characteristics of R22 and its considered alternatives are computed at AHRI (Air conditioning, Heating and Refrigeration Institute) conditions. Generally, AHRI conditions were used in the performance computation of air conditioners and these operating conditions are given in Table 11. And also in this study, performance parameters of various R22 alternatives are computed for various evaporator temperatures by keeping the condenser temperature constant. Capacity of air conditioner was taken as 5.25 kW.
In this investigation, a MATLAB program was developed to compute the thermodynamic performance characteristics of various R22 alternatives. All the governing equations used to compute performance parameters of various developed R22 alternatives were incorporated in the program. The significance of MATLAB program is that, it incorporates the saturated and superheated properties of given refrigerants and also, it includes any given operating conditions of the system in order to find the various performance parameters of the alternative refrigerants. And also in this study, thermodynamic performance results found from the MATLAB code have been validated with the results published in the literature.
Table 11. AHRI conditions for air conditioners
Operating condtions 
Temperature (^{0}C) 
Evaporator temperature 
7.2 
Condenser temperature 
54.4 
Superheating 
11.1 
Subcooling 
8.3 
5.1 Performance computations
Governing equations used to compute thermodynamic performance characteristics of standard vapour compression refrigeration cycle operating with various R22 alternatives were taken from the literature and they are given below [3942].
Refrigerant mass flow rate is computed as
$\dot{m} =\frac{Q_c}{RE}$ (31)
Pressure ratio is computed by
$P_r=\left(\frac{P_k}{P_e}\right)$ (32)
Refrigeration effect is calculated as
$RE= h_{1"}h_{4"}$ (33)
Isentropic compressor work is computed by
$W_c= h_{2"}h_{1"}$ (34)
Coefficient of performance (COP) is calculated as
$COP=RE⁄W_c $ (35)
Condenser heat rejection (CHR) is calculated as
$CHR=(h_{2"}h_{3"} )$ (36)
Heat transfer through condenser (Q_{k}) is computed as
$Q_k=\dot{m}(h_{2"}h_{3"} )$ (37)
Volumetric cooling capacity is computed by
$VCC=ρ_{1"}×RE$ (38)
Power consumed per ton of refrigeration is calculated as
$PPTR=\dot{m}W_c=3.5167\left(\frac{h_{2"}h_{1"}}{h_{1"}h_{4"}}\right)$ $PPTR=\dot{m}W_c=3.5167\left(\frac{W_c}{RE}\right)=\left(\frac{3.5167}{COP}\right)$ $\dot{m}=\frac{Q_c}{RE}=\frac{3.5167}{RE}$(39)
Discharge temperature of compressor can be found with the help of refrigerants superheated properties tables and by interpolating for the given superheating value, equivalent to difference in entropy which is known.
Table 12. Summary of results of various investigated R22 alternatives
Refrigerants 
$\dot{m}$ (kg/s) 
RE (kJ/kg) 
W_{c} (kJ/kg) 
COP 
Change in COP (%) 
P_{k} (MPa) 
P_{e} (MPa) 
P_{r} 
R22 
0.03383 
155.915 
32.940 
4.733 
0 
2.1562 
0.6284 
3.431 
RM1 
0.01962 
268.840 
57.083 
4.709 
0.507 
1.8714 
0.5841 
3.203 
RM2 
0.01989 
265.079 
57.044 
4.646 
1.838 
1.8815 
0.5838 
3.222 
RM3 
0.02031 
259.708 
57.318 
4.531 
4.267 
1.8818 
0.5800 
3.244 
RM4 
0.02049 
257.444 
57.572 
4.471 
5.535 
1.8699 
0.5756 
3.248 
RM5 
0.02092 
252.068 
58.188 
4.331 
8.49 
1.8659 
0.5678 
3.286 
RM6 
0.01887 
279.532 
59.434 
4.703 
0.63 
1.8691 
0.5582 
3.348 
RM7 
0.01902 
277.274 
58.444 
4.744 
0.23 
2.2075 
0.6954 
3.174 
RM8 
0.01879 
280.733 
59.648 
4.706 
0.57 
2.1698 
0.6737 
3.220 
RM9 
0.01859 
283.691 
60.697 
4.673 
1.26 
2.1322 
0.6506 
3.277 
RM10 
0.01841 
286.469 
61.734 
4.640 
1.96 
2.0943 
0.6287 
3.331 
RM11 
0.01824 
289.130 
62.598 
4.618 
2.42 
2.0536 
0.6050 
3.394 
RM12 
0.01804 
292.268 
63.460 
4.605 
2.70 
2.0133 
0.5918 
3.401 
R407C 
0.03477 
151.686 
36.505 
4.155 
12.21 
2.4349 
0.5860 
4.155 
Table 12. Continued…
Refrigerants 
T_{d} (^{0}C) 
PPTR (kW/TR) 
VCC (kJ/m^{3}) 
CHR (kJ/kg) 
Q_{k} (kW) 
R22 
85.19 
0.743 
3868 
188.509 
6.377 
RM1 
68.04 
0.746 
3210 
325.244 
6.381 
RM2 
68.66 
0.756 
3150 
321.958 
6.403 
RM3 
69.06 
0.776 
3103 
317.827 
6.455 
RM4 
69.36 
0.786 
3018 
314.483 
6.443 
RM5 
70.14 
0.812 
2942 
312.253 
6.532 
RM6 
71.15 
0.747 
3185 
338.882 
6.394 
RM7 
73.28 
0.741 
3833 
335.018 
6.372 
RM8 
74.08 
0.747 
3740 
340.871 
6.405 
RM9 
74.65 
0.752 
3643 
342.979 
6.375 
RM10 
74.75 
0.757 
3554 
349.604 
6.436 
RM11 
75.20 
0.761 
3486 
352.306 
6.426 
RM12 
75.84 
0.763 
3372 
356.667 
6.434 
R407C 
82.45 
0.846 
3529 
188.542 
6.555 
Results of thermodynamic performance parameters of various considered alternatives were compared with the base line refrigerant R22 and they are given in Table 12.
5.2 Validation of results based on literature
In the present study, a MATLAB program was developed to compute performance parameters of various R22 alternatives. Results obtained from present program have been validated with literature results [40]. Dalkilic and Wongwises computed performance characteristics of simple saturation vapour compression cycle using refrigerant R22 at T_{e}=10 ℃ and T_{k}=50 ℃ with no subcooling and superheating. For validation, same R22 refrigerant and operating conditions were used in the program as that of researchers. The deviation of program results when compared with literature results is within 1%. Therefore, the program which is developed in this study can be considered as reliable and thus it can be employed for the thermodynamic analysis of various alternative refrigerants considered for the study. The deviation between present work results and literature results is given in Table 13.
However, in this investigation, validation of various performance parameters like coefficient of performance (COP), power per ton of refrigeration and volumetric cooling capacity of R22 for different evaporator temperatures are validated with literature results and they are shown in Figures 5, 6 and 7 respectively [40].
From Figures 5, 6 and 7, it was observed that, present work results of COP, power per ton of refrigeration and volumetric cooling capacity of baseline refrigerant R22 for various evaporator temperatures exhibit good agreement with literature results. Therefore the present MATLAB program which is developed in this study can be considered as reliable.
Table 13. Comparison of performance parameters of R22 with literature results [40]
S.no 
Performance Parameters 
Dalkilic and Wongwises Results [40] 
Present Work Results 
Deviation (%) 
1 
P_{k} (MPa) 
1.943 
1.9427 
0.015 
2 
P_{e} (MPa) 
0.355 
0.35481 
0.053 
3 
P_{r} 
5.4732 
5.4753 
0.038 
4 
RE (kJ/kg) 
138 
137.9490 
0.036 
5 
Wc (kJ/kg) 
43.40 
43.7740 
0.861 
6 
COP 
3.180 
3.1514 
0.899 
7 
PPTR (kW/TR) 
1.101 
1.1106 
0.871 
8 
VCC (kJ/m^{3}) 
2094 
2072 
1.050 
Figure 5. Comparison of COP for R22 at different evaporator temperatures with literature results
Figure 6. Comparison of power per ton of refrigeration for R22 at different evaporator temperatures with literature results
Figure 7. Comparison of volumetric cooling capacity for R22 at different evaporator temperatures with literature results
Figure 8. Effect of evaporator temperature on the refrigeration effect at T_{k}=54.4 ℃
Figure 8 shows the effect of different evaporator temperatures on the refrigeration effect of various R22 alternatives at T_{k}=54.4 ℃. From Figure 8, it is noticed that the refrigeration effect increases with increase in evaporator temperature for all the investigated refrigerants. Particularly refrigeration effect of refrigerants (RM1 to RM12) is higher than R407C and R22, since conventional refrigerants are blends of hydrocarbons, which will have high latent heat of vapourization compared to R407C and R22.
6.2 Compressor work
Figure 9. Effect of evaporator temperature on the compressor work at T_{k}=54.4 ℃
Figure 9 shows the effect of different evaporator temperatures on the compressor work of various R22 alternative refrigerants at T_{k}=54.4 ℃. From Figure 9, it is noticed that the compressor work input decreases with increase in evaporator temperature for all the considered refrigerants. Particularly compressor work of refrigerants (RM1 to RM12) is higher than R22, since these refrigerants are blends of hydrocarbons, which will have higher vapour enthalpy compared to baseline refrigerant R22.
6.3 Coefficient of performance
Figure 10. Effect of evaporator temperature on COP at T_{k}=54.4 ℃
Figure 10 shows the effect of different evaporator temperatures on the coefficient of performance (COP) of various R22 alternatives at T_{k}=54.4 ℃. COP can be measured as an energy efficiency index of the equipment, while it is working with specific refrigerant. From Figure 10, it is evident that the COP increases with increase in evaporator temperature for all the investigated refrigerants, since COP depends upon both the refrigeration effect and work of compressor. Mainly COP of refrigerant mixture RM7 (R1270/RE170 95/5 by mass %) is the highest among twelve investigated refrigerants and it is 0.23% higher than that of COP of R22.
6.4 Pressure ratio
Figure 11 shows the effect of different evaporator temperatures on the pressure ratio of various R22 alternative refrigerants at T_{k}=54.4 ℃. From Figure 11, it is noticed that the pressure ratio decreases with increase in evaporator temperature for all the considered refrigerants. This is due to their increase in evaporator pressure with increase in evaporator temperature. From Figure 11 and Table 12, it is clear that the pressure ratio of twelve refrigerants studied (RM1 to RM12) is lower in the range of 0.88% to 7.94% compared to R22 whereas pressure ratio of R407C is the highest among twelve investigated refrigerants and it is 21.10% higher than R22. The high pressure ratio causes significant increase in discharge temperature of compressor.
Figure 11. Effect of evaporator temperature on the pressure ratio at T_{k}=54.4 ℃
6.5 Compressor discharge temperature
Figure 12 shows the effect of different evaporator temperatures on the compressor discharge temperature of various R22 alternatives at T_{k}=54.4 ℃. From Figure 12, it is noticed that the compressor discharge temperature decreases with increase in evaporator temperature for all the investigated refrigerants, since pressure ratio of refrigerants decreases with increase in evaporator temperature. Compressor discharge temperature shows the lifetime of compressor motor and hence it is important to compute the compressor discharge temperature, while it is working with various alternative refrigerants. The excessive discharge temperature causes burnt out of compressor motor windings, which in turn reduces the lifespan of compressor significantly. Therefore compressor discharge temperature should be as low as possible from the stand point of lifespan of compressor motor is concerned. From Figure 12 and Table 12 it is evident that the compressor discharge temperature of all the twelve investigated refrigerants (RM1 to RM12) is lower in the range of 9.35 ℃ to 17.15 ℃ when compared with R22. This is due to their lower pressure ratio compared to R22 and hence refrigerants (RM1 to RM12) exhibit better durability of compressor motor.
Figure 12. Effect of evaporator temperature on the compressor discharge temperature at T_{k}=54.4 ℃
6.6 Power consumed per ton of refrigeration
Figure 13. Effect of evaporator temperature on the power per ton of refrigeration at T_{k}=54.4 ℃
Figure 13 shows the effect of different evaporator temperatures on the power consumed per ton of refrigeration of various R22 alternative refrigerants at T_{k}=54.4 ℃. It indicates the power consumed by the compressor in order to produce per ton of refrigeration and it is inversely proportional to the COP of a given refrigerant. From Figure 13, it is observed that the power consumed per ton of refrigeration decreases with increase in evaporator temperature for all the considered refrigerants, since COP of refrigerants increases with increase in evaporator temperature. From Figure 13, it is evident that the power consumed by the compressor per ton of refrigeration of refrigerant mixture RM7 (R1270/RE170 95/5 by mass %) is the lowest among twelve investigated refrigerants and it is 0.27 % lower than that of R22, since COP of RM7 is the highest among all the investigated refrigerants.
6.7 Condenser heat rejection
Figure 14 shows the effect of different evaporator temperatures on the condenser heat rejection of various R22 alternatives at T_{k}=54.4 ℃. From Figure 14, it is noticed that the condenser heat rejection decreases with increase in evaporator temperature for all the investigated refrigerants. Particularly condenser heat rejection of refrigerants (RM1 to RM12) is higher than R22, since these refrigerants are the blends of hydrocarbons, which will have high latent heat of condensation compared to baseline refrigerant R22.
Figure 14. Effect of evaporator temperature on the condenser heat rejection at T_{k}=54.4 ℃
6.8 Volumetric cooling capacity
Figure 15 shows the effect of different evaporator temperatures on the volumetric cooling capacity of various R22 alternative refrigerants at T_{k}=54.4^{0}C. It indicates the size of compressor required in order to produce desired cooling effect. Volumetric cooling capacity depends upon the vapour density occurs at the outlet of evaporator and also on the cooling effect. From Figure 15, it is observed that the volumetric cooling capacity increases with increase in evaporator temperature for all the considered R22 alternative refrigerants, since increase in volumetric capacity depends on both the values of vapour density and cooling effect of refrigerants. From Figure 15, it is evident that the volumetric cooling capacity of refrigerant mixture RM7 (R1270/RE170 95/5 by mass %) is the highest among twelve studied refrigerants and it is very close to that of volumetric capacity of R22. Therefore same size of R22 compressor can be used for RM7 without modifications.
Figure 15. Effect of evaporator temperature on the volumetric cooling capacity at T_{k}=54.4 ℃
6.9 Heat transfer through condenser
Figure 16 shows the effect of different evaporator temperatures on the heat transfer through condenser of various R22 alternatives at T_{k}=54.4^{0}C. Heat transfer through condenser denotes the load taken by the condenser to reject heat to the surroundings for a given refrigerant. It depends upon both the mass flow rate and latent heat of condensation of given refrigerant. From Figure 16, it is noticed that the heat transfer through condenser decreases with increase in evaporator temperature for all the investigated refrigerants, since decrease in heat transfer depends on both the values of mass flow rate and latent heat of condensation of refrigerants. From Figure 16, it is clear that, heat transfer through condenser of refrigerant mixture RM7 (R1270/RE170 95/5 by mass %) is the lowest among twelve investigated refrigerants and it is very close to that of heat transfer through condenser of R22.
Figure 16. Effect of evaporator temperature on the heat transfer through condenser at T_{k}=54.4 ℃
This investigation presents the theoretical thermodynamic performance analysis of window air conditioner operating with various sustainable refrigerants as substitutes to R22. The conclusions drawn from the thermodynamic analysis of various R22 alternative refrigerants are presented below.
BP 
Boiling point, ℃ 
COP 
Coefficient of performance, Dimensionless 
CHR 
Condenser heat rejection, kJ/kg 
HOC 
Enthalpy of combustion, kJ/mol 
L 
Lower flammability limit, kg/m^{3} 
MW 
Molecular weight, kg/kmol 
PPTR 
Power required per ton of refrigeration, kW/TR 
RE 
Refrigeration effect, kJ/kg 
TR 
Ton of refrigeration, kW 
U 
Upper flammability limit, kg/m^{3} 
VCC 
Volumetric cooling capacity, kJ/m^{3} 
C_{p0} 
Ideal gas heat capacity, J/mol K 
h 
Enthalpy, kJ/kg 
h_{f} 
Liquid enthalpy, kJ/kg 
h_{fg} 
Enthalpy of vapourization, kJ/kg 
h_{g} 
Vapour enthalpy, kJ/kg 
h_{1”} 
Enthalpy at compressor inlet, kJ/kg 
h_{2”} 
Enthalpy at compressor outlet, kJ/kg 
h_{3”} 
Enthalpy at condenser outlet, kJ/kg 
h_{4”} 
Enthalpy at evaporator inlet, kJ/kg 
P 
Pressure, MPa 
P_{c} 
Critical pressure, Mpa 
P_{cm} 
Critical pressure of mixture, MPa 
P_{e} 
Evaporating pressure, MPa 
P_{k} 
Condensing pressure, MPa 
P_{r} 
Pressure ratio, Dimensionless 
P_{sat} 
Saturation pressure, MPa 
Q_{c} 
Refrigeration capacity, kW 
Q_{k} 
Heat transfer through condenser, kW 
R 
Universal gas constant, J/mol K 
S_{f} 
Liquid entropy, kJ/kg K 
S_{fg} 
Entropy of vapourization, kJ/kg K 
S_{g} 
Vapour entropy, kJ/kg K 
T 
Temperature, K 
T_{bub} 
Bubble point temperature, ^{0}C 
T_{c} 
Critical temperature, K 
T_{cm} 
Critical temperature of mixture, K 
T_{dew} 
Dew point temperature, ℃ 
T_{e} 
Evaporating temperature, ℃ 
T_{glide} 
Temperature glide, ℃ 
T_{k} 
Condensing temperature, ℃ 
T_{sat} 
Saturation temperature, K 
U 
Internal energy, kJ/kg 
V 
Specific volume, m^{3}/kg 
V_{cm} 
Critical volume of mixture, m^{3}/kg 
V_{g} 
Vapour volume, m^{3}/kg 
V_{f} 
Liquid volume, m^{3}/kg 
W_{c} 
Specific work of compressor, kJ/kg 
Z_{cm} 
Critical compressability factor of mixture, Dimensionless 
ρ 
Density, kg/m^{3} 
ρ_{c} 
Critical density, kg/m^{3} 
ω 
Accentric factor, Dimensionless 
Greek symbols 

ρ 
Density, kg/m^{3} 
ω 
Accentric factor, Dimensionless 
Subscripts 

c 
Critical 
f 
Liquid phase 
g 
Vapour phase 
m 
Mixture 
Abbreviations 

AHRI 
AirConditioning, Heating, and Refrigeration Institute 
ASHRAE 
American Society of Heating, Refrigerating and AirConditioning Engineers 
BP 
Boiling point 
CHR 
Condenser heat rejection 
COP 
Coefficient of performance 
GWP 
Global warming potential 
HCFCs 
Hydrochlorofluorocarbons 
HFO 
Hydrofluoroolefin 
MHEOS 
MartinHou equation of state 
ODP 
Ozone depleting potential 
PPTR 
Power required per ton of refrigeration 
RE 
Refrigeration effect 
RF 
Refrigerant flammability 
TR 
Ton of refrigeration 
VCC 
Volumetric cooling capacity 
VCR 
Vapour compression refrigeration 
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