# Evaluating Eco-Friendly Refrigerant Alternatives for Cascade Refrigeration Systems: A Thermoeconomic Analysis

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## Abstract

**:**

## 1. Introduction

_{2}–NH

_{3}. They investigated the effect of different operational parameters on the system cost rates. They noted the optimum annualized cost of the system to be USD 109242, which was 9.34% lower than the cost of the base case system. Colorado et al. [12] carried out a thermodynamic analysis of a CRS for simultaneous cooling and heating using ammonia (NH

_{3}), R134a, butane, and propane in the low-temperature cycle (LTC) and carbon dioxide (CO

_{2}) in the high-temperature cycle (HTC) as refrigerants to find out the optimum performance and optimum design parameters. They found up to 7.3% improvement in the COP using butane in the LTC compared to that with NH

_{3}–CO

_{2}. Messineo and Panno [13] performed a thermodynamic analysis of a two-stage CRS to replace synthetic refrigerants with natural refrigerants. They obtained a maximum COP of 1.71 using natural refrigerant pair R744–R717 at evaporator and condenser temperatures of −35 °C and 35 °C, respectively. Aminyavari et al. [14] performed thermoeconomic optimization on a 50 kW cooling capacity CRS. They observed that a 24% increase in exergetic efficiency led to a 164% increase in the total cost rate. Ust and Karakurt [15] performed an exergetic analysis of a CRS using different refrigerant couples, namely, R23–R290, R23–R404A, R23–R507A, and R23–R717. They noted a decrease in the rate of exergy destruction with the increase in ${T}_{eva}$ and a rise in the rate of exergy destruction with the rise in ${T}_{cond}$ and $\mathsf{\Delta}T$. Kasi [16] carried out a numerical analysis to analyze the energetic performance of a CRS separately using R-23, R508B, and R170 in the LTC and R134a, R290, R404A, R407C, and R410A in the LTC. They observed the best performances using refrigerant pair R170–R134a and worst performances using refrigerant pair R404A–R508B among all the investigated refrigerant pairs. Rawat and Pratihar [17] thermodynamically analyzed a CRS using N

_{2}O in the LTC and five different refrigerants, namely, R717, R290, R1290, R134a, and an azeotropic mixture R507A in the HTC. Gholamian et al. [18] exergetically simulated a CO

_{2}/NH

_{3}-based CRS to determine the magnitude and location of the exergy destructions. The authors of the study determined through a conventional exergy analysis that the condenser of the NH

_{3}cycle, as well as the throttling valve and compressor of the CO

_{2}cycle, experienced the most significant exergy destruction. However, the results of an advanced exergy analysis indicated that improvements were needed in the throttling valve, compressor, and cascade condenser of the CO

_{2}cycle. In fact, the authors found that improving the performances of these components could yield a 63% improved cycle performance. Patel et al. [19] performed a comparative thermoeconomic analysis and optimization of a CRS using refrigerant pairs NH

_{3}/CO

_{2}and C

_{3}H

_{8}/CO

_{2}. They concluded that the C

_{3}H

_{8}/CO

_{2}pair offered 5.33% less cost with 6.42% more exergy destruction compared to the NH

_{3}/CO

_{2}pair. Roy and Mandal [20] presented a numerical investigation on a CRS and recommended R161 as an alternative to R404A in the high-temperature cycle. Adebayo et al. [21] performed a comparative thermodynamic analysis on a CRS using CO

_{2}in the LTC and four different refrigerants, namely, NH

_{3}, R717, HFE7000, and HFE7100 in the HTC. Aktemur and Öztürk [22] thermodynamically simulated a CRS using low GWP refrigerant R41 in the LTC and R1243zf, R423A, R601, R601A, R1233zd (E), and RE170 in the HTC. They noted the worst performances in terms of COP and exergy efficiency using R41–R423A. On the other hand, they noted the maximum COP and exergy efficiency to be 1.210 and 37.18%, respectively, using R41–RE170. In another study, Aktemur et al. [23] studied the effect of ${T}_{eva}$, ${T}_{cond}$, and ΔT on the system’s thermodynamic performances using three low-GWP refrigerant pairs, namely, R41–R601, R41-cyclopentane, and R41–R602A. They noted the best thermodynamic performances using R41–R601, which exerted maximum exergetic efficiencies of 43.10%. However, they noted a very high compressors’ discharge temperature over 120 °C for all three refrigerant pairs. Zhang et al. [24] conducted an experimental investigation on a CRS using R1270/CO

_{2}as the refrigerant combination and noted a rise in the COP and cooling capacity with the increase in T

_{HE}. Chen et al. [25] investigated the influence of the subcooling degree in the LTC of a NH

_{3}/CO

_{2}CRS to find out the thermodynamic performances of the system. They reported 4.58% improvement in the COP and 4.4% improvement in the exergy efficiency of the modified system compared to the conventional CRS when the subcooling degree was kept fixed to 10 °C. Sun and Wang [26] simulated a modified CRS for industrial application using the R1150/R717 refrigerant pair to replace the three-stage CRS using R1150/R41/R717 as the refrigerant combinations. Faruque et al. [27] simulated a CRS using Trans-2-butane, Toluene, Cyclopentane, and Cis-2-butane as refrigerants and thermodynamically analyzed the system. They noted the maximum COP and exergy efficiency using the Trans-2-butane/Toluene refrigerant combination. Cabello et al. [28] experimentally analyzed the energy performance of a CRS using four alternative refrigerant pairs, namely, R290/R744, R1270/R744, R600a/R744, and R1234ze(E)/R744 and compared the results with the system using R134a/R744. They also carried out an environmental analysis and noted less CO

_{2}emission using all four refrigerant pairs. Deymi-Dashtebayaz et al. [29] presented an energy–exergoeconomic–environmental analysis on a CRS using six pairs of low-GWP refrigerants including R41–R161, R41–R1234yf, R41–R1234ze, R744–R161, R744–R1234yf, and R744–R1234ze and reported R41–R161 and R41–R1234ze as the best refrigerant pairs in terms of the COP/exergy efficiency and total cost rate, respectively. Soni et al. [30] used CO

_{2}as a high-temperature cycle refrigerant and used different low-temperature refrigerants for the low-temperature cycle for the simulation of a CRS for ultra-low-temperature applications.

_{2}-NH

_{3}as a refrigerant pair. Therefore, it will certainly be compelling to analyze a CRS using low-GWP refrigerant pairs of hydrocarbons. The use of hydrocarbon refrigerants will solve the environmental issues of ozone depletion and global warming caused by refrigeration and air conditioning systems. In this paper, the thermoeconomic performance of a CRS was analyzed and compared, using R170–R404A and R41–R404A as refrigerant combinations to find out a possible alternative of R41. A mathematical model was developed in Engineering Equation Solver (EES) [31] software using different energy-, exergy-, and economy-based equations to carry out the simulation work.

## 2. System Description

_{eva}quantity of heat at ${T}_{eva}$ and evaporates it. While entering the LTC compressor, the vapor refrigerant receives the W

_{l}to increase its temperature and pressure. When it reaches the cascade heat exchanger, the LTC refrigerant rejects heat Q

_{cc}at ${T}_{LC}$, which is subsequently absorbed by the HTC refrigerant at T

_{HE}. As a result, the HTC refrigerant evaporates, and the LTC refrigerant condenses. The liquid refrigerant then enters the throttle valve and expands to the evaporator pressure. The cascade heat exchanger releases the vaporized HTC refrigerant, which again enters the HTC compressor. The HTC refrigerant is then passed to the condenser after being compressed by the HTC compressor to the condenser pressure, which requires W

_{h}amount of work. At ${T}_{cond}$, heat Q

_{cond}is rejected by the HTC refrigerant and becomes condensed. The condensed HTC refrigerant is then entered into the HTC throttle device and becomes expanded to the HTC evaporator. The important parameters that have a significant impact on CRS performance are ${T}_{eva}$, ${T}_{cond}$, ${T}_{LC}$, and ΔT.

## 3. Mathematical Model

#### 3.1. Energy Analysis

#### 3.2. Exergy Analysis

#### 3.3. Economic Analysis

_{T}is the total compressor power, and $\alpha $ is the unit electrical cost in USD/kWh.

_{2}avoided, and ${m}_{C{O}_{2}e}$ is the amount of annual GHG emission from the system and can be calculated as [39]

## 4. Results and Discussions

#### 4.1. Effect of ${T}_{LC}$ on COP

_{l}, COP

_{h}, and overall COP of the system, with ${T}_{LC}$ using refrigerant pair R170–R404A for a ${T}_{eva}$ and ${T}_{cond}$ of −30 °C and 40 °C, respectively, are presented in Figure 2.

_{l}decreases, and the COP

_{h}increases. As a result, the overall COP reaches its peak and then decreases with the increase in ${T}_{LC}$. Therefore, an optimal ${T}_{LC}$ exists for any fixed ${T}_{eva}$ and ${T}_{cond}$ where the system gives maximum performance. It is worth noting, as shown in Figure 2, that the optimum COP is obtained at a ${T}_{LC}$ of 7 °C when the ${T}_{eva}$ and the ${T}_{cond}$ are constant at −30 °C and 40 °C, respectively, and the value of the optimum COP is noted to be 1.845. Similarly, the optimum ${T}_{LC}$ of the system was estimated while varying the ${T}_{eva}$, ranging from −60 °C to −30 °C, and the results are presented in Table 8.

#### 4.2. Effect of ${T}_{eva}$ on COP

_{l}as well as the overall COP. The system’s COP is found slightly lower when R170 is used in the LTC instead of R41. However, the differences in the COPs using R170 as the LTC refrigerant at each ${T}_{eva}$ are very small compared to the differences in the COPs using R41 as the LTC refrigerant. The maximum and minimum differences were calculated, which are noted to be 2.79% and 1.85%, respectively, at a ${T}_{eva}$ of −30 °C and −60 °C, respectively.

#### 4.3. Effect of ${T}_{eva}$ on Compressor Discharge Temperature

#### 4.4. Effect of ${T}_{eva}$ on Exergetic Efficiency

#### 4.5. Effect of ${T}_{eva}$ on Plant Cost Rate

_{T.}, as shown in Figure 8 and Figure 9, respectively.

## 5. Conclusions

- The COPs for both systems are comparable at any temperature.
- Refrigerant pair R170–R404A shows a 1.85% to 2.79% lower COP compared to refrigerant pair R41–R404A.
- The compressor discharge temperature is in favor of the system using refrigerant R170–R404A.
- The system with R170–R404A shows a 1.5% to 2.4% lower exergetic efficiency than the other system within the investigated evaporator temperature range.
- The total annual plant cost rate of the R170–R404A system is only USD 200 higher compared to that of the R41–R404A system.

## Author Contributions

## Funding

## Institutional Review Board Statement

## Informed Consent Statement

## Data Availability Statement

## Conflicts of Interest

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**Figure 1.**(

**a**) Schematic diagram, (

**b**) P-h diagram, and (

**c**) T-s diagram of cascade refrigeration system.

**Table 1.**Thermophysical properties of R41, R161, and R170 [32].

Refrigerant | Molecular Mass (gm/mole) | Critical Temperature (°C) | Boiling Point (°C) | ASHARAE Safety Code | ODP | GWP |
---|---|---|---|---|---|---|

R170 | 30.07 | 32.2 | −88.9 | A3 | 0 | 20 |

R41 | 34.03 | 44.1 | −78.1 | A2 | 0 | 97 |

R404A | 97.6 | 72.1 | −46.6 | A1 | 0 | 3800 |

Parameters | Values | Ref. |
---|---|---|

Cooling load, Q_{eva} | 10 kW | |

LTC compressor isentropic efficiency, η_{C,l} | 80% | [33] |

HTC compressor isentropic efficiency, η_{C,h} | 80% | [33] |

Condenser temperature, ${T}_{cond}$ | 40 °C | [8] |

Dead-state temperature, T_{0} | 25 °C | [34] |

Evaporator temperature, ${T}_{eva}$ | −60 °C to −30 °C | [8] |

Superheating in the LTC and HTC | 5 °C | [8] |

U_{eva} | 0.03 kW m^{−2} K^{−1} | [35] |

U_{cond} | 0.04 kW m^{−2} K^{−1} | [35] |

U_{cc} | 1 kW m^{−2} K^{−1} | [35] |

Temperature of the inlet air to the evaporator | −10 °C | |

Maintenance factor, $\epsilon $ | 1.06 | [36] |

Interest rate, i | 14% | [36] |

Plant life time, n | 15 years | [36] |

Annual operational hour, N | 4266 h | [37] |

Electrical power cost, | 0.09 USD/kWh | [38] |

Emission factor, ${\mu}_{C{O}_{2}e}$ | 0.968 kg/kWh | [39] |

Cost of CO_{2} avoided,
${C}_{C{O}_{2}}$ | 0.09 USD/kg of CO_{2} emission | [38] |

Components | |
---|---|

Evaporator | $E{D}_{eva}=E{X}_{4}-E{X}_{1}+{Q}_{eva}\times \left(1-\frac{{T}_{0}}{{T}_{ref}}\right)$ |

LTC compressor | $E{D}_{comp,l}=E{X}_{1}-E{X}_{2}+{W}_{l}$ |

LTC expansion device | $E{D}_{\mathit{exp},l}=E{X}_{3}-E{X}_{4}$ |

HTC compressor | $E{D}_{comp,h}=E{X}_{5}-E{X}_{6}+{W}_{h}$ |

HTC expansion device | $E{D}_{\mathit{exp},h}=E{X}_{7}-E{X}_{8}$ |

Cascade condenser | $E{D}_{cc}=E{X}_{2}+E{X}_{8}-E{X}_{3}-E{X}_{5}$ |

Condenser | $\mathrm{E}{\mathrm{D}}_{\mathrm{cond}}=\mathrm{E}{\mathrm{X}}_{6}-\mathrm{E}{\mathrm{X}}_{7}$ |

**Table 4.**Capital cost components of different components of the cascade refrigeration system [34].

Components | Cost Functions |
---|---|

Evaporator | ${C}_{eva}=1397\times {A}_{eva}^{0.89}$ |

LTC compressor | ${C}_{comp,l}=10167.5\times {W}_{l}^{0.46}$ |

Cascade condenser | ${C}_{cc}=383.5\times {A}_{cc}^{0.65}$ |

LTC throttle valve | ${C}_{\mathit{exp},l}=114.5\times {\dot{m}}_{LTC}$ − 4 |

HTC compressor | ${C}_{comp,h}=9624.2\times {W}_{h}^{0.46}$ |

Condenser | ${C}_{cond}=1397\times {A}_{cond}^{0.89}$ |

HTC throttle valve | ${C}_{\mathit{exp},h}=114.5\times {\dot{m}}_{HTC}$ |

**Table 5.**Details of input parameters [43].

Parameters | Values | Parameters | Values |
---|---|---|---|

${T}_{eva}$ | −26 °C | ${T}_{cond}$ | 32 °C |

${T}_{LC}$ | −9 °C | T_{HE} | −11 °C |

Degree of superheating in the LTC | 7 °C | Degree of superheating in the HTC | 0 °C |

Degree of subcooling in the LTC | 0 °C | Degree of subcooling in the HTC | 0 °C |

η_{S,LTC} | 21% | η_{S,HTC} | 76% |

η_{m,LTC} | 93% | η_{m,HTC} | 93% |

η_{elec,LTC} | 80% | η_{elec,HTC} | 80% |

**Table 6.**Validation of the simulation model by the work of Sawalha et al. [43].

Parameters | Predicted Data | Experimental Data | Error |
---|---|---|---|

W_{l} | 1.586 | 1.62 | −2.1% |

COP_{l} | 1.892 | 1.86 | +1.72% |

COP_{h} | 2.889 | 2.65 | +9.02 |

**Table 7.**Conditions at the different state points of the system using refrigerant pair R41–R404A at base case condition.

State Pt | Temperature | Enthalpy | Entropy | Exergy | Mass Flow Rate |
---|---|---|---|---|---|

(K) | (kJ/kg) | (kJ/kg-K) | (kW) | (kg/s) | |

1 | 238 | 539.8 | 2.498 | 4.072 | 0.031 |

2 | 347 | 648 | 2.562 | 6.824 | |

3 | 279 | 215.7 | 1.054 | 7.348 | |

4 | 233 | 215.7 | 1.106 | 6.873 | |

5 | 279 | 371 | 1.623 | 5.353 | 0.121 |

6 | 326.2 | 397.9 | 1.64 | 7.991 | |

7 | 313 | 259.9 | 1.2 | 7.148 | |

8 | 274 | 259.9 | 1.218 | 6.506 |

**Table 8.**Optimum LTC temperatures for different evaporator temperatures for refrigerant pairs R170–R404A and R41–R404A.

Evaporator Temperature (°C) | R170–R404A | R41–R404A |
---|---|---|

−60 | −10 | −4 |

−55 | −8 | −2 |

−50 | −6 | 0 |

−45 | −4 | 2 |

−40 | −2 | 4 |

−35 | 0 | 5 |

−30 | 2 | 6 |

**Table 9.**Compressor discharge temperature in the HTC for refrigerant pairs R170–R404A and R41–R404A.

Evaporator Temperature (°C) | R170–R404A (°C) | R41–R404A (°C) |
---|---|---|

−60 | 57.1 | 55.4 |

−55 | 56.5 | 54.9 |

−50 | 56 | 54.4 |

−45 | 55.4 | 53.9 |

−40 | 54.9 | 53.4 |

−35 | 54.4 | 53.2 |

−30 | 53.9 | 52.9 |

Exergy Loss Percentage | ${\mathit{T}}_{\mathit{e}\mathit{v}\mathit{a}}$ °C | |||||||
---|---|---|---|---|---|---|---|---|

−60 | −55 | −50 | −45 | −40 | −35 | −30 | ||

${\delta}_{Comp,l}$ | R170–R404A | 14.06 | 14.08 | 14.05 | 13.96 | 13.80 | 13.56 | 13.25 |

R41–R404A | 14.05 | 14.34 | 14.58 | 14.76 | 14.87 | 14.55 | 14.08 | |

${\delta}_{evap,l}$ | R170–R404A | 0.12 | 0.15 | 0.18 | 0.22 | 0.27 | 0.33 | 0.40 |

R41–R404A | 0.17 | 0.21 | 0.25 | 0.28 | 0.31 | 0.32 | 0.33 | |

${\delta}_{\mathit{exp},l}$ | R170–R404A | 14.10 | 13.73 | 13.33 | 12.90 | 12.44 | 11.94 | 11.40 |

R41–R404A | 12.38 | 12.17 | 11.93 | 11.66 | 11.37 | 10.44 | 9.47 | |

${\delta}_{Comp,h}$ | R170–R404A | 18.24 | 18.37 | 18.50 | 18.61 | 18.72 | 18.81 | 18.89 |

R41–R404A | 16.41 | 16.49 | 16.56 | 16.60 | 16.63 | 17.11 | 17.61 | |

${\delta}_{\mathit{exp},h}$ | R170–R404A | 24.87 | 24.36 | 23.83 | 23.29 | 22.72 | 22.13 | 21.53 |

R41–R404A | 20.52 | 20.02 | 19.48 | 18.93 | 18.35 | 18.56 | 18.79 | |

${\delta}_{cond,h}$ | R170–R404A | 18.41 | 19.23 | 20.11 | 21.05 | 22.07 | 23.17 | 24.36 |

R41–R404A | 18.55 | 19.44 | 20.39 | 21.41 | 22.51 | 23.75 | 25.09 | |

${\delta}_{cc}$ | R170–R404A | 10.21 | 10.08 | 10.00 | 9.97 | 9.99 | 10.06 | 10.17 |

R41–R404A | 17.91 | 17.33 | 16.81 | 16.36 | 15.96 | 15.26 | 14.63 |

Evaporator Temperature (°C) | Refrigerant Pair R170/R404A | Refrigerant Pair R41/R404A | ||
---|---|---|---|---|

Exergy Destruction (kW) | Total Plant Cost Rate (USD/Year) | Exergy Destruction (kW) | Total Plant Cost Rate (USD/Year) | |

−60 | 5.577 | 23001 | 5.425 | 22805 |

−55 | 5.1 | 22000 | 4.945 | 21798 |

−50 | 4.666 | 21108 | 4.511 | 20906 |

−45 | 4.27 | 20340 | 4.117 | 20139 |

−40 | 3.908 | 19725 | 3.758 | 19528 |

−35 | 3.576 | 19330 | 3.428 | 19131 |

−30 | 3.271 | 19332 | 3.127 | 19134 |

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## Share and Cite

**MDPI and ACS Style**

Shanmugasundar, G.; Logesh, K.; Čep, R.; Roy, R. Evaluating Eco-Friendly Refrigerant Alternatives for Cascade Refrigeration Systems: A Thermoeconomic Analysis. *Processes* **2023**, *11*, 1622.
https://doi.org/10.3390/pr11061622

**AMA Style**

Shanmugasundar G, Logesh K, Čep R, Roy R. Evaluating Eco-Friendly Refrigerant Alternatives for Cascade Refrigeration Systems: A Thermoeconomic Analysis. *Processes*. 2023; 11(6):1622.
https://doi.org/10.3390/pr11061622

**Chicago/Turabian Style**

Shanmugasundar, G., Kamaraj Logesh, Robert Čep, and Ranendra Roy. 2023. "Evaluating Eco-Friendly Refrigerant Alternatives for Cascade Refrigeration Systems: A Thermoeconomic Analysis" *Processes* 11, no. 6: 1622.
https://doi.org/10.3390/pr11061622