#
First and Second Law Analyses of Trans-critical N_{2}O Refrigeration Cycle Using an Ejector

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## Abstract

**:**

_{2}O ejector-expansion cycle is 63% and 53% less than for IHEC and VCRC, respectively. Furthermore, the highest COPs for the vapor-compression refrigeration, the internal heat exchanger and the ejector-expansion refrigeration cycles correspond to a high side pressure of 7.3 MPa, and the highest COPs for the three types of CO

_{2}refrigeration cycles correspond to a high side pressure of 8.5 MPa. Consequently, these lead to a lower electrical power consumption by the compressor.

## 1. Introduction

_{2}, one type of natural refrigerant, is non-flammable, non-toxic and non-corrosive fluid. Nitrous oxide (N

_{2}O) is a non-toxic fluid, albeit with a somewhat higher global warming potential than CO

_{2}[1,2]. It has thermodynamic similarities in critical temperature and pressure and molar weight with carbon dioxide, and could be replaced with CO

_{2}. N

_{2}O has a critical temperature, boiling point and triple point of 36.4, −88.5 and −90.82 °C, respectively [3]. The use of ejectors in refrigeration systems has been practiced by many investigators, in large part because they do not have moving parts and need no work for compression. These components reduce exergy destruction and can be utilized when there are three pressure levels in a system.

_{2}refrigeration cycle with an ejector, and studied the relation between ejector entrainment ratio and vapor quality at the outlet of the ejector. The results showed that the entrainment ratio decreases as compressor discharge pressure increases, which is the opposite of what was observed for vapor quality. Moreover, the authors concluded that the system has an optimum entrainment ratio corresponding to the maximum COP for the system.

_{2}O refrigeration cycle, and investigated the effect of superheating in the evaporator, internal heat exchange and the use of a recovery turbine instead of an expansion valve on cycle behavior. They also compared the cycle with a CO

_{2}refrigerant cycle, and found that the trans-critical N

_{2}O cycle has a higher cooling coefficient of performance, a lower compressor pressure ratio, and a lower discharge pressure and temperature than the carbon dioxide refrigerant cycle. Aghazadeh Dokandari et al. [8] investigated a novel configuration for ejector expansion in a CO

_{2}/NH

_{3}cascade cycle, and performed first and second law analyses of its performance; the theoretical analysis of the functional features based on the first and second laws of the thermodynamics illustrated that the maximum COP and the maximum second law efficiency are on average 7% and 5%, respectively, higher than for the conventional cycle.

_{2}O and CO

_{2}as refrigerants in various configurations for heating and refrigeration and optimized the cycles.

_{2}/NH

_{3}cascade refrigeration systems to obtain the lowest exergy destruction and the highest COP.

_{con}(43–53°C) and of an evaporator temperature T

_{eva}(3–11°C), they showed that the coefficient of performance could increase by up to 26%.

_{2}cycle in the top and bottom portions of both cycles, respectively. Yari [18] also studied the performance of a novel two-stage ejector-expansion trans-critical refrigeration cycle.

_{2}O based trans-critical refrigeration system in which an ejector is used as an expansion device. Their system is found to have a higher COP, a lower compressor discharge pressure and a higher entrainment ratio but suffers from having a lower volumetric cooling capacity. They reported that the maximum COP is about 10% higher compared to the case when CO

_{2}is used as a working fluid.

_{2}O has attracted the attention of investigators because of some advantageous features in its thermodynamic properties. To the authors’ knowledge, the use of N

_{2}O in a trans-critical refrigeration cycle with an ejector has not been investigated thermodynamically yet and the present work addresses this lack of information. This investigation aims to improve understanding of the system, and help in comparing its performance with the performances of other refrigeration system.

## 2. System Description

_{2}O refrigeration cycle are considered: vapor-compression refrigeration cycle (VCRC), internal heat exchanger cycle (IHEC) and ejector-expansion refrigeration cycle (EERC).

#### 2.1. Ejector–Expansion Refrigeration Cycle

_{2}O cycle with an ejector system. The system includes a gas cooler, compressor, ejector, evaporator, expansion valve and vapor–liquid separator.

_{2}O enters the compressor at state (1) at pressure P

_{s}and is pressurized to P

_{d}(the high-side pressure) at state (2), with an isentropic compressor efficiency, η

_{c}[4]. The ideal compression process follows process 1–2 s. Then, the refrigerant cools at constant pressure P

_{d}in the gas cooler (process 2–3). The working fluid passes through an ejector nozzle with a nozzle isentropic efficiency η

_{n}of 0.7 [4] at the following stage, expanding to a subcritical condition at state (4) at pressure P

_{e}. The saturated secondary vapor stream enters the ejector at a pressure of P

_{e}in accordance with state 9. The two streams mix at constant pressure and the final state of the mixture is state 5. Note that after the mixture goes through the ejector diffuser with a diffuser isentropic efficiency η

_{d}of 0.8 [4], it recovers to the pressure P

_{s}at state 6. Subsequently, the mixture enters the vapor–liquid separator. The vapor component enters the compressor at state (1), while the liquid enters expansion valve at state (7), where its pressure is reduced through an isenthalpic process. Then, refrigerant is evaporated isobarically in the evaporator via process 8–9.

#### 2.2. Vapor-Compression Refrigeration Cycle (VCRC), Internal Heat Exchanger Cycle (IHEC)

#### 2.3. Assumptions

_{2}O cycles, several assumptions are invoked:

- All processes are at steady state and there are no flow losses within the system.
- There is no heat loss from the compressor.
- The fluid undergoes a constant enthalpy process in the expansion valve.
- The N
_{2}O at the inlet of the compressor and the outlet of the evaporator is a saturated vapor. - The mixing process in the mixing chamber occurs at a constant evaporation pressure.
- Changes in the potential and kinetic energies at the outlet and inlet of components are negligible.
- The dead state temperature for the analysis is 35 °C.
- The heat sink temperature is 5 °C higher than the evaporation temperature.

## 3. System Modelling

#### 3.1. Energy Analysis

_{2}O entering the ejector nozzle. That is,

_{2}O which flows to the separator, the suction and motive mass flow rates, respectively, are as follows [4]:

#### 3.2. Exergy Analysis

## 4. Results and Discussion

_{2}as the working fluid for these types of cycles, to permit comparisons with the models proposed by Deng et al. [4].

_{2}cycles developed in this paper. As shown, the results are approximately in line with those from the proposed model in reference [4].

_{2}O is performed. The operating conditions for these types of cycles are given in Table 4, Table 5 and Table 6.

#### 4.1. Ejector Entrainment Ratio Analysis

_{2}O refrigerant with high-side pressure, for three evaporation temperatures. The entrainment ratio increases greatly during the first small rise in compressor discharge pressure starting at 7 MPa, and then experiences a moderate rise as the pressure increases further. This also holds true for the vapor quality at the outlet of the ejector diffuser, except that the changes are declines. In fact, the entrainment ratio and vapor quality are related and vapor quality can be calculated as a function of entrainment ratio, and vice versa. Furthermore, an increased evaporation temperature leads to a higher entrainment ratio and a lower vapor quality. The sharp variations in these results are attributed to the thermodynamic properties changes in the vicinity of N

_{2}O critical point.

#### 4.2. Comparison of Three Trans-critical N_{2}O Refrigeration Cycles

_{2}refrigerant, the high-side pressure corresponding to the highest COP is about 8.5 MPa (Figure 4). Thus, the N

_{2}O cycle compressor consumes much less power to pressurize the working fluid. Table 7 lists the exergy destructions in all systems for a high-side pressure of 7.3 MPa and 1 kg/s of N

_{2}O, for the three system configurations. The total exergy destruction for the N

_{2}O ejector-expansion cycle is 63% and 53% less than for the IHEC and the VCRC, respectively. The exergy destroyed in the expansion process is 19% and 40% lower than the total exergy destruction in than IHEC and VCRC, respectively. As a result, the ejector can decrease significantly the exergy destruction associated with the throttling process in the expansion valves by reducing the pressure difference between the inlet and outlet of the expansion valve. Furthermore, Table 7 illustrates the total exergy destruction in the same cycles using CO

_{2}as the working fluid. The exergy destructions are lower for the EERC, IHEC and VCRC employing N

_{2}O in comparison with those for the cycles using CO

_{2}which means that using N

_{2}O as a working fluid is more practical than using CO

_{2}for these refrigeration systems.

#### 4.3. Comparison of Three Cycle Types Using Trans-critical N_{2}O and CO_{2} Working Fluids

_{2}O and CO

_{2}as working fluids, Figure 8 illustrates the maximum coefficient of performance and maximum exergy efficiency of the three types of the cycles using both working fluids. Each of the cycles employing N

_{2}O exhibits a higher COP and exergy efficiency compared to the cycles employing CO

_{2}. The maximum COP and maximum exergy efficiency for the EERC using N

_{2}O are higher than the corresponding values for the IHEC employing the same working fluid by about 12% and 15%, while the maximum COP and exergy efficiency are 14% and 16.5% higher than corresponding values for the VCRC using N

_{2}O.

_{2}O and CO

_{2}refrigerants, respectively.

## 5. Conclusions

_{2}O as the working fluid was investigated, and energy and exergy analyses were carried out. The effects of key factors on system performance were determined and this system was compared with others employing the same refrigerant as well as systems using CO

_{2}as the working fluid. Furthermore, the results from this study were validated using results for similar systems proposed in other studies using CO

_{2}as the working fluid. Three types of cycles are considered: vapor-compression refrigeration cycle (VCRC), internal heat exchanger cycle (IHEC) and ejector-expansion refrigeration cycle (EERC). The results for the cycles using N

_{2}O showed that the ejector entrainment ratio, one of the important parameters in ejector-expansion cycles representing the proportion of vapor and liquid in the outlet of ejector, varies significantly with high-side pressure of the cycle the quality in the outlet of the ejector. Increasing entrainment ratio by raising compressor discharge pressure causes the coefficient of performance to increase to a peak and then sharply decrease. Moreover, coefficient of performance for the cycle exhibits an optimum value for the different evaporation cycles and a higher evaporation temperature results in a higher system COP. This variation is the opposite of that observed for the cycle using N

_{2}O. Then, the temperature and pressure at the outlet of the ejector smoothly decline to a minimum as the entrainment ratio increases. The comparison between three types of N

_{2}O refrigeration cycles shows that the maximum coefficient of performance for the cycles occurs at roughly the same high side pressure, and that the maximum value is exhibited by the EERC cycle. The same occurs for exergy efficiency of the cycles. The highest COP in this study corresponds to a high side pressure of 7.3 MPa for three types of N

_{2}O refrigeration cycles, but about 8.5 MPa for three types of CO

_{2}refrigeration cycles. Consequently, the compressor in the N

_{2}O systems requires less work to pressurize the working fluid in the system. The exergy analysis also identifies the exergy destroyed in the system in three types of cycles. The total exergy destruction in the N

_{2}O ejector-expansion cycle was seen to be 63% and 53% less than the values for the IHEC and VCRC, respectively. A comparison of the total exergy destruction is also made between the cycles using CO

_{2}and N

_{2}O working fluids. The results show that there is less exergy destruction in the EERC, IHEC and VCRC employing N

_{2}O than in those using CO

_{2}, which means that using N

_{2}O is better than CO

_{2}for these refrigeration systems. To enhance the comparison, the maximum COP and exergy efficiency of the three types of cycles using both working fluids were examined. It was seen that these parameters are higher for the EERC than the IHEC and VCRC, and that each type of cycle employing N

_{2}O has a higher COP and exergy efficiency than the corresponding cycle using CO

_{2}. Further, the maximum COP and exergy efficiency for the EERC using N

_{2}O are higher than those values for the IHEC employing the same working fluid by about 12% and 15%, respectively. Meanwhile the maximum COP and exergy efficiency are 14% and 16.5% higher than VCRC using N

_{2}O, respectively. Thus, this leads to an optimal cooling system with lower power consumption in compressor. Finally, linear regression is applied to determine the optimum high-side pressure, the maximum COP and the maximum exergy efficiency, as functions of the gas cooler and evaporator temperatures.

## Acknowledgments

## Author Contributions

## Conflicts of Interest

## Nomenclature

Ψ | Specific stream exergy (kJ/kg) |

η | Efficiency |

ε | Heat exchanger effectiveness (%) |

COP | Coefficient of performance |

ex | exergy (kJ) |

h | Specific enthalpy (kJ/kg) |

$\dot{I}$ | Exergy destruction rate (kW) |

$\dot{m}$ | Mass flow rate (kg/s) |

P | Pressure (MPa) |

$\dot{Q}$ | Heat transfer rate (kW) |

s | Specific entropy (kJ/kg K) |

$\dot{S}$ | Entropy generation rate (kW/K) |

T | Temperature (°C) |

U | Entrainment ratio of ejector |

v | Velocity (m/s) |

$\dot{\mathrm{W}}$ | Work rate (kW) |

x | Quality (kg/kg) |

Subscripts | |

amb | Ambient |

c | Compressor |

d | Discharge pressure of ejector; Diffuser |

eva /e | Evaporator |

exp | Expansion valve |

gc | Gas cooler |

gen | Generation |

Hex | Heat Exchanger |

n | Ejector nozzle |

r | Heat sink temperature |

s | Vapor–liquid separator |

0 | Dead state of system |

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**Figure 1.**(

**a**) P-h diagram; and (

**b**) schematic of the trans-critical N

_{2}O cycle with an ejector expansion system.

**Figure 4.**Validation of the proposed model based on CO

_{2}as the working fluid for three types of cycles.

**Figure 5.**Variation with compressor discharge pressure of: (

**a**) entrainment ratio; and (

**b**) vapor quality.

**Figure 6.**Variation with entrainment ratio of: (

**a**) COP; and (

**b**) pressure and temperature of ejector outlet.

**Figure 7.**Variation with compressor discharge pressure of: (

**a**) COP; and (

**b**) exergy efficiency of the system, for three system configurations.

**Table 1.**Thermodynamic equations of the trans-critical N

_{2}O cycle with an ejector expansion system.

Unit | Energy Equations | Exergy Destruction Equations |
---|---|---|

Compressor | ${\eta}_{c}=1.003-0.121\left(\raisebox{1ex}{${P}_{d}$}\!\left/ \!\raisebox{-1ex}{${P}_{s}$}\right.\right)$, ${\eta}_{c}=\frac{{h}_{2s}-{h}_{1}}{{h}_{2}-{h}_{1}}$ ${W}_{c}=\frac{{h}_{2}-{h}_{1}}{U+1}$ | ${I}_{c}=\frac{1}{U+1}\left[{T}_{0}\left({s}_{2}-{s}_{1}\right)\right]$ |

Gas cooler | ${Q}_{gc}=\frac{{h}_{2}-{h}_{3}}{U+1}$ | ${I}_{gc}=\frac{\left[\left({h}_{2}-{h}_{3}\right)-{T}_{0}\left({s}_{2}-{s}_{3}\right)\right]}{U+1}$ |

Ejector | ${\eta}_{n}=\frac{{h}_{3}-{h}_{4}}{{h}_{3}-{h}_{4s}}$, $\frac{{v}_{4}{}^{2}}{2}={h}_{3}-{h}_{4}$ ${v}_{5}=\frac{1}{U+1}{v}_{4}$, ${x}_{6}=\frac{1}{U+1}$, ${\eta}_{d}=\frac{{h}_{6s}-{h}_{5}}{{h}_{6}-{h}_{5}}$, $\frac{{v}_{5}{}^{2}}{2}={h}_{6}-{h}_{5}$, ${h}_{6}=\frac{{h}_{3}}{U+1}+\frac{U}{U+1}{h}_{9}$ | ${I}_{ejector}=[{s}_{6}-\frac{{s}_{3}}{U+1}-\frac{U}{U+1}{s}_{9}]$ |

Expansion valve | ${h}_{7}={h}_{8}$ | ${I}_{exp}=\frac{U}{U+1}\left[{T}_{0}\left({s}_{8}-{s}_{7}\right)\right]$ |

Evaporator | ${Q}_{eva}=\frac{U}{U+1}\left({h}_{9}-{h}_{8}\right)$ | ${I}_{eva}=\frac{U}{U+1}{T}_{0}\left[\left({s}_{9}-{s}_{8}\right)-\frac{\left({h}_{9}-{h}_{8}\right)}{{T}_{r}}\right]$ |

Unit | Energy Equations | Exergy Destruction Equations |
---|---|---|

Compressor | ${\eta}_{c}=1.003-0.121\left(\raisebox{1ex}{${P}_{d}$}\!\left/ \!\raisebox{-1ex}{${P}_{e}$}\right.\right)$, ${\eta}_{c}=\frac{{h}_{2s}-{h}_{1}}{{h}_{2}-{h}_{1}}$ ${W}_{c}={h}_{2}-{h}_{1}$ | ${I}_{c}={T}_{0}\left({s}_{2}-{s}_{1}\right)$ |

Gas cooler | ${Q}_{gc}={h}_{2}-{h}_{3}$ | ${I}_{gc}=\left({h}_{2}-{h}_{3}\right)-{T}_{0}\left({s}_{2}-{s}_{3}\right)$ |

Expansion valve | ${h}_{3}={h}_{4}$ | ${I}_{exp}={T}_{0}\left({s}_{4}-{s}_{3}\right)$ |

Evaporator | ${Q}_{eva}={h}_{1}-{h}_{4}$ | ${I}_{eva}={T}_{0}\left[\left({s}_{1}-{s}_{4}\right)-\frac{\left({h}_{1}-{h}_{4}\right)}{{T}_{r}}\right]$ |

Unit | Energy Equations | Exergy Destruction Equations |
---|---|---|

Compressor | ${\eta}_{c}=1.003-0.121\left(\raisebox{1ex}{${P}_{d}$}\!\left/ \!\raisebox{-1ex}{${P}_{e}$}\right.\right)$, ${\eta}_{c}=\frac{{h}_{2s}-{h}_{1}}{{h}_{2}-{h}_{1}}$ ${W}_{c}={h}_{2}-{h}_{1}$ | ${I}_{c}={T}_{0}\left({s}_{2}-{s}_{1}\right)$ |

Gas cooler | ${Q}_{gc}={h}_{2}-{h}_{3}$ | ${I}_{gc}=\left({h}_{2}-{h}_{3}\right)-{T}_{0}\left({s}_{2}-{s}_{3}\right)$ |

Internal Hex | ${\epsilon}_{Hex}=\frac{{h}_{1}-{h}_{6}}{{h}_{3}-{h}_{6}}$ | ${I}_{Hex}=\left[\left({h}_{3}-{h}_{4}\right)-{T}_{0}\left({s}_{3}-{s}_{4}\right)\right]$− $\left[\left({h}_{1}-{h}_{6}\right)-{T}_{0}\left({s}_{1}-{s}_{6}\right)\right]$ |

Expansion valve | ${h}_{4}={h}_{5}$ | ${I}_{exp}={T}_{0}\left({s}_{5}-{s}_{4}\right)$ |

Evaporator | ${Q}_{eva}={h}_{6}-{h}_{5}$ | ${I}_{eva}={T}_{0}\left[\left({s}_{6}-{s}_{5}\right)-\frac{\left({h}_{6}-{h}_{5}\right)}{{T}_{r}}\right]$ |

**Table 4.**State quantities for the vapor-compression N

_{2}O trans-critical refrigeration cycle at the following conditions: ${\mathrm{T}}_{\mathrm{eva}}=5\text{}\xb0\mathrm{C},{\mathrm{T}}_{\mathrm{gc}}=36\text{}\xb0\mathrm{C},{\mathrm{T}}_{\mathrm{amb}}=35\text{}\xb0\mathrm{C},{\text{}\mathrm{P}}_{\mathrm{gc}}=8.5\text{}\mathrm{MPa}$.

State | T (°C) | P (MPa) | h (kJ/kg) | s (kJ/kg. K) | X | $\dot{\mathit{m}}$ (kg/s) |
---|---|---|---|---|---|---|

1 | 5 | 3.53 | 396.3 | 1.54 | 1 | 1 |

2 | 80.25 | 8.5 | 444.6 | 1.58 | - | 1 |

3 | 36 | 8.5 | 254 | 0.99 | - | 1 |

4 | 5 | 3.53 | 254 | 1.03 | 0.35 | 1 |

**Table 5.**State quantities for the internal heat exchanger N

_{2}O trans-critical cycle at the following conditions: ${\mathrm{T}}_{\mathrm{eva}}=5\text{}\xb0\mathrm{C},{\mathrm{T}}_{\mathrm{gc}}=36\text{}\xb0\mathrm{C},{\mathrm{T}}_{\mathrm{amb}}=35\text{}\xb0\mathrm{C},{\text{}\mathrm{P}}_{\mathrm{gc}}=8.5\text{}\mathrm{MPa}$.

State | T (°C) | P (MPa) | h (kJ/kg) | s (kJ/kg. K) | X | $\dot{\mathit{m}}$ (kg/s) |
---|---|---|---|---|---|---|

1 | 34.4 | 3.53 | 439.9 | 1.69 | - | 1 |

2 | 119.6 | 8.5 | 502.8 | 1.74 | - | 1 |

3 | 36 | 8.5 | 254 | 0.99 | - | 1 |

4 | 21.5 | 8.5 | 210.4 | 0.85 | - | 1 |

5 | 5 | 3.53 | 210.4 | 0.87 | 0.15 | 1 |

6 | 5 | 3.53 | 396.3 | 1.54 | 1 | 1 |

**Table 6.**State quantities for the ejector expansion N

_{2}O trans-critical refrigeration cycle at the following conditions: ${\mathrm{T}}_{\mathrm{eva}}=5\text{}\xb0\mathrm{C},{\mathrm{T}}_{\mathrm{gc}}=36\text{}\xb0\mathrm{C},{\mathrm{T}}_{\mathrm{amb}}=35\text{}\xb0\mathrm{C},{\text{}\mathrm{P}}_{\mathrm{gc}}=8.5\text{}\mathrm{MPa}$.

State | T (°C) | P (MPa) | h (kJ/kg) | s (kJ/kg. K) | X | $\dot{\mathit{m}}$ (kg/s) |
---|---|---|---|---|---|---|

1 | 8.5 | 3.86 | 394.4 | 1.52 | 1 | 0.6 |

2 | 75 | 8.5 | 435.1 | 1.55 | - | 0.6 |

3 | 36 | 8.5 | 254 | 0.99 | - | 0.6 |

4 | 5 | 3.53 | 246.9 | 1.01 | 0.32 | 0.6 |

5 | 5 | 3.53 | 308.3 | 1.22 | 0.59 | 1 |

6 | 8.5 | 3.86 | 310.8 | 1.23 | 0.60 | 1 |

7 | 8.5 | 3.86 | 185 | 0.781 | 0 | 0.4 |

8 | 5 | 3.53 | 185 | 0.783 | 0.04 | 0.4 |

9 | 5 | 3.53 | 396.3 | 1.54 | 1 | 0.4 |

**Table 7.**Exergy destructions in the three types of N

_{2}O refrigeration cycles for a high-side pressure of 7.3 MPa.

Device | EERC | IHEC | VCRC | ||||||
---|---|---|---|---|---|---|---|---|---|

Loss (kJ) | (%) | Loss/ex_{eva} | Loss (kJ) | (%) | Loss/ex_{eva} | Loss (kJ) | (%) | Loss/ex_{eva} | |

Compressor | 4.02 | 32.46 | 0.58 | 10.08 | 29.77 | 0.67 | 8.74 | 32.69 | 0.77 |

Gas cooler | 1.98 | 15.96 | 0.29 | 11.30 | 33.38 | 0.75 | 4.25 | 15.92 | 0.35 |

Ejector | 4.68 | 37.79 | 0.68 | - | - | - | - | - | - |

Exp. valve | 0.19 | 1.55 | 0.03 | 7.09 | 20.94 | 0.48 | 11.25 | 42.05 | 1.03 |

Evaporator | 1.51 | 12.22 | 0.22 | 3.29 | 9.74 | 0.22 | 2.49 | 9.33 | 0.22 |

Internal HX | - | - | - | 2.08 | 6.15 | 0.14 | - | - | - |

Overall (N _{2}O cycle) | 12.40 | 100 | 1.80 | 33.85 | 100 | 2.26 | 26.75 | 100 | 2.37 |

Overall (CO _{2} cycle) * | 13.65 | 100 | 2.20 | 38.05 | 100 | 2.74 | 30.92 | 100 | 2.94 |

_{2}refrigeration cycle for a high-side pressure of 8.5 MPa.

**Table 8.**Optimum high-side pressure of ejector-expansion refrigeration cycle (EERC) for CO

_{2}refrigerant and its corresponding maximum COP and ${\mathsf{\eta}}_{\mathrm{ex}}$ for various design parameters.

T_{gc} (°C) | T_{eva} (°C) | P_{OPT,High-side} (MPa) | COP_{max} | ${\mathsf{\eta}}_{\mathbf{ex},\mathbf{max}}$ (%) |
---|---|---|---|---|

36 | 0 | 8.678 | 2.89 | 31.4 |

5 | 8.637 | 3.51 | 31.2 | |

10 | 8.658 | 4.33 | 30.3 | |

38 | 0 | 9.128 | 2.61 | 28.3 |

5 | 9.128 | 3.14 | 28.0 | |

10 | 9.120 | 3.82 | 26.7 | |

40 | 0 | 9.573 | 2.37 | 25.7 |

5 | 9.585 | 2.83 | 25.2 | |

10 | 9.584 | 3.40 | 23.8 |

**Table 9.**Optimum high-side pressure of ejector-expansion refrigeration cycle (EERC) for N

_{2}O refrigerant and its corresponding maximum COP and ${\mathsf{\eta}}_{\mathrm{ex}}$ for various design parameters.

T_{gc} (°C) | T_{eva} (°C) | P_{OPT,High-side} (MPa) | COP_{max} | ${\mathsf{\eta}}_{\mathbf{ex},\mathbf{max}}$ (%) |
---|---|---|---|---|

36 | 0 | 7.335 | 3.28 | 35.6 |

5 | 7.327 | 4.02 | 35.7 | |

10 | 7.314 | 5.01 | 35.1 | |

38 | 0 | 7.786 | 2.91 | 31.6 |

5 | 7.779 | 3.53 | 31.4 | |

10 | 7.766 | 4.34 | 30.4 | |

40 | 0 | 8.222 | 2.61 | 28.3 |

5 | 8.219 | 3.15 | 28.0 | |

10 | 8.206 | 3.82 | 26.7 |

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**MDPI and ACS Style**

Aghazadeh Dokandari, D.; S. Mahmoudi, S.M.; Bidi, M.; Haghighi Khoshkhoo, R.; Rosen, M.A.
First and Second Law Analyses of Trans-critical N_{2}O Refrigeration Cycle Using an Ejector. *Sustainability* **2018**, *10*, 1177.
https://doi.org/10.3390/su10041177

**AMA Style**

Aghazadeh Dokandari D, S. Mahmoudi SM, Bidi M, Haghighi Khoshkhoo R, Rosen MA.
First and Second Law Analyses of Trans-critical N_{2}O Refrigeration Cycle Using an Ejector. *Sustainability*. 2018; 10(4):1177.
https://doi.org/10.3390/su10041177

**Chicago/Turabian Style**

Aghazadeh Dokandari, Damoon, S. M. S. Mahmoudi, M. Bidi, Ramin Haghighi Khoshkhoo, and Marc A. Rosen.
2018. "First and Second Law Analyses of Trans-critical N_{2}O Refrigeration Cycle Using an Ejector" *Sustainability* 10, no. 4: 1177.
https://doi.org/10.3390/su10041177