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Thermodynamic analyses and optimization of a transcritical N2O refrigeration cycle Jahar Sarkar a, Souvik Bhattacharyya b,* a b

Department of Mechanical Engineering, Institute of Technology, BHU, Varanasi 221005, India Department of Mechanical Engineering, Indian Institute of Technology Kharagpur, Kharagpur 721302, India

article info

abstract

Article history:

Thermodynamic analyses as well as optimization studies based on maximum cooling COP

Received 28 April 2009

of a transcritical N2O cycle and both energetic and exergetic comparisons with CO2 cycle

Received in revised form

are presented in this article. Effects of superheating, internal heat exchanger and expan-

10 August 2009

sion turbine are studied as well. An expression for optimum discharge pressure has been

Accepted 11 September 2009

developed. Variation trends of optimal parameters for the N2O system are similar to that of

Available online 16 September 2009

a CO2 system. The N2O cycle exhibits higher cooling COP, lower compressor pressure ratio and lower discharge pressure and temperature, and higher second law efficiency when

Keywords:

compared to CO2 based systems; however, it is inferior in term of volumetric cooling

Refrigeration system

capacity at the optimum condition. Effect of superheating in evaporator is negligible and

Nitrous oxide

effect of introducing an internal heat exchanger is moderate whereas effect of employing

Transcritical cycle

a work recovery turbine is significant on COP improvement and discharge pressure

Modelling

reduction at the optimal condition for both working fluids. ª 2009 Elsevier Ltd and IIR. All rights reserved.

Simulation Performance Comparison Carbon dioxide

Analyses thermodynamiques et optimisation d’un cycle frigorifique au N2O transcritique Mots cle´s : Syste`me frigorifique ; Protoxyde d’azote ; Cycle transcritique ; Mode´lisation ; Simulation ; Performance ; Comparaison ; Dioxyde de carbone

1.

Introduction

With respect to the environmental safety and personal safety, natural refrigerants do appear more attractive than the other

synthetic refrigerants (Calm, 2008). Due to their zero ozone layer depletion potential and low global warming potential, several natural refrigerants are regaining their importance and are on a revival path, and nitrous oxide (N2O) is one of the

* Corresponding author. Tel.: þ91 3222 282038; fax: þ91 3222 255303. E-mail address: [email protected] (S. Bhattacharyya). 0140-7007/$ – see front matter ª 2009 Elsevier Ltd and IIR. All rights reserved. doi:10.1016/j.ijrefrig.2009.09.012

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international journal of refrigeration 33 (2010) 33–40

COP h i pd,opt q s t,T Te To Vc w his

coefficient of performance specific enthalpy (kJ kg-1) component irreversibility (kJ kg-1) optimum discharge pressure (MPa) specific heat transfer (kJ kg-1) specific entropy (kJ kg-1 K-1) temperature ( C, K) cooling temperature (K) reference temperature (K) volumetric cooling capacity (MJ m–3) specific work (kJ kg-1) isentropic efficiency

prospects as a future refrigerant. While CO2 based systems have already gained large acceptance in refrigeration and heat pump applications, N2O still remains mostly unexplored. However, CO2 can be used down to an evaporation temperature of 55 C and further lowering of temperature can not be achieved since the triple point of CO2 is about 56.56 C. On the other hand, N2O has a triple point temperature of 90.82 C with a boiling point temperature of 88.47 C and hence it can be used in the region below the application range of CO2. Furthermore, similarity between critical temperature, pressure and molecular weight of N2O and CO2 (Kruse and Russmann, 2006) causes nearly similar behaviour with respect to system temperature, pressure, compactness and properties. The main disadvantage of N2O is its higher GWP compared to CO2, however it is significantly more favourable in terms of toxicity. Kruse and Russmann (2006) reported a theoretical investigation on nitrous oxide refrigerant based cascade refrigeration systems and compared with an R23-R134a cascade system, and showed that the transcritical CO2 topping cycle with N2O bottoming cycle cascade system exhibited lower performance than the R134a/R23 cascade system. Di Nicola et al. (2007) experimentally determined COP of CO2 and N2O binary mixture in low temperature cycle with R404a in high temperature cycle and results were compared with R23 in low temperature cycle and R404a in high temperature cycle. Bhattacharyya et al. (2009) studied the N2O and CO2 based cascade system and showed similar behavioural trends where fluids are swapped. Recent study on transcritical N2O cycle (Sarkar and Bhattacharyya, 2008) showed higher COP and lower high side pressure compared to CO2, which motivates the further detailed study. In the present study, the first and second law analyses as well as compressor discharge pressure optimization of transcritical N2O refrigeration cycle are carried out. The performance improvement and effect on optimum parameters with the use of superheating in the evaporator, internal heat exchanger and work recovery turbine as expansion device are presented as well. Performance comparison with a transcritical CO2 cycle is also shown. Correlations are obtained for optimum discharge pressure in terms of evaporator and gas cooler exit temperatures.

second law efficiency fluid density (kg m–3) internal heat exchanger effectiveness

hII r 3ihx

Subscripts 1–6 refrigerant state points c compressor co gas cooler exit ev evaporator ed expansion device gc gas cooler t expansion turbine

2.

Theoretical modelling and simulation

The process layout and corresponding pressure-enthalpy diagram of a basic transcritical N2O refrigeration cycle are illustrated in Fig. 1. Similar to the transcritical CO2 cycle, the saturated or superheated vapour at state 1 is compressed in the compressor to the supercritical state 2 and then cooled in a gas cooler (2–3). Unlike in a condenser, in the gas cooler the heat rejection takes place with a gliding temperature. Supercritical fluid from gas cooler is expanded (3–4) and then passed through evaporator (4–1). The layout and corresponding pressure-enthalpy diagram of a basic transcritical N2O cycle with internal heat exchanger are shown in Fig. 2. As shown, the saturated vapour at state 6 is superheated to state 1 in the internal heat exchanger and then compressed in the compressor to state 2. The supercritical fluid at state 2 is cooled in the gas cooler to state 3 by rejecting heat to the external fluid. Fluid at high pressure is further cooled from 3 to 5 in the internal heat exchanger and then expanded through an expansion device to state 4, which is the inlet to the

3

Gas cooler

2 Compressor

Expansion device 4

1 Evaporator

2 3

Pressure

Nomenclature

4

1

With superheat

Specific enthalpy Fig. 1 – Layout and p–h diagram of basic expansion transcritical N2O cycle.

international journal of refrigeration 33 (2010) 33–40

3

Gas cooler

Compressor

Expansion device 6 Evaporator 5

1 2 3

Pres s ure

4

3. Processes in evaporation, gas cooler and internal heat exchanger are isobaric. 4. Unless otherwise stated, the refrigerant at evaporator exit is dry saturated.

2

Internal heat exchanger

5

35

For the basic expansion cycle (Fig. 1), in which, the expansion process is assumed to be isenthalpic, the specific work input to the compressor is given by: wc ¼ h2 h1

(1)

The specific refrigerating effect obtained from the evaporator is expressed as:

4

1

6

qev ¼ h1 h4

Hence, the cooling COP and volumetric cooling capacity are given by, respectively,

Specific enthalpy Fig. 2 – Layout and p–h diagram of transcritical N2O cycle with internal heat exchanger.

evaporator. The state of the refrigerant changes from 4 to 6 as it evaporates by extracting heat to give useful cooling effect. The layout and corresponding pressure-enthalpy diagram of a work recovery turbine expansion transcritical N2O refrigeration cycle are illustrated in Fig. 3, where the turbine is used as an expansion device (3–4) to improve the cycle performance and the other processes are similar to the basic cycle. The theoretical modelling of the transcritical N2O cycle has been performed based on the steady flow energy and exergy balances (first and second laws of thermodynamics). The following common assumptions have been made in the thermodynamic analysis: 1. Heat transfer with the ambient has been neglected except in gas cooler. 2. Compression process is adiabatic but non-isentropic with given efficiency.

qev wc

(3)

Vc ¼ qev r1

(4)

COP ¼

Employing the second law of thermodynamics to each component, the following relations can be established for exergetic performance: i) Compressor irreversibility, ic ¼ To ðs2 s1 Þ

Gas cooler

ied ¼ To ðs4 s3 Þ

4

Compressor

iii) Gas cooler irreversibility, igc ¼ ðh2 h3 Þ To ðs2 s3 Þ

(7)

iv) Evaporator irreversibility,

1 Evaporator

(6)

To Te

(8)

Second law (exergy) efficiency for the system is given by the ratio of net exergy output and the work input to the compressor, i.e.:

2

Expansion turbine

(5)

ii) Throttling device irreversibility,

iev ¼ To ðs1 s4 Þ qev

3

(2)

2

hII ¼

wc ic þ ied þ iev þ igc wc

(9)

3 Pressure

For the internal heat exchanger cycle (Fig. 2), the effectiveness and energy balance, respectively, for internal heat exchanger can be written as (Sarkar et al., 2004):

3ihx ¼

4

1

Specific enthalpy Fig. 3 – Layout and p–h diagram of turbine expansion transcritical N2O cycle.

T1 T6 T3 T6

h1 h6 ¼ h3 h5

(10)

(11)

The specific compressor work and refrigerating effect are given by:

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international journal of refrigeration 33 (2010) 33–40

wc ¼ h2 h1

tev =

qev ¼ h6 h4

(13)

qev wc

(14)

For the work recovery turbine expansion cycle (Fig. 3), in which, the expansion process is assumed to be adiabatic but non-isentropic with given isentropic efficiency, specific compressor work, refrigerating effect and volumetric cooling capacity are given by Eqs. (1)–(2),(4). The specific turbine work and the cooling COP are given by, respectively, wt ¼ h3 h4

3.51

tev = −75oC 2.69 1.87 1.05

(15)

0.23 30

qev COP ¼ wc wt

(16)

Based on the theoretical model presented above, a simulation code was developed to study the cycle performances for the given design and operating parameters. This code was integrated with the thermodynamic property subroutine ‘N2OPROP’ to estimate thermodynamic properties of nitrous oxide in subcritical and supercritical region, which has been developed exclusively for this study based on literature, reported earlier (Lemmon and Span, 2006). Efficient iterative procedures have been used to predict assorted state properties with reasonable accuracy. It is noted that a similar numerical model was reported for a transcritical CO2 cycle integrated with the thermodynamic property subroutine ‘CO2PROP’ (Sarkar et al., 2004). The percentage irreversibility has also been estimated to represent the contribution of each component to the total irreversibility in the system and is given by the ratio of the irreversibility of each component to the compressor work.

3.

Results and discussion

35 34 33 32 31 30 -75

Subcritical cycle p d,opt > p saturation

Subcritical cycle p d,opt = p saturation

60

been estimated for various operating conditions with a 0.05 MPa step increase in compressor discharge pressure. Compressor isentropic efficiency is assumed to be 75% in the present study. Due to the unique behavioural pattern of N2O properties around the critical point and beyond, the COP of the transcritical N2O cycle varies non-monotonically with the gas cooler pressure and there exists an optimum pressure leading to maximum COP similar to the transcritical CO2 cycle, which is a very important factor in system design. Present results are based on optimum discharge pressure leading to maximum COP and important parameters at optimum conditions are suitably plotted to illustrate the various performance trends. Previous study (Sarkar et al., 2007) showed that in the subcritical cycle operation (condenser exit temperature is lower than critical temperature), in general, the isotherms are nearly vertical in p–h diagram and hence the saturation

20.6 Optimum discharge pressure (MPa)

Condenser/Cooler exit temperature (oC)

Transcritical cycle

40 50 Gas cooler exit temperature (oC)

Fig. 5 – Maximum cooling COP variation for basic expansion cycle.

The performance of the transcritical N2O refrigeration cycle have been evaluated on the basis of cooling COP, which have

36

5oC

tev = −15oC tev = −35oC tev = −55oC

4.33 Maximum cooling COP

The cooling COP is given by: COP ¼

5.15

(12)

tev = 5oC tev = −15oC

18.2

tev = −35oC tev = −55oC

15.8

tev = −75oC

13.4 11 8.6 6.2

-55 -35 -15 Evaporation temperature (oC)

5

Fig. 4 – Temperature zones for subcritical and transcritical N2O cycle.

30

40 50 60 Gas cooler exit temperature (oC)

Fig. 6 – Optimum discharge pressure variation for basic expansion cycle.

37

international journal of refrigeration 33 (2010) 33–40

Table 1 – Energetic comparison of N2O and CO2 basic expansion cycles. Refrigerant

tev ( C) tco ( C) Pd,opt (MPa) pd,opt/pev t2 ( C) qev (kJ/kg) COP COP deviation (%) Vc (MJ/m3)

N2O

CO2

N2O

CO2

N2O

CO2

N2O

CO2

5 40 8.52 2.406 78.9 122.8 2.689 11.50 12.14

5 40 10.1 2.545 83.2 115.9 2.412 – 13.28

5 50 11.28 3.185 105.0 108.7 1.744 8.65 10.74

5 50 13.2 3.325 108.0 103.1 1.605 – 11.81

–55 40 9.90 18.67 198.2 131.7 0.649 6.84 1.850

–55 40 12.1 21.84 224.1 135.4 0.607 – 1.987

–55 50 13.94 26.288 236.5 119.1 0.505 5.17 1.674

55 50 16.52 29.82 260.8 122.6 0.480 – 1.800

pressure corresponding to the condenser exit temperature is optimum one leading to maximum COP; however the isotherms show S-shape pattern (similar to the transcritical cycle) at near critical zone and hence the optimum high side pressure is higher than the saturation pressure corresponding to the condenser exit temperature, which can observed for certain combinations of evaporation and condenser exit temperatures depending on the working fluid nature. Similar behaviour is observed for N2O cycle also. As shown in Fig. 4 (representing three combination zones), the cycle is the subcritical one and the optimum high side pressure is higher than the saturation one for some combinations of evaporation and gas cooler/condenser exit temperatures (at middle zone), whereas the lower side zone represent the combinations for subcritical cycle with saturation pressure as optimum one and the higher zone represents the combination for transcritical cycle (optimum high side pressure is higher than critical pressure). The variations of maximum cooling COP and corresponding optimum discharge pressure with gas cooler exit temperature for various evaporator temperatures are shown in Figs. 5 and 6, respectively. At lower condenser/gas cooler exit temperature, where the temperature combinations are in the lower zone in Fig. 4, optimum discharge pressure, invariant with evaporator temperature (Fig. 6), indicates the saturation pressure in a subcritical cycle. With the increase in evaporation temperature and decrease in gas cooler exit temperature, the maximum cooling COP increases and the optimum discharge pressure decreases, which are similar to the variation in case of transcritical CO2 cycle (Sarkar et al., 2004). This behaviour of optimum discharge pressure is due to the divergent nature of the compression process lines with evaporation temperature and flattening nature of isotherm with gas cooler exit temperature. Hence, design of transcritical N2O cycle at highest

possible evaporation temperature and lowest possible gas cooler exit temperature is better in terms of higher cooling COP as well as lower optimum discharge pressure. Effect of gas cooler exit temperature on optimum pressure variation is more significant compared to evaporator temperature, however equally significant on maximum COP. Variation trends show that the maximum cooling COP varies least towards the minimum evaporator temperature and maximum gas cooler exit temperature whereas the optimum discharge pressure varies least towards the maximum evaporator temperature and minimum gas cooler exit temperature. Previously reported results (Sarkar and Bhattacharyya, 2008) showed that the isentropic efficiency of the compressor has a strong influence on maximum cooling COP, while the optimum compressor discharge pressure is usually not dependent on the compressor performance. Hence, the optimum discharge pressure for basic expansion cycle is given by, pd,opt ¼ f(tev, tco). Performing a regression analysis on the data obtained from the cycle simulation for evaporation temperatures ranging between 75 C and 5 C and cooler exit temperatures ranging between 30 C and 60 C, the following relation has been established to predict estimates of the optimum discharge pressure (MPa): pd;opt ¼

tco 28:885 176:08 þ 2:661 þ 0:01192tev tco

(17)

Table 1 represents the comparison of N2O cycle with its counterpart CO2 cycle at optimum condition for various evaporation and gas cooler exit temperatures. Results show that N2O cycle is better compared to CO2 cycle in terms of lower compressor discharge pressure, lower compressor pressure ratio, lower discharge temperature, higher specific cooling capacity and higher cooling COP, however, inferior in term of volumetric cooling capacity. Lower compressor

Table 2 – Exergetic comparison of N2O and CO2 basic expansion cycles. Refrigerant

tev ( C) tco ( C) Compressor (%) Expansion device (%) Gas cooler (%) Evaporator (%) 2nd law efficiency (%)

N2O

CO2

N2O

CO2

N2O

CO2

N2O

CO2

5 40 21.73 29.08 25.01 10.18 14.00

5 40 21.46 30.63 26.23 9.12 12.56

5 50 20.28 31.18 32.86 6.60 9.08

5 50 20.12 32.84 32.61 6.07 8.36

–55 40 16.82 30.56 27.34 3.95 21.33

–55 40 15.99 29.70 30.64 3.71 19.96

–55 50 15.58 30.96 33.78 3.08 16.60

55 50 14.92 30.40 35.97 2.93 15.78

38

international journal of refrigeration 33 (2010) 33–40

2.4 Discharge pressure reduction (%)

pressure ratio leads to higher isentropic and volumetric efficiencies resulting in further increase in performance. Some remarkable properties of N2O leading to performance deviation with CO2 are higher critical temperature, lower saturation pressure variation with temperature and lower vapour density. Deviation of maximum cooling COP increases with increase in evaporation temperature due to the fact that the evaporative latent heat of N2O is lower at lower temperature and higher at higher temperature compared to CO2 whereas deviation decreases with increase in gas cooler exit temperature. Deviation of optimum discharge pressure increases with decrease in evaporator as well as gas cooler exit temperatures and deviation of volumetric cooling capacity show opposite trends. Table 2 represents the comparison based on second law of thermodynamics (Second law efficiency and exergetic comparison for different components) at optimum discharge pressure for various evaporation and gas cooler exit temperatures. In the evaporator, the secondary fluid temperature (Te) is assumed to be 10 C above the evaporator temperature and ambient temperature is assumed to be 30 C. Componentwise irreversibility distributions show similar trends for both transcritical N2O and CO2 cycles. Compressor and evaporator irreversibilities increase with increase in evaporator temperature and with decrease in gas cooler exit temperature whereas gas cooler and expansion irreversibilities show opposite trends. Compressor irreversibility as well as expansion device irreversibility of transcritical N2O is moderately lower compared to that of CO2 due to lower cycle pressure difference. Gas cooler irreversibility for N2O is also moderately lower due to high critical temperature whereas evaporator irreversibility is similar. As shown in table, overall cycle irreversibility for both N2O is lower and hence the second law efficiency is higher than that for CO2. Second law efficiency increases with decrease in evaporator as well as gas cooler exit temperatures. Comparison of transcritical N2O cycle with CO2 cycle in terms of effect of superheating in the evaporator on the maximum cooling COP and the optimum discharge pressure are shown in Figs. 7 and 8, respectively, at the evaporator

Nitrous oxide

Carbon dioxide

2 1.6 1.2 0.8 0.4

tev

0

Superheat

5oC 5oC

5oC 10oC

–55oC 5oC

–55oC 10oC

Fig. 8 – Effect of superheating in evaporator on optimum discharge pressure.

temperatures of 5 C and 55 C, and gas cooler exit temperature of 40 C. Results show that the maximum cooling COP increases and corresponding optimum discharge pressure decreases with increase in superheat insignificantly but monotonically for both N2O and CO2 cycles. It can be noted that the slope of compression process is given by dp/ dh¼his,c r and hence the compressor work as well as the cooling COP is dependent on density variation with superheating as well as compression process and the pressure difference for given isentropic efficiency, which are dependent on fluid properties. Less decrease in density by superheating leads to less increase in compressor work and hence higher COP improvement. The improvement of maximum COP by superheating is more significant for higher evaporation temperature whereas as the optimum discharge pressure reduction is more significant for lower evaporation temperature in case of both N2O and CO2. The improvement of maximum COP over basic saturation cycle for transcritical N2O cycle by superheating is less significant for higher

3 12

Nitrous oxide Carbon dioxide

COP improvement (%)

COP improvement (%)

2.5 2 1.5 1 0.5

Carbon dioxide

tco = 40 oC

8 6 4 2

0

tev

10

Nitrous oxide

5 C

o

5 C

o

o

Superheat 5 C

o

10 C

o

–55 C o

5 C

o

–55 C o

10 C

Fig. 7 – Effect of superheating in evaporator on maximum COP improvement.

tev ε ihx

0 o

5C 60%

o

5C 80%

o

55 C 60%

o

55 C 80%

Fig. 9 – Effect of using internal heat exchanger on maximum COP improvement.

39

international journal of refrigeration 33 (2010) 33–40

12

18

10

Dis c harge pres s ure reduc t ion (% )

Discharge pressure reduction (%)

Nitrous oxide Carbon dioxide

tco = 40 oC 8 6 4 2 0

tev ε ihx

5oC 60%

5oC 80%

55oC 60%

55oC 80%

Fig. 10 – Effect of using internal heat exchanger on optimum discharge pressure.

evaporation temperature and more significant for lower evaporation temperature whereas the optimum discharge pressure reduction is less significant compared to that for CO2 cycle. The COP improvement and discharge pressure reduction at optimum condition by using internal heat exchanger in transcritical N2O cycle and CO2 cycle compared to basic expansion cycle for various effectiveness are shown in Figs. 9 and 10, respectively at the evaporator temperatures of 5 C and 55 C, and gas cooler exit temperature of 40 C. Results show that the effects of using internal heat exchanger on the maximum cooling COP improvement and corresponding optimum discharge pressure reduction are moderate for both N2O and CO2 cycles. The maximum cooling COP improvement increases and optimum discharge pressure reduction decreases monotonically with the internal heat exchanger effectiveness for both N2O and CO2. Effects of using internal heat exchanger on the maximum cooling COP improvement and optimum discharge

60 Nitrous oxide

COP improv ement (% )

50

Carbon dioxide tco = 40 oC

40 30 20 10

tev ηis ,t

0

5oC 60%

5oC 80%

55oC 60%

55oC 80%

Fig. 11 – Effect of using expansion turbine on maximum cooling COP.

15

Nitrous oxide Carbon dioxide

tco = 40 oC 12 9 6 3

tev ηis ,t

0 5oC 60%

5oC 80%

55oC 60%

– 55oC 80%

Fig. 12 – Effect of using expansion turbine on optimum discharge pressure.

pressure reduction increase with decrease in evaporator temperature, although the compressor discharge temperature increases and the volumetric capacity decreases significantly. It appears that the effect of using internal heat exchanger in the transcritical N2O cycle on the maximum cooling COP improvement is more significant at lower evaporator temperature and less significant for higher evaporator temperature, whereas the effect on optimum discharge pressure reduction is less significant compared to its counterpart CO2 cycle. The COP improvement and discharge pressure reduction at optimum condition by using expansion turbine replacing basic expansion device in transcritical N2O and CO2 cycles for various turbine isentropic efficiencies are shown in Figs. 11 and 12, respectively at the evaporator temperatures of 5 C and 55 C, and gas cooler exit temperature of 40 C. Similar to the transcritical CO2 cycle, the optimum discharge pressure reduces while employing an expansion turbine replacing an expansion valve due to the distinct behaviour of isotherm and expansion lines. Results show that the effect of using expansion turbine in both N2O and CO2 cycles is very significant on the maximum cooling COP improvement as well as optimum discharge pressure reduction. However, the effect is more predominant for CO2 cycle compared to N2O cycle due to more turbine pressure drop. By using turbine, cooling capacity increases and net work input decreases and hence the cooling COP increases significantly. The maximum cooling COP improvement and optimum discharge pressure reduction increases monotonically with increase in turbine isentropic efficiency for both N2O and CO2. Effect of using work recovery turbine as expansion device on the maximum cooling COP improvement and optimum discharge pressure reduction is more significant at lower evaporator temperature.

4.

Conclusions

Thermodynamic analyses as well as optimization studies for a transcritical N2O refrigeration cycle and effects of

40

international journal of refrigeration 33 (2010) 33–40

superheating, use of internal heat exchanger and expansion turbine along with comparison with a transcritical CO2 refrigeration cycle have been presented in this study. Based on the results and optimization of the cycle, following conclusions can be drawn. 1. Effects of evaporation and gas cooler exit temperatures on maximum cooling COP and corresponding optimum discharge pressure are similar for both N2O and CO2 based transcritical cycles. Variation trends of other parameters at optimum conditions are also similar for both working fluids. 2. For higher evaporation temperature and lower gas cooler exit temperature, the system is profitable in terms of cooling COP as well as optimum discharge pressure for both working fluids. Expression for optimum discharge pressure has been developed, which offers useful guidelines for optimal system design. 3. Transcritical N2O cycle is better than CO2 cycle in terms of higher cooling COP, lower compressor pressure ratio and lower discharge pressure and temperature; however, it is inferior in term of volumetric cooling capacity. 4. Component-wise irreversibility distribution as well as variation with evaporator and gas cooler exit temperature show similar trends for both fluids, however second law efficiency of N2O cycle is higher than that of CO2. 5. The maximum cooling COP improvement and discharge pressure reduction by using superheat are marginal; however, the COP improvement for N2O is higher for low temperature whereas lower for high temperature applications and the discharge pressure reduction is lower, although marginally, compared to CO2. 6. The influence of internal heat exchanger effectiveness on COP and optimum discharge pressure is modest for both working fluids; however, the effect on COP improvement and optimum discharge pressure reduction are greater,

although marginally, for the CO2 cycle. Hence, the use of internal heat exchanger is less profitable in case of N2O compared to CO2 in terms of both COP and high side pressure. 7. The effect of using of turbine as expansion device on cooling COP improvement and discharge pressure reduction at optimum condition is significant for both working fluids; however, use is marginally less profitable in case of N2O compared to CO2 in terms of both COP and high side pressure.

references

Bhattacharyya, S., Garai, A., Sarkar, J., 2009. Thermodynamic analysis and optimization of a novel N2O–CO2 cascade system for refrigeration and heating. Int. J. Refrigeration 32 (5), 1077–1084. Calm, J.M., 2008. The next generation of refrigerants – historical review, considerations, and outlook. Int. J. Refrigeration 31 (7), 1123–1133. Di Nicola, G., Giuliani, G., Polonara, F., Santori, G., Stryjek, R., 2007. Cascade cycles operating with CO2 þ N2O binary systems as low temperature fluid: experimental results. In: Int. Congress of Refrigeration, Beijing. Kruse, H., Russmann, H., 2006. The natural fluid nitrous oxide – an option as substitute for low temperature synthetic refrigerants. Int. J. Refrigeration 29 (5), 799–806. Lemmon, E.E., Span, R., 2006. Short fundamental equations of state for 20 industrial fluids. J. Chem. Eng. Data 51 (3), 785–850. Sarkar, J., Bhattacharyya, S., 2008. Optimization of a transcritical N2O refrigeration/heat pump cycle (NIK-06-T1-14). In: 8th IIR G Lorentzen Conference on Natural Working fluids, Copenhagen. Sarkar, J., Bhattacharyya, S., Ramgopal, M., 2004. Optimization of a transcritical CO2 heat pump cycle for simultaneous cooling and heating applications. Int. J. Refrigeration 27 (8), 830–838. Sarkar, J., Bhattacharyya, S., Ramgopal, M., 2007. Natural refrigerant-based subcritical and transcritical cycles for high temperature heating. Int. J. Refrigeration 30 (1), 3–10.