International Journal for Research in Engineering Application & Management (IJREAM) ISSN : 2454-9150 Vol-04, Issue-03, June 2018 A Thermodynamic Performance Study of Simple Cooling and Combined Power Transcritical N2O Cycles
1Anjan Kumar Sahu, 2Prasant Nanda 1Asst. Professor, Dept. of Mechanical Engineering, Aalim Muhammed Salegh College of Engineering, Chennai, India, [email protected] 2 Professor, Department of Training and Placement, Veer Surendra Sai University of Technology, Burla, Odisha, India, [email protected]
Abstract - Thermodynamic performances of transcritical N2O simple cooling and combined power cycles have been studied under steady state condition. The effect of influential parameters like evaporator temperature, gas cooler outlet temperature, turbine inlet temperature and gas heater pressure on COP and exergetic efficiency have been carried out.
It has been found that the thermodynamic performance of transcritical N2O combined power cycle is better than simple cooling cycle under all operating conditions. It is also noticed there is optimum exergetic efficiency at a particular evaporator temperature for both the cycles. A Grassmann diagram for N2O cooling and combined power cycles presented at a given operating condition.
Keywords - Cooling and power cycle, energy, exergy, grassmann diagram, N2O, transcritical.
I. INTRODUCTION It also provides several environmental advantages in Due to global warming, ozone layer depletion and present refrigeration and air conditioning systems. atmospheric pollution problems, the halogenated Kim et al. [4] presented the use of CO in refrigerant cycle refrigerants like CFCs and HCFCs need to be replaced by 2 giving more attention to the system design and cycle natural refrigerants like hydrocarbon, water, carbon modification of a vapor compression refrigeration system. dioxide, ammonia and nitrous oxide etc. In the recent years The attractive features like low pressure ratio and the researchers and environmentalist around the globe tried volumetric capacity of carbon dioxide also given more to revive carbon dioxide and nitrous oxide as promising impotence. refrigerants over hazardous halogenated refrigerants. Kauf [5] showed the existence of an optimum gas cooler pressure in a transcritical CO2 vapor compression II. LETERATURE REVIEW refrigeration system for maximum COP at a particular gas Various researchers have well reported their work on cooler outlet temperature. system design, cycle modification, multi stage operation, Liao et al. [6] explained the value of optical heat rejection optimization of gas cooler pressure, use of different pressure mainly depends on gas cooler outlet temperature, mixture of gases and comparison of thermodynamic evaporator temperature and performance of compressor in performances in CO2 and N2O vapor compression cycles. simple transcritical carbon dioxide cycle based on cycle A brief literature review of few researcher are presented simulation, correlation of optical heat rejection pressure at below. different operating parameters. Lorenzen [1] successfully demonstrated the use of natural Sarkar et al. [7] optimized transcritical CO2 heat pump refrigerants like ammonia, propane and carbon dioxide system for simultaneous cooling and heating applications, over halogenated refrigerants like CFCs and HCFCs due to similarly a correlation was presented to find at optimum certain excellent thermo physical and heat transfer gas cooler pressure. properties for practically all conventional refrigeration and heat pump systems. Bhattacharyya et al. [8] carried out optimization study and showed the existence of optimum gas cooler pressure of Lorenzen [2] revived CO2 as most preferred and attractive CO2-C3H8 cascade system for both cooling and heating alternate natural refrigerant, which coupled with low cost, applications. Agrawal et al. [9] carried out optimization easily available, non-toxicity, non-flammability for energy studies of different two stage transcritical CO2 heat pump performance heat pump systems. systems. A correlation was prepared to find out for Lorenzen and Patterson [3] showed carbon dioxide is an optimum value of inter stage and gas cooler pressure excellent refrigerant by solving the environment related respectively. problems associated with car air conditioning systems.
408 | IJREAMV04I033936 DOI : 10.18231/2454-9150.2018.0355 © 2018, IJREAM All Rights Reserved. International Journal for Research in Engineering Application & Management (IJREAM) ISSN : 2454-9150 Vol-04, Issue-03, June 2018
Agrawal et al. [10] carried out thermodynamic energy and motivation to do research work on transcritical N2O analysis for steady state operating condition for combined cycles. The present work is aimed to carry out power and refrigeration transcritical CO2 cycles. Stegou- thermodynamic analysis of simple and combined power transcritical N O cycles and a comparison is made between Stagia et al. [11] presented exergy loss analysis of using 2 them. different mixtures (R-404 A, R 410-A, R 410-B and R 507). Yang et al. [12] performed a comparative study on III. MATHEMATICAL MODELING energy and exergy analysis of a simple transcritical CO2 refrigeration cycle using expander and throttle valve. 3.1. PROCESS ANALYSIS OF SIMPLE COOLING CYCLE Agrawal et al. [13] demonstrated comparative study of two Fig.1 and Fig. 2 exhibit the simple transcritical N O stage transcritical N2O and CO2 cycles. Results are shown 2 for simultaneous optimization of intercooler and gas cooler cooling cycle consisting of four components such as pressure based on cycle simulation. compressor, gas cooler, throttle valve and evaporator and corresponding T-s diagram, respectively. Sarkar [14] optimized discharge pressure for transcritical The following assumptions have been made in the N2O refrigeration system with an internal heat exchanger. A comparative study has been done for counterpart analysis. transcritical CO2 system. It is observed that the heat 1. Heat transfer with ambient is negligible exchanger of N2O refrigeration system is less profitable in 2. Compression process in compressor is adiabatic but not comparison to CO2 system considering both COP improvement and high side pressure reduction at an isentropic. optimum condition. 3. Expansion process in throttle valve is constant isenthalpic. Chen et al. [15] carried out energetic analysis of a 4. Evaporation and gas cooling processes are isobaric. transcritical CO2 combined cooling and power cycle using engineering equation solver for automobile air Gas Cooler conditioning applications. In gas heater the exhausted 3 gases from automobile are used as a heat generating source. It is observed that optimum gas cooler pressure for combined cooling and power cycle more or less remains Throttle valve Compressor same as that for basic cooling cycle. It is also noticed that 2 there exists an optimum gas heater pressure in both basic cycle and combined cooling and power cycle. There is significant improvement of COP of 40% in case of combined cooling and power cycle. 4 1 Evaporator Kruse et al. [16] observed that a transcritical CO2 topping cycle and combined with NO2 bottoming cycle shows poor performance in comparison with R23 - R134 cascade Fig.1. Schematic Diagram of Transcritical N2O cooling cycle [20] system. They also studied two stage N2O system with flash intercooler and with intermediate liquid injection and T observed that using flash intercooler for mean temperature below -10 0C system COP improved in comparison with 2S 2 cascade system. 3 Di Nicola et al. [17] studied an experimental study CO2 and N2O binary mixture in a low temperature cycle and R404a in the high temperature cycle. Sarkar et al. [18] in his experimental study observed that - performances of transcritical N2O cycle better than CO2 cycles with respect to energetic, exergetic analysis and system operating pressure. Bhattacharyya et al. [19] comprehensively analyzed and modeled N2O - CO2 4 1 cascade systems.
Anjan et al. [20] carried out parametric steady state s thermodynamic analysis of transcritical CO2 cooling and Fig.2. T-s Diagram of Transcritical N2O simple cooling cycle combined power cycles. Anjan et al. [21] demonstrated an [20] analytical study of the energetic and exergetic analysis of transcritical CO2 and N2O combined power refrigeration 3.2. SPECIFIC ENERGY ANALYSIS OF SIMPLE cycles. Prasant et al. [22] comprehensively studied and COOLING CYCLE compared energetic and exergetic analysis of CO2 and N2O combined cooling heating and power cycles. Refrigerating effect of evaporator
In the literature it is reported little study has been done on qevap = h1 - h4 (1) (1) exclusively N2O transcritical cycles. So there is a scope
409 | IJREAMV04I033936 DOI : 10.18231/2454-9150.2018.0355 © 2018, IJREAM All Rights Reserved. International Journal for Research in Engineering Application & Management (IJREAM) ISSN : 2454-9150 Vol-04, Issue-03, June 2018
Work input to compressor
wcomp = h2 – h1 (2)
Heat rejected in gas cooler Gas Cooler 3 10 qgc = h2 – h3 (3) COP of simple cooling cycle 9 Turbine
qevap IHEX COPcool = II w comp (4) Gas Heater Isentropic efficiency of the compressor
hh Compressor is. comp = 21S 2 7 8 hh 21 (5) 3.3. SPECIFIC EXERGY LOSS IN SIMPLE COOLING CYCLE 1 Exergy loss in the compressor i = T (s – s ) (6) IHEX I comp o 2 1 Exergy loss in the evaporator 4 Throttle Valve T i = T (s – s ) – o (h – h ) (7) evap o 1 4 1 4 Tevap 6 Evaporator Exergy loss in the throttle valve i = T (s – s ) (8) tv o 4 3 5 Exergy loss in the gas cooler Fig.3. Schematic Diagram of Transcritical N2O Combined power To cycle [20] igc = (h2 – h3) - To (s2 – s3) (9) T Tgc 3.4. SPECIFIC EXERGY CHANGE IN SIMPLE 8 COOLING CYCLE 2 7 Exergy change in gas cooler 2S egc = (h2 – h3) – To(s2 – s3) (10) 9 3 Exergy change in the evaporator 4 10 9S eevap = (h1 – h4) – To(s1 – s4) (11) Actual exergy supplied to the compressor 1
hh 5 6 eact.comp. = 21 mech (12)
II law efficiency (II ) is defined as s
Fig.4. T-s Diagram of Transcritical N2O Combined power cycle [20] 1 – Total Exergy loss The following assumptions have been made in the Total Exergy in (13) analysis. Heat transfer with ambient is negligible 3.5. PROCESS ANALYSIS OF COMBINED POWER Compression process in compressor is adiabatic but CYCLE not isentropic.
Fig. 3 shows the schematic diagram of combined power Expansion process in turbine is adiabatic but not isentropic. system comprising of two internal heat exchangers, turbine, gas heater and other conventional refrigeration Expansion process in throttle valve is constant cycle components. Corresponding T-s diagram of isenthalpic. combined power cycle is shown in Fig. 4. Evaporation, gas cooling, gas heating and heat exchanger processes are isobaric.
410 | IJREAMV04I033936 DOI : 10.18231/2454-9150.2018.0355 © 2018, IJREAM All Rights Reserved. International Journal for Research in Engineering Application & Management (IJREAM) ISSN : 2454-9150 Vol-04, Issue-03, June 2018
3.6. SPECIFIC ENERGY ANALYSIS OF COMBINED Exergy loss in internal heat exchanger I
POWER CYCLE ihex I = T0{(s1 + s4) – (s3+s6)} (30)
Exergy loss in internal heat exchanger II ihex II = T0{(s7 + s10) – (s2+s9)} (31)
Refrigerating effect of evaporator 3.8. SPECIFIC EXERGY CHANGE IN COMBINED qevap = h6 - h5 (14) POWER CYCLE Work input to compressor Exergy change in gas cooler wcomp = h2 – h1 (15) egc = (h10 – h3) – To(s10 – s3) (32) Heat rejected in gas cooler Exergy change in the evaporator qgc = h10 – h3 (16) eevap = (h6 – h5) – To(s6 – s5) (33) Heat supplied in gas heater Exergy change in the compressor qgh = h8 - h7 (17) ecomp= (h2 – h1) – To(s2 – s1) (34) Work output in turbine Exergy change in turbine wtur = h8 - h9 (18) etur= (h8 – h9) – To(s8 – s9) (35) Isentropic efficiency of compressor Exergy change in the gas heater hh egh = (h8 – h7) – To(s8 – s7) (36) is. comp = 21S hh Actual energy supplied to the compressor 21 (19) wact,comp.. = hh21 Isentropic efficiency of turbine mech (37)
= hh89 is. tur hh II law efficiency ( ) in % 89S (20) II = Exergy output Effectiveness of internal heat exchanger II 100 % Exergy input to the system hh 72 ihex II eevap e tur hh92 100 % (21) ()w w e = act.. comp tur gh (38) Energy balance in Internal heat exchanger I The working condition of both simple and combined power refrigeration cycles are listed in Table.1. h3 – h4 = h1 – h6 (22) Table1 Basic N2O and combined power Cycles Operating COP of combined power and refrigeration cycle Parameters [21] q Operating Value Unit COPcomb = evap Ambient Temperature (To) 303 K ()wcomp w tur q gh (23) Evaporator Temperature for simple cycle 270 K 3.7. SPECIFIC EXERGY LOSS IN COMBINED (Tevap) POWER CYCLE Super heat after evaporator for Combined 5 (Fixed K cycles Value)
Exergy loss in the compressor Gas cooler exit Temperature (Tgc) 308 K
Gas Heater Pressure (Pgh) 140 Bar
icomp = To (s2 – s1) (24) Gas Heater temperature (Tgh) 460 K Compressor Isentropic efficiency for 0.75 --- Exergy loss in the evaporator Combined cycle ()is. comp T ievap = To (s6 – s5) – o (h6 – h5) (25) Turbine isentropic efficiency 0.8 --- ()is. tur Tevap
Exergy loss in the throttle valve Mechanical efficiency of 0.8 --- ()mech Effectiveness of IHEX II ( ε ) 0.7 itv = To (s5 – s4) (26) ihex ii --- Exergy loss in the gas cooler A computer code has been developed for the energetic and exergetic analysis of simple refrigeration cycle and T i = (h10– h3) o - To (s10 – s3) (27) combined power refrigeration cycle for various operating gc Tgc parameters. Sub-critical and super-critical thermodynamic and transport properties of N O are calculated employing Exergy loss in the gas heater 2 N2O property code N2OPROP. To igh = To (s8 – s7) – (h8 – h7) (28)
Tgh IV. RESULTS AND DISCUSSION The performance comparison of the N O simple cooling Exergy loss in turbine 2 cycle and combined power refrigeration cycles are i T() s s presented based on the operating conditions such as tur o 98 (29) evaporator temperature, gas cooer outlet temperature,
411 | IJREAMV04I033936 DOI : 10.18231/2454-9150.2018.0355 © 2018, IJREAM All Rights Reserved. International Journal for Research in Engineering Application & Management (IJREAM) ISSN : 2454-9150 Vol-04, Issue-03, June 2018 turbine inlet temperature and gas heater pressure. The isentropic efficiency of the compressor and the turbine are taken as 75% and 80%, respectively. The effectiveness of IHEX-II is taken to be 0.7 and the mechanical efficiency of the compressor is taken as 80%. It is assumed that the gas is heated by 50 C in the IHEX-I as referred to [15]. The intermediate gas cooler pressure is taken as the geometric mean of inlet and outlet pressure of compressor. 4.1. EFFECT OF EVAPORATOR TEMPERATURE ON THE PERFORMANCES OF SIMPLE AND COMBINED POWER N O CYCLES 2 Fig.5. Variation of COP with evaporator temperature for T3 = 308 K, T8 = 500 K and P = 170 Bar. Fig. 5 presents the variation of COP with evaporator temperature considering T3 = 308 K, T8 = 500 K and P = 170 bar respectively. As evaporator temperature increases the work input to compressor decreases and simultaneously cooling effect increases, hence COP increases for both N2O cooling and combined power cycles. However, the increase in COP of combined power cycle is better in comparison with simple cooling cycle due to power generation at turbine. Fig. 6 shows the variation of second law efficiency of N2O simple cooling and combined power cycles. The overall component exergy loss decreases up to optimum Fig.6. Percentage Exergetic efficiency with evaporator temperature evaporator temperature and then starts increasing, hence it for T3 = 308 K, T8 = 500 K and P = 170 Bar. is noticed that the second law efficiency initially increases and attains maximum values of 36.48% at evaporator 0 temperature of 276 C for N2O combined power cycle and 0 35.3% at 278 C evaporator temperatures for N2O cooling cycle. The exergetic efficiency starts decreasing with further increase of optimum evaporator temperature for both simple and combined power N2O cycles. 4.2. EFFECT OF GAS COOLER OUTLET
TEMPERATURE ON THE PERFORMANCES OF N2O CYCLES Fig.7. presents the variation of COP with gas cooler outlet temperature. Reduction in refrigerant effect Fig.7. Variation of COP with gas cooler outlet temperature for Tevap = decreases the COP with increase in gas cooler outlet 275 K, T8 = 500 K and P = 170 Bar. temperature for both cooling and combined power N2O cycles. However, the variation in COP for N2O cooling cycle is relatively less in comparison to N2O combined power cycle due to lower refrigerating effect. Fig.8 presents the variation of exergetic efficiency with gas cooler outlet temperature for N2O cooling and combined power cycles. As overall component exergy loss increases with gas cooler outlet temperature, it is observed that the exergetic efficiency decreases with gas cooler outlet temperature for both the cycles. In N2O combined power cycle, the exergetic efficiency percentage decreases significantly in comparison to N2O cooling cycle.
Fig.8. Percentage Exergetic efficiency with gas cooler outlet temperature for Tevap = 275 K, T8 = 500 K and P = 170 Bar.
412 | IJREAMV04I033936 DOI : 10.18231/2454-9150.2018.0355 © 2018, IJREAM All Rights Reserved. International Journal for Research in Engineering Application & Management (IJREAM) ISSN : 2454-9150 Vol-04, Issue-03, June 2018
4.3. EFFECT OF TURBINE INLET TEMPERATURE ON THE PERFORMANCES OF N2O COMBINED POWER CYCLE Fig. 9 presents the variation of COP with turbine inlet temperature. As turbine inlet temperature increases 460 K to 600 K, the turbine output increases in turn COP increases monotonically for N2O combined power cycle. Fig. 10 shows the effect of turbine inlet temperature on the exergetic efficiency on the N2O combined power cycle. The trend of the cycle increases monotonically with turbine inlet temperature due to additional exergy output at turbine, hence it improves exergetic efficiency. 4.4. EFFECT OF GAS HEATER PRESSURE ON THE PERFORMANCES OF N2O COMBINED POWER CYCLE Fig.11. COP variation with Gas heater pressure for Tevap = 275 K, Gas heater pressure which is the discharge pressure of T3 = 308 K and T8 =500 K. the compressor is an important parameter. Fig. 11 depicts the variation of COP with gas heater pressure. At higher gas heater pressure turbine works at high inlet temperature, it increases the turbine output and reduces the compressor work input, so COP increases monotonically with gas heater pressure for the N2O combined power cycle. Due to improvement of turbine work with gas heater pressure, the exergy output at turbine increases; hence it increases exergetic efficiency for N2O combined power cycle as shown in Fig. 12.
Fig.12. Variation of percentage exergetic efficiency with gas heater pressure for Tevap = 275 K, T3 = 308 K and T8 = 500 K.
4.5. GRASSMANN DIAGRAM FOR N2O COOLING AND COMBINED POWER CYCLES Figs. 13 and 14 represent the percentage exergy loss in components for N2O cooling and combined power cycles, respectively. It is defined as the % exergy loss in components with respect to exergy input (100 %) to the compressor for simple cooling cycle and compressor, gas heater for combined power cycle. The operating conditions Fig.9. COP with turbine inlet temperature for Tevap = 275 K, T3 = 308 K and P = 175 Bar are taken as Tevap = 275 K, T3 = 308 K, T8=500 K and P = 175 bar for both the cycles. It is observed that 31.92 % exergy recovered at evaporator in simple cooling cycle, whereas 53.69% exergy recovered at evaporator and turbine in combined power cycle.
Tev = 275 K Pht = 175 bar T3 = 308 K Exergy to motor 100% Exergy to room 31.92%
Evaporator Throttle valve 5.32 % Compressor 12.36 % Gas Cooler 21.59 % 28.81 %
Fig. 13. Grassmann diagram for N2O simple cooling cycle
Fig.10. Percentage exergetic efficiency with turbine inlet temperature
for Tevap = 275 K, T3 = 308 K and P = 175 Bar
413 | IJREAMV04I033936 DOI : 10.18231/2454-9150.2018.0355 © 2018, IJREAM All Rights Reserved. International Journal for Research in Engineering Application & Management (IJREAM) ISSN : 2454-9150 Vol-04, Issue-03, June 2018
comb combined cycle gh gas heater tv throttle valve is isentropic mech mechanical efficiency o ambient II second law ht high temperature
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415 | IJREAMV04I033936 DOI : 10.18231/2454-9150.2018.0355 © 2018, IJREAM All Rights Reserved.