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  1. 1. USAGE OF A TWO-PHASE EJECTOR AS AN EXPANSION DEVICE TO IMPROVE THE PERFORMANCE OF R-134A USED REFRIGERATION SYSTEM 1. K.GANESH BABU, 2. K.RAVI KUMAR B.Tech., M.Tech -R&A/C. Assistant Professor, Department of Mechanical Engineering, All- Habeeb College of Engineering, Chevella, Ranga Reddy – 515002, Andhra Pradesh, India. Email: katam.ganeshbabu@gmail.com Email: ravikumarkotturi@gmail.com RAJESH B.Tech., M.Tech -R&A/C. Department of Mechanical Engineering, JNTUA College of Engineering, Anantapur – 515002, Andhra Pradesh, India. Email: stanlyrajesh@gmail.com Abstract— several preceding researches have evidenced that the simple vapour compression refrigeration system using expansion device has an inherent inefficiency resulting in degraded system performance. In the present work Refrigerant R-134a is used as working fluid. A computer simulation of the Simple vapour compression system and Ejector used vapour compression system is carried out using Engineering equation solver software (EES) commercial version 6.883, by using 1-Dimensional Two-phase ejector model. Under the optimal values of Ejector area ratio (Ar) = 14, entrainment ratio (U) = 0.53, nozzle efficiency (nn) = 85%, diffuser efficiency (nd) = 85% at operating conditions of evaporator temperature (Te) = -15°C and condenser temperature (Tc) = 30°C. The Ejector used vapour compression refrigeration (EVCR) system COP = 4.649, at the same operating conditions, simple vapor compression refrigeration system COP = 3.733. The percentage improvement of COP of Ejector used vapour compression refrigeration system 19.70% higher than that of the Simple vapour compression refrigeration system. Keywords— Vapour compression refrigeration system; Refrigerant; EES software; COP; R134a I INTRODUCTION With the development of society, approximately 15% of the world electricity consumption is used for refrigeration and air conditioning applications. How to reduce the energy consumption has become a worldwide research topic. Compared with other refrigeration systems, Ejector used vapour compression refrigeration system has some special advantages such as the simplicity in construction, high reliability and low cost. Throttling loss is one of the thermodynamic losses in a conventional vapor compression refrigeration system. In order to reduce this loss, various devices and techniques have been attempted to use instead of the conventional devices like Capillary tube, Thermostatic expansion valve. Ejector is a device that uses a high-pressure fluid to pump a lowpressure fluid to a higher pressure at a diffuser outlet. Its low cost, no moving parts and ability to handle two-phase flow without damage make it attractive for being the expansion device in the refrigeration system. In order to reduce the throttling losses of vapour compression cycle we replace the throttling
  2. 2. device. Ejector as an expander in that, the process is isentropic. By using ejector devices we are able to reduce the throttling looses and reduces the load on evaporator so that the COP of the combined compression refrigeration cycle is increases. The descriptions of ejector are as given below 1.1 Two phase ejector: Ejector is an expander which uses an expansion device in vapour compression cycle and replaces the throttling loss by replacing throttling device in vapour compression cycle. The appropriate installation of the ejector in vapour compression cycle increases the COP of the refrigeration system by raising the compression suction pressure to a level higher than that in the evaporator and consequently, to reduce the load on the compressor and motor. In the ejector geometry there are three sections of ejector 1. 2. 3. II.EJECTOR USED SYSTEM DESCRIPTION: Figure.3 Schematic diagram of the Ejector used Vapour compression refrigeration system Motive Nozzle, Mixing tube, Diffuser section. Figure.4 P-h diagram of the Ejector used vapour compression system Figure.1 Configuration of ejector (1-Dimentional) Figure.2 Pressure variation along the ejector length In Ejector used Vapour compression refrigeration cycle comprises of the following Ejector, Evaporator, Compressor, Condenser, Separator, etc. Figure.1 shows the configuration of ejector. Ejector is an expression device which replaces the throttling valve in order to reduce the throttling losses of expansion device. The ejector is installed at the outlet of the condenser (6 to 1), and the motive fluid (liquid from the condenser) enters into the nozzle at a relatively high pressure. Reduction of the pressure of the liquid in the nozzle provides the potential energy for conversion to kinetic energy of the liquid. The driving flow entrains vapour out of the evaporator. The two phases are mixed in mixing chamber (at point 2) and leave it after a recovery of pressure in the diffuser part of the ejector (at point 3). The liquid portion is directed to the evaporator through a small pressure-drop expansion device (7 to 8) while the vapour portion enters the compressor suction (3 to 4). The lines from points 4, 5, 6 are a series process in the compressor and the condenser. The lines from points 7, 8, 9 are a series process in
  3. 3. the expander and the evaporator. Points 6 and 1 are the state of the flow at the exit of the primary nozzle and in the mixing area (point 2) of the ejector while point 2–3 is a compression process in Diffuser. The appropriate installation of the ejector increases COP of the refrigeration system by raising the compression suction pressure to a level higher than that in the evaporator and consequently, to reduce the load on the compressor and motor. R-134a (CH2FCF3- Tetrafluoro Ethane) is from Hydro Fluoro Carbons (HFCs) family. It is used as working fluid in the vapour compression system and Ejector used vapour compression refrigeration system because it has Higher safety level i.e., A1, zero - Ozone Depletion Potential (ODP), and it has 1300 - Global Warming Potential (GWP). exiting numerical equations solving programs. First, EES automatically identifies and groups equations that must be solved simultaneously. This feature simplifies the process for the users and ensures that the solver will always operate at optimum efficiency. Second, EES provides many built in mathematical and thermo physical property functions useful for engineering calculations. The basic function provided by the engineering equation solver (EES) is the numerical solution of the non-linear algebraic and differentials equations, EES provides built in thermodynamics and transport property functions for many fluids including water, dry and moist air. Included in the property database are thermodynamics properties for H2O-LiBr and NH3-H2O mixture. Any information between quotation marks [“] or [{}] is an optional comment. Variable names must start with a letter. A code containing a good library of working fluid properties suitable for heat pumps is the Engineering Equations Solver (EES). Here the user must write the equations governing the cycle and make sure the set is well-defined. In the case of a non-linear set of equations, the user must check the results to make sure that the mathematical solutions are also a physical One. 3.2 Simulation analysis on compression refrigeration system: Figure.5 Comparison between both the systems III. SIMULATION ANALYSIS: Based on the Thermodynamic analysis, a steady-state simulation program for the vapour compression refrigeration cycle, using EES software (Klein and Alvarda (2006)) is developed. 3.1 Solution methodology in engineering equation solver (EES) (Klein and Alvarado (2007)) EES is a software package developed by Dr. Sanford Klein of the University of Wisconsin. EES incorporates the programming structure of C and FORTRAN with a built-in iterate, thermodynamics and transport property relations, graphical capacities, numerical integration, and many others useful mathematical functions. By grouping equations there are to be solved simultaneously, EES is able to function at a high rate of computational speed. Ammonia-water mixture properties are calculated in EES using the correlation developed by Ibrahim and Klein (1993). There are two major differences between EES and Vapour 3.2.1 Simulation was performed to evaluate the COP of vapour compression cycle with the following assumptions 1. 2. 3. 4. 5. 6. The refrigerant was at all times in thermodynamic quasi-equilibrium. Characteristics and velocities were constant over cross section Negligible pressure drop. There is no wall friction. The processes in compressor, expansion valve area assumed to be adiabatic. Saturated state at the evaporator and the condenser outlet. 3.2.2 Input parameters for Vapour compression refrigeration system simulation: Evaporator temperature (Te) Condenser temperature (Tc) Mass flow rate (m) Refrigerant : : : : -15 0C 30 0C 1/60 (kg/sec) R-134a
  4. 4. 3.2.3 Flow chart for Vapour compression system simulation analysis: 3.3 Simulation was performed to evaluate the ejector used vapour compression cycle with the following assumptions: 1. 2. 3. 4. 5. 6. 7. 8. 9. 3.2.4 Vapour compression refrigeration system simulation results: Evaporator temperature (Te) Condenser temperature (Tc) = = = = 1.64 bar 7.706 bar Compressor inlet enthalpy (h1) Compressor outlet enthalpy (h2) Evaporator inlet enthalpy (h4) = = = 241.5 kJ/kg 281.1kJ/kg 93.58 kJ/kg Mass flowrate (m) 3.3.1 Flow chart for Ejector used vapour compression refrigeration system simulation analysis: -15 ˚C 30 ˚C Evaporator pressure (Pe) Condenser pressure (Pc) The refrigerant was at all times in thermodynamic quasi-equilibrium. Characteristics and velocities were constant over cross section (1-dimensional model). All fluid characteristics are uniform over the cross section after complete mixing at the exit of the mixing tube. There is no external heat transfer to the system. There is no wall friction. Negligible pressure drop. The processes in compressor, expansion valve and ejector area assumed to be adiabatic. Saturated state at the evaporator and the condenser outlet. One dimensional flow in the ejector. 0.01667 kg/sec = Coefficient of performance (COP) = = = = =  m   m    1h  2h     0.01667 *  241.5  93.58   0.01667 *  281.1  241.5   1 4 7 . 9  3 9 . 6  3.733 2     4 h 1h        3.3.2 Input parameters for Ejector used system simulation analysis: Evaporator temperature ( ) Condenser temperature ( ) Ejector convergent nozzle efficiency ( ) : : -15 ºC 30 ºC : 85%
  5. 5. Ejector diffuser efficiency ( Ejector area ratio (Ar) Entrainment ratio (U) Refrigerant ) : : : : 85% 8 1 R134a 3.3.3 ALGORITHM TO SIMULATE EJECTOR USED VAPOUR COMPRESSION REFRIGERATION SYSTEM: 3.3.3.1 Flow in the Convergent nozzle: The exit velocity from the nozzle is calculated from )) = √( ∗ ( − (3.1) h1a = h6 = Nozzle outlet actual enthalpy Condenser or Nozzle inlet enthalpy = Nozzle outlet velocity V1 h1 is the enthalpy, at the outlet of the motive nozzle = ( , ) (3.2) = − ∗( − ) For an isentropic process actual enthalpy (3.3) For an isentropic process actual pressure = − ∗( − ) (3.4) The density, at the outlet of the motive nozzle = ( , ) (3.5) Actual density, at the outlet of the motive nozzle = ( , ) (3.6) The mass flow rate, at the outlet of nozzle = ∗ ∗ (3.7) 3.3.3.2 Flow in the mixing tube: Using the continuity equation, the total mass flow through the mixing tube is computed as + = ∗ ∗ (3.8) A momentum balance of the mixing tube yields ( − )∗ = ∗ −( + )∗ (3.9) The density ratio, can be approximated by Chen (1988) as = ∗ + (3.10) Where U = represents the flow entrainment ratio (m2/m1) = Is the refrigerant’s vapour density at the evaporator outlet The mixing velocity is defined as = = ∗ + ∗ (3.11) ∗ −( / ) (3.12) The enthalpy, at the outlet of the mixing section The entropy, at the outlet of the mixing section = ( , ) (3.13) 3.3.3.3 Flow in the Diffuser: The enthalpy, at the outlet of the diffuser, by conservation of energy = + / (3.14) The exit diffuser actual enthalpy is computed as = + ∗( / ) (3.15) The exit diffuser pressure is defined as (s2 = s3) = ( , ) (3.16) The exit diffuser actual enthalpy is computed as =( − )/( − ) (3.17) The isentropic compressor efficiency proposed by Mouna Elakdhar (2007) = . − . ∗ (3.18) Where Pr = Compressor pressure ratio =( / ) (3.19) The isentropic compressor efficiency =( − )/( − ) (3.20) The cooling capacity is defined by = ∗( − ) (3.21) Where ℎ = superheated vapour enthalpy at ℎ = saturated vapour enthalpy at ℎ = Isentropic vapour enthalpy at Where ℎ = saturated vapour enthalpy at ℎ = saturated liquid enthalpy at The coefficient of the performance of the Ejector used Vapour compression system, is determined by the following equation = = ∗ ( ( ) ) (3.22) 3.3.4 Ejector used vapour compression refrigeration system simulation results: Condenser temperature (Tc) Evaporator temperature (Te) Condenser pressure (Pc) = = = 30˚C -15˚C 7.706 bar
  6. 6. Evaporator pressure (Pe=P1) = 1.64 bar, Enthalpy (h1) = 88.25 kJ/kg Nozzle outlet actual pressure (P1a) = 2.55 bar, Enthalpy (h1a) = 89.05 kJ/kg Mixing tube pressure (P2) = 2.794 bar, Enthalpy (h2) = 140.2 kJ/kg Diffuser actual outlet pressure (P3a) = 3.263 bar, Enthalpy (h3a) = 141.6 kJ/kg Area ratio (Ar) = 14 Entrainment ratio (U) = (m2/m1) Where Mass flow rate in Evaporator (m2) = 0.005773 kg/sec Compressor (m1) = 0.01089 kg/sec Entrainment ratio (U) = 0.53 Nozzle efficiency (nn) = 85% Diffuser efficiency (nd) = 85% Compressor inlet enthalpy (h4) = 252.2 kJ/kg Compressor outlet enthalpy(h5) = 273.3 kJ/kg Evaporator inlet enthalpy (h7 = h8) = 55.9 kJ/kg Evaporator outlet enthalpy (h9) = 241.5 kJ/kg Coefficient of performance (COP) = ((m2*Refrigerating effect)/(m1*Work supplied)) Compressor Work (Wc) in kJ/kg 10 5 0 1 1.1 1.2 1.3 1.4 1.5 1.6 1.7 1.8 Area ratio (Ar) Graph 1: Area ratio (Ar=4 to 18) Vs compressor work (Wc in kJ/kg) Area ratio (Ar) = (A2/A1) Graph.2: Entrainment ratio Vs Compressor work (Wc in kJ/kg) at evaporator temperature (Te=-15˚C) & condenser temperature (Tc=30˚C) Entrainment ratio (U) = (m2/m1) U= (Mass flow rate from evaporator outlet) / (mass flow rate from nozzle outlet) Entrainment ratio increasing means the mass flow rate from the Evaporator Increasing i.e., m2 or by decreasing mass flow rate from motive nozzle i.e., m1. When we are increasing mass flow rate in evaporator automatically the cooling capacity will increase because of that coefficient of performance (COP) of the system will increase. 4.3 Effect of variation of Nozzle and Diffuser efficiencies: Coefficient of Performance (COP) Compressor Work (Wc) 34.84 31.66 33.16 Compresso r work (Wc) in kJ/kg 9 10 11 12 13 14 15 16 17 18 = (area at the inlet of the diffuser) / (area at the outlet of nozzle) 29.08 24.19 24.14 24.16 24.11 24.12 24.12 24.16 24.23 24.36 25 8 24.59 25.71 27.03 29.74 36.56 7 30.31    Area ratio(Ar) Vs Compressor work(Wc) 6 26.91 4.649 5 15 Entrainment Ratio (U) 4.1 Effect of variation of Area ratio: 4 Compressor Work (Wc) in kJ/kg 20 0.4 0.5 0.6 0.7 0.8 0.9 IV RESULTS AND DISCUSSION: 40 35 30 25 20 15 10 5 0 27.96     25     25.93 25 23.17 = 30 21.19  1 .0 7 1   0 .2 3 0 4 35 24.09    22.23 = Entrainment ratio (U) Vs Compressor Work (Wc) in kJ/kg 18.49 h8) h 4 ))   0.005773 * (185.6)     0.01089*  21.15   = 4.2 Effect of variation of Entrainment ratio: 40   0.005773 * (241.5  55.9)    0.01089*  273.3  252.2   = Pressure ratio (Pr) = = (Pressure at the outlet of compressor) / (pressure at the inlet of compressor). The increase in area ratio up to optimum level the suction pressure of the compressor is increased so that the pressure ratio is decreased. After the optimum level any increase in area ratio the suction pressure is decrease so the pressure ratio increases. 19.99   m 2 * ( h9     ( m1 * ( h 5 = Area ratio increasing means the area of the mixing tube is increasing. Area of the mixing tube increases the pressure from the diffuser outlet is increases, at optimum area ratio (Ar = 14) the compressor pressure ratio (Pr) minimum because of that compressor work also minimum at that optimal point but COP is maximum. Nozzle Efficiency (nn) Vs COP at Tc=30, Te=-15 ºC 90 80 70 60 50 40 30 20 10 0 78.78 5.761 10 20 30 40 50 60 Nozzle Efficiency (nn) 70 80 90 100
  7. 7. 7.725 7.622 7.673 7.52 7.571 7.419 7.469 7.4 7.32 7.6 7.369 7.8 COP 7.2 7 Compressor work (Wc) in kJ/kg Graph.4 Diffuser efficiency (nn=10 to 100 %) Vs COP Diffuser efficiency(nd) Vs Compressor work (Wc) in kJ/kg 26.6 26.38 26.16 25.95 26 25.73 25.52 25.5 25.31 25.1 24.9 25 24.69 Compressor Work (Wc) in kJ/kg 24.5 24 23.5 10 20 30 40 50 60 70 80 90 100 Diffuser efficiency (nd) Graph.5 Diffuser efficiency (nd) Vs Compressor work (Wc in kJ/kg) Above Graph.3 shows the decreasing of COP by increasing the nozzle efficiency. Diffuser efficiency is mainly affecting the COP of the system, because the diffuser efficiency increasing means the pressure of the refrigerant from diffuser inlet to outlet pressure is increasing with respect to increase in efficiency shown in Graph.4. But it is going to affect on the refrigerating effect i.e. enthalpies (h9-h8) starts to decrease, again the expansion valve throttling losses coming to matter. For better results the nozzle efficiency is little bit less than or equal to diffuser efficiency. 4.4 Effect of variation in Evaporator and Condenser temperature: Coefficient of Performance (COP) 5 5.013 4.819 4.649 4.501 4.372 4.258 4.158 4.071 3.995 4 3 4.382 Both systems COP comparison by changing Evaporator temperature(Te) 7 6.56 6.496 6.059 6 5.368 5.911 5.094 4.856 5.354 4.47 4.649 5 4.314 4.87 4.445 4.069 4 3.733 3.431 3.157 3 5.687 -22.5 -20 -17.5 -15 -12.5 -10 -7.5 -5 -2.5 Evaporator temperature (Te in ˚C) 4.04 2 3.733 3.457 3.206 2.977 2.767 2.574 1 2.395 Ejector Used VCR system COP VCR system COP 0 27.5 30 32.5 35 37.5 40 Condenser temperature (Tc in ˚C) 42.5 45 Graph.7 Both systems COP comparison by changing Condenser temperature(Tc) Diffuser Efficiency (nd) 26.5 6 25 10 20 30 40 50 60 70 80 90 100 27 Both systems COP comparison by changing Condenser temperature(Tc) Coefficient of performance (COP) Diffuser Efficiency (nd) Vs COP at Tc=30, Te=-15 ºC 7.271 Coefficient of Performance (COP) Graph.3 Nozzle efficiency (nn=10 to 100 %) Vs COP Ejector used system COP 0 Graph.6 Both systems COP comparison by changing Evaporator temperature(Te) Graphs 6 & 7 showing the performance of the vapour compression refrigeration system and ejector used vapour compression refrigeration system with respect to variation of evaporator temperature (Te in ˚C) and condenser temperature (Tc in ˚C). When the evaporator temperature decreasing (i.e.,-20 to 0), the performance of the vapour compression system is very less compare to ejector used vapour compression refrigeration system. When the condenser temperature increasing (i.e., 25 to 45), the performance of the vapour compression system is very less compare to ejector used vapour compression refrigeration system. Based on simulation analysis the condenser temperature is mainly affecting on the performance of vapour compression refrigeration system and ejector used vapour compression refrigeration system. The Ejector used refrigeration system is especially suitable for low temperature refrigeration applications. V: CONCLUSIONS In the present simulation analysis the results has been computed for vapour compression refrigeration cycle & ejector used vapour compression refrigeration cycle. The effect of the geometry of the ejector Area ratio with the refrigerant R134a has been analyzed. The maximum COP = 4.649 is obtained for optimum area ratio whose value is around Ar = 14 for ejector used refrigeration system because, at this area ratio the COP is maximum and compressor pressure ratio (Pr) is minimum. Optimal entrainment ratio(U) = 0.53 because the compressor pressure ratio (Pr) = 2.362 low compared to simple vapour compression system compressor pressure ratio (Pr) = 4.698, the quality of vapour refrigerant at outlet of ejector x3=0.4366 is optimal at entrainment ratio (U) =
  8. 8. 0.53 compare to quality of vapour refrigerant at outlet of ejector x3 = 0.8331 at entrainment ratio (U) = 2. The simulation analysis results shows that for a given evaporator temperature (Te) = -15˚C and the condenser temperature (Tc) =30 ˚C, the vapour compression refrigeration cycle COP = 3.733 is less than that of the ejector used vapour compression refrigeration cycle COP = 4.649. The system performance mainly depends on the diffuser efficiency, because the diffuser efficiency increase is increasing the pressure at the outlet of diffuser. This pressure increase is decreasing the pressure ratio of the compressor. Compressor pressure ratio decreases the compressor work also decreases. If work decreases the COP of the system will increase. The system performance also depends on the nozzle efficiency. The nozzle efficiency increase is decreasing the pressure at the outlet of nozzle. This pressure decrease is increasing the pressure ratio (Pr) of the compressor. Compressor pressure ratio increases the compressor work (Wc) also increases. If work increases the COP of the system will decrease. For better results the nozzle efficiency is little bit less than or equal to diffuser efficiency. The performance increase of the ejector used vapour compression refrigeration system is 19.70%. VI: Comparison of result: Graphs.6 and 7 shows the COP comparison of vapour compression refrigeration (VCR) system & ejector used vapour compression refrigeration (EVCR) system at evaporator temperature (Te) =-15˚C, condenser temperature (Tc) = 30˚C. Table No: 1 m = 0.01667, m2 = 0.01386 VCR EVCR System System Refrigerating effect 147.9 185.6 (RE in kJ/kg) Cooling capacity 2.465 2.571 (Qe in kW) Compressor pressure 4.698 2.362 ratio (Pr) Compressor input work 39.62 21.15 (Wc in kJ/kg) Compressor input 0.6603 0.5531 power (PowerC in kW) Efficiency of the 81.06 84.21 compressor (nc) Refrigerant quality at 0.294 0.1141 evaporator inlet (x8) COP 3.733 4.649 VII: References: 1. A.Pongtornkulpanich, S.Thepa and M.Amornkitbamburg: exergy analysis: absorption heat transformer cycle with a combining ejector using lithium bromide/water as working fluid. 2. S.A.Klein: Development and integration of an equation solving program for engineering thermodynamic courses: Mech engg department, university of visconsin, Madison – 53706. 3. Chen, LT. (1988) ‘A new ejector–absorber cycle to improve the COP of an absorption refrigeration system’, Applied Energy, Vol.30, pp.37–51. 4. Elbel, S., Hrnjak, P. (2006) ‘Experimental validation of a prototype ejector Designed to reduce throttling losses encountered in transcritical R744 system operation’ International journal of refrigeration, Vol.32, pp.411-412. 5. Huang, B.J., Hu, S.S., Lee, S.H. (2005) ‘Development of an ejector cooling system with thermal pumping effect’ International journal of refrigeration, Vol.29, pp.476-484. 6. Klein, S.A., Alvarda, F., (2003) Engineering Equation Solver, Version 6.883. F-chart software, Middleton, WI. 7. Li, D., Groll, A. (2005) ‘Transcritical CO 2 refrigeration cycle with ejector expansion device’, International Journal of Refrigeration, Vol. 28, pp.766–773. 8. Nehdi, E., Kairouani, L., Bouzaina, M. (2007) ‘Performance analysis of the vapour compression cycle using ejector as an expander’, International Journal of Energy Research, Vol. 31, pp.364–375. 9. Refrigeration and Air-conditioning by R.S.Kurmi 10. Refrigeration and Air-conditioning by Domkundwar, Arora, Domkundwar.

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