Performance comparison of vapour jet refrigeration system with environment friendly working fluids

Applied Thermal Engineering 21 (2001) 585±598

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Performance comparison of vapour jet refrigeration system with environment friendly working ¯uids K. Cizungu a, A. Mani b, M. Groll a,* a

Institut f ur Kernenergetik und Energiesysteme (IKE), Universit at Stuttgart, Pfa€enwaldring 31, D-70550 Stuttgart, Germany b Department of Mechanical Engineering, Indian Institute of Technology, Chennai 600 036, India Received 28 October 1999; accepted 15 May 2000

Abstract A computer simulation of a vapour jet refrigeration system is carried out using a one-dimensional model based on mass, momentum and energy balances. The simulated performance of the system is in good agreement with the available experimental performance from the literature. A comparison of system performance is carried out for the same ejector geometry using the environmentally friendly working ¯uids R123, R134a, R152a and R717 (ammonia). The results suggest that, for di€erent boiler temperatures, the entrainment ratio and the system eciency (COP) depend mainly on the ejector geometry and the compression ratio. Ó 2001 Elsevier Science Ltd. All rights reserved. Keywords: Simulation; Jet compressor; Ejector; Refrigerator; Solar energy

1. Introduction Heat powered refrigeration systems are attractive for the rational use of energy and could have signi®cant savings in the consumption of electrical energy and hence in the reduction of pollutant gases. Both absorption and ejector refrigerators can be powered by low grade heat energy. Compared with the absorption system, an ejector system is cheaper due to its simplicity in construction, installation and maintenance, it also does not need to use a two-component working ¯uid (working ¯uid pair). The main component of such systems is the ejector, which is robust and has no moving parts. Many researchers have investigated vapour jet refrigeration systems (VJRS) with water as working ¯uid. The main disadvantage of the steam VJRS is its inability to realise *

Corresponding author. Tel.: +49-711-685-2481; fax: +49-711-685-2010. E-mail address: [email protected] (M. Groll).

1359-4311/01/$ - see front matter Ó 2001 Elsevier Science Ltd. All rights reserved. PII: S 1 3 5 9 - 4 3 1 1 ( 0 0 ) 0 0 0 7 0 - 3

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Nomenclature

A COP d f h L Ma m_ P Q_ T w

area (m2 ) coecient of performance diameter (m) friction factor speci®c enthalpy (kJ/kg) length (m) Mach number mass ¯ow rate (kg/s) pressure (bar) heat transfer rate (kW) temperature (°C) velocity (m/s)

Greek g eciency l ˆ m_ s =m_ p entrainment ratio q density (kg/m3 ) / ˆ Am =At area ratio between mixing tube and primary nozzle throat n ˆ Pb =Pc driving pressure ratio w ˆ Pc =Pe compression ratio Subscripts b boiler c condenser d di€user e evaporator id ideal is isentropic m mixing tube p primary (¯ow, nozzle) s secondary (¯ow, nozzle) 1, 2, 3,. . ., x, y states in respective cross-sections of the ejector (Fig. 2)

subzero evaporator temperatures. Subsequent studies by various authors identi®ed R11 as a candidate to replace water to overcome the above disadvantage [1]. This refrigerant has been banned under the Montreal and subsequent International Protocols. Sun and Eames [2] have carried out an analysis with R123 as working ¯uid to use it in VJRS as a replacement for R11. In this paper, an attempt is made to consider other environmentally safe working ¯uids, viz. R134a, R152a, R717. The VJRS operating with the above working ¯uids is simulated on a computer using the equations derived from mass, momentum and energy balances. The simulated results are compared with the experimental data from the literature.

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2. System description Fig. 1 shows the schematic diagram of the VJRS. Fig. 2 depicts the con®guration of the ejector. The low grade heat, Q_ b is supplied to the boiler to generate high pressure vapour as motive ¯uid. This high pressure saturated vapour is expanded through the convergent±divergent nozzle in the ejector to get a high velocity stream, which entrains the saturated vapourised refrigerant from the evaporator. Both ¯uids mix together in the mixing chamber section and the pressure is recovered, while ¯owing through the di€user part of the ejector. It undergoes condensation by rejecting heat, Q_ c in the condenser. A portion of the condensate is expanded through an expansion device to the evaporator for realising the refrigerating e€ect, Q_ e . The remaining liquid is pumped to the boiler by the pump.

Fig. 1. Schematic diagram of vapour jet refrigeration system.

Fig. 2. Con®guration of the ejector.

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3. Analysis of the system The governing equations required for the analysis are obtained by applying the mass, momentum and energy balances across the respective control volumes. The following assumptions are made to simplify the complexity of the problem: · One-dimensional steady state ¯ow of saturated vapour. · Adiabatic ejector. · Losses in the mixing chamber are represented by a friction factor; the losses in the primary nozzle, secondary nozzle and the di€user are taken into account by assuming respective sub-component eciencies. · At the entrance of the cylindrical mixing tube, the suction ¯uid velocity reaches the speed of sound (choking condition) [3±5,11]. The velocity at the exit of the primary nozzle is evaluated as q wp3 ˆ 2gp …hb ÿ hp3is †; and the velocity of the suction ¯uid at the entrance of the mixing tube is q ws3 ˆ 2gs …he ÿ hs3is †:

…1†

…2†

From the continuity equation, the entrainment ratio is determined as lˆ

qs3 ws3 As3 : qp3 wp3 Ap3

…3†

Using the continuity equation, the total mass ¯ow through the mixing tube is computed as m_ m ˆ m_ p3 ‡ m_ s3 :

…4†

From the de®nition of the entrainment ratio, by applying the momentum balance, the velocity of the ¯uid in the mixing chamber is obtained as wm ˆ 

wp3 ‡ lws3 …P3 ÿ Pm †Am   : ‡ 1 ‡ f2m Ldmm …1 ‡ l† 1 ‡ f2m Ldmm m_ m

The mixing enthalpy is obtained by the energy balance as   w2 w2 hp3 ‡ 2p3 ‡ l hs3 ‡ 2s3 w2 hm ˆ ÿ m: 2 1‡l

…5†

…6†

From mass, momentum and energy balances, the following equations are obtained across the normal shock (section xy in Fig. 2) before di€usion takes place: qm wm ˆ qy wy ;

…7†

py ÿ pm ˆ qm w2m ÿ qy w2y ;

…8†

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  w2m ÿ w2y

: 2 The governing equations across the di€user are obtained as given below: m_ p ‡ m_ s ; qd ˆ Ad wd hy ÿ hm ˆ

589

…9†

…10†


gd ÿ 2 wm ÿ w2d : …11† 2 By assuming constant pressure mixing and neglecting velocities of the ¯uid at the di€user exit, the entrainment ratio can be deduced as r   hb ÿhp3is fm Lm g g ÿ 1 ‡ 2 dm hdis ÿhm p d : …12† lˆ   r he ÿhs3is fm Lm 1 ‡ 2 dm ÿ hd ÿhm gs gd hdis ÿ hm ˆ

is

Neglecting the pump work, the coecient of performance (COP) is de®ned as the ratio of refrigerating e€ect to heat supplied to the boiler Q_ e he ÿ hc COP ˆ ˆl : …13† hb ÿ hc Q_ b When the working ¯uid is assumed to be an ideal gas and gas dynamic relations [6,7] are introduced, then, the system is greatly simpli®ed. These equations are used for simulating the ideal performance of the system for comparison purposes. 4. Computational procedure Fig. 3 gives the ¯ow diagram of the computational procedure. For the given geometry of the ejector and operating conditions Tb , Tc , Te , Tb , Pc and Pe , the above governing equations are solved simultaneously. The suction pressure Ps3 can be determined from the choking condition of the suction ¯ow at the entrance of the constant-area mixing chamber. Assuming uniform nozzle exit pressure Pp3 ˆ P3 ˆ Ps3 , the exit thermodynamic state for both primary and secondary nozzles can be found. Hence, the nozzle exit velocities wp3 and ws3 can be determined. The mixed stream velocity wm and the enthalpy hm are evaluated by iteration assuming Pm . For a supersonic ¯ow condition in the mixing chamber, the existence of a normal shock is checked and the required parameters are calculated across the shock. Finally, using the di€user enthalpy hd , the di€user velocity wd is evaluated from an energy balance. This computation process is repeated till the continuity equation is satis®ed. Finally, the entrainment ratio and COP are calculated. 5. Computational input data The following ejector geometry data have been used for the calculations: a0 ˆ 20 ;

a ˆ 4 ;

b ˆ 0 ;

h ˆ 3 ;

Lm =dm ˆ 10;

Lx ˆ 0 m:

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Fig. 3. Flow diagram for the computational procedure.

For the area ratio /, which is the main geometric parameter of the ejector, di€erent values have been used, viz. / ˆ 4:0, 5.76 and 7.84. The other geometry data are calculated by the programme. To take into account the losses in both primary and secondary nozzles and the di€user in the calculations, the isentropic eciencies gp ˆ 0:95, gs ˆ 0:95, and gd ˆ 0:85, respectively, were taken from the available literature [8]. For accounting the losses in the mixing section, the friction factor is taken as fm ˆ 0:03 [8,9]. 6. Results and discussion 6.1. Reference calculations for R11 For validation of the theoretical model, the performance for R11 was simulated on the computer by varying the evaporator temperature from 4°C to 16°C, the boiler temperature from 63°C to 85°C, the condensing temperature from 25°C to 30°C and by choosing the area ratios / ˆ 4:0, 5.76 and 7.84. The theoretical performance computed using the model is compared with that of experimental data available in the literature [10]. Fig. 4 shows the variation of entrainment ratio l with compression ratio w for di€erent geometries / ˆ 4:0 (curves set 1) and / ˆ 5:76 (curves set 2). The corresponding temperatures are Tb ˆ 67 C, and Tc ˆ 30 C for set 1 and Tb ˆ 78:5 C, and Tc ˆ 30 C for set 2. Since the condensing

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Fig. 4. E€ect of compression ratio on entrainment ratio for R11.

temperature Tc is ®xed, the evaporator temperature Te is the variable (Te ˆ 0±10 C and Te ˆ ÿ0:5±13:5 C, respectively). It can be seen that the simulated performance of the model agrees well with the experimental performance in the range of the operating conditions and geometry considered. In addition, the entrainment ratio is shown for ideal conditions without friction, and for the area ratio / ˆ 5:76 (curve theoid ). In this ideal case, the entrainment ratio is signi®cantly higher. Fig. 5 shows the variation of COP of the VJRS with boiler temperature. The COP of the system is computed corresponding to the experimental maximum performance ejector con®gurations of / ˆ 4:0, / ˆ 5:76 and / ˆ 7:84 [11] at Tc ˆ 27:7 C, and Te ˆ 8:8 C; the respective boiler temperatures are Tb ˆ 63 C, 76 C and 84:5 C. It can be observed that the theoretical COP of the simulated model agrees fairly well with the experimental data from the literature [11]. Fig. 6 depicts the variation of COP with the evaporator temperature for the given conditions of Tb ˆ 80 C, Tc ˆ 29 C and / ˆ 5:76. It can be seen that the COP increases as the evaporator temperature increases. The computed COP of the model agrees well with the experimental COP from the literature [11]. 6.2. Calculations for R123, R134a, R152a, and R717 Some properties of the selected refrigerants are given in Table 1. Comparative calculations have been made for R123, R134a, R152a and R717. In all cases, the following operating temperatures have been assumed: · boiler temperatures: Tb ˆ 60°C, 70°C, and 80°C, · condenser temperatures: Tc in the range from 25°C to 40°C, · evaporator temperatures: Te in the range from ÿ5°C to 5°C.

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Fig. 5. E€ect of boiler temperature on COP for R11.

Fig. 6. E€ect of evaporator temperature on COP for R11.

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Table 1 Some properties of the selected working ¯uids (R11 for comparison only) Name

Critical pressure (bar)

NBPa (°C)

ODPb

GWPc

ALTd (a)

LFTe (volume % in dry air)

TOXf (ppm)

CATg

R717 R134a R152a R123 R11

113.4 40.6 45.2 36.6 44.7

ÿ33.6 ÿ26.4 ÿ24.3 ÿ27.5 23.8

0 0 0 0.02 1

1 420 47 29 1500

<1 16 2 2 60

15.0 None 3.7 None None

25 1000 1000 10 1000

A A A B C

a

Normal boiling point. Ozone depletion potential (relative to R11). c Global warming potential (relative to CO2 : GWP ˆ 1; integration time ˆ 500 a). d Atmospheric lifetime. e Lower ¯ammability limit. f Toxicity. g A: environmentally acceptable replacements, B: transitional replacements, and C: banned by the Montreal Protocol. b

Fig. 7(a±c). E€ect of compression ratio on COP.

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Fig. 7(a)±(c) shows the e€ect of the compression ratio on the COP for di€erent refrigerants at the boiler temperatures of 60°C, 70°C and 80°C, respectively, and ®xed / ˆ 4:0 …/ ˆ 4:0 is optimum for R11 at Tb ˆ 63 C, but not for higher Tb and not for other ¯uids). The corresponding ranges of evaporator temperature and condenser temperature are from Te ˆ ÿ5 C to 5°C and from Tc ˆ 25±40°C, respectively. R11 is included only for comparison in Fig. 7(b). The COP of the system decreases as the compression ratio increases because the increase in compression ratio increases the pressure di€erence between the condenser and the evaporator. The COPs of all ®ve refrigerants are not very di€erent. Speci®cally; R717 has a somewhat lower COP at Tb ˆ 60 C, but has the highest COP at Tb ˆ 70 C and 80°C. This is because of its higher pressure. For the ejector geometry parameter / ˆ 4:0, and the ranges of the considered evaporator and condenser temperatures, the optimum boiler temperature for R717 is around Tb ˆ 76 C, but it lies lower for the other refrigerants (around Tb ˆ 67 C for example). The decrease of the COP with increasing compression ratio as shown in Fig. 7 can be explained as follows: apart from the losses within the ejector compressor, the kinetic energy of the primary ¯ow at the driving nozzle exit serves on the one hand to draw in the secondary ¯ow and on the other hand to compress the mixed ¯ow to the di€user pressure. For ®xed condenser temperature, an increase of the evaporator temperature reduces the compression ratio w ˆ Pc =Pe . This means that only a small part of the available kinetic energy of the primary ¯ow will be used for the pressure lift between the evaporator and the di€user. All the rest of the kinetic energy will serve to induce the secondary ¯ow in the mixing chamber, that is the entrainment ratio will increase. The same observation can be made in Figs. 8 and 9, wherein the geometry parameter / has been changed. Fig. 8 illustrate the performance characteristics of the system for the ejector geometry parameter / ˆ 5:76 and Tb ˆ 60 C, 70°C, and 80°C. The higher the ejector geometry parameter, the higher the boiler temperature and hence the higher the performance characteristics (entrainment ratio and COP). For this reason, only the refrigerant R123 operates at Tb ˆ 60°C and / ˆ 5:76 (Fig. 8(a)). For the other working ¯uids, the COP is practically zero and higher boiler temperatures are necessary. For Tb ˆ 70 C (Fig. 8(b)), all refrigerants (including R11 for comparison) have similar COPs. However, R717 needs a boiler temperature higher than 70°C. For 70°C, its COP would be very small. The COPs are generally higher than for the case / ˆ 4:0. For Tb ˆ 80 C (Fig. 8(c)) the situation is similar. However, the COPs become smaller for w 6 2. The COP for R717 is still small, but it has been included for comparison. Fig. 9 shows the COP of the system vs. compression ratio for the ejector geometry parameter / ˆ 7:84. In this case, the COPs are noticeably higher than for Tb ˆ 70 C=/ ˆ 5:76; this holds for all refrigerants. R717 requires a minimum boiler temperature considerably higher than 80°C. For Tb ˆ 90 C and / ˆ 7:84, the system performance would be COP ˆ 0:43. The e€ect of evaporator temperature on COP is shown in Fig. 10, for ®xed values Tb ˆ 90 C, and / ˆ 6:68. As expected, the higher the evaporator temperature, the higher the COP. For the condenser temperature Tc ˆ 30 C (Fig. 10(a)), high values of the COP can be achieved, specially by using the refrigerants R134a and R717 for which the given boiler temperature comes near to the optimum for the above ejector geometry parameter and evaporator temperature range. For Tc ˆ 35 C (Fig. 10(b)), the ranking of the working ¯uids is the same. However, the COP is noticeably smaller, showing that besides the ejector geometry, the compression ratio has a considerable in¯uence on the VJRS performance [10±12].

K. Cizungu et al. / Applied Thermal Engineering 21 (2001) 585±598

Fig. 8(a±c). E€ect of compression ratio on COP.

Fig. 9. E€ect of compression ratio on COP.

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Fig. 10(a,b). E€ect of evaporator temperature on COP.

As a summary, Fig. 11 shows a three-dimensional presentation of the COP as a function of ejector area ratio / and driving pressure ratio n, for a ®xed compression ratio w ˆ 2:2. Since only

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Fig. 11. Variation of COP versus driving pressure ratio n and geometry ratio /, for a ®xed compression ratio w ˆ 2:2.

non-dimensional parameters are used in the presentation, it holds for all working ¯uids considered. 7. Conclusions A comparison has been made of di€erent environmentally friendly working ¯uids for the same operating conditions and the same ejector geometry using a one-dimensional model which is shown to accurately predict the VJRS performance. Based on this comparative study, it can be concluded that all the selected working ¯uids give more or less the same performance characteristics, depending on the ejector geometry and the operating conditions. For given ejector geometry parameter / and compression ratio w, the optimum driving temperatures Tb of the alternative ¯uids R123, R134a and R152a are noticeably lower than the one of R717. Therefore they perform better than R717 for low boiler temperatures (e.g. Tb < 70°C for / ˆ 4:0). For higher temperatures (Tb P 70 C and the same small / ˆ 4:0), R123, R134a and R152a are far away from their respective optimum, while R717 comes close to its optimum, giving the highest performance. For higher ejector geometry parameters and boiler temperatures, e.g. /  7:0 and Tb P 90 C, R717 and R134a have about the same performance. This suggests that in a VJRS operating with low grade heat source such as solar energy, waste heat, etc. in the temperature range 70 C 6 Tb 6 85 C, the working ¯uids R134a and R152a should be used in combination with high ejector geometry parameters 5 6 / 6 8 to achieve high COP. When using R717 in the same range of driving temperatures, the ejector geometry parameter / should be chosen smaller, e.g. / 6 5; for greater ejector geometry parameters, e.g. / P 6±7, higher driving temperatures, e.g. Tb P 90 , are necessary to achieve good performance.

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References [1] P. Paliwoda, A review paper on the experimental study on low grade heat and solar energy operated halocarbon vapour jet refrigeration system, Topical Studies, IIR Bull. 1968, p. 1003. [2] D.W. Sun, I.W. Eames, Performance characteristics of HCFC-123 ejector refrigeration cycles, Int. J. Energy Res. 20 (1996) 871. [3] J. Fabri, R. Siestrunck, Supersonic air ejectors, Advances in Applied Mechanics, vol. 5, Academic Press, New York, 1958. [4] K. Matsuo, K. Sasaguchi, K. Tasaki, H. Mochizuki, Investigation of supersonic air ejectors. Part 1: Performance in the case of zero-secondary ¯ow, Bull. JSME 24 (1981) 2090. [5] K. Matsuo, K. Sasaguchi, K. Tasaki, H. Mochizuki, Investigation of supersonic air ejectors. Part 2: E€ects of throat-area-ratio on ejector performance, Bull. JSME 25 (1982) 1898. [6] H. Christensen, Application of gas-dynamic functions for steam ejector design, Heat Trans. Engng. 4 (1983) 83. [7] K. Cizungu, M. Groll, Leistungscharakteristik eines k uhlsystems mit dampfstrahlverdich, Ki-Luft- und K altetechnik 7 (1999) 348. [8] J.-J. Henzler, Design of ejectors for single-phase material systems, Ger. Chem. Engng. 6 (1983) 292. [9] H.-P. Schlag, Experimentelle und theoretische Untersuchungen zur Berechnung der Kennvon gasbetriebenen Einphaseninjektoren und Gutaufgabeninjektoren. Fortschritt-BeVDI, Reihe 3, Nr. 313: Verfahrenstechnik. VDI Verlag GmbH, D usseldorf, 1993. [10] E. Nahdi, J.-C. Champoussin, G. Hostache, J. Cheron, Optimal geometric parameters of a cooling ejectorcompressor, Int. J. Refrig. 16 (1993) 67. [11] L.T. Lu, J.-C. Champoussin, E. Nahdi, Performances optimales et utilisation du systeme  a ejecteur en production de froid. RGF-Octobre (1988) 529. [12] B.J. Huang, et al., A 1-D analysis of ejector performance, Int. J. Refrig. 22 (1999) 354.