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Performance estimation of ejector cycles using heavier hydrocarbon refrigerants
Applied Thermal Engineering 71 (2014) 197e203
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Applied Thermal Engineering journal homepage: www.elsevier.com/locate/apthermeng
Performance estimation of ejector cycles using heavier hydrocarbon refrigerants Jacek Kasperski*, Bartosz Gil Wrocław University of Technology, Institute of Power Engineering and Fluid Mechanics, Poland
h i g h l i g h t s Advantages of use of higher hydrocarbons as ejector refrigerants were presumed. Computer software basing on theoretical model of Huang et al. (1999) was prepared. Optimal temperature range of vapor generation for each hydrocarbon was calculated.
a r t i c l e i n f o
a b s t r a c t
Article history: Received 31 January 2014 Accepted 27 June 2014 Available online 4 July 2014
Computer software basing on theoretical model of Huang et al. with thermodynamic properties of hydrocarbons was prepared. Investigation was focused on nine hydrocarbons: propane, butane, iso-butane, pentane, iso-pentane, hexane, heptane and octane. A series of calculations was carried out for the generator temperature between 70 and 200 C, with assumed temperatures of evaporation 10 C and condensation 40 C. Calculation results show that none of the hydrocarbons enables high efficiency of a cycle in a wide range of temperature. Each hydrocarbon has its own maximal entrainment ratio at its individual temperature of optimum. Temperatures of entrainment ratios optimum increase according to the hydrocarbon heaviness with simultaneous increase of entrainment ratio peak values. Peak values of the COP do not increase according to the hydrocarbons heaviness. The highest COP ¼ 0.32 is achieved for iso-butane at 102 C and the COP ¼ 0.28 for pentane at 165 C. Heptane and octane can be ignored. © 2014 Elsevier Ltd. All rights reserved.
Keywords: Ejector Environmentally friendly refrigerants Entrainment ratio Simulation
1. Introduction Constant increase of the use of electricity to power refrigeration and air conditioning systems, as well as the struggle for environmental considerations contribute to looking for new green technologies in the field of refrigeration. An ejector system driven by thermal energy is an attractive solution (e.g. obtained from solar collectors), which does not have a compressor to force a working cycle. The key element of the device is an ejector consisting of a motive nozzle, suction nozzle, mixing chamber, and a diffuser. High-pressure and temperature motive stream expand in the motive nozzle (geyp process in Fig. 1), where its internal energy converts into kinetic energy. High speed of motive steam in the
* Corresponding author. E-mail address: [email protected] (J. Kasperski). http://dx.doi.org/10.1016/j.applthermaleng.2014.06.057 1359-4311/© 2014 Elsevier Ltd. All rights reserved.
motive nozzle entrains low pressure sucked steam into the suction chamber (eeys). Both streams enter the mixing chamber, in which the exchange of both momentum and kinetic and internal energies take place (yseyp), and become one stream with almost uniform pressure and speed. A re-conversion of kinetic energy back into internal energy occurs in the diffuser, where the pressure of the stream gradually increases to a pressure higher than the suction pressure of the ejector (c). Literature describes a number of different issues on ejector refrigeration systems. An unquestionable advantage of this solution is the lack of movable parts in the device (except a condensate pump which, however, may be replaced by a thermal pumping effect, presented by Huang et al. [1]), the ability to work with environmentally friendly refrigerants, working devoid of vibration and noise, and low investment costs. The disadvantage is a relatively low value of the COP, several times lower than in conventional compressor systems.
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Nomenclature A cp Cr COPth GWP h DHpump NBP ODP P Q RCL T
area (m2) specific heat (kJ kg1 K1) compression ratio coefficient of performance (¼Qe/Qg) Global Warming Potential specific enthalpy (kJ kg1) total pump head (m) normal boiling point ozone depletion potential pressure (kPa), heat rejected or supplied (W) refrigerant concentration limit (ppm v/v) temperature ( C)
h u
efficiency entrainment ratio
Subscripts 1…7 points of ejector and cycle analysis c condenser crit critical point d diffuser e evaporator g generator m mixing n nozzle N normal boiling s secondary or shock t throat y location of choking for the entrained flow
Greek symbols heat capacity ratio
g
2. Refrigerant selection High efficiency of the energy conversion in ejector system depends on a type of refrigerant. Over the last twenty years refrigerants applied were practically completely replaced. New generation of working fluids from the group of HCF was applied and the use of natural liquids such as CO2 (R744) and hydrocarbons was introduced. Based on observation, it was also discovered that the biggest increase of global temperatures since the mid-20th century has very likely been due to the increase of anthropogenic greenhouse gas concentrations [2]. Ban on the use of some refrigerants and a constant research to replace them have also influenced the research on ejector refrigerators. Detailed review of over a hundred international research papers in the last decade revealed data collected and presented by the authors in Fig. 2. A percentage share of refrigerants used or considered for use in ejector refrigeration is demonstrated. It is worth mentioning that the refrigerant R11, despite the fact that it is out of operation, occurs in comparative studies in historical context. Study conducted by Selvaraju and Mani [3] on the influence of operational parameters on the performance of the system with environmentally friendly refrigerants, i.e. R134a, R152a, R290, R600a and R717, revealed that a rising driving pressure ratio increases both the entrainment ratio and the COP. Among considered working fluids, the best performance was obtained for a system
using R134a. In the last years, ejector systems with flammable pure fluid have been more frequently considered as the replacement for the previous one. Roman and Hernandez [4] considered a theoretical behavior of an ejector cooling system using hydrocarbons e propane (R290), butane (R600) and iso-butane (R600a), and halocarbons e R123, R134a and R152a as refrigerants. The best results were obtained for the hydrocarbon R290, because of the maximum COP and u value, which indicates the lowest primary fluid mass flow rates. Dahmani et al. [5] also studied the performance of an ejector system using R134a, R152a, R290 and R600a as refrigerants. Once again, the best performance was observed for refrigerant R290 with a value of 0.725. The proper choice of a working fluid, which optimally matches discussed ejector cycle, allows for achieving high performance at low system charge and low refrigerant cost. Considering the use of flammable refrigerants, such as hydrocarbons or alcohols, it is necessary to draw attention to safety concerns. For environmental impact, all ozone-depleting fluids should be avoided. Furthermore, the Global Warming Potential (GWP) should be considered. A useful data regarding the above mentioned issues is presented in Table 1. Depending on the hydrocarbon heaviness, its operating pressure can be higher or lower than the atmospheric one. If the refrigerant operating pressure is higher than the atmospheric one (propane, butane), a little leak usually does not cause a problem of
Fig. 1. Ejector cycle in peh graph and configuration diagram of ejector refrigeration system.
J. Kasperski, B. Gil / Applied Thermal Engineering 71 (2014) 197e203
Fig. 2. Percentage share of refrigerants used (or considered for use) in ejector refrigeration, described in research papers over the last few years.
refrigerator inefficiency. Great refrigerant leakage creates a possibility of ignition or explosion, according to the exceeded concentration limit of refrigerant. If the refrigerant operating pressure is lower than the atmospheric one (pentanes, hexanes, heptane, octane), even a little amount of sucked atmospheric air can destabilize processes of boiling and condensation. In case of flammable mediums, such as hydrocarbons, sucking the air causes a significant danger of explosion. The authors are aware that the issue of using flammable refrigerants such as hydrocarbons deserves a serious investigation. To avoid risk of damage, refrigerators charged with hydrocarbons should be exploited outside of habitable floor area. That demand is easy to fulfill for air conditioning chillers. Secondly, if flammable refrigerants prove their advantages, safety systems will probably develop (for example the oxygen/nitrogen/moisture sensors could be used for refrigerant space control, fire blocking powders as refrigerant additives, decompression covers). The aim of the work presented is rather to look for advantages of hydrocarbons application than to investigate explosion risk and its prevention systems. 3. Ejector model and its properties First theoretical model of ejector performance was proposed by Keenen and Newman [9]. It was a one-dimensional model based on
Table 1 Thermodynamic and environmental properties of refrigerants considered. Fluid
Tcrit
Pcrit [8] NBP
kPa
4059 4212 4247 3796 3629
26.1 32.0 42.1 0.5 11.8
Safety classification RCL1) [6] ODP GWP ASHRAE [6] EN [7]
R134a R141b Propane (R290) Butane (R600) Iso-butane (R600a) Pentane (R601) Iso-pentane (R601a) Hexane (R602) Heptane (R603) Octane (R604)
C
101.1 204.4 96.7 152.0 134.7
196.6 3370 187.2 3378 234.7 3034 267.0 2736 296.2 2497
e
e
ppm v/v e
A1 A3 A3 A3
A1 A2 A3 A3 A3
50,000 2600 5300 1000 4000
36.1 27.8 A3
A3 A3
1000
C
68.7 98.4 125.6
e
0 1300 0.11 600 0 3 0 3 0 3 0 0
3 3
199
ideal gas dynamics and on principles of mass, momentum and energy maintenance. In this model friction and heat loses were ignored. A computer simulation model was developed by Sun and Eames [10] for an ejector refrigerating system which used R123 as a working fluid. This model was also based on the ideal gas theory, but allowed for internal irreversibility of processes within the ejector. Huang et al. [11] developed a concept of one-dimensional analysis of ejector performance at critical-mode operation. They assumed a constant-pressure mixing occurring inside a mixing section of the ejector and the choking of entrained flow. By combining two distinct models of a mixing section, good agreement with experimental results has been reached. This is by far the simplification of processes, which are in fact more complicated. Although this model does not take into account a number of issues, such as the scale effect of ejector on the velocity profile of working fluid, pressure losses due to friction at the media border (wall-fluid and fluidefluid) or phenomena taking place laterally to the axis of the nozzle, the present papers still refer to it, e.g. Refs. [12], due to simplicity of use in the form of calculation program and its relevance to preliminary analysis. Apart from theoretical models, several empirical and semiempirical models were created (Selvaraju and Mani [13], Yapici et al. [14], Jia and Wenjiam [15], Chen et al. [16], Kumar and Ooi [17]). According to Chen et al. [18], working fluids for an ejector refrigeration cycle can be classified in two groups: wet vapor and dry vapor. For the dry vapor fluid, there is no phase change occurring in the expansion process across the primary nozzle of the ejector. Therefore, it is possible to use the ideal gas dynamics principle to modeling the ejector. In the case of wet vapor, fluid will partially condense in the nozzle and small drops may be formed. The governing equations for the wet and dry vapor should be different, especially the energy equation, because of the condensation process. In this study, a 1-D model of ejector performance at the criticalmode operation, developed by Huang et al. [11], was applied for the following reasons: Simplicity of a numerical model and the ease to be applied as a tool for the estimation of entrainment ratio and COP; Superheated vapor region is considered for processes occurring within primary nozzle and mixing chamber; Ejector cooling systems with low capacity are considered. Primary nozzle and mixing chamber dimensions have been retained in the similar size. The schematic view of modeled processes taking place inside the ejector is shown in Fig. 3. After exhausting from the motive nozzle (state 1), the primary stream enters into the mixing section of the ejector (state 2). At the yey section both flows reach choking conditions and starts to mix with a uniform pressure. Next, the supersonic shock takes place in section ses accompanied by a sudden pressure increase. Assuming isentropic process undergoing after the shock, the mixed flow between section mem and the end of the mixing section (state 3) has a uniform pressure P3. In the case of applying a refrigerant classified in the dry vapor group, Huang's model assumes four levels of refrigerant pressure (see Fig. 1). The highest pressure is obtained in the vapor generator and the lowest in yey section of the ejector. No phase change occurs during both the expansion process g-2 of the primary flow along the primary nozzle and the expansion process 6-e inside the suction chamber. Moreover, the compression process mec takes place in the superheated vapor region. Computer software of ejector performance estimation was prepared by the authors. According to the 1-D ejector theory
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Fig. 3. Schematic view of the ejector sectors (redrawn after Huang et al. [11]).
described by Huang et al. [11], the following assumptions were made: Working fluid is semi-ideal gas with properties cp and g varying according to temperature; Flow inside the ejector is steady and one-dimensional; Kinetic energies at the secondary flow inlet, as well as the exit of the diffuser, are negligible; Isentropic relations are used for model simplicity, but efficiency coefficients are accounted to specify their non-ideality; Ejector's inner walls are adiabatic. Thermodynamic properties of working fluids were obtained from the REFPROP v9.0 package [8]. Assumed temperatures are: in the evaporator þ10 C, in the condenser þ40 C, in the generator between þ70 and 200 C (if available). Aiming to simplify comparisons, an assumption was made that a process occurring in evaporator is completed without vapor superheating. By analogy, it was assumed that a process in condenser is ended without liquid subcooling. These assumptions resulted from an aspiration to create universality of comparisons for many versions of evaporators and condensers applied. Transcritical cycle was not considered. In the risk of wet vapor while calculating the expansion process, an indispensable overheating in generator was assumed. Assumed temperature level is a result of the ejector system use for air conditioning purpose. Combustion chamber, solar system, electrical heating could be the heat source for a vapor generator. Considering a solar drive, several types of collectors can be applied according to the temperature demand [19]. Flat solar collectors are
suitable for temperature up to 100 C, evacuated tube collectors for a temperature range 100e150 C, and compound parabolic concentrating as well as parabolic-through concentrating collectors e for more than 150 C. A wide range of operational temperature available for the collectors was the reason of the range of generator temperature assumption. The diameter of the primary nozzle was established as equal to 1 mm. The rest of ejector's geometry was calculated according to model demands. The COP's value of the ejector system depends strongly on the efficiency of ejector's components. Unfortunately, there are very few papers that recognize this issue. In most cases researchers assume values of the ejector parts efficiency at 0.7e1.0 (see Table 2). In the present paper the efficiency coefficients were assumed as reported by Huang: 0.95 of primary nozzle, 0.85 of secondary nozzle, and 0.88 of primary flow arbitrary coefficient efficiency. The value of frictional loss coefficient has the most important influence on the model result. In the model, it was calculated as a relation of Am/At, according to a proposition described in the original paper. The main purpose of this paper is to predict ejector's performance in critical-mode operation for fixed ejector geometry and for various refrigerants. Nine working fluids from the group of hydrocarbons were selected for the presented analysis: propane
Table 2 Assumed isentropic efficiency of ejector particular components e selected works from the last decade. Authors
Refrigerant hn
[20] Alexis and Karayiannis (2005) [21] Yapici and Ersoy (2005) [22] Yu et al. (2006)
R141b
0.90
R134a
0.85
R134a R152a R141b R142b R123 R142b R600a R744
0.85
0.95
0.85
0.90 0.85 0.85 0.95
0.85 0.95 0.95 0.95
0.85 0.85 0.85 0.85
0.70
0.70
0.95
0.80
[23] Yu and Li (2007) [24] Yu et al. (2007) [25] Ersoy et al. (2007) [26] Boumaraf and Lallemand (2009) [27] Eskandari Manjili and Yavari (2012) [28] Cardemil and Colle (2012)
R141b R717 R744 [29] Grazzini et al. (2012) R245fa [30] Liu and Groll (2013) R744
hs
hm
hd
0.80 0.85
0.95 0.85 0.95 0.95 0.98 0.50e0.93 0.37e0.90
0.85
G(Am/At) 0.95 0.77 0.95 1.00 0.95 0.95 0.50e1.00
Fig. 4. Comparison of saturation curves of selected hydrocarbons collected in peh graph.
J. Kasperski, B. Gil / Applied Thermal Engineering 71 (2014) 197e203
201
Fig. 5. Compression ratio of various refrigerants. Fig. 6. Calculated value of entrainment ratio of the ejector versus generator temperature.
(R290), butane (R600), iso-butane (R600a), pentane (R601), isopentane (R601a), hexane (R602), heptane (R603), octane (R604). For hydrocarbons heavier than propane, vapor saturation curve bends to the right. This favors the expansion process towards the area of superheated vapor. The heavier a hydrocarbon is, the more sloped its saturation curve, as it is presented in Fig. 4. To show the comparison, a dashed line for R141b and a dotted line for R134a were used. 4. Results and discussion 4.1. Efficiency The values of ejector compression ratio Cr for selected refrigerants were shown in Fig. 5. Refrigerants commonly used in ejector cooling systems were marked with filled contour, while fluids analyzed as likely to cooperate with that system were marked with unfilled contour. It should be noticed that hydrocarbons cover the entire range of the X-axis. The heavier a hydrocarbon is, the higher the value of Cr, what in practice means that ejector has to overcome greater pressure differential. The value of Cr, however, does not exceed the value obtained for water, which is the main refrigerant used in ejector cooling system. In addition, points defining both pentanes have almost the same Cr as refrigerants R141b and R123a, which are very popular synthetic refrigerants. The values of Cr for considered hydrocarbons and for assumed temperatures level were summarized in Table 3. Calculated values of entrainment ratio of the ejector versus generator temperature are presented in Fig. 6. There is none hydrocarbon enabling a high value of entrainment ratio in a wide range of generator temperature. For the temperature of the
generator at range 65e115 C, the highest value of entrainment ratio is obtained for iso-butane, at range 115e130 C for butane, at range 130e160 C for iso-pentane, at range 160e175 C for pentane, and iso-hexane for over 175 C. For heavier hydrocarbons, curves bend to the right towards the increasing temperature of the generator. Generally, peak values of the entrainment ratio grow up according to the generator temperature increase for the propane/ butane(s)/pentane(s)/hexane(s) set. A relatively low value of u for propane is the result of overheating demand at the beginning of the expansion process. In the range of 65e90 C, the overheating parameters 5e10 C were accordingly needed. The investigation of the calculated model parameters showed the maximum of entrainment ratio according to the maximum speed of primary flow in section yey of the ejector. A slightly sloped curve visible to the left side of pentane is a result of a calculation formula of frictional loss coefficient. Calculated COP of the ejector cycle versus its generator temperature is presented in Fig. 7. The same temperatures assignment for hydrocarbons is observed, as it was described for the entrainment ratio results. Unexpectedly, the COP peaks do not increase according to the propane/butane(s)/pentane(s)/hexane(s) set as it was visible for the entrainment ratio. The investigation reveals that the slope and constriction of saturation lines are the main reason. A heavier hydrocarbon is featured by bigger slopes of lines and a decreasing phase change heat.
Table 3 Compression ratio of considered refrigerants. Fluid
propane (R290) Iso-butane (R600a) n-Butane (R600) Iso-pentane (R601a) n-Pentane (R601) Iso-hexane (R602a) n-Hexane (R602) Heptane (R603) Octane (R604)
Condensing pressure at 40 C
Evaporating pressure at 10 C
Cr
kPa
kPa
e
1370 532 379 152 116 51.1 37.5 12.4 4.2
637 221 149 52.4 37.9 14.6 10.1 2.8 0.8
2.15 2.41 2.55 2.90 3.07 3.49 3.71 4.51 5.51
Fig. 7. Calculated performance of the ejector cycle versus generator temperature.
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The maximum COP ¼ 0.32 is observed for iso-butane at temperature 102 C in the generator. All other hydrocarbons present worse COP. For butane, the maximum COP ¼ 0.28 is observed at temperature 95 C, for iso-pentane 0.26 at 145 C, for pentane 0.27 at temperature 165 C, and for iso-hexane 0.26 at 190 C. Heptane and octane use can be eliminated because of a low value of entrainment ratio and the COP. For most of refrigerants the maximum value of COP and u is observed, and occurs over a dozen Celsius degrees below the temperature of critical point. The occurrence of this maximum of function is caused by the rapid growth of cp and g, which leads to decrease in ejector efficiency. In other words, approaching the critical point causes an increase of vapor density and impedes the realization of the processes involved. Comparing hydrocarbons with halogenated refrigerants R134a and R141b showed that value of both u and COPth achieved for R134a and R141b were similar to propane. Obtained results showed that in the temperature tg < 100 C using butane and iso-butane instead of synthetic refrigerants R134a and R141b as a working fluid can bring measurable benefits. 4.2. Suitability for gravitational and thermal pumping Depending on the way, in which refrigerant returns from condenser to generator, there are two main configurations: pump, gravitational and/or thermal pumping. In a typical pump version of ejector, refrigerator heat exchangers can be placed on different levels. The flow of liquid refrigerant from condenser to steam generator is forced by pump. An alternative for pump ejector's refrigerator is gravitational and/or thermal pumping. Vertical arrangement of heat exchangers on different levels allows equalizing the pressure differences between exchangers with help of refrigerant hydrostatic pressure. The highest pressure is obtained in steam generator, which forces the lowest liquid level. The lowest pressure obtained in evaporator causes the inflow of liquid to the highest installation level. The concept of gravitational feeding of generator was presented by Nguyen et al. [31]. The use of thermal pumping effect with additional feeding vessel was proposed by Huang et al. [1]. The concept of multi-function generator using thermal pumping effect was presented by Srisastra and Aphornratana [32] and Srisastra et al. [33]. The application of rotary motion can replace hydrostatic pressure by more effective difference of radius of liquids, taking place inside roto-gravitational cycle, presented by Kasperski [34]. The coaxial arrangement of heat exchangers on different radius allows setting pressure difference of vapors. Calculated indispensable difference of levels for gravitational cycle versus its generator temperature is presented in Fig. 8. The same height is visible for R134a as well as for n-butane at temperature range 70e110 C, and also for R141b as well as for nhexane at temperature range 70e190 C. The same height is visible for both water and heptane at temperature range 130e180 C. Considering the use of hydrocarbons instead of synthetic refrigerants, only iso-pentane, heptane, and octane can minimize the height of gravitational ejector installation. 5. Conclusions In this paper, a theoretical simulation of ejector performance with nine heavier hydrocarbons as refrigerants is presented. The results achieved in the application of the 1-D computational model have been discussed. The following conclusions have been drawn: 1. There is no hydrocarbon enabling a high value of entrainment ratio in a wide range of generator temperature. Each
Fig. 8. Indispensable difference of levels for gravitational cycle versus generator temperature.
2.
3.
4.
5.
6.
7.
hydrocarbon has its own maximum entrainment ratio at its individual temperature of the optimum. Optimal temperature of vapor generation and maximum values of the entrainment ratio increase according to the hydrocarbon heaviness. According to the temperature of vapor generation, the best results are achieved by iso-butane at the temperature range 65e115 C, butane at the range 115e130 C, iso-pentane at the range 130e160 C, pentane at the range 160e175 C, iso-hexane at over 175 C. The highest values of entrainment ratio for specific fluids are as follows: u ¼ 0.45 for iso-butane at the temperature of 102 C, u ¼ 0.47 for iso-pentane at 100 C, u ¼ 0.5 for pentane at 165 C, u ¼ 0.56 for iso-hexane at 195 C. Peak values of the COP do not increase according to the hydrocarbon heaviness. The highest COP ¼ 0.32 is achieved for isobutane at the temperature of 102 C. Heptane and octane are characterized by the low COP value at the temperature range 65e200 C, thus can be eliminated from further consideration. Considering the use of hydrocarbons instead of synthetic refrigerants, only iso-pentane, heptane and octane can minimize the height of gravitational ejector installation.
Acknowledgements Project of Wroclaw University of Technology number MK/SN/ 490/IX/2013/U co-financed by the European Union as part of the European Social Fund. References [1] B.J. Huang, S.S. Hu, S.H. Lee, Development of an ejector cooling system with thermal pumping effect, Int. J. Refrig. 29 (2006) 476e484. [2] International Panel of Climate Change (ICCP), 2007. [3] A. Selvaraju, A. Mani, Analysis of a vapour ejector refrigeration system with environment friendly refrigerants, Int. J. Therm. Sci. 43 (2004) 915e921. [4] R. Roman, J.I. Hernandez, Performance of ejector cooling systems using low ecological impact refrigerants, Int. J. Refrig. 34 (2011) 1707e1716. [5] A. Dahmani, Z. Aidoun, N. Galanis, Optimum design of ejector refrigeration systems with environmentally benign fluids, Int. J. Therm. Sci. 50 (2011) 1562e1572. [6] ASHRAE Standards 34-2004: Designation and Safety Classification of Refrigerants. [7] EN 378-1:2008: Refrigerating Systems and Heat Pumps. Safety and Environmental Requirements. Basic Requirements, Definitions, Classification and Selection Criteria. [8] REFPROP-Reference Fluid Thermodynamic and Transport Properties, Version 9.0, 2010.
J. Kasperski, B. Gil / Applied Thermal Engineering 71 (2014) 197e203 [9] J.H. Keenan, E.P. Neumann, A simple air ejector, J. Appl. Mech. Trans. ASME 64 (1942) 75e81. [10] D.W. Sun, I.W. Eames, Performance characteristics of HCFC-123 ejector refrigeration cycles, Int. J. Energ. Res. 20 (1996) 871e885. [11] B.J. Huang, J.M. Chang, C.P. Wang, V.A. Petrenko, A 1-D analysis of ejector performance, Int. J. Refrig. 22 (1999) 354e364. [12] J. Chen, H. Havtun, B. Palm, Parametric analysis of ejector working characteristics in the refrigeration system, Appl. Therm. Eng. 69 (2014) 130e142. [13] A. Selvaraju, A. Mani, Experimental investigation on R134a vapour ejector refrigeration system, Int. J. Refrig. 29 (2006) 1160e1166. [14] R. Yapici, H.K. Ersoy, A. Aktoprakoǧlu, H.S. Halkaci, O. Yiǧit, Experimental determination of the optimum performance of ejector refrigeration system depending on ejector area ratio, Int. J. Refrig. 31 (2008) 1183e1189. [15] Y. Jia, C. Wenjian, Area ratio effects to the performance of air-cooled ejector refrigeration cycle with R134a refrigerant, Energ. Convers. Manage. 53 (2012) 240e246. [16] J. Chen, H. Havtun, B. Palm, Investigation of ejectors in refrigeration systems: optimum performance evaluation and ejector area ratio perspectives, Appl. Therm. Eng. 64 (2014) 182e191. [17] N.S. Kumar, K.T. Ooi, One dimensional model of an ejector with special attention to fanno flow within the mixing chamber, Appl. Therm. Eng. 65 (2014) 226e235. [18] S.L. Chen, J.Y. Yen, M.C. Huang, An experimental investigation of ejector performance based upon different refrigerants, ASHRAE Trans. 104 (1998) 153e160. [19] W. Pridasawas, Solar-driven Refrigeration Systems with Focus on the Ejector Cycle (Doctoral thesis), Royal Institute of Technology, KTH, Stockholm, 2006. [20] G.K. Alexis, E.K. Karayiannis, A solar ejector cooling system using refrigerant R134a in the Athens area, Renew. Energ. 30 (2005) 1457e1469. [21] R. Yapici, H.K. Ersoy, Performance characteristics of the ejector refrigeration system based on the constant area ejector flow model, Energ. Convers. Manage. 46 (2005) 3117e3135.
203
[22] J. Yu, H. Chen, Y. Ren, Y. Li, A new ejector refrigeration system with an additional jet pump, Appl. Therm. Eng. 26 (2006) 312e319. [23] J. Yu, Y. Li, A theoretical study of a novel regenerative ejector refrigeration cycle, Int. J. Refrig. 30 (2007) 464e470. [24] J. Yu, Y. Ren, H. Chen, Y. Li, Applying mechanical subcooling to ejector refrigeration cycle for improving the coefficient of performance, Energ. Convers. Manage. 48 (2007) 1193e1199. [25] H.K. Ersoy, S. Yalcin, R. Yapici, M. Ozgoren, Performance of a solar ejector cooling-system in the southern region of Turkey, Appl. Energ. 84 (2007) 971e983. [26] L. Boumaraf, A. Lallemand, Modeling of an ejector refrigerating system operating in dimensioning and off-dimensioning conditions with the working fluids R142b and R600a, Appl. Therm. Eng. 29 (2009) 265e274. [27] F. Eskandari Manjili, M.A. Yavari, Performance of a new two-stage multiintercooling transcritical CO2 ejector refrigeration cycle, Appl. Therm. Eng. 40 (2012) 202e209. [28] J.M. Cardemil, S. Colle, A general model for evaluation of vapor ejectors performance for application in refrigeration, Energ. Convers. Manage. 64 (2012) 79e86. [29] G. Grazzini, A. Milazzo, D. Paganini, Design of an ejector cycle refrigeration system, Energ. Convers. Manage. 54 (2012) 38e46. [30] F. Liu, E.A. Groll, Study of ejector efficiencies in refrigeration cycles, Appl. Therm. Eng. 52 (2013) 360e370. [31] V.M. Nguyen, S.B. Riffat, P.S. Doherty, Development of a solar-powered passive ejector cooling system, Appl. Therm. Eng. 21 (2001) 157e168. [32] P. Srisastra, S. Aphornratana, A circulating system for a steam jet refrigeration system, Appl. Therm. Eng. 25 (2005) 2247e2257. [33] P. Srisastra, S. Aphornratana, T. Sriveerakul, Development of a circulating system for a jet refrigeration cycle, Int. J. Refrig. 31 (2008) 921e929. [34] J. Kasperski, Rotational type of a gravitational ejector refrigerator e a system balance of the refrigerant analysis, Int. J. Refrig. 33 (2010) 3e11.
Contents lists available at ScienceDirect
Applied Thermal Engineering journal homepage: www.elsevier.com/locate/apthermeng
Performance estimation of ejector cycles using heavier hydrocarbon refrigerants Jacek Kasperski*, Bartosz Gil Wrocław University of Technology, Institute of Power Engineering and Fluid Mechanics, Poland
h i g h l i g h t s Advantages of use of higher hydrocarbons as ejector refrigerants were presumed. Computer software basing on theoretical model of Huang et al. (1999) was prepared. Optimal temperature range of vapor generation for each hydrocarbon was calculated.
a r t i c l e i n f o
a b s t r a c t
Article history: Received 31 January 2014 Accepted 27 June 2014 Available online 4 July 2014
Computer software basing on theoretical model of Huang et al. with thermodynamic properties of hydrocarbons was prepared. Investigation was focused on nine hydrocarbons: propane, butane, iso-butane, pentane, iso-pentane, hexane, heptane and octane. A series of calculations was carried out for the generator temperature between 70 and 200 C, with assumed temperatures of evaporation 10 C and condensation 40 C. Calculation results show that none of the hydrocarbons enables high efficiency of a cycle in a wide range of temperature. Each hydrocarbon has its own maximal entrainment ratio at its individual temperature of optimum. Temperatures of entrainment ratios optimum increase according to the hydrocarbon heaviness with simultaneous increase of entrainment ratio peak values. Peak values of the COP do not increase according to the hydrocarbons heaviness. The highest COP ¼ 0.32 is achieved for iso-butane at 102 C and the COP ¼ 0.28 for pentane at 165 C. Heptane and octane can be ignored. © 2014 Elsevier Ltd. All rights reserved.
Keywords: Ejector Environmentally friendly refrigerants Entrainment ratio Simulation
1. Introduction Constant increase of the use of electricity to power refrigeration and air conditioning systems, as well as the struggle for environmental considerations contribute to looking for new green technologies in the field of refrigeration. An ejector system driven by thermal energy is an attractive solution (e.g. obtained from solar collectors), which does not have a compressor to force a working cycle. The key element of the device is an ejector consisting of a motive nozzle, suction nozzle, mixing chamber, and a diffuser. High-pressure and temperature motive stream expand in the motive nozzle (geyp process in Fig. 1), where its internal energy converts into kinetic energy. High speed of motive steam in the
* Corresponding author. E-mail address: [email protected] (J. Kasperski). http://dx.doi.org/10.1016/j.applthermaleng.2014.06.057 1359-4311/© 2014 Elsevier Ltd. All rights reserved.
motive nozzle entrains low pressure sucked steam into the suction chamber (eeys). Both streams enter the mixing chamber, in which the exchange of both momentum and kinetic and internal energies take place (yseyp), and become one stream with almost uniform pressure and speed. A re-conversion of kinetic energy back into internal energy occurs in the diffuser, where the pressure of the stream gradually increases to a pressure higher than the suction pressure of the ejector (c). Literature describes a number of different issues on ejector refrigeration systems. An unquestionable advantage of this solution is the lack of movable parts in the device (except a condensate pump which, however, may be replaced by a thermal pumping effect, presented by Huang et al. [1]), the ability to work with environmentally friendly refrigerants, working devoid of vibration and noise, and low investment costs. The disadvantage is a relatively low value of the COP, several times lower than in conventional compressor systems.
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Nomenclature A cp Cr COPth GWP h DHpump NBP ODP P Q RCL T
area (m2) specific heat (kJ kg1 K1) compression ratio coefficient of performance (¼Qe/Qg) Global Warming Potential specific enthalpy (kJ kg1) total pump head (m) normal boiling point ozone depletion potential pressure (kPa), heat rejected or supplied (W) refrigerant concentration limit (ppm v/v) temperature ( C)
h u
efficiency entrainment ratio
Subscripts 1…7 points of ejector and cycle analysis c condenser crit critical point d diffuser e evaporator g generator m mixing n nozzle N normal boiling s secondary or shock t throat y location of choking for the entrained flow
Greek symbols heat capacity ratio
g
2. Refrigerant selection High efficiency of the energy conversion in ejector system depends on a type of refrigerant. Over the last twenty years refrigerants applied were practically completely replaced. New generation of working fluids from the group of HCF was applied and the use of natural liquids such as CO2 (R744) and hydrocarbons was introduced. Based on observation, it was also discovered that the biggest increase of global temperatures since the mid-20th century has very likely been due to the increase of anthropogenic greenhouse gas concentrations [2]. Ban on the use of some refrigerants and a constant research to replace them have also influenced the research on ejector refrigerators. Detailed review of over a hundred international research papers in the last decade revealed data collected and presented by the authors in Fig. 2. A percentage share of refrigerants used or considered for use in ejector refrigeration is demonstrated. It is worth mentioning that the refrigerant R11, despite the fact that it is out of operation, occurs in comparative studies in historical context. Study conducted by Selvaraju and Mani [3] on the influence of operational parameters on the performance of the system with environmentally friendly refrigerants, i.e. R134a, R152a, R290, R600a and R717, revealed that a rising driving pressure ratio increases both the entrainment ratio and the COP. Among considered working fluids, the best performance was obtained for a system
using R134a. In the last years, ejector systems with flammable pure fluid have been more frequently considered as the replacement for the previous one. Roman and Hernandez [4] considered a theoretical behavior of an ejector cooling system using hydrocarbons e propane (R290), butane (R600) and iso-butane (R600a), and halocarbons e R123, R134a and R152a as refrigerants. The best results were obtained for the hydrocarbon R290, because of the maximum COP and u value, which indicates the lowest primary fluid mass flow rates. Dahmani et al. [5] also studied the performance of an ejector system using R134a, R152a, R290 and R600a as refrigerants. Once again, the best performance was observed for refrigerant R290 with a value of 0.725. The proper choice of a working fluid, which optimally matches discussed ejector cycle, allows for achieving high performance at low system charge and low refrigerant cost. Considering the use of flammable refrigerants, such as hydrocarbons or alcohols, it is necessary to draw attention to safety concerns. For environmental impact, all ozone-depleting fluids should be avoided. Furthermore, the Global Warming Potential (GWP) should be considered. A useful data regarding the above mentioned issues is presented in Table 1. Depending on the hydrocarbon heaviness, its operating pressure can be higher or lower than the atmospheric one. If the refrigerant operating pressure is higher than the atmospheric one (propane, butane), a little leak usually does not cause a problem of
Fig. 1. Ejector cycle in peh graph and configuration diagram of ejector refrigeration system.
J. Kasperski, B. Gil / Applied Thermal Engineering 71 (2014) 197e203
Fig. 2. Percentage share of refrigerants used (or considered for use) in ejector refrigeration, described in research papers over the last few years.
refrigerator inefficiency. Great refrigerant leakage creates a possibility of ignition or explosion, according to the exceeded concentration limit of refrigerant. If the refrigerant operating pressure is lower than the atmospheric one (pentanes, hexanes, heptane, octane), even a little amount of sucked atmospheric air can destabilize processes of boiling and condensation. In case of flammable mediums, such as hydrocarbons, sucking the air causes a significant danger of explosion. The authors are aware that the issue of using flammable refrigerants such as hydrocarbons deserves a serious investigation. To avoid risk of damage, refrigerators charged with hydrocarbons should be exploited outside of habitable floor area. That demand is easy to fulfill for air conditioning chillers. Secondly, if flammable refrigerants prove their advantages, safety systems will probably develop (for example the oxygen/nitrogen/moisture sensors could be used for refrigerant space control, fire blocking powders as refrigerant additives, decompression covers). The aim of the work presented is rather to look for advantages of hydrocarbons application than to investigate explosion risk and its prevention systems. 3. Ejector model and its properties First theoretical model of ejector performance was proposed by Keenen and Newman [9]. It was a one-dimensional model based on
Table 1 Thermodynamic and environmental properties of refrigerants considered. Fluid
Tcrit
Pcrit [8] NBP
kPa
4059 4212 4247 3796 3629
26.1 32.0 42.1 0.5 11.8
Safety classification RCL1) [6] ODP GWP ASHRAE [6] EN [7]
R134a R141b Propane (R290) Butane (R600) Iso-butane (R600a) Pentane (R601) Iso-pentane (R601a) Hexane (R602) Heptane (R603) Octane (R604)
C
101.1 204.4 96.7 152.0 134.7
196.6 3370 187.2 3378 234.7 3034 267.0 2736 296.2 2497
e
e
ppm v/v e
A1 A3 A3 A3
A1 A2 A3 A3 A3
50,000 2600 5300 1000 4000
36.1 27.8 A3
A3 A3
1000
C
68.7 98.4 125.6
e
0 1300 0.11 600 0 3 0 3 0 3 0 0
3 3
199
ideal gas dynamics and on principles of mass, momentum and energy maintenance. In this model friction and heat loses were ignored. A computer simulation model was developed by Sun and Eames [10] for an ejector refrigerating system which used R123 as a working fluid. This model was also based on the ideal gas theory, but allowed for internal irreversibility of processes within the ejector. Huang et al. [11] developed a concept of one-dimensional analysis of ejector performance at critical-mode operation. They assumed a constant-pressure mixing occurring inside a mixing section of the ejector and the choking of entrained flow. By combining two distinct models of a mixing section, good agreement with experimental results has been reached. This is by far the simplification of processes, which are in fact more complicated. Although this model does not take into account a number of issues, such as the scale effect of ejector on the velocity profile of working fluid, pressure losses due to friction at the media border (wall-fluid and fluidefluid) or phenomena taking place laterally to the axis of the nozzle, the present papers still refer to it, e.g. Refs. [12], due to simplicity of use in the form of calculation program and its relevance to preliminary analysis. Apart from theoretical models, several empirical and semiempirical models were created (Selvaraju and Mani [13], Yapici et al. [14], Jia and Wenjiam [15], Chen et al. [16], Kumar and Ooi [17]). According to Chen et al. [18], working fluids for an ejector refrigeration cycle can be classified in two groups: wet vapor and dry vapor. For the dry vapor fluid, there is no phase change occurring in the expansion process across the primary nozzle of the ejector. Therefore, it is possible to use the ideal gas dynamics principle to modeling the ejector. In the case of wet vapor, fluid will partially condense in the nozzle and small drops may be formed. The governing equations for the wet and dry vapor should be different, especially the energy equation, because of the condensation process. In this study, a 1-D model of ejector performance at the criticalmode operation, developed by Huang et al. [11], was applied for the following reasons: Simplicity of a numerical model and the ease to be applied as a tool for the estimation of entrainment ratio and COP; Superheated vapor region is considered for processes occurring within primary nozzle and mixing chamber; Ejector cooling systems with low capacity are considered. Primary nozzle and mixing chamber dimensions have been retained in the similar size. The schematic view of modeled processes taking place inside the ejector is shown in Fig. 3. After exhausting from the motive nozzle (state 1), the primary stream enters into the mixing section of the ejector (state 2). At the yey section both flows reach choking conditions and starts to mix with a uniform pressure. Next, the supersonic shock takes place in section ses accompanied by a sudden pressure increase. Assuming isentropic process undergoing after the shock, the mixed flow between section mem and the end of the mixing section (state 3) has a uniform pressure P3. In the case of applying a refrigerant classified in the dry vapor group, Huang's model assumes four levels of refrigerant pressure (see Fig. 1). The highest pressure is obtained in the vapor generator and the lowest in yey section of the ejector. No phase change occurs during both the expansion process g-2 of the primary flow along the primary nozzle and the expansion process 6-e inside the suction chamber. Moreover, the compression process mec takes place in the superheated vapor region. Computer software of ejector performance estimation was prepared by the authors. According to the 1-D ejector theory
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Fig. 3. Schematic view of the ejector sectors (redrawn after Huang et al. [11]).
described by Huang et al. [11], the following assumptions were made: Working fluid is semi-ideal gas with properties cp and g varying according to temperature; Flow inside the ejector is steady and one-dimensional; Kinetic energies at the secondary flow inlet, as well as the exit of the diffuser, are negligible; Isentropic relations are used for model simplicity, but efficiency coefficients are accounted to specify their non-ideality; Ejector's inner walls are adiabatic. Thermodynamic properties of working fluids were obtained from the REFPROP v9.0 package [8]. Assumed temperatures are: in the evaporator þ10 C, in the condenser þ40 C, in the generator between þ70 and 200 C (if available). Aiming to simplify comparisons, an assumption was made that a process occurring in evaporator is completed without vapor superheating. By analogy, it was assumed that a process in condenser is ended without liquid subcooling. These assumptions resulted from an aspiration to create universality of comparisons for many versions of evaporators and condensers applied. Transcritical cycle was not considered. In the risk of wet vapor while calculating the expansion process, an indispensable overheating in generator was assumed. Assumed temperature level is a result of the ejector system use for air conditioning purpose. Combustion chamber, solar system, electrical heating could be the heat source for a vapor generator. Considering a solar drive, several types of collectors can be applied according to the temperature demand [19]. Flat solar collectors are
suitable for temperature up to 100 C, evacuated tube collectors for a temperature range 100e150 C, and compound parabolic concentrating as well as parabolic-through concentrating collectors e for more than 150 C. A wide range of operational temperature available for the collectors was the reason of the range of generator temperature assumption. The diameter of the primary nozzle was established as equal to 1 mm. The rest of ejector's geometry was calculated according to model demands. The COP's value of the ejector system depends strongly on the efficiency of ejector's components. Unfortunately, there are very few papers that recognize this issue. In most cases researchers assume values of the ejector parts efficiency at 0.7e1.0 (see Table 2). In the present paper the efficiency coefficients were assumed as reported by Huang: 0.95 of primary nozzle, 0.85 of secondary nozzle, and 0.88 of primary flow arbitrary coefficient efficiency. The value of frictional loss coefficient has the most important influence on the model result. In the model, it was calculated as a relation of Am/At, according to a proposition described in the original paper. The main purpose of this paper is to predict ejector's performance in critical-mode operation for fixed ejector geometry and for various refrigerants. Nine working fluids from the group of hydrocarbons were selected for the presented analysis: propane
Table 2 Assumed isentropic efficiency of ejector particular components e selected works from the last decade. Authors
Refrigerant hn
[20] Alexis and Karayiannis (2005) [21] Yapici and Ersoy (2005) [22] Yu et al. (2006)
R141b
0.90
R134a
0.85
R134a R152a R141b R142b R123 R142b R600a R744
0.85
0.95
0.85
0.90 0.85 0.85 0.95
0.85 0.95 0.95 0.95
0.85 0.85 0.85 0.85
0.70
0.70
0.95
0.80
[23] Yu and Li (2007) [24] Yu et al. (2007) [25] Ersoy et al. (2007) [26] Boumaraf and Lallemand (2009) [27] Eskandari Manjili and Yavari (2012) [28] Cardemil and Colle (2012)
R141b R717 R744 [29] Grazzini et al. (2012) R245fa [30] Liu and Groll (2013) R744
hs
hm
hd
0.80 0.85
0.95 0.85 0.95 0.95 0.98 0.50e0.93 0.37e0.90
0.85
G(Am/At) 0.95 0.77 0.95 1.00 0.95 0.95 0.50e1.00
Fig. 4. Comparison of saturation curves of selected hydrocarbons collected in peh graph.
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201
Fig. 5. Compression ratio of various refrigerants. Fig. 6. Calculated value of entrainment ratio of the ejector versus generator temperature.
(R290), butane (R600), iso-butane (R600a), pentane (R601), isopentane (R601a), hexane (R602), heptane (R603), octane (R604). For hydrocarbons heavier than propane, vapor saturation curve bends to the right. This favors the expansion process towards the area of superheated vapor. The heavier a hydrocarbon is, the more sloped its saturation curve, as it is presented in Fig. 4. To show the comparison, a dashed line for R141b and a dotted line for R134a were used. 4. Results and discussion 4.1. Efficiency The values of ejector compression ratio Cr for selected refrigerants were shown in Fig. 5. Refrigerants commonly used in ejector cooling systems were marked with filled contour, while fluids analyzed as likely to cooperate with that system were marked with unfilled contour. It should be noticed that hydrocarbons cover the entire range of the X-axis. The heavier a hydrocarbon is, the higher the value of Cr, what in practice means that ejector has to overcome greater pressure differential. The value of Cr, however, does not exceed the value obtained for water, which is the main refrigerant used in ejector cooling system. In addition, points defining both pentanes have almost the same Cr as refrigerants R141b and R123a, which are very popular synthetic refrigerants. The values of Cr for considered hydrocarbons and for assumed temperatures level were summarized in Table 3. Calculated values of entrainment ratio of the ejector versus generator temperature are presented in Fig. 6. There is none hydrocarbon enabling a high value of entrainment ratio in a wide range of generator temperature. For the temperature of the
generator at range 65e115 C, the highest value of entrainment ratio is obtained for iso-butane, at range 115e130 C for butane, at range 130e160 C for iso-pentane, at range 160e175 C for pentane, and iso-hexane for over 175 C. For heavier hydrocarbons, curves bend to the right towards the increasing temperature of the generator. Generally, peak values of the entrainment ratio grow up according to the generator temperature increase for the propane/ butane(s)/pentane(s)/hexane(s) set. A relatively low value of u for propane is the result of overheating demand at the beginning of the expansion process. In the range of 65e90 C, the overheating parameters 5e10 C were accordingly needed. The investigation of the calculated model parameters showed the maximum of entrainment ratio according to the maximum speed of primary flow in section yey of the ejector. A slightly sloped curve visible to the left side of pentane is a result of a calculation formula of frictional loss coefficient. Calculated COP of the ejector cycle versus its generator temperature is presented in Fig. 7. The same temperatures assignment for hydrocarbons is observed, as it was described for the entrainment ratio results. Unexpectedly, the COP peaks do not increase according to the propane/butane(s)/pentane(s)/hexane(s) set as it was visible for the entrainment ratio. The investigation reveals that the slope and constriction of saturation lines are the main reason. A heavier hydrocarbon is featured by bigger slopes of lines and a decreasing phase change heat.
Table 3 Compression ratio of considered refrigerants. Fluid
propane (R290) Iso-butane (R600a) n-Butane (R600) Iso-pentane (R601a) n-Pentane (R601) Iso-hexane (R602a) n-Hexane (R602) Heptane (R603) Octane (R604)
Condensing pressure at 40 C
Evaporating pressure at 10 C
Cr
kPa
kPa
e
1370 532 379 152 116 51.1 37.5 12.4 4.2
637 221 149 52.4 37.9 14.6 10.1 2.8 0.8
2.15 2.41 2.55 2.90 3.07 3.49 3.71 4.51 5.51
Fig. 7. Calculated performance of the ejector cycle versus generator temperature.
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The maximum COP ¼ 0.32 is observed for iso-butane at temperature 102 C in the generator. All other hydrocarbons present worse COP. For butane, the maximum COP ¼ 0.28 is observed at temperature 95 C, for iso-pentane 0.26 at 145 C, for pentane 0.27 at temperature 165 C, and for iso-hexane 0.26 at 190 C. Heptane and octane use can be eliminated because of a low value of entrainment ratio and the COP. For most of refrigerants the maximum value of COP and u is observed, and occurs over a dozen Celsius degrees below the temperature of critical point. The occurrence of this maximum of function is caused by the rapid growth of cp and g, which leads to decrease in ejector efficiency. In other words, approaching the critical point causes an increase of vapor density and impedes the realization of the processes involved. Comparing hydrocarbons with halogenated refrigerants R134a and R141b showed that value of both u and COPth achieved for R134a and R141b were similar to propane. Obtained results showed that in the temperature tg < 100 C using butane and iso-butane instead of synthetic refrigerants R134a and R141b as a working fluid can bring measurable benefits. 4.2. Suitability for gravitational and thermal pumping Depending on the way, in which refrigerant returns from condenser to generator, there are two main configurations: pump, gravitational and/or thermal pumping. In a typical pump version of ejector, refrigerator heat exchangers can be placed on different levels. The flow of liquid refrigerant from condenser to steam generator is forced by pump. An alternative for pump ejector's refrigerator is gravitational and/or thermal pumping. Vertical arrangement of heat exchangers on different levels allows equalizing the pressure differences between exchangers with help of refrigerant hydrostatic pressure. The highest pressure is obtained in steam generator, which forces the lowest liquid level. The lowest pressure obtained in evaporator causes the inflow of liquid to the highest installation level. The concept of gravitational feeding of generator was presented by Nguyen et al. [31]. The use of thermal pumping effect with additional feeding vessel was proposed by Huang et al. [1]. The concept of multi-function generator using thermal pumping effect was presented by Srisastra and Aphornratana [32] and Srisastra et al. [33]. The application of rotary motion can replace hydrostatic pressure by more effective difference of radius of liquids, taking place inside roto-gravitational cycle, presented by Kasperski [34]. The coaxial arrangement of heat exchangers on different radius allows setting pressure difference of vapors. Calculated indispensable difference of levels for gravitational cycle versus its generator temperature is presented in Fig. 8. The same height is visible for R134a as well as for n-butane at temperature range 70e110 C, and also for R141b as well as for nhexane at temperature range 70e190 C. The same height is visible for both water and heptane at temperature range 130e180 C. Considering the use of hydrocarbons instead of synthetic refrigerants, only iso-pentane, heptane, and octane can minimize the height of gravitational ejector installation. 5. Conclusions In this paper, a theoretical simulation of ejector performance with nine heavier hydrocarbons as refrigerants is presented. The results achieved in the application of the 1-D computational model have been discussed. The following conclusions have been drawn: 1. There is no hydrocarbon enabling a high value of entrainment ratio in a wide range of generator temperature. Each
Fig. 8. Indispensable difference of levels for gravitational cycle versus generator temperature.
2.
3.
4.
5.
6.
7.
hydrocarbon has its own maximum entrainment ratio at its individual temperature of the optimum. Optimal temperature of vapor generation and maximum values of the entrainment ratio increase according to the hydrocarbon heaviness. According to the temperature of vapor generation, the best results are achieved by iso-butane at the temperature range 65e115 C, butane at the range 115e130 C, iso-pentane at the range 130e160 C, pentane at the range 160e175 C, iso-hexane at over 175 C. The highest values of entrainment ratio for specific fluids are as follows: u ¼ 0.45 for iso-butane at the temperature of 102 C, u ¼ 0.47 for iso-pentane at 100 C, u ¼ 0.5 for pentane at 165 C, u ¼ 0.56 for iso-hexane at 195 C. Peak values of the COP do not increase according to the hydrocarbon heaviness. The highest COP ¼ 0.32 is achieved for isobutane at the temperature of 102 C. Heptane and octane are characterized by the low COP value at the temperature range 65e200 C, thus can be eliminated from further consideration. Considering the use of hydrocarbons instead of synthetic refrigerants, only iso-pentane, heptane and octane can minimize the height of gravitational ejector installation.
Acknowledgements Project of Wroclaw University of Technology number MK/SN/ 490/IX/2013/U co-financed by the European Union as part of the European Social Fund. References [1] B.J. Huang, S.S. Hu, S.H. Lee, Development of an ejector cooling system with thermal pumping effect, Int. J. Refrig. 29 (2006) 476e484. [2] International Panel of Climate Change (ICCP), 2007. [3] A. Selvaraju, A. Mani, Analysis of a vapour ejector refrigeration system with environment friendly refrigerants, Int. J. Therm. Sci. 43 (2004) 915e921. [4] R. Roman, J.I. Hernandez, Performance of ejector cooling systems using low ecological impact refrigerants, Int. J. Refrig. 34 (2011) 1707e1716. [5] A. Dahmani, Z. Aidoun, N. Galanis, Optimum design of ejector refrigeration systems with environmentally benign fluids, Int. J. Therm. Sci. 50 (2011) 1562e1572. [6] ASHRAE Standards 34-2004: Designation and Safety Classification of Refrigerants. [7] EN 378-1:2008: Refrigerating Systems and Heat Pumps. Safety and Environmental Requirements. Basic Requirements, Definitions, Classification and Selection Criteria. [8] REFPROP-Reference Fluid Thermodynamic and Transport Properties, Version 9.0, 2010.
J. Kasperski, B. Gil / Applied Thermal Engineering 71 (2014) 197e203 [9] J.H. Keenan, E.P. Neumann, A simple air ejector, J. Appl. Mech. Trans. ASME 64 (1942) 75e81. [10] D.W. Sun, I.W. Eames, Performance characteristics of HCFC-123 ejector refrigeration cycles, Int. J. Energ. Res. 20 (1996) 871e885. [11] B.J. Huang, J.M. Chang, C.P. Wang, V.A. Petrenko, A 1-D analysis of ejector performance, Int. J. Refrig. 22 (1999) 354e364. [12] J. Chen, H. Havtun, B. Palm, Parametric analysis of ejector working characteristics in the refrigeration system, Appl. Therm. Eng. 69 (2014) 130e142. [13] A. Selvaraju, A. Mani, Experimental investigation on R134a vapour ejector refrigeration system, Int. J. Refrig. 29 (2006) 1160e1166. [14] R. Yapici, H.K. Ersoy, A. Aktoprakoǧlu, H.S. Halkaci, O. Yiǧit, Experimental determination of the optimum performance of ejector refrigeration system depending on ejector area ratio, Int. J. Refrig. 31 (2008) 1183e1189. [15] Y. Jia, C. Wenjian, Area ratio effects to the performance of air-cooled ejector refrigeration cycle with R134a refrigerant, Energ. Convers. Manage. 53 (2012) 240e246. [16] J. Chen, H. Havtun, B. Palm, Investigation of ejectors in refrigeration systems: optimum performance evaluation and ejector area ratio perspectives, Appl. Therm. Eng. 64 (2014) 182e191. [17] N.S. Kumar, K.T. Ooi, One dimensional model of an ejector with special attention to fanno flow within the mixing chamber, Appl. Therm. Eng. 65 (2014) 226e235. [18] S.L. Chen, J.Y. Yen, M.C. Huang, An experimental investigation of ejector performance based upon different refrigerants, ASHRAE Trans. 104 (1998) 153e160. [19] W. Pridasawas, Solar-driven Refrigeration Systems with Focus on the Ejector Cycle (Doctoral thesis), Royal Institute of Technology, KTH, Stockholm, 2006. [20] G.K. Alexis, E.K. Karayiannis, A solar ejector cooling system using refrigerant R134a in the Athens area, Renew. Energ. 30 (2005) 1457e1469. [21] R. Yapici, H.K. Ersoy, Performance characteristics of the ejector refrigeration system based on the constant area ejector flow model, Energ. Convers. Manage. 46 (2005) 3117e3135.
203
[22] J. Yu, H. Chen, Y. Ren, Y. Li, A new ejector refrigeration system with an additional jet pump, Appl. Therm. Eng. 26 (2006) 312e319. [23] J. Yu, Y. Li, A theoretical study of a novel regenerative ejector refrigeration cycle, Int. J. Refrig. 30 (2007) 464e470. [24] J. Yu, Y. Ren, H. Chen, Y. Li, Applying mechanical subcooling to ejector refrigeration cycle for improving the coefficient of performance, Energ. Convers. Manage. 48 (2007) 1193e1199. [25] H.K. Ersoy, S. Yalcin, R. Yapici, M. Ozgoren, Performance of a solar ejector cooling-system in the southern region of Turkey, Appl. Energ. 84 (2007) 971e983. [26] L. Boumaraf, A. Lallemand, Modeling of an ejector refrigerating system operating in dimensioning and off-dimensioning conditions with the working fluids R142b and R600a, Appl. Therm. Eng. 29 (2009) 265e274. [27] F. Eskandari Manjili, M.A. Yavari, Performance of a new two-stage multiintercooling transcritical CO2 ejector refrigeration cycle, Appl. Therm. Eng. 40 (2012) 202e209. [28] J.M. Cardemil, S. Colle, A general model for evaluation of vapor ejectors performance for application in refrigeration, Energ. Convers. Manage. 64 (2012) 79e86. [29] G. Grazzini, A. Milazzo, D. Paganini, Design of an ejector cycle refrigeration system, Energ. Convers. Manage. 54 (2012) 38e46. [30] F. Liu, E.A. Groll, Study of ejector efficiencies in refrigeration cycles, Appl. Therm. Eng. 52 (2013) 360e370. [31] V.M. Nguyen, S.B. Riffat, P.S. Doherty, Development of a solar-powered passive ejector cooling system, Appl. Therm. Eng. 21 (2001) 157e168. [32] P. Srisastra, S. Aphornratana, A circulating system for a steam jet refrigeration system, Appl. Therm. Eng. 25 (2005) 2247e2257. [33] P. Srisastra, S. Aphornratana, T. Sriveerakul, Development of a circulating system for a jet refrigeration cycle, Int. J. Refrig. 31 (2008) 921e929. [34] J. Kasperski, Rotational type of a gravitational ejector refrigerator e a system balance of the refrigerant analysis, Int. J. Refrig. 33 (2010) 3e11.