Your SlideShare is downloading. ×
The jet pump cycle-a low cost refrigerator
Upcoming SlideShare
Loading in...5

Thanks for flagging this SlideShare!

Oops! An error has occurred.


Introducing the official SlideShare app

Stunning, full-screen experience for iPhone and Android

Text the download link to your phone

Standard text messaging rates apply

The jet pump cycle-a low cost refrigerator


Published on

1 Like
  • Be the first to comment

No Downloads
Total Views
On Slideshare
From Embeds
Number of Embeds
Embeds 0
No embeds

Report content
Flagged as inappropriate Flag as inappropriate
Flag as inappropriate

Select your reason for flagging this presentation as inappropriate.

No notes for slide


  • 1. Printed in n THE JET-PUMP CYCLE-A LOW COST REFRIGERATOR OPTION POWERED BY WASTE HEAT I. W. EAMES, S. APHORNRATANA and DA-WEN SUN Department of Mechanical and Process Engineering, University of Sheffield, Mappin Street, Sheffield, S14DU, U.K. (Received 4 December 1994) Abstract-A perennial problem with waste heat is the capital cost of plant required to make its utilisation justifiable. A good example of this is the use of waste heat to power absorption refrigerators. The capital cost of absorption refrigerators rises sharply as the temperature of the heat source falls, making waste heat recovery and use uneconomic. This paper describes and evaluates the potential of the jet-pump cycle as a low capital cost option for providing refrigeration utilising low grade waste heat. A brief literature review is provided. An experimental jet-pump refrigerator is described, experimental results are presented and evaluated and the cost benefits of jet-pump refrigerators compared with vapour compression systems are discussed. INTRODUCTION A perennial problem with waste heat is the capital cost of plant required to make its utilisation justifiable. A good example of this is the use of waste heat to power absorption refrigerators. The capital cost of absorption refrigerators rises sharply as the temperature of the heat source falls, making waste heat recovery and use uneconomic. An example of this is the problem of utilising the vast quantities of low temperature waste heat generated by municipal incinerators during summer months. During the winter this heat can be sold to warm offices, homes, public buildings and so on, through a district heating company. However, during the summer vast quantities of heat generated by incineration are wasted to the environment because of the prohibitively high cost of absorption refrigerators, which can be twice that of conventional electrically-powered vapour compression systems, that could convert some of the waste heat into useful refrigeration for building cooling. This paper addresses how this wasted heat could be directed in an energy efficient and environmentally-beneficial way to provide refrigeration. The paper describes and evaluates the results of an experimental study undertaken to assess the potential application of low capital cost refrigerators designed on the jet-pump principle powered by waste heat. Water was selected as the refrigerant for the experimental refrigerator. There has been little published research on steam jet-pump refrigeration in recent years. Most published work has concentrated on the use of halocarbon refrigerants. Because of the harmful environmental effects of these compounds and the growing interest in utilizing waste heat, it was felt to be an opportune time to investigate the potential of low-temperature, steam-powered, jet-pump refrigeration systems. It is hoped that this contribution will encourage debate in this important area of research. THE JET-PUMP REFRIGERATOR CYCLE Figure 1 shows a schematic view of a jet-pump refrigerator. In principle a jet-pump refrigerator operates in much the same way as a conventional vapour compression unit, except that the compressor is replaced by a boiler and a jet-pump. The operation of jet-pump refrigerators is described in most standard thermodynamics text books. However, in operation the jet-pump refrigerator cycle is similar to that of the conventional vapour compression cycle, in which the mechanical compressor in the latter is replaced by a boiler and a jet-pump. The schematic diagram in Fig. 1 shows the main components of the jet-pump cycle. High pressure and temperature 711
  • 2. JET PUMP CONDENSER I THROTTLE VALVE II I PUMP Fig. 1. Schematic representation of a jet-pump cycle-a low cost refrigerator. refrigerant vapour is evolved in the boiler to produce the primary (motive) fluid. This enters the jet-pump nozzle, where it expands to produce a supersonic flow that creates a low pressure region within the mixing or entrainment chamber. This region of low pressure is fed by a secondary vapour flow from the evaporator to produce the required refrigeration effect. Down-stream of the mixing region the combined primary and secondary flows enter a diffuser, where the static pressure rises to equal that in the condenser. The refrigerant is liquefied in the condenser, from where some is pumped at pressure to the boiler whilst the remainder is expanded across a throttle valve before returning to the evaporator. The Coefficient of Performance (COP) of a jet-pump refrigerator is defined as the ratio between the cooling capacity (evaporator heat input) and the energy input to the boiler. The power input to the mechanical pump is normally negligible compared with the heat input to the boiler and, therefore, is not usually included in thermodynamic performance calculations. A LITERATURE SURVEY The first steam jet-pump refrigeration system dates back to 1901 and was designed by LeBlanc of France and Parsons of England [l]. During the 193Os, steam jet refrigeration units experienced their first wave of popularity for air-conditioning large buildings. However, these units were later supplanted by mechanical vapour-compression systems. This was encouraged by compressor developments, particularly centrifugal machines. Since the 1930s there has been little published research on steam-driven jet-pump refrigeration. The most recent paper concerned with steam systems is by Decker [2] who considered the application of steam jet-pump vacuum cooling systems in paper bleaching operations and in the pharmaceutical and chemical industries for various manufacturing processes. In addition to providing chilled water, steam jet-pump systems may also be used for quick chilling of process fluids. Direct vacuum or ‘ flash’ cooling of food is a further application. In this process water is evaporated quickly from produce, such as leafy vegetables like lettuce and cabbage, to produce the cooling effect. Flash cooling may not be practical with mechanical compression systems due to the large volumes of low pressure water vapour needed to be pumped, since this would require compressors with high compression ratios and large displacements. Therefore, steam jet-pump refri- geration systems compete successfully with mechanical compressor systems in these applications. Bowrey ef al. [3] produced an analysis designed to minimise the energy consumption of a three-stage steam ejector deodoriser-cooling system for use in the food industry. The analysis, though simple, demonstrated the need to set operating conditions correctly to reduce energy costs. Other papers reporting research on steam jet-pumps and worthy of mention are those by Addy et al. [4], Munday and Bagster [5-l and Fabri and Siestrunck [8].
  • 3. The jet-pump cycle-low cost refrigeration 713 A limitation of the steam jet-pump cycle is that cooling temperatures can only be above the freezing point of water. This limitation has previously encouraged researchers to investigate the use of halocarbon fluids as alternative refrigerants. As with water, halocarbon compounds such as CFCs, HCFCs and HFCs can be used to utilise low grade heat energy to power the jet-pump cycle, at temperatures ranging upwards from 60°C. This energy is available from a plate solar collector, waste steam, exhaust from automobiles and flue gases. In some cases the cost of the heat supply is negligible and, therefore, the operating costs can be significantly lower than for conventional vapour compression systems. The earliest reported research on jet-pump refrigeration using a refrigerant other than water was that carried out by Mizrahi et al. [9], who undertook a theoretical study to determine the performance of a jet-pump cycle operating on a number of different refrigerants. They assumed a boiler temperature of 60°C supplied by a solar collector. Of the conventional refrigerants tested, R22 and R12 proved to provide the best all-round performance. Misrahi et al.‘ work was extended by Heymann and Resnick [lo] who used Keenan et al.3 [l l] s method for analysing jet-pumps. They argued that a design point boiler temperatures of 90°C would be more appropriate for solar collector heat sources. However, this temperature is still well within the sphere of ‘ waste’ heat supply. An interesting application was devised by Chen [12], who optimised the design of an automobile air-conditioning cooling system, powered by a jet-pump refrigerator, using Elrod’ jet-pump theory s [ 131.This refrigerator was powered using waste heat from an engine. He used R113 as the working fluid. The optimum performance of the cooling cycle at the design point was determined, however, off-design performance data were not provided. Hamner [14, 151 investigated the use of Rl 1 for his jet-pump compression heat pump. His experimental and theoretical investigation concentrated on the overall performance of the refrigeration cycle and did not consider the performance of the jet-pump within this system. Like Chen [ 121, Hamner also suggested the use of jet-pump refrigerators powered by waste engine heat to air-condition vehicles. Other research papers worthy of mention here include: Faithful1 [16], who constructed a Rll ‘ combined Rankine and vapour compression cycle heat pump’ for teaching purposes; Nahdi et al. [I 71. who undertook an experimental study of an (Rl 1) jet-pump refrigerator and concluded that its thermodynamic performance was very sensitive to the design of the jet-pump and the thermal conditions at the boiler, condenser and evaporator; Tyagi and Murty [18], who investigated the use of R 11 in jet-pump refrigerators and conducted similar tests using R113. Huang et al. [ 191also used R113 in their study. They redefined the jet-pump choking theory of Munday and Bagster [6,7] and used it to analyse the performance of their jet-pump. Huang et al. [19] found that the COP of the jet-pump cycle can be increased if the liquid refrigerant returning to the generator is preheated by the ‘ hot’ refrigerant vapour coming from the ejector exhaust, hence reducing the heat input to the generator. Similar improvements in COP were reported when liquid refrigerant was pre-cooled on leaving the condenser, before it enters the expansion valve, by using the cold refrigerant vapour leaving the evaporator. Chen and Hsu [20] performed a simulation study on an ejector refrigerator using RI 1. In this case Elrod’ method [ 131 was used to calculate jet-pump performance. Their results showed that s halocarbon refrigerants, particularly Rl 1, R113 and R114, are most suitable for jet-pump applications and with the addition of a regenerator and precooler into the cycle, as suggested by Huang et al. [ 191and described previously; the COP of an R 11 system could be increased by 17% when operating at boiling, condensing and evaporating temperatures of 93.3, 43.3 and 10°C respectively. Sokolov and Hershgal [21-231 carried out a detailed theoretical and experimental study on possible improvements to the jet-pump refrigeration cycle powered by low grade heat. For the jet-pump analysis, they modified Keenan et al’ [l l] method by using real gas (refrigerant) data s instead of ideal gas data. A comparison of the refrigerants they tested indicated that R114 was the most suitable from a thermodynamic standpoint. They compared the performance of a booster- assisted jet-pump cycle, a hybrid compressor and jet-pump cycle, and a combined booster-compressor-jet-pump cycle. They also reported experiments on a double jet-pump with compression-enhanced cycle. Double jet-pumps were used to improve the part-load performance.
  • 4. 714 I. W. EAMES al. et One was designed to provide high entrainment ratios when condenser pressures were low, whilst the other was designed to provide low entrainment ratios when condenser pressures were high. This need for two or more jet-pumps indicates the sensitivity of the jet-pump cycle to changes in ambient temperature, particularly at the condenser. An alternative approach is to use a variable geometry jet-pump. It should be pointed out that although jet-pump refrigeration systems using halocarbon compounds as refrigerants have some advantages over steam systems, most halocarbon refrigerants damage the ozone layer and are ‘ greenhouse’ gases. Also, the production and import of CFCs is to cease shortly and HCFCs are expected to be subject to end-use controls from 1995, with total phase-out by no later than 2015. Clearly, it is becoming important that research into the use of environment-friendly alternative refrigerants for jet-pump refrigeration systems be undertaken, particularly with regard to their application in the utilisation of low-grade waste heat for refrigeration. JET-PUMP REFRIGERATOR EXPERIMENTS The aim of the experiments was to determine the performance of a laboratory-scale jet-pump refrigerator operating with water as the refrigerant and using a low-temperature heat source of a super heater I w -- I7 ,3 to vacuum pump cooling water Fig. 2. Schematic view of experimental jet-pump refrigerator.
  • 5. The jet-pump cycle--low cost refrigeration 715 Fig. 3. Experimental jet-pump refrigerator. degree commonly available as waste heat from an industrial process. The experimental refrigerator is shown schematically in Fig. 2 and a photograph of the unit is included in Fig. 3. The boiler design was based on the thermosiphon principle with baffle plates located at its upper end to prevent liquid droplets being carried over with the saturated vapour. The maximum heating capacity of the boiler was 7 kW, provided by two 3.5 kW electric heaters. A 500W electrically- powered superheater was positioned in the steam line between the boiler and the jet-pump, to dry the vapour prior to entering the primary nozzle. In practice the addition of this super-heater was found to be unnecessary as its effect on the performance of the refrigerant was insignificant. The evaporator design was based on the flash-evaporation principle. A single 3.5 kW electric heater located within the evaporator vessel was used to provide a cooling load. The output of all electric heaters was controlled using variable transformers. The condenser was a shell and coil type cooled by water taken from the laboratory’ cooling tower. s The test jet-pump was designed using a one-dimensional compressible flow method reported by Keenan and Neumann [l I]. A sectional drawing showing the dimensions of the test’ jet-pump is s shown in Fig. 4. The position of the nozzle was fixed at 26 mm in from the bell mouth entry to the mixing chamber.
  • 6. 716 I. W. EAMFSet al. Primary nozzle throat diameter 2mm Primary nozzle exit diameter gmm 4Omm 1OOmm 40mm 2lOmm I nozzle exit positive (NXP) Fig. 4. REFRIGERATOR PERFORMANCE CHARACTERISTICS Experiments on the jet-pump refrigerator were carried out over a range of boiler, evaporator and condenser temperatures. The electric power input to the boiler and the evaporator were measured and the coefficient of performance (COP) of the cycle was calculated using the following equation: Cop = Electrical power consumption at evaporator Electrical power consumption at boiler ’ The results of these experiments are shown in Figs 5 and 6 for various operating temperatures. Referring to these results it can be seen that the COP is dependent on boiler and evaporator temperatures only and independent of condenser temperature. However, at a certain value of condenser temperature the COP was found to fall sharply to zero. If the condenser temperature was further increased above this critical value, the COP remained at zero. The condenser pressure at which the COP just began to fall was named by Huang et al. [19] as “the critical condenser pressure”. The performance of the experimental refrigerator at the critical condenser operating temperature is shown in Figs 7 and 8. T cond (“‘I 25 28 31 34 37 I I I I I nozzle exit Position 26.15mm evaporator temperature 10.0 ‘C A Tboiler = 120°C. Pbbtler = 1.98 bar o Tbiler = 12S°C, Pboilar = 2.32 bar 130°c. Pboiler = 2.70 bar 3S”c. Pboiler = 3.13 bar ??Tboilcr = 140°C. Pboiler = 3.61 bar 30 40 50 60 P cond (mbar) Fig. 5. Experimental results showing the variation in COP with condenser pressure over a range of boiler pressures and with an evaporator temperature of 10°C.
  • 7. The jet-pump cycle-low cost refrigeration 717 T cond (“‘ ) 25 28 31 34 37 0.5 nozzleexit posttion 26.15mm evaporator temperature 5.0 OC A *boiler = 120°C. Pboiler= 1.98 bar = 125°C. Pboiler= 2.32 bar ? ?*boiler s 130°C. Pboilcr = 2.70 bar O %oiler ’ *boiler = 135°C. Pboilcr= 3.13 bar ??*boiler = 14OT. Pboilcr= 3.61 bar 0.0 30 40 50 60 P mod cmbar) Fig. 6. Experimental results showing the variation in COP with condenser pressure over a range of boiler temperatures and with an vaporator temperature of 5°C. The reason for the sudden cut-off in COP, shown in Figs 5 and 6, is not yet fully understood. However, a possible explanation comes from the behaviour of supersonic flows through conver- gent-divergent nozzles. The pressure ratio across a nozzle (Pboiler/Pevaporator) above which supersonic flow at its outlet can be expected to occur for steam is approximately 1.86. Throughout these experiments this pressure ratio was always between 160 and 300, producing an estimated nozzle outlet Mach number of between 3.5 and 4.5. It is certain, therefore, that during all experiments the primary nozzle always operated in a choked condition. With a fixed nozzle throat area the primary flow was, therefore, only a function of boiler temperature (assuming the steam to be saturated at entry to the nozzle). In other words, for a fixed boiler temperature the primary flow was constant and independent of both evaporator and condenser temperatures. For the same reasons the diffuser was always choked and, therefore, the structure of the flow up-stream of the diffuser throat was independent of the condenser temperature up to its critical value, as indicated by the step characteristic of the results shown in Figs 5 and 6. This meant that, as long as evaporator and boiler temperatures remained constant, the primary flow from the nozzle would entrain the same quantity of secondary flow refrigerant from the evaporator, regardless of the 1000 900 1ooc 800 7.YC 3 PC g 700 Q f 600 1 & 500 nozzle(l): 2mm throat dia. 400 i 8mm dia. outlet 300 1 I I 1 I I I 2s 27 29 31 (OF) 35 37 T(,,,) Fig. 7. Experimental results showing the variation in cooling capacity with condenser temperature over a range of boiler and evaporator temperatures.
  • 8. 718 I. W. EAMESef al. 0.40 0.35 = 10°C 0.30 = 7.5v 0.25 = 3°C % ” 0.20 I 0.15 0.10 0.05 I 0.00 1 25 I 27 I 29 I 31 c: I I 35 I 31 T(co.d) Fig. 8. Experimental results showing the variation in COP with condenser temperature over a range of boiler and evaporator temperatures. conditions in the condenser, so long as its temperature did not exceed the critical value. Therefore, at condenser temperatures less than critical the refrigeration capacity and the COP of the cycle were dependent only on boiler and evaporator temperatures. When operating with condenser temperatures (pressures) less than critical, it was thought that a (normal) shock wave was produced in the divergent section of the diffuser, thus allowing the static pressure of the flow to rise at the diffuser outlet to equal the conditions in the condenser. As the condenser pressure increased the shock wave would have moved up-stream towards the throat and at the critical pressure it would enter the throat. It was noticeable during experimentation that there was a sharp rise in steam temperature at a section of the diffuser throat when operating the refrigerator at a critical condenser condition. This is thought to have resulted from the sudden compression effect of the shock wave. Any further increase in condenser pressure, above its critical value, caused a sudden rise in evaporator temperature. This could only be explained by ‘ hot’ steam from the primary nozzle flowing directly to the evaporator from the mixing chamber. At this point the refrigeration capacity and COP fell sharply to zero. Refrigeration could then only be re-established by increasing the boiler temperature, however, as shown in Figs 5 and 6, the COP was reduced. This behaviour of the refrigerator was interesting because, as shown in Figs 7 and 8, the greater the boiler temperature the lower the COP and refrigeration capacity. This was clearly contrary to what would normally be expected. A reason for this result might be that the flow from the primary (fixed geometry) nozzle was progressively more under-expanded as boiler temperature and pressure were increased. This was confirmed by calculation. Therefore, as the flow became more under- expanded with increasing boiler temperature, a network of oblique expansion waves would have projected progressively further into the mixing chamber from the nozzle outlet. It is thought that this expansion region tended to resist mixing and entrainment between the primary and secondary flow streams because the pressure within the (still expanding) primary stream would, by definition, be greater than that within the secondary flow stream. In effect the secondary flow would have been pushed aside by the primary stream. If this hypothesis is correct then mixing and entrainment would be delayed until the pressure in the primary flow stream had fallen to that within the secondary and consequently there would be less area and time for refrigerant to be entrained into the primary stream before it entered the diffuser throat, thus resulting in the characteristic curves shown in Figs 7 and 8. It was observed that the critical condenser temperature increased with boiler temperature. There- fore, in order to operate the refrigerator at high condenser (ambient) temperatures it would seem necessary to have a high temperature boiler heat source. At first this appeared to limit the usefulness of low temperature waste heat for refrigeration purposes to conditions when the condenser ambient temperature is low, for example less than 25°C. However, it is believed, and experiments have shown, that if the geometry of the nozzle and diffuser were to be made variable then the critical condenser temperature can be increased whilst maintaining a constant boiler temperature.
  • 9. The jet-pump cycle-low cost refrigeration 719 Table I. Estimated variation in COP values based on mass evaporated and electrical energy input T,bOLlW, T,CO”d) (“(3 (“(3 cop,c,c4 cop,rnaw 120 28.3 0.37 0.54 125 29.8 0.33 0.49 130 31.9 0.29 0.44 135 34.0 0.24 0.36 140 26.3 0.19 0.29 EXPERIMENTAL ERRORS The COP characteristics of the experimental refrigerator shown in Figs 5-8 were based on electric power input to the generator and evaporator. These values, therefore, describe the net performance of the refrigerator because they include unwanted heat losses and gains. Estimates of COP based on mass flows have shown that the unwanted heat losses at the boiler were in the order of 25%, whilst the heat gains at the evaporator were 20%. On average the combination of these two effects produced a 30% underestimate of COP values shown in Figs 5-8. Taking this underestimate into account, the best COP value recorded was a respectable 0.544 with a boiler temperature of 120°C evaporator temperature of 10°C and condenser temperature of 28°C. A summary of typical comparisons is shown in Table 1. GENERAL DISCUSSION The experimental results described in the previous section have shown that it is technically feasible to design a refrigerator based on the jet-pump principle and that this will provide a reasonably efficient performance when powered by low grade waste heat. It is recognised that the COP values given for the jet-pump refrigerator are not as high as those commonly reported for absorption systems (typically from 0.6 to 0.9). However, the capital cost of a refrigerator designed to the jet-pump cycle principle will be significantly less than that of a similar capacity absorption refrigerator, making the jet-pump cycle particulary attractive when the cost of heat energy is low. In order that refrigerators based on the jet-pump principle become economically feasible, it is necessary that their life-cycle costs at least equal those of conventional vapour compression units. Of importance here is the comparison between the two systems in terms of capital, operating and maintenance costs throughout the life cycle of the plant. Capital costs In order to compare the capital costs of the two systems it is only necessary to consider the cost of a compressor with that of the boiler and ejector. For example, a 200 kW jet-pump refrigerator would require a 600 kW boiler. If a plate heat exchanger arrangement were used the cost of this will be approximately El800 and the cost of a suitable ejector will be f 1500 making a total cost of f3300. A recent quotation from a leading refrigerator manufacturer gave the purchase price of f5000 for an electrically-powered compressor that would supply 200 kW of cooling. Assuming the cost of the remaining system components in each case to be similar, then it could be expected that the cost of a chiller designed to the jet-pump principle will be slightly less than that of the conventional vapour compression system. Also, because water can be used both as the refrigerant and the primary driving fluid, low cost materials can be used to construct the jet-pump refrigerator because of the low pressures involved. Operating costs It is clear from the experiments described in this paper that the COP values of the jet-pump refrigerator fall short of those of electrically-powered refrigerators. However, electricity to drive a compressor is normally more expensive to the user in terms of kWh than waste heat. In the UK the current general commercial tariff for electricity is approximately f72.2 per MWh. In order to make economic use of waste heat for refrigeration purposes, it is necessary for the cost of the heat to be less than f9.6 per MWh. This figure assumes that a conventional vapour compression refrigerator has a COP of 3.0 and the jet-pump system has a COP of 0.4
  • 10. 720 I. W. EAMES al. et Maintenance costs and reliability Reliability and low maintenance costs incurred are important factors in the selection of refrigeration units. Conventional absorption systems are more reliable and have lower maintenance costs than vapour compression systems because they have fewer highly-stressed moving parts. The jet-pump refrigerator too has few moving parts and, therefore, such machines should present no increase in maintenance cost or reductions in reliability. In all probability maintenance costs will be much less for the jet-pump refrigerator. Energy savings The use of heat-powered refrigerators offers the potential for savings in electrical energy associated with refrigeration by utilising waste heat energy normally rejected to atmosphere. This, in turn, will save burning fossil fuels and reduce COz emissions. There are also additional environmental advantages if water is used as the refrigerant for higher temperature applications. CONCLUSIONS The potential of the jet-pump refrigeration cycle to utilise low-grade waste heat as its power source is now recognised. A theoretical study undertaken by Bevilacqua provided further confirmation that the jet-pump refrigeration cycle is suitable for the utilisation of waste heat at temperatures above 60°C. The use of water as a refrigerant has a number of obvious environmental advantages over halocarbon compounds. Also, the relatively low system pressures and low corrosion potential of water offer the possibility of using low-cost materials in the construction of a refrigerator. An experimental refrigerator designed on the jet-pump cycle was described. Tests were carried out with boiler temperatures between 120 and 140°C and evaporator temperatures between 5 and 10°C. Results indicated that COP values in excess of 0.5 are possible from this type of machine. The results have also shown that refrigerators based on the same principle are economically feasible, particularly if powered by waste heat. It is also concluded that the capital cost of a jet-pump refrigerator could be less than that of a conventional vapour-compression unit for the same cooling capacity. Also, because the jet-pump refrigerator has no moving parts, it is potentially more reliable than conventional systems. REFERENCES ASHRAE, Steam-jet refrigeration equipment. 1979 Equipment Handbook, Chap. 13, pp. 13.1-13.6. ASHRAE, Atlanta, GA, U.S.A. (1979). L. 0. Decker, Consider the cold facts about steam-jet vacuum cooling. Chem. Engng. Prog. 89 (l), 7477 (1993). R. G. Bowrey, V. B. Dang and G. D. Sergeant, An energy model to minimise energy consumption in a low-temperature operation, steam ejector-cooling system. J. Inst. Energy 45, 45-48 (1986). A. L. Addy, J. C. Dutton and C. D. Mikkelsen, Supersonic ejector-diffuser theory and experiments. Report No. UILU-ENG-82-4001, Department of Mechanical and Industrial Engineering, University of Illonois at Urbana- Champaign, Urbana, Illonois, U.S.A. (1981). 5. J. T. Munday and D. F. Bagster, Design and performance of a steam-jet refrigeration system. ThermofIuids Conf of National Committee on Thermodynamics and Fluid Mechanics of Institution of Engineers of Australia, Melbourne, December 1974, pp. 57-61. National Conference Publication, Australia (1974). 6. J. T. Munday and D. F. Bagster, The choking phenomena in ejector with particular reference to steam-jet refrigeration, ThermoJluirls Conf of National Committee on Thermodynamics and Fluid Mechanics of Institution of Engineers of Australia, Hobart, December 1976, pp. 8488, National Conference Publication, Australia (1976). *-__ . .--- 7. J. I. Munday and D. r‘ Bagster, A new ejector theory applied to steam-jet refrigeration. Ind. Engng. Chem. Proc. Res. . Deu. 16, 442&!49 (1977). 8. J. Fabri and R. Siestrunck, Supersonic air ejectors. In Aduances in Applied Mechanics, Vol. V., pp. l-34, edited by von Mises and T. von Karman. Academic Press, New York, U.S.A. (1958). 9. J. Mizrahi, M. Solomiansky, T. Zisner and W. Resnick, Ejector refrigeration from low temperature energy sources. BUN Res. Council of Israel, 6C, l-8 (1957). 10. M. Heymann and W. Resnick, Optimum ejector design for ejector operated refrigeration cycles. Israel J. Technol. 2, 242-247 c 1964). 11. J. H. Keenan, E. P. Neumann and F. Lustwerk, An investigation of ejector design by analysis and experiment. ASME J. Appt. Me& September, 299-309 (1950). 12. L.-T. Chen, A heat-driven mobile refrigeration cycle analysis. Energy Comers. 18 (1) 25-29. 13. H. G. Elrod Jr., The theory of ejectors. J. Appl. Mech. Trans. ASME 67, AllO-A114.
  • 11. The jet-pump cycle-low cost refrigeration 721 14. R. H. Hamner, An investigation of an ejector-compression refrigeration cycle and its applications to heating, cooling and energy conservation, Ph.D. Thesis, University of Alabama, Birmingham. 15. R. H. Hamner, An alternate source of cooling: the ejector-compression heat pump. ASHRAE J. 22, 62-66. 16. D. C. Faithfull, A combined Rankine and vapour compression cycle heat pump for teaching purposes. In Directly Fired Heat Pump-For use in Domestic and Commercial Premises, Proc. of hr. Conf., 19-21 Sept 1984, University of Bristol, U.K., Paper No. 3.1, pp. (3.1) 1-7, edited by P. W. Fitt and R. T. Moses (1984). 17. E. Nahdi, J. C. Champoussin, G. Hostache and J. Cheron, Optimal geometric parameters of a cooling ejector- compressor. hr. J. R&g. 16, (1) 67-72 (1993). 18. K. P. Tyagi and K. N. Murty, Ejector-compression systems for cooling: utihsing low grade waste heat. Heal Recovery Systems & CHP 5 (6), 545-550 (1985). 19. B. J. Huang, C. B. Juang and F. L. Hu, Ejector performance characteristics and design analysis of jet refrigeration system. J. Engng. Gas Turbines and Power, Trans. ASME 187, 792-802 (1985). 20. F. C. Chen and C.-T. Hsu, Performance of ejector heat pumps. Energy Res. 11, 289-300. 21. M. Sokolov and D. Hershgal, Enhanced ejector refrigeration cycles powered by low grade heat. Part 1. Systems characterisation. Inr. J. Refrig. 13, 351-356 (1990). 22. M. Sokolov and D. Hershgal, Enhanced ejector refrigeration cycles powered by low grade heat. Part 2. Design procedures. Znr. J. Refrig. 13, 357-363 (1990). 23. M. Sokolov and D. Hershgal, Enhanced ejector refrigeration cycles powered by low grade heat. Part 3. Experimental results. ZnZ.J. Refrig. 14, 24-31 (1991).