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Analysis of a jet-pump-assisted vacuum desalination system using power plant waste heat
DESALINATION
ELSEVIER
Desalination 179 (2005) 345-354
www.elsevier.com/locate/desal
Analysis of a jet-pump-assisted vacuum desalination system using power plant waste heat R. Senthil Kumar, A. Mani*, S. Kumaraswamy Department of Mechanical Engineering, Indian Institute of Technology Madras, Chennai 600 036, India Tel. +91 (44) 2257-8534; Fax +91 (44) 2257-0509; emai# [email protected] Received 30 September 2004; accepted 22 November 2004
Abstract
With ever-increasing population and rapid growth of industrialization, there is a great demand for fresh water, especially for drinking, as natural resources are becoming limited. In view of the above, different desalination technologies are evolving with a thrust for utilization of renewable energy sources like solar energy, ocean thermal energy, geothermal energy and waste heat. Vacuum desalination is one such technology in which fresh water is produced from brackish water by evaporation and subsequent condensation. This desalination technique involves different processes like pressurization of brackish water by a pump, creation and maintenance of a vacuum using jet pumps, and evaporation of brackish water at reduced pressure using waste heat from a power plant such as water from condenser. In this paper an analysis of a vacuum desalination system is presented. By applying the mass, momentum and energy balances across the various components, the governing equations are obtained for the analysis. These equations are solved using simulation. Validation of the simulated performance is made with the experimental data available in the literature. The study was carried out by varying operational parameters such as evaporator temperature, condenser temperature, evaporator flow rate, condenser flow rate and chamber pressure. Yield of fresh water obtained from the system increased as condenser temperature decreased and the evaporator temperature increased. Further, the yield increased as chamber pressure decreased. Keywords: Desalination; Jet pump; Vacuum; Seawater
I. Introduction
The world is becoming more and more aware o f its shortage of fresh water. In more than 50
*Corresponding author.
countries in the world, the shortage o f water is already creating a critical situation. A dramatic increase o f the world's population will worsen this scenario, In order to overcome this problem, new economical ways for production o f potable water at socially acceptable costs have to be
Presented at the conference on Membranes in Drinking and Industrial Water Production, L "Aquila, Italy, 15-17 November 2004. Organized by the European Desalination Society. 0011-9164/05/$- See front matter © 2005 Elsevier B.V. All rights reserved doi: I0.1016/j.desal.2004.11.081
346
R. Senthil Kumar et al. / Desalination 179 (2005) 345-354
established. A techno-economic evaluation of the available desalination technologies shows that more advanced desalination systems to reduce investment and maintenance costs are needed. Growing industrialization of the arid zones does not only mean a steady increase of water consumption, but also new sources of energy that can be used for desalination of seawater. For this reason, some are trying to reclaim waste heat for industrial use while others, located near coastlines, may obtain their water through desalination of seawater using solar energy, although costs can be high. Converting seawater into potable water is an energy-intensive process. It is only viable if the source of energy is practically free, such as from celestial sources, geothermal sources or industrial waste heat. On the energy side, safe, jet-pumpassisted vacuum desalination system can be coupled with a thermal power plant for utilizing waste heat, and this might be less expensive and viable. There is an abundant amount of waste hot/ warm water available from thermal power stations. The process reported here is vacuum desalination. This technique takes advantage of a drop in the water boiling point at reduced pressure. This vacuum is created by a jet pump. By dropping the saturation pressure exerted on the seawater to about 0.05622 bar, a condenser outlet temperature could be used as an extemal heat source for boiling the seawater, and the condenser inlet temperature could be used for condensing water vapour. This paper presents a comprehensive review and description of design, logistics and installation aspects of the desalination system in a power station together with an evaluation of capital costs. Table 1 shows major flow rates and temperatures of the system. Tay et al. [1] conducted a pilot study on a laboratory-scale system and concluded that the system performance depends on how efficiently the losses of heat are eliminated. Utilization of waste heat from a steam turbine for production of
Table 1 Major flow rates and temperatures S1. Stream no.
Flow rate (lps)
Temperature (°C)
1
3333.33
44
3333.33
33
0.56
43
2 3
Seawater through evaporator Seawater through condenser Seawater through jet pump
fresh water through a vacuum desalination process was first reported by Low and Tay [2]. A detailed experimental study was made by Mani [3,4] to probe to the effect of water depth and slope of a single-sloped solar still. Mani et al. [5] reported on the utilization of an ocean thermal gradient for production of fresh water through a vacuum desalination process and presented the design details of the system. Simulation of the desalination system was also carried out by Kudish et al. [6], and their work was validated with experimental measurements. The development of a desalination system using solar energy was discussed by Chafik [7]. Experimental performance for a typical operating condition for the vacuum desalination plant was given by Senthil Kumar et al. [8]. Hoefer [9] discussed the application of the liquid jet gas (LJG) pump for evacuation of steam condensers. A number of similar application references were included in the British Hydromechanics Research Association's literature survey of jet pumps, but in many cases the information provided has been inadequate to describe pump operation. Rammingen [ 10] first reported the rapid mixing phenomenon and sudden rise in throat pressure within a few times diameters distance that was noted by most investigators in subsequent years. A one-dimensional analysis of the LJG pump mixing process in a
R. Senthil Kumar et al. / Desalination 179 (2005) 345-354
cylindrical throat (no diffuser) was reported by Folsom [11 ]. Takashima [ 12] developed a similar expression for a throat as well as a diffuser, and also reported experimental data. A somewhat better performance was achieved with a four-hole nozzle compared with single jet nozzles. Witte's [ 13,14] contributions are noteworthy, particularly his study of flow processes in the throat and the "mixing shock" as he termed it. Experimentally he demonstrated high volumetric entrainment ratios and accompanying high isothermal compression efficiencies, up to 40%, by means of multi-hole nozzles, and a relatively long mixing throat. Witte [15] was the first to plan LJG pump tests using a dimensionless Euler number. Betzler [16] improved the analysis, particularly for the diffuser. He demonstrated good agreement between theory and experiments, provided the mixing zone remained in the throat. Design refinements - - particularly throat lengths up to 23 times diameters - - resulted in isothermal efficiencies as high as 19%. Bonnington [17] showed that measured efficiencies decreased with increasing jet velocities. The best performance was obtained when the mixing zone was positioned in the cylindrical throat section. Cunningham [18] and Cunningham and Dopkin [19] presented the performance of various flow obstruction devices and reported that the one with an orifice gives maximum efficiency for maximum flow rate. Schmitt [20] presented onedimensional relations for the LJG pump including friction loss coefficients, and compared experimental results with theory. Isothermal compression efficiencies of about 10% were measured using liquid-liquid pump configurations with short (four diameters) throat lengths. Cunningham [21] developed a one-dimensional model for liquid jet pump handling two-phase gas, in-liquid bubbly mixtures, and examined the characteristics of this liquid jet gas liquid (LJGL) pump. The LJGL model also encompassed the secondaryflow extremes of a liquid and a gas, i.e., the LJGL
347
model bridged and linked two established jet pumps. Sherif et al. [22] used two-phase primary fluid to entrain single-phase secondary liquid injected into the jet pump mixing chamber. The model was capable of incorporating the effects of the temperature and pressure dependency in the analysis.
2. Principle of operation Fig. 1 gives a schematic layout of the jetpump-assisted vacuum desalination system coupled with power station using seawater for its condenser. The system consists of a warm water pump, condenser, evaporator, expansion device, condensate pump, cold-water pump andj et pump. The warm-water pump pressurizes the warm seawater from the condenser of the power station. This water is supplied through the evaporator to give out sensible heat. Seawater from the evaporator is divided into three streams. One portion of the water is sent out, after giving out required heat through valve V 3. A second portion of water passes through the booster pump to the jet pump, which is used for creating the required vacuum. A third portion of water expands through a capillary tube through valve V4 to the chamber intermittently depending on the requirements. The expanded brackish water in the flash chamber undergoes a phase change due to flashing and subsequently by absorbing sensible heat. The water vapour coming from the lower part of the chamber is condensed in the condenser by giving out heat to the cold water pumped through it and then passes to the power station for condensation of the turbine exhaust. The condensate obtained in the desalination system is pumped out periodically by a condensate pump. The concentrated seawater available at the bottom of the flash chamber is periodically removed.
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R. Senthil Kumar et al. / Desalination 179 (2005) 345-354
DESALINATION PLANT [
/
34'C
TT
(
_
_
POWER ] PLANT_ }
)
0.06624 bar "38'C E..
C.W.P. ~5~'~0.56 Ips B.P.
-
=L-~D
O.S.W.
O
Pump
3333.33Ips SEA LEVEL
Fig. 1. Layout of desalination plant utilising power plant waste heat. BP, booster pump; B, boiler; C, condenser; CWP, cold water pump; CT, capillary tube; D, drain plug; E, evaporator; FW, fresh water; G, generator; JP, jet pump; OSW, overflow seawater; PDP, positive displacement pump; SW, seawater; T, turbine; V1-Vs, valves; WWP, warm water pump. 3. D e s i g n o f the s y s t e m
3.1. Waterjet pump The water jet pump required for producing the required vacuum in the desalination system is designed based on principles of continuity and momentum. The dimensional details of the water jet pump components are shown in Table 2. The seawater coming as a primary jet at high velocity from the orifice passes through a converging section where it comes in contact with the gas/ vapour coming from chamber. Due to exchange of momentum, it entrains the suction gas/vapour and moves into the mixing tube followed by a diffuser. From the diffuser, the liquid flows through the evaporator into the discharge line.
3.2. Expansion device The expansion device selected is a stainless steel capillary tube of 3 mm ID and 1 m length. The portion of seawater coming from the evaporator expands through the capillary tube to the
Table 2 Design details of jet pump SI no. Name Orifice diameter, mm Suction chamber diameter, mm Mixing tube diameter, mm Length of the mixing tube, mm Diffuser semi cone angle Length of jet pump, mm
Dimension 7 50 I0 262 2°30 ' 798
chamber by dropping pressure due to friction and momentum change. The needle valve with a size of 1/4" is also used in parallel with the capillary tube so as to match the pressure existing in the flash chamber by regulating the flow rate of brackish water.
3.3. Evaporator Flashed low-temperature water available at the lower part of the chamber undergoes a phase
349
R. Senthil Kumar et al. / Desalination 179 (2005) 345-354
Table 3 Design details of the evaporator
Table 4 Design details of the condenser
SI. no. Name
Dimension
SI no. Name
Dimension
1 2 3 4 5 6 7 8
21 25 2 64 15 55 1 9/inch
1 2 3 4 5 6 7 8
50 55 2 62 15 55 1 9/inch
Inner diameter of the tube, mm Outer diameter of the tube, mm Length of each tube, m No. of tubes Height of the fin, mm Diameter of the fin, mm Thickness of the fin, mm No. of fins
Inner diameter of the tube, mm Outer diameter of the tube, mm Length of each tube, m No. of tubes Height of the fin, mm Diameter of the fin, mm Thickness of the fin, mm No. of fins
change by absorbing heat from the water flowing through the evaporator. The design details of the evaporator are shown in Table 3.
Table 5 Design details of the chamber SI no. Name
Dimension
3.4. Condenser
1 2 3 4 5
2.5 1.2 8 2.5 2.5
The water vapour coming from the lower part of the chamber through the filter is condensed in the condenser by giving its latent heat to the cold water flowing through it. The design details of the condenser are shown in Table 4. 3.5. C h a m b e r
The chamber consists o f the following components, namely condenser, evaporator, angle plates, filter, etc. The chamber length and diameter are evaluated based on condenser and evaporator dimensions housed in it. The chamber is separated into three parts by using angle plates. The evaporator is placed on the left bottom side of the chamber, and the condenser is placed on the right top side o f chamber. A filter is placed above the evaporator before the condenser on supports. This filter is used to prevent the carryover of the salt or brackish water droplets. Using an angle plate, which is fixed at an angle of 45 ° in the chamber, helps to collect the condensate. The thickness of the chamber is evaluated by considering internal pressure, Young's
Chamber length, m Chamber diameter, m Thickness of chamber wall, mm Angle plate length, m Filter length, m
modulus, Poisson's ratio and the yield strength of the material. The chamber material selected was stainless steel 316L based on corrosion resistance. The chamber should be insulated with polyurethane foam insulation material to prevent heat conduction through walls from the ambient. The design details of the chamber are shown in Table 5.
4. Simulation
Momentum and energy balance equations are arrived at by making the momentum and energy balance. These equations have been used for simulation. Validation o f the simulated performance was made with the experimental data available from the literature. The model includes operating parameters such as evaporator temperature and flow rate, con-
350
R. Senthil Kumar et aL / Desalination 179 (2005) 345-354
denser temperature and flow rate and chamber pressure as the input for solving the equations. The physical properties of seawater are significantly influenced by the temperature, pressure and salinity. These properties are required for design and performance calculations. As all the property equations are not available in the literature at various operating pressures and temperatures for seawater, saturated water and steam, and superheated steam, a regression analysis was made to find the equations for the properties of
density, viscosity, thermal conductivity, specific heat capacity, specific enthalpy, specific entropy, enthalpy of evaporation, and internal energy of seawater and steam. The properties of seawater and enthalpy of evaporation of vapour at the assumed chamber pressure are obtained with property equations. Assuming the quantity of hot seawater flowing through the evaporator, the quantity of heat supplied by waste hot water is calculated. From that, the amount of vapour generation for the
l --~
Input Tc
Initial guess At,
! i=h
Input p, Cp, Q¢
L
Energy balance
l Compute m~va,Tv
No
Compute moo,,
Fig. 2. Simulation flow chart for performance analysis.
L
F
l t l i=to
R. Senthil Kumar et al. / Desalination 179 (2005) 345-354
given chamber pressure is computed. By equating the heat lost by vapour and heat gained by the condenser seawater, the mass of fresh water is calculated. The procedure followed for analysis is shown in the flow chart (Fig. 2). The computational procedure yields the output of fresh water.
sation rate of the system, any rise in waste heat temperature would not lead to any further improvement in the distillate yield rate. In this situation, the yield rate could only be increased through an increase in cooling capacity of the system. Fig. 4a and 4b present the effect of chamber pressure on yield while other parameters including evaporator and condenser temperature difference are maintained constant. As the chamber pressure increases the fresh water yield decreases. This is because any increase in chamber pressure affects the saturation pressure inside the system. Hence, the boiling point of the brackish water increases, resulting is decrease in evaporation rate. Thus resulting in decrease in yield. Fig. 5 shows the effect of condenser temperature difference on the yield. When the condenser temperature difference increases, fresh water yield decreases. This is due to reduction in the heat removal rate from the condenser when condensing temperature increases.
5. Results and discussion
Fig. 3 shows the effect of the evaporator temperature difference on yield while other parameters such as chamber pressure, flow rates of water through condenser and evaporator are maintained constant. As the evaporator temperature difference increases, the yield increases. As the evaporator temperature difference increases, the heat input to the system increases leading to an increase in generation rate, thus resulting in an increase in yield. Also a continuous increase in the hot water temperature inlet to the evaporator leads to an increase in the evaporation rate. If the evaporation rate exceeds the maximum conden-
..... Qe = Qc = 3333.33 lps, Pe=0.05622 bar . . . . . . Qe = Qc = 3333.33 Ips, Pc=0.08639 bar
5
' * .... =
4
351
Qe = Qc = 2777.78 Ips, Pe--0.05622 bar Qe = Qc = 2777.78 ips, Pt~0.08639 bar Qe = Qc = 2222.22 Ips, Pe=0.05622 bar
~,- " j " ~,'~" ,,~" , ~
2
l
0
........................................................................................................
0
0.2
0.4
0.6
0.8
Evaporator temperature difference (Ate) Fig. 3. E f f e c t o f e v a p o r a t o r t e m p e r a t u r e d i f f e r e n c e o n y i e l d for d i f f e r e n t f l o w rates.
1
1.2
R. Senthil Kumar et al. /Desalination ] 79 (2005) 345-354
352 5.5
(a)
~.
.
.
.
5.25
.
.
.
.
.
.
.
.
.
.
.
.
.
.
.
.
.
.
.
.
Qe = Qc = 3333.33 Ips -I-
Qe = Qc = 2777.78 tps Qe = Qc = 2222.22 Ips
~"
4.75
"~
4.5
A|~ = 1°C .._
~,
.
......
~
_-
__
A
....
A
,lk . . . . . . . .
: 4.25 i 4P
3.75
[ J
3.5 0.06
0.05
0.07
0.08
0.09
Chamber pressure (Pc), bar 0.6
(b) 0.55
~
41~
O
• ........
4-
~
•
----
,L
J-
J-
Qe = Qc = 3333.33 Ips 0.5
-I-
Qe = Qc = 2777.78 lps
, L Q¢ = Qc = 2222.22 Ips
e~ 0.45
m
~
k
"-
......
Ate = 0.1°C
0.4
0.35
0.3 0.05
0.055
0.06
0.065
0.07
0.075
0.08
0.085
0.09
Chamber pressure (Pc), bar
Fig. 4a, b. Effect of chamber pressure on yield. Most of the energy for producing the potable water depends mainly on heat losses. The cost for producing every kilogram of potable water can be reduced if the system is properly insulated. The flow rate of water in the desalination system could be increased by using larger evaporators and condensers whose heat transfer areas are larger. This would allow the same quantity of heat transfer within a shorter duration. However,
this would increase the capital cost for building the system. The rate of heat transfer for the same exposed surface area and system temperature could be reduced if the system is properly insulated. Again, the gain from the reduction in heat losses would add costs to the system. The breakdown of capital costs of the desalination system is indicated in Table 6.
353
R. Senthil Kumar et al. / Desalination 179 (2005) 345-354 6
...........................................................................
%,. ~
5
*
~ 4
~
"~ ~
e~
~,~
~-~ ~.
i
3
~
~ ~ _
~ " ~
"~ ~
~ ~ .
~
2
Qe = Qc = 3333.33 Ips, Pc~ 0.05622 b a r . . . . . . Qe = Qc = 3333.33 Ips, Pc~ 0.08639 bar A Qe = Qc = 2777.78 lps, Pc= 0.05622 bar ---Qe=Qc=2777.781ps, Pc= 0.08639 bar • Qe = Qc = 2222.22 Ips, Pc= 0.05622 bar . . . . Qe = Qc = 2222.22 lps, ~ . . bar
0 ' ...........................................................................................................................................................................
0
0.2
0.4
0.6
0.8
C o n d e n s e r t e m p e r a t u r e difference
1
1.2
(Ate)
Fig. 5. Effect of condenser temperature difference on yield for different flow rates.
Table 6 Capital cost summary SI no.
Item
1 2 3
Jet pump 1 Booster pump 1 Condenser with two I headers Evaporator with two 1 headers Chamberwith angle 1 plates Reciprocating pump 1 Capillarytube 1 Drain plugs 2 Filter (SS frame with 1 plastic honey comb packing material) Piping,valves and -materials Miscellaneous hardware, - structure, paints, glass tubes, skid, insulation etc., Fabrication charges -Total
4 5 6 7 8 9
10 11
12
Nos.
Cost (Rs.) 15,000 10,000 200,000 200,000 400,000 10,000 1,000 500 5,000
200,000 200,000
150,000 1,391,500
Prior to the construction of the reported desalination system, experiments were conducted by Senthil Kumar et at. [8]. Based on the experience leamed from the previous work, the system proposed here uses ambient temperature of 33°C from the sea surface. The hot wastewater at 44°C from the power station is used as the heat source. The increase in hot wastewater temperature causes a greater difference between the evaporator and condenser temperatures. This increases the heat transfer rate, and, subsequently, raises the distillate production rate of the system. For a parametric study the hot wastewater temperature has been taken as 44°C with a quarter degree of intervals of heat rejection, say 0.75°C, 0.5°C, 0.25°C and 0.1 °C, and the condenser temperature has been taken as 33°C. The effect of the evaporation temperature below 33°C and the condenser temperature have no obvious effect on the yield rate of the desalination system utilizing power plant waste heat. For evaporation temperatures above 33°C, the greater the difference between the evaporator and condenser temperature, the greater the distillate yield rate from the system.
R. Senthil Kumar et aL / Desalination 179 (2005) 345-354
354
6. Conclusions The vacuum desalination system has been designed for utilizing p o w e r plant waste heat and was studied theoretically to produce fresh water. A jet pump was designed and selected to create the vacuum inside the desalination system instead o f a vacuum pump. A m o n g the several orifices and nozzle contours tested, a jet pump with an orifice is best in terms o f producing liquid jets with superior air induction effectiveness, ease o f manufacture and pump operating stability at all flow ratios. The yield increases with a decrease in chamber pressure and condenser temperature. Also the yield increases with an increase in evaporator temperature. The energy efficiency o f the system depends very much on how well the heat losses from the system could be reduced.
[6]
[7] [8]
[9]
[10] [11] [12]
Symbols
[13]
Cp
--
hfg moon mow Pc Qc Qe Tc T~ p
----------
Specific heat o f seawater, kJ/kg K Enthalpy, kJ/kg Yield, lps Mass o f evaporation (kg) Chamber pressure, bar Condenser flow rate, Ips Evaporator flow rate, lps Condenser temperature, °C Vapour temperature, °C Density o f seawater, kg/m 3
[14] [15] [16] [17]
[18]
[19]
References [1] S.C. Low and J.H. Tay, Vacuum desalination using waste heat from a steam turbine, Desalination, 106 (1991) 321-331. [2] J.H. Tay, S.C. Low and S. Jeyaseelan, Vacuum desalination for water purification using waste heat, Desalination, 106 (1996) 131-135. [3] A. Mani, Studies on single sloped solar still, National Solar Energy Convention, I.I.T., New Delhi, 1982, pp. 17.4-17.7. [4] A. Mani, Experimental studies on single sloped solar still, M.Teeh Thesis, liT, Madras, 1982. [5] A. Mani., S. Kumaraswamy and R. Senthil Kumar,
[20]
[21]
[22]
Utilisation of ocean thermal energy for desalination of brackish water, Technical Report, National Insti-tute of Technology, Chennai, 2002. A.I. Kudish, E.G. Evseev, G. Walter and T. Priebe, Simulation study on a solar desalination system utilizing an evaporator/condenser chamber, Energy Conv. Manage., 44 (2003) I653-1670. E. Chafik, A new seawater desalination process using solar energy, Desalination, 153 (2003) 25-37. R. Senthil Kumar, A. Mani and S. Kumaraswamy, Utilisation of ocean thermal gradient for desalination, International Conference on Coastal and Ocean Technology (supplementary), 2003, pp. 101-108. K. Hoefer, Experiments on vacuum pumps for condensers, VDI Forshung, Geb. Ihg. Weseng, 1922, No. 253. G. Von Pawel-Rammigen, Experiments on ejectors, MS Thesis, Pennsylvania State University, 1936. R.G. Folsom, Jet pumps with liquid drive, Chem. Eng. Prog., 44(10) (1952) 765-770. Y. Takashima, Studies on liquid jet gas pumps, J. Sci. Res. Institute, 46 (1952) 230-246. J.H. Witte, Mixing shocks and their influence on the design of liquid-gas ejectors, MS Thesis, Delft, 1962. J.H. Witte, Efficiency and design of liquid-gas ejectors, Brit. Chem. Eng., 32 (1966) 602-607. J.H. Witte, Mixing shocks in two phase flow, J. Fluid Mech., 36(4) (1969) 639-655. R.L. Betzter, The liquid-gas jet pump analysis and experimentalresults, M.S. Thesis, Braunschweig, 1969. S.T. Bonnington, Jet pumps and ejectors, A state ofthe art review and bibliography, BHRA Fluid Engineering, Cranfield, Bedford, England, 1972. R.G. Cunningham, Gas compression with the liquidjet pump, Trans. ASME, J. Fluids Eng., Series 1, 94(3) (1974) 203-215. R.G. Cunningham and R.J. Dopkin, Jet breakup and mixing throat lengths for the liquid-jet gas pump, Trans. ASME, J. Fluids Eng., Series 1, 96(3) (1974) 216-226. H. Schmitt, Diversity of jet pumps and ejector techniques, Second Symp, Jet pumps and ejectors and gas lift techniques, 1974, pp. A4-35-50. R.G. Cunningham, Liquid jet pumps for two-phase flows, Trans. ASME, J. Fluids Eng., Series 5, 117 (1995) 309-316. S.A. Sherif, W.E. Lear, J.M. Steadham, P.L. Hunt and J.B. Holladay, Analysis and modeling of a two-phase jet pump of a thermal management system for aerospace applications, Intemat. J. Mech. Sci., 42 (2000) 85-198.
ELSEVIER
Desalination 179 (2005) 345-354
www.elsevier.com/locate/desal
Analysis of a jet-pump-assisted vacuum desalination system using power plant waste heat R. Senthil Kumar, A. Mani*, S. Kumaraswamy Department of Mechanical Engineering, Indian Institute of Technology Madras, Chennai 600 036, India Tel. +91 (44) 2257-8534; Fax +91 (44) 2257-0509; emai# [email protected] Received 30 September 2004; accepted 22 November 2004
Abstract
With ever-increasing population and rapid growth of industrialization, there is a great demand for fresh water, especially for drinking, as natural resources are becoming limited. In view of the above, different desalination technologies are evolving with a thrust for utilization of renewable energy sources like solar energy, ocean thermal energy, geothermal energy and waste heat. Vacuum desalination is one such technology in which fresh water is produced from brackish water by evaporation and subsequent condensation. This desalination technique involves different processes like pressurization of brackish water by a pump, creation and maintenance of a vacuum using jet pumps, and evaporation of brackish water at reduced pressure using waste heat from a power plant such as water from condenser. In this paper an analysis of a vacuum desalination system is presented. By applying the mass, momentum and energy balances across the various components, the governing equations are obtained for the analysis. These equations are solved using simulation. Validation of the simulated performance is made with the experimental data available in the literature. The study was carried out by varying operational parameters such as evaporator temperature, condenser temperature, evaporator flow rate, condenser flow rate and chamber pressure. Yield of fresh water obtained from the system increased as condenser temperature decreased and the evaporator temperature increased. Further, the yield increased as chamber pressure decreased. Keywords: Desalination; Jet pump; Vacuum; Seawater
I. Introduction
The world is becoming more and more aware o f its shortage of fresh water. In more than 50
*Corresponding author.
countries in the world, the shortage o f water is already creating a critical situation. A dramatic increase o f the world's population will worsen this scenario, In order to overcome this problem, new economical ways for production o f potable water at socially acceptable costs have to be
Presented at the conference on Membranes in Drinking and Industrial Water Production, L "Aquila, Italy, 15-17 November 2004. Organized by the European Desalination Society. 0011-9164/05/$- See front matter © 2005 Elsevier B.V. All rights reserved doi: I0.1016/j.desal.2004.11.081
346
R. Senthil Kumar et al. / Desalination 179 (2005) 345-354
established. A techno-economic evaluation of the available desalination technologies shows that more advanced desalination systems to reduce investment and maintenance costs are needed. Growing industrialization of the arid zones does not only mean a steady increase of water consumption, but also new sources of energy that can be used for desalination of seawater. For this reason, some are trying to reclaim waste heat for industrial use while others, located near coastlines, may obtain their water through desalination of seawater using solar energy, although costs can be high. Converting seawater into potable water is an energy-intensive process. It is only viable if the source of energy is practically free, such as from celestial sources, geothermal sources or industrial waste heat. On the energy side, safe, jet-pumpassisted vacuum desalination system can be coupled with a thermal power plant for utilizing waste heat, and this might be less expensive and viable. There is an abundant amount of waste hot/ warm water available from thermal power stations. The process reported here is vacuum desalination. This technique takes advantage of a drop in the water boiling point at reduced pressure. This vacuum is created by a jet pump. By dropping the saturation pressure exerted on the seawater to about 0.05622 bar, a condenser outlet temperature could be used as an extemal heat source for boiling the seawater, and the condenser inlet temperature could be used for condensing water vapour. This paper presents a comprehensive review and description of design, logistics and installation aspects of the desalination system in a power station together with an evaluation of capital costs. Table 1 shows major flow rates and temperatures of the system. Tay et al. [1] conducted a pilot study on a laboratory-scale system and concluded that the system performance depends on how efficiently the losses of heat are eliminated. Utilization of waste heat from a steam turbine for production of
Table 1 Major flow rates and temperatures S1. Stream no.
Flow rate (lps)
Temperature (°C)
1
3333.33
44
3333.33
33
0.56
43
2 3
Seawater through evaporator Seawater through condenser Seawater through jet pump
fresh water through a vacuum desalination process was first reported by Low and Tay [2]. A detailed experimental study was made by Mani [3,4] to probe to the effect of water depth and slope of a single-sloped solar still. Mani et al. [5] reported on the utilization of an ocean thermal gradient for production of fresh water through a vacuum desalination process and presented the design details of the system. Simulation of the desalination system was also carried out by Kudish et al. [6], and their work was validated with experimental measurements. The development of a desalination system using solar energy was discussed by Chafik [7]. Experimental performance for a typical operating condition for the vacuum desalination plant was given by Senthil Kumar et al. [8]. Hoefer [9] discussed the application of the liquid jet gas (LJG) pump for evacuation of steam condensers. A number of similar application references were included in the British Hydromechanics Research Association's literature survey of jet pumps, but in many cases the information provided has been inadequate to describe pump operation. Rammingen [ 10] first reported the rapid mixing phenomenon and sudden rise in throat pressure within a few times diameters distance that was noted by most investigators in subsequent years. A one-dimensional analysis of the LJG pump mixing process in a
R. Senthil Kumar et al. / Desalination 179 (2005) 345-354
cylindrical throat (no diffuser) was reported by Folsom [11 ]. Takashima [ 12] developed a similar expression for a throat as well as a diffuser, and also reported experimental data. A somewhat better performance was achieved with a four-hole nozzle compared with single jet nozzles. Witte's [ 13,14] contributions are noteworthy, particularly his study of flow processes in the throat and the "mixing shock" as he termed it. Experimentally he demonstrated high volumetric entrainment ratios and accompanying high isothermal compression efficiencies, up to 40%, by means of multi-hole nozzles, and a relatively long mixing throat. Witte [15] was the first to plan LJG pump tests using a dimensionless Euler number. Betzler [16] improved the analysis, particularly for the diffuser. He demonstrated good agreement between theory and experiments, provided the mixing zone remained in the throat. Design refinements - - particularly throat lengths up to 23 times diameters - - resulted in isothermal efficiencies as high as 19%. Bonnington [17] showed that measured efficiencies decreased with increasing jet velocities. The best performance was obtained when the mixing zone was positioned in the cylindrical throat section. Cunningham [18] and Cunningham and Dopkin [19] presented the performance of various flow obstruction devices and reported that the one with an orifice gives maximum efficiency for maximum flow rate. Schmitt [20] presented onedimensional relations for the LJG pump including friction loss coefficients, and compared experimental results with theory. Isothermal compression efficiencies of about 10% were measured using liquid-liquid pump configurations with short (four diameters) throat lengths. Cunningham [21] developed a one-dimensional model for liquid jet pump handling two-phase gas, in-liquid bubbly mixtures, and examined the characteristics of this liquid jet gas liquid (LJGL) pump. The LJGL model also encompassed the secondaryflow extremes of a liquid and a gas, i.e., the LJGL
347
model bridged and linked two established jet pumps. Sherif et al. [22] used two-phase primary fluid to entrain single-phase secondary liquid injected into the jet pump mixing chamber. The model was capable of incorporating the effects of the temperature and pressure dependency in the analysis.
2. Principle of operation Fig. 1 gives a schematic layout of the jetpump-assisted vacuum desalination system coupled with power station using seawater for its condenser. The system consists of a warm water pump, condenser, evaporator, expansion device, condensate pump, cold-water pump andj et pump. The warm-water pump pressurizes the warm seawater from the condenser of the power station. This water is supplied through the evaporator to give out sensible heat. Seawater from the evaporator is divided into three streams. One portion of the water is sent out, after giving out required heat through valve V 3. A second portion of water passes through the booster pump to the jet pump, which is used for creating the required vacuum. A third portion of water expands through a capillary tube through valve V4 to the chamber intermittently depending on the requirements. The expanded brackish water in the flash chamber undergoes a phase change due to flashing and subsequently by absorbing sensible heat. The water vapour coming from the lower part of the chamber is condensed in the condenser by giving out heat to the cold water pumped through it and then passes to the power station for condensation of the turbine exhaust. The condensate obtained in the desalination system is pumped out periodically by a condensate pump. The concentrated seawater available at the bottom of the flash chamber is periodically removed.
348
R. Senthil Kumar et al. / Desalination 179 (2005) 345-354
DESALINATION PLANT [
/
34'C
TT
(
_
_
POWER ] PLANT_ }
)
0.06624 bar "38'C E..
C.W.P. ~5~'~0.56 Ips B.P.
-
=L-~D
O.S.W.
O
Pump
3333.33Ips SEA LEVEL
Fig. 1. Layout of desalination plant utilising power plant waste heat. BP, booster pump; B, boiler; C, condenser; CWP, cold water pump; CT, capillary tube; D, drain plug; E, evaporator; FW, fresh water; G, generator; JP, jet pump; OSW, overflow seawater; PDP, positive displacement pump; SW, seawater; T, turbine; V1-Vs, valves; WWP, warm water pump. 3. D e s i g n o f the s y s t e m
3.1. Waterjet pump The water jet pump required for producing the required vacuum in the desalination system is designed based on principles of continuity and momentum. The dimensional details of the water jet pump components are shown in Table 2. The seawater coming as a primary jet at high velocity from the orifice passes through a converging section where it comes in contact with the gas/ vapour coming from chamber. Due to exchange of momentum, it entrains the suction gas/vapour and moves into the mixing tube followed by a diffuser. From the diffuser, the liquid flows through the evaporator into the discharge line.
3.2. Expansion device The expansion device selected is a stainless steel capillary tube of 3 mm ID and 1 m length. The portion of seawater coming from the evaporator expands through the capillary tube to the
Table 2 Design details of jet pump SI no. Name Orifice diameter, mm Suction chamber diameter, mm Mixing tube diameter, mm Length of the mixing tube, mm Diffuser semi cone angle Length of jet pump, mm
Dimension 7 50 I0 262 2°30 ' 798
chamber by dropping pressure due to friction and momentum change. The needle valve with a size of 1/4" is also used in parallel with the capillary tube so as to match the pressure existing in the flash chamber by regulating the flow rate of brackish water.
3.3. Evaporator Flashed low-temperature water available at the lower part of the chamber undergoes a phase
349
R. Senthil Kumar et al. / Desalination 179 (2005) 345-354
Table 3 Design details of the evaporator
Table 4 Design details of the condenser
SI. no. Name
Dimension
SI no. Name
Dimension
1 2 3 4 5 6 7 8
21 25 2 64 15 55 1 9/inch
1 2 3 4 5 6 7 8
50 55 2 62 15 55 1 9/inch
Inner diameter of the tube, mm Outer diameter of the tube, mm Length of each tube, m No. of tubes Height of the fin, mm Diameter of the fin, mm Thickness of the fin, mm No. of fins
Inner diameter of the tube, mm Outer diameter of the tube, mm Length of each tube, m No. of tubes Height of the fin, mm Diameter of the fin, mm Thickness of the fin, mm No. of fins
change by absorbing heat from the water flowing through the evaporator. The design details of the evaporator are shown in Table 3.
Table 5 Design details of the chamber SI no. Name
Dimension
3.4. Condenser
1 2 3 4 5
2.5 1.2 8 2.5 2.5
The water vapour coming from the lower part of the chamber through the filter is condensed in the condenser by giving its latent heat to the cold water flowing through it. The design details of the condenser are shown in Table 4. 3.5. C h a m b e r
The chamber consists o f the following components, namely condenser, evaporator, angle plates, filter, etc. The chamber length and diameter are evaluated based on condenser and evaporator dimensions housed in it. The chamber is separated into three parts by using angle plates. The evaporator is placed on the left bottom side of the chamber, and the condenser is placed on the right top side o f chamber. A filter is placed above the evaporator before the condenser on supports. This filter is used to prevent the carryover of the salt or brackish water droplets. Using an angle plate, which is fixed at an angle of 45 ° in the chamber, helps to collect the condensate. The thickness of the chamber is evaluated by considering internal pressure, Young's
Chamber length, m Chamber diameter, m Thickness of chamber wall, mm Angle plate length, m Filter length, m
modulus, Poisson's ratio and the yield strength of the material. The chamber material selected was stainless steel 316L based on corrosion resistance. The chamber should be insulated with polyurethane foam insulation material to prevent heat conduction through walls from the ambient. The design details of the chamber are shown in Table 5.
4. Simulation
Momentum and energy balance equations are arrived at by making the momentum and energy balance. These equations have been used for simulation. Validation o f the simulated performance was made with the experimental data available from the literature. The model includes operating parameters such as evaporator temperature and flow rate, con-
350
R. Senthil Kumar et aL / Desalination 179 (2005) 345-354
denser temperature and flow rate and chamber pressure as the input for solving the equations. The physical properties of seawater are significantly influenced by the temperature, pressure and salinity. These properties are required for design and performance calculations. As all the property equations are not available in the literature at various operating pressures and temperatures for seawater, saturated water and steam, and superheated steam, a regression analysis was made to find the equations for the properties of
density, viscosity, thermal conductivity, specific heat capacity, specific enthalpy, specific entropy, enthalpy of evaporation, and internal energy of seawater and steam. The properties of seawater and enthalpy of evaporation of vapour at the assumed chamber pressure are obtained with property equations. Assuming the quantity of hot seawater flowing through the evaporator, the quantity of heat supplied by waste hot water is calculated. From that, the amount of vapour generation for the
l --~
Input Tc
Initial guess At,
! i=h
Input p, Cp, Q¢
L
Energy balance
l Compute m~va,Tv
No
Compute moo,,
Fig. 2. Simulation flow chart for performance analysis.
L
F
l t l i=to
R. Senthil Kumar et al. / Desalination 179 (2005) 345-354
given chamber pressure is computed. By equating the heat lost by vapour and heat gained by the condenser seawater, the mass of fresh water is calculated. The procedure followed for analysis is shown in the flow chart (Fig. 2). The computational procedure yields the output of fresh water.
sation rate of the system, any rise in waste heat temperature would not lead to any further improvement in the distillate yield rate. In this situation, the yield rate could only be increased through an increase in cooling capacity of the system. Fig. 4a and 4b present the effect of chamber pressure on yield while other parameters including evaporator and condenser temperature difference are maintained constant. As the chamber pressure increases the fresh water yield decreases. This is because any increase in chamber pressure affects the saturation pressure inside the system. Hence, the boiling point of the brackish water increases, resulting is decrease in evaporation rate. Thus resulting in decrease in yield. Fig. 5 shows the effect of condenser temperature difference on the yield. When the condenser temperature difference increases, fresh water yield decreases. This is due to reduction in the heat removal rate from the condenser when condensing temperature increases.
5. Results and discussion
Fig. 3 shows the effect of the evaporator temperature difference on yield while other parameters such as chamber pressure, flow rates of water through condenser and evaporator are maintained constant. As the evaporator temperature difference increases, the yield increases. As the evaporator temperature difference increases, the heat input to the system increases leading to an increase in generation rate, thus resulting in an increase in yield. Also a continuous increase in the hot water temperature inlet to the evaporator leads to an increase in the evaporation rate. If the evaporation rate exceeds the maximum conden-
..... Qe = Qc = 3333.33 lps, Pe=0.05622 bar . . . . . . Qe = Qc = 3333.33 Ips, Pc=0.08639 bar
5
' * .... =
4
351
Qe = Qc = 2777.78 Ips, Pe--0.05622 bar Qe = Qc = 2777.78 ips, Pt~0.08639 bar Qe = Qc = 2222.22 Ips, Pe=0.05622 bar
~,- " j " ~,'~" ,,~" , ~
2
l
0
........................................................................................................
0
0.2
0.4
0.6
0.8
Evaporator temperature difference (Ate) Fig. 3. E f f e c t o f e v a p o r a t o r t e m p e r a t u r e d i f f e r e n c e o n y i e l d for d i f f e r e n t f l o w rates.
1
1.2
R. Senthil Kumar et al. /Desalination ] 79 (2005) 345-354
352 5.5
(a)
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.
.
.
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.
.
.
.
.
.
.
.
.
.
.
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.
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.
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.
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~"
4.75
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4.5
A|~ = 1°C .._
~,
.
......
~
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....
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,lk . . . . . . . .
: 4.25 i 4P
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[ J
3.5 0.06
0.05
0.07
0.08
0.09
Chamber pressure (Pc), bar 0.6
(b) 0.55
~
41~
O
• ........
4-
~
•
----
,L
J-
J-
Qe = Qc = 3333.33 Ips 0.5
-I-
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, L Q¢ = Qc = 2222.22 Ips
e~ 0.45
m
~
k
"-
......
Ate = 0.1°C
0.4
0.35
0.3 0.05
0.055
0.06
0.065
0.07
0.075
0.08
0.085
0.09
Chamber pressure (Pc), bar
Fig. 4a, b. Effect of chamber pressure on yield. Most of the energy for producing the potable water depends mainly on heat losses. The cost for producing every kilogram of potable water can be reduced if the system is properly insulated. The flow rate of water in the desalination system could be increased by using larger evaporators and condensers whose heat transfer areas are larger. This would allow the same quantity of heat transfer within a shorter duration. However,
this would increase the capital cost for building the system. The rate of heat transfer for the same exposed surface area and system temperature could be reduced if the system is properly insulated. Again, the gain from the reduction in heat losses would add costs to the system. The breakdown of capital costs of the desalination system is indicated in Table 6.
353
R. Senthil Kumar et al. / Desalination 179 (2005) 345-354 6
...........................................................................
%,. ~
5
*
~ 4
~
"~ ~
e~
~,~
~-~ ~.
i
3
~
~ ~ _
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~ ~ .
~
2
Qe = Qc = 3333.33 Ips, Pc~ 0.05622 b a r . . . . . . Qe = Qc = 3333.33 Ips, Pc~ 0.08639 bar A Qe = Qc = 2777.78 lps, Pc= 0.05622 bar ---Qe=Qc=2777.781ps, Pc= 0.08639 bar • Qe = Qc = 2222.22 Ips, Pc= 0.05622 bar . . . . Qe = Qc = 2222.22 lps, ~ . . bar
0 ' ...........................................................................................................................................................................
0
0.2
0.4
0.6
0.8
C o n d e n s e r t e m p e r a t u r e difference
1
1.2
(Ate)
Fig. 5. Effect of condenser temperature difference on yield for different flow rates.
Table 6 Capital cost summary SI no.
Item
1 2 3
Jet pump 1 Booster pump 1 Condenser with two I headers Evaporator with two 1 headers Chamberwith angle 1 plates Reciprocating pump 1 Capillarytube 1 Drain plugs 2 Filter (SS frame with 1 plastic honey comb packing material) Piping,valves and -materials Miscellaneous hardware, - structure, paints, glass tubes, skid, insulation etc., Fabrication charges -Total
4 5 6 7 8 9
10 11
12
Nos.
Cost (Rs.) 15,000 10,000 200,000 200,000 400,000 10,000 1,000 500 5,000
200,000 200,000
150,000 1,391,500
Prior to the construction of the reported desalination system, experiments were conducted by Senthil Kumar et at. [8]. Based on the experience leamed from the previous work, the system proposed here uses ambient temperature of 33°C from the sea surface. The hot wastewater at 44°C from the power station is used as the heat source. The increase in hot wastewater temperature causes a greater difference between the evaporator and condenser temperatures. This increases the heat transfer rate, and, subsequently, raises the distillate production rate of the system. For a parametric study the hot wastewater temperature has been taken as 44°C with a quarter degree of intervals of heat rejection, say 0.75°C, 0.5°C, 0.25°C and 0.1 °C, and the condenser temperature has been taken as 33°C. The effect of the evaporation temperature below 33°C and the condenser temperature have no obvious effect on the yield rate of the desalination system utilizing power plant waste heat. For evaporation temperatures above 33°C, the greater the difference between the evaporator and condenser temperature, the greater the distillate yield rate from the system.
R. Senthil Kumar et aL / Desalination 179 (2005) 345-354
354
6. Conclusions The vacuum desalination system has been designed for utilizing p o w e r plant waste heat and was studied theoretically to produce fresh water. A jet pump was designed and selected to create the vacuum inside the desalination system instead o f a vacuum pump. A m o n g the several orifices and nozzle contours tested, a jet pump with an orifice is best in terms o f producing liquid jets with superior air induction effectiveness, ease o f manufacture and pump operating stability at all flow ratios. The yield increases with a decrease in chamber pressure and condenser temperature. Also the yield increases with an increase in evaporator temperature. The energy efficiency o f the system depends very much on how well the heat losses from the system could be reduced.
[6]
[7] [8]
[9]
[10] [11] [12]
Symbols
[13]
Cp
--
hfg moon mow Pc Qc Qe Tc T~ p
----------
Specific heat o f seawater, kJ/kg K Enthalpy, kJ/kg Yield, lps Mass o f evaporation (kg) Chamber pressure, bar Condenser flow rate, Ips Evaporator flow rate, lps Condenser temperature, °C Vapour temperature, °C Density o f seawater, kg/m 3
[14] [15] [16] [17]
[18]
[19]
References [1] S.C. Low and J.H. Tay, Vacuum desalination using waste heat from a steam turbine, Desalination, 106 (1991) 321-331. [2] J.H. Tay, S.C. Low and S. Jeyaseelan, Vacuum desalination for water purification using waste heat, Desalination, 106 (1996) 131-135. [3] A. Mani, Studies on single sloped solar still, National Solar Energy Convention, I.I.T., New Delhi, 1982, pp. 17.4-17.7. [4] A. Mani, Experimental studies on single sloped solar still, M.Teeh Thesis, liT, Madras, 1982. [5] A. Mani., S. Kumaraswamy and R. Senthil Kumar,
[20]
[21]
[22]
Utilisation of ocean thermal energy for desalination of brackish water, Technical Report, National Insti-tute of Technology, Chennai, 2002. A.I. Kudish, E.G. Evseev, G. Walter and T. Priebe, Simulation study on a solar desalination system utilizing an evaporator/condenser chamber, Energy Conv. Manage., 44 (2003) I653-1670. E. Chafik, A new seawater desalination process using solar energy, Desalination, 153 (2003) 25-37. R. Senthil Kumar, A. Mani and S. Kumaraswamy, Utilisation of ocean thermal gradient for desalination, International Conference on Coastal and Ocean Technology (supplementary), 2003, pp. 101-108. K. Hoefer, Experiments on vacuum pumps for condensers, VDI Forshung, Geb. Ihg. Weseng, 1922, No. 253. G. Von Pawel-Rammigen, Experiments on ejectors, MS Thesis, Pennsylvania State University, 1936. R.G. Folsom, Jet pumps with liquid drive, Chem. Eng. Prog., 44(10) (1952) 765-770. Y. Takashima, Studies on liquid jet gas pumps, J. Sci. Res. Institute, 46 (1952) 230-246. J.H. Witte, Mixing shocks and their influence on the design of liquid-gas ejectors, MS Thesis, Delft, 1962. J.H. Witte, Efficiency and design of liquid-gas ejectors, Brit. Chem. Eng., 32 (1966) 602-607. J.H. Witte, Mixing shocks in two phase flow, J. Fluid Mech., 36(4) (1969) 639-655. R.L. Betzter, The liquid-gas jet pump analysis and experimentalresults, M.S. Thesis, Braunschweig, 1969. S.T. Bonnington, Jet pumps and ejectors, A state ofthe art review and bibliography, BHRA Fluid Engineering, Cranfield, Bedford, England, 1972. R.G. Cunningham, Gas compression with the liquidjet pump, Trans. ASME, J. Fluids Eng., Series 1, 94(3) (1974) 203-215. R.G. Cunningham and R.J. Dopkin, Jet breakup and mixing throat lengths for the liquid-jet gas pump, Trans. ASME, J. Fluids Eng., Series 1, 96(3) (1974) 216-226. H. Schmitt, Diversity of jet pumps and ejector techniques, Second Symp, Jet pumps and ejectors and gas lift techniques, 1974, pp. A4-35-50. R.G. Cunningham, Liquid jet pumps for two-phase flows, Trans. ASME, J. Fluids Eng., Series 5, 117 (1995) 309-316. S.A. Sherif, W.E. Lear, J.M. Steadham, P.L. Hunt and J.B. Holladay, Analysis and modeling of a two-phase jet pump of a thermal management system for aerospace applications, Intemat. J. Mech. Sci., 42 (2000) 85-198.