Performance Evaluation of a Model Thermocompressor using Computational Fluid Dynamics

Transcription

1 Performance Evaluaton of a Model Thermocompressor usng Computatonal Flud Dynamcs Kavous Arafar Lecturer, Mechancal Engneerng Department, Islamc Azad Unversty, Neyrz Branch, Neyrz, Iran Abstract Thermocompressors are wdely used n a large number of ndustres that use steam as ther heatng medum or as a power generatng utlty. They are devces that use the energy of a hgh pressure flud to move a low pressure flud and enable t to be compressed to a hgher pressure accordng to the prncple of energy converson. They work lke a vacuum pump but wthout usage of any movng part and so they can save energy. The performance of a thermocompressor hghly depends on ts geometry and operatng condtons. Ths paper frst descrbes the flow behavor wthn a desgned model of a thermocompressor usng the computatonal flud dynamcs code, FLUENT. Snce the flow s turbulent and supersonc, CFD s an effcent tool to reveal the phenomena and mxng process at dfferent part of the thermocompressor whch are not smply obtaned through an expermental work. Then ts performance s analyzed by choosng dfferent operatng condtons at the boundares and also dfferent area ratos whch s one of the sgnfcant geometrcal factors to descrbe the thermocompressor performance. Fnally, the effect of varous nozzle ext plane dameters whch cause dfferent Mach numbers at the nozzle ext s nvestgated on the thermocompressor performance. The results ndcate that these varables can affect both the entranment rato and crtcal back pressure. Ths devce uses water vapor as the workng flud and operates at 7.5 bar motve pressure, 63 C and 80 C for sucton and dscharge temperatures, respectvely. Keywords Thermocompressor, ejector, performance evaluaton, entranment rato, convergng-dvergng nozzle, CFD. L I. INTRODUCTION arge ndustral plants often vent sgnfcant quanttes of low-pressure steam to the atmosphere, wastng energy, water, and water-treatment chemcals. Recovery of the latent heat content of low-pressure steam reduces the boler load, resultng n energy and fuel cost savngs. Low-pressure steam s potental uses nclude drvng evaporaton and dstllaton processes, producng hot water, space heatng, producng vacuum, or chllng water. If the steam pressure s too low for the ntended applcaton, a steam jet thermocompressor or ejector can boost the pressure and temperature to the requred level. Thermocompressors and ejectors operate on the same thermodynamc and physcal prncple: energy contaned n hgh-pressure steam can be Lecturer, Mechancal Engneerng Department, Islamc Azad Unversty, Neyrz Branch, Neyrz, Iran. (Emal: kavous.arafar@gmal.com) transferred to a lower pressure vapor or gas to produce a mxed dscharge stream of ntermedate pressure. These devces are known for: smple constructon, easy nstallaton, low captal and nstallaton costs, easy mantenance wth no movng parts, long useful operatng lfe. A hgh performance thermocompressor n an ndustry leads to the hgher recovery of low pressure steam and save more energy. In order to desgn a hgh performance thermocompressor, a better understandng of the flow behavor such as shock nteractons and mxng process nsde t s necessary. It s also needed to have a good knowledge of the effects of varous parameters on the thermocompressor performance. Prevously, desgners reled prmarly on the past experence, analyss based on emprcal formulatons and test results to make the fnal desgn decsons. However, complex flud flow behavor wthn the thermocompressor results n a flow feld where comprehensve expermental data are too dffcult and expensve to obtan. Keen and Neumann developed a classcal one-dmensonal theory based on the gas dynamc rule n order to desgn the ejectors [1]. Ths theory then was modfed to consder the loss coeffcents at dfferent part of ejector by Eames et al. [2]. But, ths theory was used to predct the performance only when the thermocompressor operates at ts desgn condton (at crtcal back pressure). A number of researchers nvestgated the effect of jet ejector geometry on jet ejector performance. For example, Kroll [3] nvestgated the effect of convergence, dvergence, length, and dameter of the throat secton, nozzle poston, nduced flud entrance, and motve velocty. Croft and Llley nvestgated the optmum length and dameter of the throat secton, nozzle poston, and angle of dvergence [4]. Eames [5] carred out an expermental work to assess the effect of ejector geometry on ts performance such as nozzle desgn and nozzle ext locaton. El-Dessouky et al. [6] developed a smple emprcal model to desgn and evaluate the performance of steam jet ejectors based on a large database extracted from several ejector manufacturers and a number of expermental lteratures. Ther model was smple and elmnated the need for teratve procedures. Recently, Because of the ablty of the computatonal flud dynamcs and numercal smulaton to explan the flow feld nsde the complex geometres, researchers tend to apply CFD n order to model and desgn the thermocompressors. The advantages of ths method are that t takes less tme and cost than expermental method for predctng the performance of a Issue 1, Volume 6,

2 thermocompressor. It also gudes desgners to the drectons n whch a desgn should be modfed to meet desgn objectves. CFD software has been proved by a number of researchers, Rffat and Evertt [7], Hoggarth [8], Rffat et al., [9], as a powerful tool for predctng flow felds nsde jet ejectors. Rffat and Omer [10] and Da-Wen and Eames [11] studed the effect of nozzle poston on jet ejector performance on both desgns; constant pressure and constant-area. As a result, they found that t greatly affects jet ejector performance, as t determnes the dstance over whch the motve and propelled stream are completely mxed. ESDU suggested that the nozzle should be placed between 0.5 and 1.0 throat dameters before the entrance of the throat secton [12]. Rusly et al. smulated the flow through an R141b ejector. They nvestgated the effects of ejector geometres and valdated ther results wth expermental data provded by other researchers [13]. Bartosewcz et al. compared expermental pressure dstrbuton data for an ejector, whch used ar as workng flud, wth results of smulaton usng dfferent turbulence models [14]. Later they extended ther work usng R142b as the workng flud [15]. In ths study, the flow behavor s frst nvestgated by showng the pressure dstrbuton profle along the axs of a desgned model thermocompressor. Then ts performance s analyzed by choosng dfferent operatng condtons at the boundares and also dfferent area ratos whch s one of sgnfcant geometrcal factor. Fnally the effect of varous nozzle ext plane dameters whch cause dfferent Mach numbers at the nozzle ext s nvestgated on the thermocompressor performance. The CFD software package (FLUENT) s used for numercal smulatons. where provdes a channel to the low pressure sucton steam to reach the jet and the subsequent dffuser secton. The functon of the dffuser s to convert the knetc energy to pressure energy. The mxng of the hgh velocty jet and the slow movng low pressure steam s completed n the convergng secton of the dffuser (pont 3). The pressure recovery of the mxed stream begns at the later secton of the convergng dffuser and contnues tll the outlet of the thermocompressor. As the mxed stream leaves the convergng part of the dffuser, t enters the smallest secton of the dffuser, throat. At ths stage, the flud stream undergoes a sudden pressure rse due to occurrence of a normal shock (pont 4). It leads to a sudden pressure recovery and flud attans subsonc velocty after shock. As the flud further moves to the dvergng part of dffuser, the velocty drops further and pressure s more recovered. Fnally, the flud reaches desgn outlet pressure at the ext of the subsonc dffuser. One of the mportant parameters to descrbe the performance of a thermocompressors s the Entranment Rato (R m ) whch s defned as the rato of the recovered Sucton steam quantty to motve steam quantty. Ths s the rato of mass flow rate (kg/h) of sucton steam to flow rate (kg/h) of motve steam. The hgh entranment rato for a thermocompressor sgnfes the hgh quantty of steam whch can be recovered and so, hgh performance thermocompressor has a hgh entranment rato. II. HOW THERMOCOMPRESSOR OPERATES Thermocompressor essentally conssts of three man parts: Nozzle, Sucton Chamber and Dffuser. The nozzle and dffuser have the geometry of convergng-dvergng ventur or deleval nozzle. The hgh pressure steam that enters the nozzle s referred to as the Motve steam or Prmary flud n some lteratures. The low pressure steam that s recovered s known as the Sucton steam or Secondary flud n some lteratures and the steam that exts the thermocompressor from the dffuser s termed as the Dscharge steam. Fgure 1 shows a schematc of the thermocompressor wth pressure-velocty profle along ts axs. The functon of the nozzle s to convert the motve steam enterng at hgh pressure and low velocty to a very hgh velocty and pressure lower than the low pressure sucton steam. The velocty of steam as t enters the nozzle (at pont p) ncreases n the convergng porton. At the end of the convergng porton s the throat, the smallest secton of the nozzle where the velocty reaches sonc velocty. Beyond the throat, the velocty of steam ncreases untl the tp of the nozzle (pont 2) where supersonc veloctes are reached and a very low pressure regon s created. The steam jet leavng the nozzle meets the surroundng low pressure steam (at pont e) and causes t to enter the sucton chamber. The mxng process begns n the sucton chamber Fg. 1 schematc of the thermocompressor wth pressure-velocty profle along the axs As mentoned earler, the thermocompressor conssts of two convergng-dvergng nozzles. So, a bref descrpton of the performance of ths knd of nozzle s mentoned for a better understandng of the flow behavor. Fgure 2 shows the mass flow rate through the nozzle. P b s the downstream pressure or back pressure, P a s the upstream pressure of a delaval nozzle and m s the mass flow rate whch enters the nozzle. When P b s equal to P a, no mass flow rate of m s expected to enter the nozzle. Wth a constant quantty for P a, as the back pressure Issue 1, Volume 6,

3 decreases, the mass flow starts to enter the nozzle and t s expected that further decreasng the back pressure causes the more mass flow rate to enter. Fg. 2 mass flow rate through the nozzle But t happens just up to a specfc pont and after that the mass flow rate stops ncreasng and remans constant. It doesn t matter how much lower the back pressure s. At ths pont t s sad that the nozzle has become choked where the flow speed at the throat reaches the speed of the sound (Mach number = 1). Each thermocompressor has the same trend for the mass flow rate accordng to ts own operatng back pressure whch wll be explaned n the results secton. III. COMPUTATIONAL MODEL A. Governng Equatons Flud flow n the thermocompressor s typcally compressble and turbulence. The governng equatons to descrbe the compressble flow are conversaton of energy, momentum and contnuty whch are as follows: D Contnuty: dvv 0 (1) Dt Dv Momentum: g p j (2) Dt Dh D v Energy: dv ( k T ) j (3) Dt Dt x The term of j s wrtten as: v v j 2 ( j Idvv ) x x 3 j The realzable k-ε turbulence model s selected to govern the turbulence characterstcs n ths smulaton. The term realzable means that the model satsfes certan mathematcal constrants on the Reynolds stresses, consstent wth the physcs of turbulent flows. An mmedate beneft of the realzable k-ε model s that t more accurately predcts the spreadng rate of both planar and round jets [16]. Ths model s a relatvely recent development and dffers from the standard k-ε model n two mportant ways: j (4) 1. The realzable k-ε model contans a new formulaton for the turbulent vscosty. 2. A new transport equaton for the dsspaton rate, ε, has been derved from an exact equaton for the transport of the mean-square vortcty fluctuaton. Bascally, the k-ε turbulence model s a sem-emprcal model based on model transport equatons for the turbulence knetc energy (k) and ts dsspaton rate (ε). The model transport equaton for k s derved from the exact equaton, whle the model transport equaton for ε s obtaned usng physcal reasonng and s dfferent for standard and realzable models. The numercal soluton of the above mentoned set of mean equatons s obtaned by ntroducng addtonal transport equatons for the Reynolds stresses, uu j. These equatons ntroduce sx varables and ncrease the dffculty of solvng the system. Also these equatons contan hgher-order correlatons that represent the processes of dffuson transport, vscous dsspaton, and fluctuatng pressure velocty nteractons. The Reynolds stresses are presumed lnearly related to the mean rate of stran va a scalar eddy vscosty: u u 2 U U j k ( ) 3 x x j j t j Where k s the turbulence knetc energy and υ t s the turbulent vscosty that s calculated from: 2 k t C ( ) (6) Where ε s the rate of dsspaton of k and C μ s one of the model constants. The local values of k and ε for realzable model are obtaned from soluton of ther modeled transport equatons k 2 [( t k u ) ] ts x x x j j k j [( t u j ) ] C1 S C2 x j x j x j k Where S n the modulus of the mean rate-of-stran tensor and s defned as: S 2S j S j (9) C 1ε s not constant and s descrbed n detals wth other varables n reference [16]. The specfcaton of the fve emprcal constants completes and closes the equaton set; these have been assgned by the followng values: C μ = 0.09, C 2 = 1.9, σ k = 1.0, σ ε = 1.2 B. Numercal Soluton Procedure The governng equatons are solved numercally by usng 2 (5) (7) (8) Issue 1, Volume 6,

4 the FLUENT code [16] whch s a commercal CFD software package that uses the control volume based methods to convert these equatons nto algebrac equatons. The nonlnear governng equatons are solved usng the coupled mplct solver and the standard wall functon s appled n the near wall treatment. The workng flud s water vapor whch s supposed to be an deal gas. Although t seems to be an unrealstc assumpton, but n the thermocompressor where the operatng pressure s relatvely low, t has been proved by other researchers [17-19]. The workng flud propertes are selected n the FLUENT database but wth choosng the deal gas relaton for the densty. Convergence of the soluton s assumed when two crtera are satsfed: the mass flow rate through at the outlet face n the model must be stable and every type of the calculaton resdual must be less than IV. RESULTS A. Flow behavor wthn the thermocompressor One of the FLUENT advantages s to descrbe the flow behavor n the models, no matter how complex the flud flow s. Fgure 4 descrbes the contours of Mach number and fgure 5 shows the statc pressure profle along the thermocompressor axs. C. Geometry and Boundary Condtons Fgure 3 shows the geometrcal detals of the thermocompressor whch was desgned bases on the methods provded n lteratures [20, 21]. In order to create the calculaton doman and grd elements of the model, the GAMBIT software [22] s used. The model s created n a 2-D doman but choosng the axsymetrc solver consders the effects of 3-D model for smulaton. Fg. 3 Schematc vew of the desgned thermocompressor ndcatng the dmensons It requres less tme and memory for calculaton than that for 3-D model. Furthermore, t s proven that the results from smulaton of 2-D axsymetrc model are very smlar to the 3- D model [23]. To optmze the CFD model, several mesh denstes are generated. The fnal structured mesh conssts of around quadrlateral elements. Two pressure nlet boundary condtons are selected for the motve and sucton steams enterng the thermocompressor and one pressure outlet boundary condton for the dscharge steam leavng t. They are set based on the saturaton condtons dependng on the characterstcs of the nlet and outlet steams. For the current model the motve steam pressure s 7.5 bar wth correspondng saturaton temperature and the sucton and dscharge steam temperatures are 63 C and 80 C respectvely, wth correspondng saturaton pressures. Fg. 4 contours of Mach number wthn the thermocompressor It s clear from the fg. 5 that the hgh Mach number at the nozzle ext leads to a sudden drop n the statc pressure at ths area. Ths pressure s lower than the sucton pressure at the sucton nlet. So ths pressure dfference causes the sucton steam to enter the sucton chamber of the thermocompressor. The sucton flow velocty at the entrance s too low, but when t enters the thermocompressor and s mxed wth the hgh velocty motve steam whch leaves the nozzle ext, t accelerates through the thermocompressor. The mxed flow reduces ts velocty at the dffuser. At ths part the statc pressure ncreases gradually wth fluctuatons. At the convergng secton of the dffuser the hgh velocty dfference between the motve flud at the nozzle ext and the sucton steam near the wall can create a separate layer. Fg. 5 statc pressure profle along the thermocompressor axs Issue 1, Volume 6,

5 The hgh speed of motve steam also acts lke a wall and the condtons for chokng the sucton flud can be provded. At the specfc dstance n the throat or at the begnnng of the dvergng part of dffuser, the flow experences a normal shock and the statc pressure gradually recovers to the dscharge pressure value. The flow velocty, on the other hand, decrease to a subsonc level as t travels through the dffuser ext. whle the sucton pressure s constant. At each case, the performance dagram of the thermocompressor can be dvded nto 3 regons as dscussed earler. It s clear from the fgure that decreasng the motve pressure causes the entranment rato to ncrease but crtcal back pressure to decrease. B. Effect of operatng condtons on thermocompressor performance 1) Effect of back pressure The effect of the back pressure on the entranment rato for the thermocompressor model accordng to ts operatng condtons s depcted n fg. 6. The motve and sucton pressures are constant whle the dscharge pressure changes. Accordng to ths pcture, there are three regons for the thermocompressor performance: choked flow, unchoked flow and reversed flow. At the choked flow regon, the thermocompressor operates wth a constant entranment rato R m and ncreasng n the back pressure does not affect ts quantty. At ths regon both motve and sucton flud are choked and for ths reason ths s named as double chokng regon n some lteratures. When the back pressure reaches to a pont whch s called crtcal back pressure, the R m starts decreasng untl a zero value. Fg. 7 effect of motve steam pressure on the entranment rato Ths behavor can be related to the sze of jet core and effectve area for each operatng pressure. These conceptons are shown n fgure 8. As mentoned earler, the jet core of the motve flud whch leaves the nozzle ext plane acts lke a wall. The area between ths vrtual wall and the thermocompressor wall s called effectve area. Fg. 8 concepton of effectve area Fg. 6 effect of back pressure on the entranment rato Accordng to the Munday and Bagster theory [24], n the mxng process of the motve and sucton fluds n the thermocompressor, these fluds do not mx untl ther veloctes reach the sonc condton. It means that the mxng process starts when the sucton flud chokes. The area where the sucton flud chokes s called effectve area. The thermocompressor operates at the unchoked regon at ths stage. At ths regon just motve flud s choked and sucton flud s no longer choked at the dffuser throat. So ths regon s named as sngle chokng regon n some lteratures. The pont that the entranment rato reaches a zero magntude s named break down pressure. Fnally, further ncreasng n the back pressure causes the thermocompressor to operate at the reversed flow regon whch the reverse flow phenomena occurs n the sucton nlet and results n falure n the thermocompressor operaton. 2) Effect of motve steam pressure Fgure 7 depcts the varaton of entranment rato for dfferent motve pressures vared n the range of 6.5 to 7.5 bar a) motve pressure: 6.5 bar Issue 1, Volume 6,

6 The sucton temperature s vared from 63 C to 67 C, whle the motve pressure s unchanged. As s clear from the profles, ncreasng the sucton flud temperature causes to ncrease both entranment rato and crtcal back pressure. C. Effect of geometrcal factors on the thermocompressor performance 1) Effect of area rato An mportant geometrcal factor whch affects the thermocompressor performance s the area rato between the dffuser throat and nozzle throat, whch s defned as: b) motve pressure: 7 bar c) motve pressure: 7.5 bar r A = (d th /d Nozz ) 2 (5) Where d th s the dffuser throat dameter and d Nozz s the nozzle throat dameter. Fgure 11 descrbes how the change of area rato nfluences the entranment rato. In ths nvestgaton three values of 25, 27 and 29 are consdered for r A by choosng the dffuser throat dameter of 130, 135 and 140 mm, respectvely. At each case the nozzle throat dameter s constant and equal to 26 mm. Ths dameter for the nozzle throat causes the motve steam to be choked when passes through t. The operatng condtons at the boundares reman constant. In general, when r A ncreases, t causes to rase the entranment rato and decrease the crtcal back pressure. It should be menton that there s always an optmal area rato for each thermocompressor based on ts operatng condtons. Fg. 9 effect of motve pressure on the effectve area The smaller jet core gves the bgger effectve area and t means that a hgher amount of sucton flud can be entered the thermocompressor. In other words, the smaller jet core causes the hgher entranment rato. The operatng pressure can affect the sze of jet core and subsequently, the sze of effectve area and entranment rato. As fgure 9 shows, the lower motve pressure gves a smaller jet core at the nozzle ext whch results n the hgher entranment rato. 3) Effect of sucton temperature The effect of sucton steam temperature on the thermocompressor performance has been llustrated n fg. 10. Fg. 10 effect of sucton temperature on the entranment rato Fg. 11 effect of area raton on the entranment rato 2) Effect of nozzle ext plane dameter In ths part, three nozzles wth equal throat but dfferent ext plane dameters are selected to gve dfferent Mach numbers of 2.3, 2.8 and 3.3 at the nozzle exts. The motve saturaton pressure and sucton temperature are fxed at 7.5 bar and 63, respectvely. Fgure 12 shows that the entranment rato n choked flow regon s not nfluenced by the change of Mach number at the nozzle ext. All three nozzles entran the same amount of secondary flud but the crtcal back pressure ncreases wth the Mach number. Ths s due to the momentum of motve flud whch ncreases wth the rsng of the Mach number and t causes the thermocompressor to operate at a hgher crtcal back pressure. It should be mentoned that the maxmum Mach number at the nozzle ext s lmted by the dameter of the Issue 1, Volume 6,

7 nozzle ext plane. Nozzle wth hgh ext Mach number wll have a large ext plane dameter whch can obstruct the sucton flud to be entraned. Moreover, hgh Mach number at the ext of the nozzle may create the nose polluton n the ndustry where n the thermocompressor operates. t U u v x ε μ ρ τ j υ υ t σ k σ ε tme mean velocty fluctuatng velocty velocty tensor general space coordnate Greek letters turbulent knetc energy dsspaton rate dynamc vscosty flud densty shear stress tensor knematc vscosty turbulent vscosty turbulent Prandtl numbers for k turbulent Prandtl numbers for ε Fg. 12 effect of nozzle ext plane dameter on the entranment rato V. CONCLUSION Flow behavor wthn a desgned model of a thermocompressor was nvestgated by usng the CFD method and the effects of operatng condtons and geometrcal factors on ts performance were evaluated. It has been already verfed that the CFD s an effcent tool to estmate the entranment rato ad crtcal back pressure of the thermocompressor for dfferent operatng condtons. It also helps to reveal the phenomena nsde the thermocompressor n detals. Accordng to the obtaned results, for practcal purposes, the best way to ncrease the entranment rato for an nstalled thermocompressor n an ndustry s to decrease the pressure of motve steam. However, t should be taken nto consderaton that decreasng the motve pressure wll decrease the crtcal back pressure. Other varables such as ncreasng the throat dameter or sucton temperature cannot be easly applcable for an nstalled thermocompressor n the ndustry, even though they ncrease the entranment rato. Results also ndcated that ncreasng the Mach number at the nozzle ext plane does not affect the thermocompressor performance but t wll ncrease the crtcal back pressure. d Nozz d th g h I k m P R m r A S T NOMENCLATURE nozzle throat dameter dffuser throat dameter acceleraton of gravty vector specfc enthalpy dentty matrx thermal conductvty, turbulent knetc energy mass flow rate pressure entranment rato area rato modulus of the mean rate-of-stran tensor temperature ACKNOWLEDGMENT Ths work was supported by the Islamc Azad Unversty, Neyrz Branch, Iran. REFERENCES [1] J.H. Keenan and E.P. Neumann, A smple ar ejector, ASME Journal of Appled Mechancs, 64, 1942, pp [2] I.W. Eames, S. Aphornratana and H. Hader, A theoretcal and expermental study of a small-scale steam jet refrgerator, Internatonal Journal of Refrgeraton, 18 (6), 1995, pp [3] A.E. Kroll, The desgn of jet pumps, Chemcal Engneerng Progress, 1 (2), [4] D.R. Croft and D.G. Llley, Jet pump desgn and performance analyss, AIAA 14 th Aerospace Scence Meetng, AIAA Paper 76183, New York, [5] I.W. Eames, A new prescrpton for the desgn of supersonc jet pumps: the constant rate of momentum change method, Appled Thermal Engneerng, 22, 2002, pp [6] H. El-Dessouky, H. Ettouney, I. Alatq and G. Al-Nuwabt, Evaluaton of steam jet ejectors, Chemcal Engneerng and Processng, 41, 2002, pp [7] S.B. Rffat and P. Evertt, Expermental and CFD modelng of an ejector system for vehcle ar condtonng, Journal of the Insttute of Energy, 72, 1999, pp [8] M.L. Hoggarth, The desgn and performance of hgh-pressure njectors as gas jet boosters, Proceedngs of the Insttuton of Mechancal Engneers, 185, 1970, pp [9] S.B. Rffat, G. Gam and S. Smth, Computatonal flud dynamcs appled to ejector pumps, Appled Thermal Engneerng, 16 (4), 1996, pp [10] S.B. Rffat and S.A. Omer, CFD modelng and expermental nvestgaton of an ejector refrgeraton system usng methanol as the workng flud, Internatonal Journal of Energy Reservaton, 25, 2000, pp [11] S. Da-Wen and I.W. Eames, Recent developments n the desgn theores and applcatons of ejectors, Journal of the Insttute of Energy, 68, 1995, pp [12] ESDU (Engneerng Scences Data Unt), Ejector and jet pump; desgn for steam drven flow, Item number 86030, ESDU Internatonal Ltd., London, [13] E. Rusly, A. Lu, W.W.S. Charters, A. Oo and K. Panthong, Ejector CFD modelng wth real gas model, Mechancal Engneerng Network of Thaland the 16th Conference, 2002, pp [14] Y. Bartosewcz, Z. Adoun, P. Desevaux and Y. Mercader, Numercal and expermental nvestgatons on supersonc ejectors, Internatonal Journal of Heat Flud Flow, 26, 2005, pp Issue 1, Volume 6,

8 [15] Y. Bartosewcz, Z. Adoun and Y. Mercader, Numercal assessment of ejector operaton for refrgeraton applcatons based on CFD, Appled Thermal Engneerng, 26, 2006, pp [16] FLUENT user s manual, verson 6.1, [17] S. Aphornratana, Theoretcal and expermental nvestgaton of a combne ejector-absorpton refrgerator, PhD thess, Unversty of Sheffeld, UK, [18] T. Srveerakul, S. Aphornratana and K. Chunnanond, Performance predcton of steam ejector usng computatonal flud dynamcs: Part 1. Valdaton of the CFD results, Internatonal Journal of Thermal Scences, 46, 2007, pp [19] S. Vargaa, A.C. Olveraa and B. Daconub, Influence of geometrcal factors on steam ejector performance - A numercal assessment, Internatonal journal of refrgeraton, 32, 2009, pp [20] S. Aphornratana, Theoretcal study of a steam-ejector refrgerator, Internatonal Energy Journal, 18 (1), 1996, pp [21] J. H. Keenan, E.P. Neumann and F. Lustwerk, An nvestgaton of ejector desgn by analyss and experment, ASME Journal of Appled Mechancs, Sep. 1950, pp [22] GAMBIT user s manual, verson 2.0.4, [23] K. Panthong, W. Seehanam, M. Behna, T. Srveerakul and S. Aphornratana, Investgaton and mprovement of ejector refrgeraton system usng computatonal flud dynamcs technque, Energy Converson and Management, 48, 2007, pp [24] J.T. Munday and D.F. Bagster, A new theory appled to steam jet refrgeraton, Industral & Engneerng Chemstry, Process Desgn and Development, 1997, 16(4), pp Issue 1, Volume 6,