流量为240t-h汽－水浮头式换热器的设计【过程装备与控制工程类】【说明书+CAD】
收藏
资源目录
压缩包内文档预览：
编号：122571990
类型：共享资源
大小：1.19MB
格式：ZIP
上传时间：2021-04-20
上传人：221589****qq.com
认证信息
个人认证
李**（实名认证）
湖南
IP属地：湖南
40
积分
- 关 键 词：
-
过程装备与控制工程类
流量
240
头式
换热器
设计
过程
装备
控制工程
说明书
CAD
- 资源描述：
-
流量为240t-h汽－水浮头式换热器的设计【过程装备与控制工程类】【说明书+CAD】,过程装备与控制工程类,流量,240,头式,换热器,设计,过程,装备,控制工程,说明书,CAD
- 内容简介：
-
ORIGINALWorachest Pirompugd Somchai WongwisesChi-Chuan WangA tube-by-tube reduction method for simultaneous heat and masstransfer characteristics for plain fin-and-tube heat exchangersin dehumidifying conditionsReceived: 19 August 2004/ Accepted: 24 November 2004/Published online: 4 March 2005? Springer-Verlag 2005Abstract This study proposed a new method, namely atube-by-tube reduction method to analyze the perfor-mance of fin-and-tube heat exchangers having plain finconfigurationunderdehumidifyingconditions.Themass transfer coefficients which seldom reported in theopen literature, are also presented. For fully wet con-ditions, it is found that the reduced results for bothsensible heat transfer performance and the mass transferperformance by the present method are insensitive tochange of inlet humidity. Unlike those tested in fully drycondition, the sensible heat transfer performance underdehumidification is comparatively independent of finpitch. The ratio of the heat transfer characteristic tomass transfer characteristic (hc,o/hd,oCp,a) is in the rangeof 0.6?1.0, and the ratio is insensitive to change of finspacing at low Reynolds number. However, a slight dropof the ratio of (hc,o/hd,oCp,a) is seen with the decrease offin spacing when the Reynolds number is sufficient high.This is associated with the more pronounced influencedue to condensate removal by the vapor shear. Corre-lations are proposed to describe the heat and massperformance for the present plate fin configurations.These correlations can describe 89% of the ChiltonColburn j-factor of the heat transfer (jh) within 15% andcan correlate 81% of the Chilton Colburn j-factor of themass transfer (jm) within 20%.Keywords Fin-and-tube heat exchanger Dehumidifying Sensible heat transfer performance Mass transfer performanceNomenclatureAfSurface area of finAoTotal surface areaAp,iInside surface area of tubesAp,oOutside surface area of tubesbpSlope of the air saturation curved between theoutside and inside tube wall temperaturebrSlope of the air saturation curved between themean water temperature and the inside walltemperaturebw,mSlope of the air saturation curved at the meanwater film temperature of the fin surfacebw,pSlope of the air saturation curved at the meanwater film temperature of the tube surfaceCp,aMoist air specific heat at constant pressureCp,wWater specific heat at constant pressureDcTube outside diameter (include collar)DiTube inside diameterfiIn-tube friction factors of waterFCorrection factorGmaxMaximum mass velocity based on minimumflow areahc,oSensible heat transfer coefficienthd,oMass transfer coefficienthiInside heat transfer coefficientho,wTotal heat transfer coefficient for wet externalfinIoModified Bessel function solution of the firstkind, order 0I1Modified Bessel function solution of the firstkind, order 1iaAir enthalpyia,inInlet air enthalpyia,mMean air enthalpyia,outOutlet air enthalpyigSaturated water vapor enthalpyW. Pirompugd S. Wongwises (&)Fluid Mechanics, Thermal Engineering andMultiphase Flow Research Lab. (FUTURE),Department of Mechanical Engineering,King Mongkuts University of Technology,Thonburi, Bangmod, Bangkok, 10140, ThailandE-mail: somchai.wonkmutt.ac.thTel.: +66-2-4709115Fax: +66-2-4709111C.-C. WangEnergy and Resources Lab.,Industrial Technology Research Institute,Hsinchu, Taiwan, ROCHeat Mass Transfer (2005) 41: 756765DOI 10.1007/s00231-004-0581-ximMean enthalpyir,inSaturated air enthalpy at the inlet water tem-peratureir,mMean saturated air enthalpy at the meanwater temperatureir,outSaturated air enthalpy at the outlet watertemperatureis,fmSaturated air enthalpy at the fin mean tem-peratureis,fbSaturated air enthalpy at the fin base tem-peratureis,p,i,mMean saturated air enthalpy at the mean in-side tube wall temperatureis,p,o,mMean saturated air enthalpy at the meanoutside tube wall temperatureis,wSaturated air enthalpy at the water film tem-peratureis,w,mMean saturated air enthalpy at the meanwater film temperature of the fin surfacejhChilton-Colburn j-factor of the heat transferjmChilton-Colburn j-factor of the mass transferK0Modified Bessel function solution of the sec-ond kind, order 0K1Modified Bessel function solution of the sec-ond kind, order 1kfThermal conductivity of finkiThermal conductivity of waterkpThermal conductivity of tubekwThermal conductivity of water filmLpTube length_ maAir mass flow rate_ mwWater mass flow rateNNumber of tube rowPPressurePlLongitudinal tube pitchPrPrandtl numberPtTransverse tube pitch_QHeat transfer rate_QaAir side heat transfer rate_QavgAverage heat transfer rate_QtotalTotal heat transfer rate_QwWater side heat transfer rateRRatio of heat transfer characteristic to masstransfer characteristicRHRelative humidityriDistance from the center of the tube to the finbaseroDistance from the center of the tube to the fintipReDiReynolds number based on inside diameterReDcReynolds number based on outside diameter(include collar)ScSchmidt numberSpFin spacingTaAir temperatureTwWater temperatureTw,mMean temperature of the water filmTp,i,mMean temperature of the inner tube wallTp,o,mMean temperature of the outer tube wallTr,mMean temperature of watertFin thicknessUo,wOverall heat transfer coefficientVAverage velocityWaHumidity ratio of moist airWa,mMean air humidity ratioWs,p,o,mMean saturated air humidity ratio at the meanoutside tube wall temperatureWs,wSaturated air humidity ratio at the water filmtemperatureWs,w,mMean saturated air humidity ratio at the meanwater film temperature of the fin surfaceywThickness of condensate water filmeFin factorgf,wetWet fin efficiencylDynamic viscosityqMass density1 IntroductionThe most widely used heat exchangers take the form offin-and-tubeconfigurationinassociationwiththeapplication of air-conditioning and refrigeration sys-tems. The heat exchangers can be applicable to con-denser and evaporators. In the evaporators whichtypically use aluminum fins and the surface temperatureof the fins is generally below the dew point temperature.As a result, simultaneous heat and mass transfer occursalong the fin surfaces. In general, the complexity of themoist air flow pattern across the fin-and-tube heatexchangers under dehumidifying conditions makes thetheoretical simulations very difficult. Accordingly, it isnecessary to resort to experimentation.Many experimental studies have been carried out tostudy the heat and mass transfer characteristics of thefin-and-tube heat exchangers under dehumidifying con-ditions. For instance, McQuiston 11, 12 presentedexperimental data for five plate fin-and-tube heatexchangers, and developed a well-known heat transferand friction correlation for both dry and wet surfaces.Mirth and Ramadhyani 13, 14 investigated the heatand mass characteristics of wavy fin heat exchangers.Their results showed that the Nusselt numbers were verysensitive to change of inlet dew point temperatures, andthe Nusselt number decreases with an increase of dewpoint temperatures. Similar results were reported by Fuet al. 7 in dehumidifying heat exchangers having alouver fin configuration. They reported a pronounceddecrease of the wet sensible heat transfer coefficientswith increases of inlet relative humidity. On the con-trary, the experimental data of Seshimo et al. 19 indi-catedthattheNusseltnumberwasrelativelyindependent of inlet conditions. Wang et al. 23 study757the effect of the fin pitch, the number of tube row, andinlet relative humidity on the heat transfer performanceunder dehumidification, and concluded that the sensibleheat transfer performance is relatively independent ofinlet humidity. The difference in the existing literatures isattributed to the different reduction methodology.Even though many efforts have been devoted to thestudy of the wet-coils, the available literature on thedehumidifying heat exchangers still offers limited infor-mation to assist the designer in sizing and rating a fin-and-tube heat exchanger. This can be made clear fromthe reported data were mainly focused on the study ofthe sensible heat transfer characteristics, little attentionwas paid to the mass transfer characteristics. Therefore,the objective of the present study is to provide furthersystematic experimental information relevant to themass transfer performance and propose a new reductionmethod to determine the air-side performance of fin-and-tube heat exchangers under dehumidifying condi-tions. The effects of fin spacing and the inlet relativehumidity on the mass transfer characteristics are exam-ined in this study.2 Experimental apparatusThe schematic diagram of the experimental air circuitassembly is shown in Fig. 1. It consists of a closed-loopwind tunnel in which air is circulated by a variable speedcentrifugal fan (7.46 kW, 10 HP). The air duct is madeof galvanized sheet steel and has an 850550 mm cross-section. The dry-bulb and wet-bulb temperatures of theinlet air are controlled by an air-ventilator that canprovide a cooling capacity up to 21.12 kW (6RT). Theair flow-rate measurement station is an outlet chamberset up with multiple nozzles. This setup is based on theASHRAE 41.2 standard 3. A differential pressuretransducer is used to measure the pressure differenceacross the nozzles. The air temperatures at the inletand exit zones across the sample heat exchangers aremeasured by two psychrometric boxes based on theASHRAE 41.1 standard 2.The working medium or the tube side is cold water. Athermostatically controlled reservoir provides the coldwater at selected temperatures. The temperature differ-ences on the water side are measured by two precali-bratedRTDs.Thewatervolumetricflowrateismeasured by a magnetic flow meter with a 0.001 L/sprecision. All the temperature measuring probes areresistance temperature devices (Pt100), with a calibratedaccuracy of 0.05?C. In the experiments, only the datathat satisfy the ASHRAE 3378 1 requirements,(namely, the energy balance condition,_Qw?_Qavg?=_Qavg; is less than 0.05, where_Qwis the water-side heattransfer rate for_Qwand air-side heat transfer rate_Qa),are considered in the final analysis. Detailed geometryused for the present plain fin-and-tube heat exchangersis tabulated in Table 1. The test fin-and-tube heatexchangers are tension wrapped having a L type fincollar. The test conditions of the inlet air are as follows:The test conditions approximate those encounteredwith typical fan-coils and evaporators of air-condition-ing applications. Uncertainties reported in the presentinvestigation, following the single-sample analysis pro-posed by Moffat 15, are tabulated in Table 2.3 Data reduction3.1 Heat transfer coefficient (hc,o)Basically, the present reduction method is based on theThrelkeld 20 method. Some important reduction pro-Fig. 1 Schematic ofexperimental setupDry-bulb temperatures of the air:270.5?CInlet relative humidity for theincoming air:50% and 90%Inlet air velocity:From 0.3 m/s to 4.5 m/sInlet water temperature:70.5?CWater velocity inside the tube:1.51.7 m/s758cedures for the original Threlkeld method is described asfollows.The total heat transfer rate used in the calculation isthe mathematical average of_Qaand_Qw; namely,_Qa _ ma(ia;in? ia;out),1_Qw _ mwCp;wTw;out? Tw;in;2_Qavg_Qa_Qw2:3The overall heat transfer coefficient, Uo,w, is based onthe enthalpy potential and is given as follows:_Qavg Uo;wAoDimF;4where Dimis the mean enthalpy difference for counterflow coil,Dim ia;m? ir;m:5According to Bump 4 and Myers 16, for thecounter flow configuration, the mean enthalpy isia;m ia;inia;in? ia;outlnia;in? ir;out?ia;out? ir;in?ia;in? ia;outia;in? ir;outia;in? ir;out ? (ia;out? ir;in;6ir;m ir;outir;out? ir;inlnia;in? ir;out?ia;out? ir;in?ir;out? ir;in)(ia;in? ir;out)ia;in? ir;out) ? ia;out? ir;in;7where F in Eq. 4 is the correction factor accounting forthe present cross-flow unmixed/unmixed configuration.The overall heat transfer coefficient is related to theindividual heat transfer resistance 16 as follows:1Uo;wb0rAohiAp;ib0pAoln Dc=Di2pkpLp1ho;wAp;o.b0w;pAo?Afgf;wet.b0w;mAo?;8whereho,w1Cp;a.b0w;mhc;o? yw=kw;9ywin Eq. 9 is the thickness of the water film. Aconstant of 0.005 in. was proposed by Myers 16. Inpractice, (yw/kw) accounts for only 0.55% compared to(Cp,a/bw,mhc,o), and has often been neglected by previ-ous investigators. As a result, this term is not included inthe final analysis.In this study, we had proposed a row-by-row andtube-by-tube reduction method for detailed evaluationof the performance of fin-and-tube heat exchanger in-stead of conventional lump approach. Hence analysis ofthe fin-and-tube heat exchanger is done by dividing itinto many tiny segments (number of tube row numberof tube per row number of fin) as shown in Fig. 2. Inthe analysis, F is the correction factor accounting for asingle-pass, cross-flow heat exchanger for one fluidmixed, other fluid unmixed that was shown by Threlkeld20.The tube-side heat transfer coefficient, hievaluatedwith the Gnielinski correlation 8,Fig. 2 Dividing of the fin-and-tube heat exchanger into the smallpiecesTable 2 Summary of estimated uncertaintiesPrimary measurementsDerived quantitiesParameterUncertaintyParameterUncertaintyReDc=400UncertaintyReDc=5,000_ ma0.31%ReDc1.0%0.57%_ mw0.5%ReDi0.73%0.73%DP0.5%_Qw3.95%1.22%Tw0.05?C_Qa5.5%2.4%Ta0.1?Cj11.4%5.9%Table 1 Geometric dimensions of the sample plain fin-and-tubeheat exchangersNo.Fin thickness(mm)Sp(mm)Dc(mm)Pt(mm)Pl(mm)Rowno.10.1151.088.5125.419.05120.1201.6310.3425.422.00130.1151.938.5125.419.05140.1152.1210.2325.419.05150.1202.3810.3425.422.00160.1151.128.5125.419.05270.1201.588.6225.419.05280.1151.958.5125.419.05290.1203.018.6225.419.052100.1302.1110.2325.422.002110.1151.1210.2325.419.054120.1151.4410.2325.419.054130.1152.2010.2325.419.054140.1302.1010.2325.422.004150.1301.7210.2325.422.006160.1302.0810.2325.422.006170.1303.0310.2325.422.006759hifi=2ReDi? 1000Pr1:07 12:7ffiffiffiffiffiffiffiffifi=2pPr2=3? 1?kiDi;10and the friction factor, fiisfi11:58ln ReDi? 3:282:11The Reynolds number used in Eqs. 10 and 11 is basedon the inside diameter of the tube and ReDi qVDi=l:In all case, the water side resistance is less than 10% ofthe overall resistance.In Eq. 8 there are four quantities (bw,m, bw,p, bpandbr) involving enthalpy-temperature ratios that must beevaluated. The quantities of bpand brcan be calculatedasb0ris;p;i;m? ir;mTp;i;m? Tr;m;12b0pis;p;o;m? is;p;i;mTp;o;m? Tp;i;m:13The values of bw,pand bw,mare the slopes of satu-rated enthalpy curve evaluated at the outer mean waterfilm temperature at the base surface and at the fin sur-face. Without loss of generality, bw,pcan be approxi-matedbytheslopeofsaturatedenthalpycurveevaluated at the base surface temperature 23. The wetfin efficiency (gf,wet) is based on the enthalpy differenceproposed by Threlkeld 20. i.e.,gf,weti ? is,fmi ? is,fb;14where is,fmis the saturated air enthalpy at the meantemperature of fin and is,fbis the saturated air enthalpyat the fin base temperature. The use of the enthalpypotential equation, greatly simplifies the fin efficiencycalculation as illustrated by Kandlikar 10. However,the original formulation of the wet fin efficiency byThrelkeld20wasforstraightfinconfiguration(Fig. 2a). For a circular fin (Fig. 2b), the wet finefficiency is 23,gf;wet2riMT(r2o? r2i)?K1(MTri)I1(MTro) ? K1(MTro)I1(MTri)K1(MTro)I0(MTri) K0(MTri)I1(MTro)?;15whereMTffiffiffiffiffiffiffiffiffiffiffi2ho;wkftr;16The test heat exchangers are of Fig. 3c configura-tion. Hence, the corresponding fin efficiency is calcu-latedbytheequivalentcircularareamethodasdepicted in Fig. 4.Evaluation of bw,mrequires a trial and error proce-dure. For the trial and error procedure, is,w,mmust becalculated using the following equation:is;w;m ia;m?Cp;aho;wgf;wetb0w;mhc;o?1 ? Uo;wAob0rhiAp;ib0pln Dc=Di2pkpLp# !? ia;m? ir;m:17An algorithm for solving the sensible heat transfercoefficient hc,ofor the present row-by-row and tube-by-tube approach is given as follows:1. Based on the measurement information, calculate thetotal heat transfer rate_Qtotalusing Eq. (3).2. Assume a hc,ofor all elements.3. Calculate the heat transfer performance for eachsegment with the following procedures.3.1. Calculate the tube side heat transfer coefficient ofhiusing Eq. 10.3.2. Assume an outlet air enthalpy of the calculatedsegment.3.3. Calculate ia,mby Eq. 6 and ir,mby Eq. 7.3.4. Assume Tp,i,mand Tp,o,m.3.5. Calculate b0rAo?= hiAp;i?andb0pAoln Dc=Dihi=h2pkpLp?.3.6. Assume a Tw,m.3.7. Calculate the gf,wetusing Eq. 15.3.8. Calculate Uo,wfrom Eq. 8.3.9. Calculate is,w,mby Eq. 17.3.10. Calculate Tw,mfrom is,w,m.Fig. 3 Type of fin configurationFig. 4 Approximation method for treating a plate fin of uniformthickness7603.11. If Tw,mderived in step 3.10 is not equal that isassumed in step 3.6, the calculation step 3.73.10 will be repeated with Tw,mderived in step3.10 until Tw,mis constant.3.12. Calculate_Q of this segment.3.13. Calculate Tp,i,mand Tp,o,mfrom the insideconvection heat transfer and the conductionheat transfer of tube and collar.3.14. If Tp,i,mand Tp,o,mderived in step 3.13 are notequal that is assumed in step 3.4, the calculationstep 3.53.13 will be repeated with Tp,i,mandTp,o,mderived in step 3.13 until Tp,i,mand Tp,o,mare constant.3.15. Calculate the outlet air enthalpy by Eq. 1 andthe outlet water temperature by Eq. 2.3.16. If the outlet air enthalpy derived in step 3.15 isnot equal that is assumed in step 3.2, the cal-culation step 3.33.15 will be repeated with theoutlet air enthalpy derived in step 3.15 until theoutlet air enthalpy is constant.4. If the summation of_Q for all elements is not equal_Qtotal, hc,owill be assumed a new value and the cal-culation step 3 will be repeated until the summationof_Q for all elements is equal_Qtotal.3.2 Mass transfer coefficient (hd,o)For the cooling and dehumidifying of moist air by a coldsurface involves simultaneously heat and mass transfer,and can be described by the process line equation fromThrelkeld 20:diadWa Ria? is;wWa? Ws;w ig? 2;501R;18Where R represent the ratio of sensible heat transfercharacteristics to the mass transfer performance.R hc;ohd;oCp;a:19However, for the present fin-and-tube heat ex-changer, Eq. 18 did not correctly describe the dehu-midification process on the psychrometric chart. This isbecause the saturated air enthalpy (is,w) at the meantemperature at the fin surface is different from that at thefin base. In this regard, a modification of the process lineon the psychrometric chart corresponding to the fin-and-tube heat exchanger is made. The derivation is as fol-lows.From the energy balance of the dehumidification onecan arrive at the following expression:_ madiahc;oCp;adAp;oia;m? is;p;o;m hc;oCp;adAfia;m? is;w;m:20Note that the first term on the right-hand side de-notes the sensible heat transfer whereas the second termis the latent heat transfer. Conservation of the watercondensate gives:_ madWa hd;odAp;oWa;m? Ws;p;o;m hd;odAfWa;m? Ws;w;m:21Dividing Eq. 20 by Eq. 21 yieldsdiadWaR ? ia;m? is;p;o;m R ? e ? 1 ? ia;m? is;w;mWa;m? Ws;p;o;m e ? 1 ? Wa;m? Ws;w;m;22wheree AoAp;o:23By assuming a value of the ratio of heat transfer tomass transfer, R and by integrating Eq. 22 with aniterative algorithm, the mass transfer coefficient can beobtained. Analogous procedures for obtaining the masstransfer coefficients are given as:1. Obtain Ws,p,o,mand Ws,w,mfrom is,p,o,mand is,w,mfrom those calculation of heat transfer.2. Assume a value of R.3. Calculations is performed from the first element tothe last element, employing the following procedures:3.1. Assume an outlet air humidity ratio.3.2. Calculate the outlet air humidity ratio of eachelement by Eq. 22.3.3. If the outlet air humidity ratio obtained fromstep 3.2 is not equal to the assumed value of step3.1, the calculation steps 3.1 and 3.2 will be re-peated.4. If the summation of the outlet air humidity ratio foreach element of the last row is not equal to themeasured outlet air humidity ratio, assuming a new Rvalue and the calculation step 3 will be repeated untilthe summation of the outlet air humidity ratio of thelast row is equal to the measured outlet air humidityratio.3.3 Chilton-Colburn j-factor for heat and mass transfer(jhand jm)The heat and mass transfer characteristics of the heatexchanger is presented by the following non-dimensionalgroup:jhhc;oGmaxCp;aPr2=3;24jmhd;oGmaxSc2=3:257614 Results and discussionsHeat transfer performance of the fin-and-tube heatexchangers is in terms of dimensionless parameter jh. Atypical plot for examination of the influence of fin pitchis shown in Fig. 5. In this figure, the reduced results bythe present tube-by-tube method and those by the ori-ginal Threlkeld method having N=2 is shown. For heattransfer performance, reduced results from both meth-ods are nearly the same. This is somehow expected be-cause the present tube-by-tube approach is originatedfrom the Threlkeld method. From the results, one cansee that the heat transfer performance is relativelyinsensitive to the fin pitch. Notice that this phenomenonis quite different from that tested in fully dry conditions.As reported by Wang et al. 22 and Rich 17, the heattransfer performance is independent of fin pitch whenN 4 operated at fully dry conditions. However, forN=1 or 2, Wang and Chi 21 reported that theheat transfer performance drops with the increase offinspacing.ThisisespeciallypronouncedwhenReDc5,000. For ReDc5,000, theheat transfer performance increases with decrease of finpitch. This phenomenon is seen for N 2, and is espe-cially pronounced for N=1. By contrast, the presentsensible heat transfer performance exhibits a compara-tively insensitive influence to the change of fin spacingfor N=1 and 2. Apparently, the results are attributed tothe presence of condensate under dehumidification. Thisis because the appearance of condensate plays a role toalter the airflow pattern, roughening the fin surface andproviding a better mixing of the airflow. As a conse-quence, the influence of fin pitch is reduced accordingly.This phenomenon is analogous to using the enhanced finsurface in fully dry condition. For enhanced surfacessuch as slit and louver fin geometry, Du and Wang 5and Wang et al. 24, 25 reported a negligible effect of finpitch even for N=1 or 2.Mass transfer performance of the present dehumidi-fying coils is termed as dimensionless jmfactor. Forexamination of the influence of inlet humidity on themass transfer characteristics between the present methodand that of original Threlkeld method, a typical com-parison for sample no. 5 and 10 is illustrated in Fig. 6.As seen in the figure, results using the present tube-by-tube method show relatively small influence of the inletrelative humidity. This is applicable for both 1-row and2-row configuration. By contrast, for the reduced resultsby the original Threlkeld method, one can see about 2040% increase of mass transfer performance when theinlet relative humidity is increased from 50% to 90%.For the heat transfer performance, as aforementionedpreviously, the effect of inlet relative humidity is almostnegligible regardless the reduction method is chosen.Hence, it is expected that the associated influence on themass transfer performance is also small. With the ori-ginal procedures of Threlkeld method that was appli-cable to the counter-cross flow arrangement and ofexclusive of the effect of primary surface, the reducedresults are somewhat misleading. Hence the presenttube-by-tube method is more appropriate than the ori-ginal procedures of Threlkeld method in reducing themass transfer coefficient under fully wet conditions. Thedeparture of the reduced results between Threlkeldmethod and the present method increases with the masstransfer rate. This can be made clear from Fig. 7 with aFig. 5 Effect of the fin pitch on jhbetween those derived byThrelkeld method and by present methodFig. 6 Effect of the inlet relative humidity on jmbetween thosederived by Threlkeld method and by present method for samplesno. 5 and 10762very close fin spacing of 1.08 mm. As seen in Fig. 7 atReDc1,000, the results indicate a departure of the re-duced results for more than 50% between these twomethods. Moreover, there is negligible influence of inlethumidity for the present method when ReDc1,000 when RH=50%.This is in connection with the blow-offof condensate atlarger ReDcwhich make more zoom for water vapor tocondensate along the surface and even resultis in apartially dry consitions due to the rise of dew pointtemperature.Thisphenomenonbecomeslesspro-nounced with the rise of the number of tube row forcondensate blow-offmay be blocked by the subsequenttube row.The dehumidifying process involves heat and masstransfersimultaneously,ifmasstransferdataareunavailable, it is convenient to employ the analogy be-tween heat and mass transfer. The existence of the heatand mass analogy is because the fact that conductionand diffusion in a liquid are governed by physical laws ofidentical mathematical form. Therefore, for air-watervapor mixture, the ratio of hc,o/hd,oCp,ais generallyaround unity, i.e.,hc;ohd;oCp;a? 1:26The term in Eq. 19 approximately equals to unity fordilute mixtures like water vapor in air near the atmo-spheric pressure (temperature well-below correspondingboiling point). The validity of Eq. 26 relies heavily onthe mass transfer rate. The experimental data of Hongand Webb 9 indicated that this value is between 0.7 and1.1, Seshimo et al. 19 gave a value of 1.1. Eckels andRabas 6 also reported a similar value of 1.11.2 fortheir test results of fin-and-tube heat exchangers havingplain fin geometry. The aforementioned studies allshowed the applicability of Eq. 26. In the present study,we notice that the values of hc,o/hd,oCp,awere generallybetween 0.6 and 1.0. There are two differences betweenthe original Threlkeld method and the present row-by-row and tube-by-tube approach. First, larger deviationoccurs via using the original Threlkeleds methods. Thisis associated with the considerable influence of inlethumidity of the original Threlkelds method. For thepresent reduction method, the ratio is insensitive tochange of inlet humidity provided that the surface isfully wet. Second, reduction by the present methodindicates that the ratio of hc,o/hd,oCp,aslightly decreaseswith the Reynolds number whereas the original Threl-keld method shows the opposite trend (slightly increasewith the Reynolds number). As aforementioned in pre-vious section, with the rise of inlet flow inertia, thecondensate can be easily removed for providing moreroom for further condensation. The condensate removalbecomes even pronounced with smaller fin spacing. Asshown in Figs. 8 and 9 the condensate retention becomesmore severe when the fin spacing is reduced. In thatregard, the removal of condensate subject to larger flowinertia help to improve the mass transfer performanceconsiderably once the condensate retention phenomenonis eliminated at higher flow inertia. Therefore, one cansee the ratio of hc,o/hd,oCp,ais slightly decreased with thefin spacing. Notice that the effect of fin spacing on theratio of hc,o/hd,oCp,ais only effective when Reynoldsnumber is sufficient high. This is associated with the highair flow rate would increase the vapor shear, and wipeaway the condensate. Conversely, the effect of fin spac-ing on the ratio of hc,o/hd,oCp,ais comparatively small atlow Reynolds number when the condensate retention iscomparatively severe.It is obvious from the shown test results that no singlecurve can be expected to describe the complex behaviorsfor both jhand jmfactors. As a result, using a multiplelinear regression technique in a practical range ofFig. 7 Effect of the inlet relative humidity on jmbetween thosederived by Threlkeld method and by present method for sample no.1Fig. 8 Effect of the fin spacing on R for one row763experimental data (300ReDc5,500), the appropriatecorrelation form of jhand jmfor the present data arejh 1:49N?0:0575Re0:002061N?0:625Dc0:00583N 0:825e?0:001921N0:068;27jm 0:84N?0:428Re0:0219N?0:553Dc?0:104N 0:769e0:107N0:01132;28hc;ohd;oCp;a 1:54N0:343Re?0:0196N?0:0664Dc;29wheree AoAp,o:30As shown in Figs. 10, 11, and 12, Eq. 27 can describe88.9% of the jhfactors within 15%, Eq. 28 can correlate81.2% of the jmfactors within 20%, and Eq. 29 can cor-relate85.5%oftheratioofhc,o/hd,oCp,atobewithin20%.5 ConclusionsThis study examines the heat and mass characteristics of17 fin-and-tube heat exchangers having plain fin geom-etry experimentally. On the basis of previous discus-sions, the following conclusions are made:1. A tube-by-tube reduction method based on theThrelkeld method is proposed in this study forreducing the test results. For fully wet conditions, it isfound the reduced results for both sensible heattransfer performance and the mass transfer perfor-mance by the present method is insensitive to changeof inlet humidity. The original Threlkeld method alsoshows negligible influence of inlet humidity for heattransferperformancebutindicatesadetectableinfluence of inlet humidity on the mass transfercharacteristics.2. Unlike those tested in fully dry condition, the sensibleheat transfer performance under dehumidification iscomparatively independent of fin pitch. This is be-cause the presence of condensate plays a role inaltering the air flow pattern within the heat ex-changer, resulting in better mixing characteristics.3. The ratio of hc,o/hd,oCp,ais in the range of 0.61.0and is insensitive to change of fin spacing at lowReynolds number. However, a slight drop of the ratioof hc,o/hd,oCp,ais seen with the decrease of fin spacingwhen the Reynolds number is sufficient high. This isFig. 10 Comparison of jhbetween those derived by experimentaldata and by correlationFig. 9 Effect of the fin spacing on R for two rowsFig. 11 Comparison of jmbetween those derived by experimentaldata and by correlation764associated with the more pronounced influence due tocondensate removal.4. A correlation is proposed for the present plate finconfiguration. This correlation can describe 88.9% ofthe jhfactors within 15% and can correlate 81.2% ofthe jmfactors within 20%.Acknowledgments The authors are indebted to the Thailand Re-search Fund (TRF) and the Energy R&D foundation funding fromthe Bureau of Energy of the Ministry of Economic Affairs, Taiwanfor supporting this study.References1. ASHRAE Standard 3378 (1978) Method of testing forcedcirculation air cooling and air heating coils. American Societyof Heating, Refrigerating and Air-Conditioning Engineers,Atlanta2. ASHRAE Standard 41.1-1986 (1986) Standard method fortemperaturemeasurement.AmericanSocietyofHeating,Refrigerating and Air-Conditioning Engineers, Atlanta3. ASHRAE Standard 41.2-1987 (1987) Standard methods forlaboratory air-flow measurement. American Society of Heat-ing, Refrigerating and Air-Conditioning Engineers, Atlanta4. BumpTR(
- 温馨提示:
1: 本站所有资源如无特殊说明，都需要本地电脑安装OFFICE2007和PDF阅读器。图纸软件为CAD,CAXA,PROE,UG,SolidWorks等.压缩文件请下载最新的WinRAR软件解压。
2: 本站的文档不包含任何第三方提供的附件图纸等，如果需要附件，请联系上传者。文件的所有权益归上传用户所有。
3.本站RAR压缩包中若带图纸，网页内容里面会有图纸预览，若没有图纸预览就没有图纸。
4. 未经权益所有人同意不得将文件中的内容挪作商业或盈利用途。
5. 人人文库网仅提供信息存储空间，仅对用户上传内容的表现方式做保护处理，对用户上传分享的文档内容本身不做任何修改或编辑，并不能对任何下载内容负责。
6. 下载文件中如有侵权或不适当内容，请与我们联系，我们立即纠正。
7. 本站不保证下载资源的准确性、安全性和完整性, 同时也不承担用户因使用这些下载资源对自己和他人造成任何形式的伤害或损失。
人人文库网所有资源均是用户自行上传分享，仅供网友学习交流，未经上传用户书面授权，请勿作他用。
川公网安备: 51019002004831号