35KWCO2跨临界热泵机组设计【含CAD图纸+PDF文档】
收藏
资源目录
压缩包内文档预览：
编号：209825481
类型：共享资源
大小：9.17MB
格式：ZIP
上传时间：2022-04-27
上传人：机****料
认证信息
个人认证
高**（实名认证）
河南
IP属地：河南
50
积分
- 关 键 词：
-
含CAD图纸+PDF文档
35
KWCO2
临界
机组
设计
CAD
图纸
PDF
文档
- 资源描述：
-
35KWCO2跨临界热泵机组设计【含CAD图纸+PDF文档】,含CAD图纸+PDF文档,35,KWCO2,临界,机组,设计,CAD,图纸,PDF,文档
- 内容简介：
-
Water-coupled carbon dioxide microchannel gas coolerfor heat pump water heaters: Part I - ExperimentsBrian M. Fronka, Srinivas Garimellaa,*aGeorge W. Woodruff School of Mechanical Engineering, Georgia Institute of Technology, Atlanta, GA 30332-0405, USAa r t i c l e i n f oArticle history:Received 13 January 2010Received in revised form15 April 2010Accepted 2 May 2010Available online 21 May 2010Keywords:Heat pumpHeatingHot waterCarbon dioxideTranscritical cycleExperimentGas coolerMicrochannela b s t r a c tAn experimental and analytical study on the performance of a compact, microchannelwater- carbon dioxide (CO2) gas cooler was conducted. The gas cooler design underinvestigation used an array of serpentine refrigerant microchannel tubes wrapped aroundwater passages containing offset strip fins, resulting in a generally counterflow configu-ration between the two fluids. Part I of this two-part paper addresses the experimentalaspects. Data were obtained using an experimental heat pump facility at varying inletconditions for three gas coolers of the same design, but different sizes. Measured heatingcapacity for the three gas coolers ranged from 1.5 to 6.5 kW. The results of this study areused in the companion paper (Part II) to develop a predictive heat exchanger model tooptimize gas cooler design over a wide range of operating conditions, eliminating the needfor expensive prototype development and testing. 2010 Elsevier Ltd and IIR.Refroidisseur de gaz a microcanaux au dioxyde de carbonepour les pompes a chaleur utilise es pour chauffer de leausanitaire : Partie I - Expe riencesMots cle s : Pompe a chaleur ; Chauffage ; Eau sanitaire ; Dioxyde de carbone ; Cycle transcritique ; Expe rimentation ; Refroidisseur de gaz ;Microcanal* Corresponding author. Tel.: 1 404 894 7479; fax: 1 404 894 8496.E-mail address: sgarimella (S. Garimella).www.iifiavailable at journal homepage: /locate/ijrefriginternational journal of refrigeration 34 (2011) 7e160140-7007/$ e see front matter 2010 Elsevier Ltd and IIR.doi:10.1016/j.ijrefrig.2010.05.0041.IntroductionAs the recognition of human-induced global climate changeincreases, nontoxic, naturally occurring, widely available andinexpensive carbon dioxide (CO2), with a global warmingpotential (GWP) of 1 and favorable transport properties, hasemerged as an attractive replacement for hydroflurocarbons(HFC). With a relatively low critical temperature (31.1?C) andpressure (73.7 bar), CO2is often a supercritical fluid on thehigh-side of a vapor compression cycle under normal oper-ating conditions, resulting in a transcritical cycle. While theabsolute pressures at which the transcritical cycle operatesare much higher than those of conventional refrigerants, thepressure ratio is much lower, leading to potentially highercompressor efficiencies. The volumetric heat capacity of CO2is five to eight times higher than that of conventional refrig-erants (Groll and Garimella, 2000), allowing for more compactequipment and systems. In supercritical operation, instead ofa constant temperature condensation process, the refrigerantis cooled from a vapor-like state to a liquid-like state in thecomponent now known as a gas cooler. The temperatureprofile associated with this non- isothermal cooling is illus-trated in a T-h plot shown in Fig. 1 from points 3 to 4.Groll and Kim (2007) provide a comprehensive review ofCO2transcriticalcycletechnology.Theyidentifymanyadvanced technologies that can be incorporated into the cycleto increase system efficiency in a variety of applicationsincluding mobile air conditioning, residential AC/heat pumpsystems and military environmental control units (ECUs).They also identified heating of water as one of the mostpromisingapplicationsoftheCO2transcriticalcycle.Commercial systems are already on the market in Japan andprovide substantial energy savings over fossil-fired systemswith COPs in the range of 3e4 (Kim et al., 2004). Water heatingrequires outlet water temperatures of 70e90?C. CO2heatpump systems have been shown to provide water up to 90?Cwithout operational problems or major losses in system effi-ciency (Kim et al., 2004). The large temperature glide in theheating of tap water matches well with the supercriticaltemperature glide of CO2. Unlike in a condensation process,here the non- isothermal heat rejection can be used toadvantage in a counterflow gas cooler, in which the wateroutlet temperature can rise to the desired high value. Thisminimizes temperature “pinch” and keeps gas cooler sizeeconomical. Fig. 2 shows water being heated by CO2andR134a. A pinch effect is observed in the R134a case, whena water outlet temperature of 70?C is desired. Increasing thehigh-side pressure of the R134a system would allow higherwater delivery temperatures; however, the correspondingnarrowing of the vapor dome and increased pressure ratiowould be detrimental to system performance. The tempera-ture profile of the CO2, on the other hand, matches well withthe high temperature lift required by the water.The gas cooler requires special design considerations dueto the high operating temperature and the temperature glideduring supercritical cooling of CO2. To achieve maximumsystem COP, the gas cooler must be designed in such a way asto minimize the approach temperature between the heat sinkand refrigerant. As CO2gas cooler design moves froma conventional tube and fin geometry to more compactmicrochannel designs, which allow high operating pressures(in excess of 120 bar) and improved heat transfer coefficients,experimental investigations and models for heat transfer andpressure drop of supercritical CO2in these heat exchangergeometries are required.Many experimental and analytical studies have been con-ducted on the performance of transcritical CO2cycles forwater heating and space conditioning. Using a tube-in-tubegas cooler design, Neksa et al. (1998) demonstrated COPs up to4.3 when heating water from 9 to 60?C. Richter et al. (2003)developed a prototype split air-to-air CO2system intendedfor residential heat pump applications. The system wasdesigned to fit in the footprint of a commercial R410a systemand used an aluminum microchannel heat exchanger for thegas cooler component. They compared the experimentalsystem to a commercial R410a system and noted that theincreasedcapacityoftheCO2systematlowambienttemperatures resulted in higher annual heating efficiency.Ortiz et al. (2003) developed and validated a detailed simula-tion program for characterizing the performance of air-to-airCO2systems. The air-coupled evaporator and gas cooler wereboth extruded aluminum microchannel designs (D 1 mm).Fig. 1 e Transcritical cycle T-h diagram.NomenclatureCvValve flow coefficient (-)GMass flux (kg m?2s?1)hSpecific enthalpy (kJ kg?1)_ mMass flow rate (g s?1)_QHeat Duty (kW)TTemperature (?C)UxUncertainty of measured variableUyUncertainty of calculated variableinternational journal of refrigeration 34 (2011) 7e168The predicted heat duties of the air-coupled heat exchangerswerewithintheexperimentalerror.Thepredictedcompressor power and system COP were within 5% of exper-imental values. The validated model was used to run a varietyof operational case studies including different compressordesigns, heating and cooling mode, seasonal variations andthe inclusion of a suction line heat exchanger and workrecovery expansion devices. In general it was found thatimproving compressor efficiency was the most critical factorfor making CO2systems competitive with R410a split systems.Laipradit et al. (2008) developed a simulation program of an airsource CO2heat pump for water heating. System performancewas evaluated under varying gas cooler water inlet tempera-ture, compressor speed and ambient evaporator air condi-tions. The gas cooler was assumed to be a tube-in-tube design.Sarkar et al. (2006; 2009) developed and validated a simulationmodel for simultaneous water heating and cooling. Similar toprevious studies, the water-coupled gas cooler was a tube-in-tube design, with an inner tube OD of 6.35 mm and thicknessof0.8 mm,outertube ODof12mmandthicknessof1 mm.Therefrigerant heat transfer coefficient during gas cooling wasmodeled using the correlation developed by Pitla et al. (2002).The model showed a maximum deviation of 15% fromexperimental results.In addition to system level simulations and results, otherstudies have focused exclusively on the performance of thegas cooler component. Yin et al. (2001) conducted an experi-mental study on an air-coupled microchannel gas cooler anddeveloped a model for predicting the pressure drop and heattransfer performance of the heat exchanger. The gas coolerconsisted of three refrigerant passes of 13, 11 and 10 tubes.Each tube had 11 channels of 0.79 mm diameter. The modelwas compared with 48 data points with varying refrigerantand air flow rates, temperatures and pressures. The modelagreed with the measured capacity within 2%, and withmeasured outlet refrigeration temperature within 0.5?C. Bothof thesevalueswerewithinthe experimental error.The modelsystematically under predicted pressure drop for every datapoint. After adjusting the model to account for ports blockedor deformed during manufacturing, the data were found to bein good agreement (Yin et al., 2001). The authors state that thesensitivity of pressure drop to channel diameter and massflux,coupledwiththeuncertaintyinmanufacturingmicrochannel heat exchangers, make it difficult to predict gascooler pressure drop for very small channel diameters.Zhao and Ohadi (2004) also conducted an experimentalstudy of an air-coupled microchannel gas cooler. The gascooler considered used microchannel tubes with a hydraulicdiameter of 1.0 mm. The gas cooler was composed of severalmicrochannel “slabs,” each with a refrigerant-side heattransfer area of 0.46 m2. Two parallel rows of five slabs areconnected in series. Tests were conducted at refrigerant massflow rates from 15 to 40 g s?1, refrigerant inlet pressurebetween69and 125bar and refrigerantinlettemperaturefrom79 to 120?C. The air inlet temperature was set at 21?C and themass flow at 520 g s?1. Experimental heat duties ranged from 4to 8 kW, with air and refrigerant energy balances within ?3%.Hwang et al. (2005) conducted a performance evaluation ofan air-coupled gas cooler similar to that of Zhao and Ohadi(2004). Instead of a microchannel heat exchanger, a moreconventional tube and fin heat exchanger with tube ID of7.9 mm was tested. The heat exchanger had 3 rows of 18 tubesin cross flow with the incoming air. Air inlet temperatureswere set at 29.4 and 35?C with frontal velocities of 1.0, 2.0 and3.0 m s?1. Refrigerant mass flow was set at 38 and 76 g s?1withgas cooler inlet pressures of 90, 100 and 110 bar. The refrig-erant inlet temperature to the gas cooler was not fixed andwas allowed to vary with high-side pressure and other systemoperating conditions. The heating capacity of the gas coolerranged from 6 to 14 kW. For every refrigerant mass flow andpressure, the capacity increased as the frontal air velocityincreased. However at the 38 g s?1refrigerant flow condition,the air side showed signs of temperature pinch as the velocityincreased from 2.0 to 3.0 m s?1. With the fixed gas cooler size,higher average approach temperature differences were seenat the higher refrigerant mass flow rates. Due to the higherrefrigerant outlet temperature, specific enthalpy differencesacross the gas cooler for the 76 g s?1cases were seen to be57e81%ofthoseat 38g s?1. Bydoublingthemassflowrate,theheating capacity increased by 14e62%.Much research has been conducted on transcritical CO2cycles for heating, cooling and water heating at the systemlevel; however, less attention has been devoted to compact,microchannel gas coolers specificallyfor water heatingapplications. Much of the experimental gas cooler work hasfocused on development of air-coupled heat exchangers forFig. 2 e CO2vs. R134a temperature pinch ernational journal of refrigeration 34 (2011) 7e169use in automotive and space conditioning applications.However, in these heat exchangers, the air-side resistance isthe limiting factor, and the contribution of the CO2to heattransfer performance is less important than would be expec-ted in a water-coupled gas cooler. Thus, there is a need forfurther experimentation on compact, water-coupled heatexchangers. A counterflow gas cooler is the key enablingcomponent to take advantage of the unique water heatingcapabilities of transcritical CO2cycles. By incorporatinga compact, microchannel gas cooler such as the one in thepresent study, the footprint of water heating systems can beminimized, which is of great importance in portable waterheatingapplications, orwhenusedtoheat asecondaryfluidinmobile applications. A detailed study of the heat transfermechanisms of a water-coupled gas cooler is therefore con-ducted in this two-part study.In the present work, an experimental transcritical heatpump system is developed to evaluate the heat duty andpressure drop of a compact, microchannel, water-CO2gascooler under a wide array of test conditions. The gas cooler isa cross-counter flow configuration, in which water makesseveral cross flow passes over a bank of serpentine refrigeranttubes. The experimental aspects and the correspondingresults are presented here in Part I. Part II focuses on theanalytical model development based on these results.2.Experimental approach and data analysisThe primary focus of the current study is on the heat transferand pressure drop performance of a compact water-coupledgas cooler. The gas cooler is a prototype aluminum brazed-plate heat exchanger developed by Modine ManufacturingCompany. Refrigerant flows through microchannel tubes,which are wrapped around a series of water plates. Waterflows through each plate of the heat exchanger through a fin-ned passage, with the total flow proceeding in a generallycounterflow orientation with respect to the refrigerant. Aphotograph and cross section schematic of two of the gascoolers under investigation are shown in Fig. 3. Importantdimensionsof the seven-plate gas cooler are shownin Table 1.The cross section in Fig. 3 shows a gas cooler consisting of fiveplates, that is, four refrigerant passes and five water channels.Refrigerant flow is depicted proceeding from bottom to top ina serpentine fashion, while the water flow is shown alter-nating in and out of the page. Each water-side pass containsan offset strip fin insert for structural stability and enhancedheat transfer. The refrigerant-side of the gas cooler consists of16 tubes brazed together resulting in a row of 64 circularmicrochannels with a diameter of 0.89 mm Fig. 4 showsa schematic of a single refrigerant tube cross section anda section of water channel fin insert with relevant dimensionsincluded.Experimentswereconductedonthreedifferentgascoolers: a five, seven and simulated twelve-plate gas cooler.The five-plate gas cooler has four refrigerant passes and fivewater passages, while the seven-plate has six refrigerantpasses and seven water passages. A simulated twelve-plategas cooler was tested by connecting the five and seven-plategas coolers in series. Thus, this gas cooler had twelve totalwater passes, but only ten refrigerant tube passes. All threegas cooler configurations had the same fin and microchanneltube dimensions, as shownin Table 1, with the only differencebetween the gas coolers being the addition of more waterplates and longer refrigerant tubes.The gas coolers were tested in a prototype CO2heat pumpsystem constructed specifically for this study. The test facilitywas designed to be flexible, allowing different gas cooler, evap-orator and compressor designs to be easily moved in and out ofthe system. An overall system schematic is shown in Fig. 5.A brazed-plate water-coupled heat exchanger served asthe evaporator. Like the gas cooler, the evaporator is of allaluminum construction, with extruded microchannel refrig-erant tubes, and an offset strip fin insert on the water-side.However, instead of the cross-counter flow geometry of thegas cooler, the evaporator is a one-pass cross flow device. Inthis evaporator, the water enters one side of the heatexchanger, splits into seven water channels and makes onepass across the bank of refrigerant tubes, exiting on the sameside it entered.Fig. 3 e Gas cooler photograph and ernational journal of refrigeration 34 (2011) 7e1610To provide the required refrigerant mass flow rates, twodifferentcompressorarrangementswereused.Formassflowsbetween 8 and 13 g s?1, a single reciprocating compressor wasused. For mass flows from 16 to 24 g s?1, two reciprocatingcompressors were run in parallel. Each compressor ran on 120VAC at 60 Hz. The compressors used in the parallel setup wereof a newer design than that used in the single compressorsetup. The compressors were early versions of the DanfossTN1416 compressor. Each of the newer compressors featuredsuction and discharge mufflers and different suction anddischarge line connections. The other characteristics of thecompressorswerethe same.Basedon recommendationsfromthe supplier, the compressor was cooled by ambient air ata measured velocity between 2.0 and 4.6 m s?1. Cooling wasprovided by an external fan, manually controlled by theoperators. Air flow was varied to maintain motor casetemperature below 70?C for all operating points. Eachcompressor was lubricated by a minimum charge of 80 mL ofpolyolester glycol (POE) oil. The compressor power in thesingle and dual compressor setups was measured with a wattmeter with a range from 0 to 4 kW and an uncertainty of?0.02 kW. To prevent liquid refrigerant from entering thecompressor, a specially designed high pressure “U-tube”accumulator was installed between the evaporator and thecompressor suction line. The expansion device was a manualmeteringvalvewithastainlesssteelvalvebodyanda maximum pressure rating of 344 bar. The valve Cvrangedfrom0 at fullclosedto 0.040at tenturns openin anearlylinearfashion.The maximum operating pressure of the refrigerant loopwas 120 bar at a temperature of 125?C. To safely contain thesepressures, seamless stainless steel tubing of outer diameter6.35 mm and wall thickness 0.89 mm was used for a majorportion of the refrigerant lines. To reduce heat transfer to theambient, all tubing on the high temperature side of the systemwas insulated with 9.53 mm thick silicon foam tube insulationwith 6.35 mm inner diameter and a thermal conductivity of0.056 W m?1K?1, capable of withstanding temperatures up to260?C. All lower temperature tubing was insulated with9.53 mm thick neoprene foam rubber tube insulation with6.35 mm inner diameter and a thermal conductivity of0.035 W m?1K?1. Irregular shapes, such as tees, valves andother fittings were wrapped with 12.7 mm thick fiberglasswrapinsulationwithathermalconductivityof0.039 W m?1K?1.Refrigerant pressuremeasurements at the gas cooler outletand compressor outlet were taken with Setra (P/N:206/207)pressure transducers. The pressure transducer range wasfrom 0 to 207 bar with an uncertainty of ?0.275 bar. The low-side pressuremeasurements at the evaporator inlet and outletused a Setra (P/N: 206/207) with a range from 0 to 69 bar and anuncertainty of ?0.090 bar. All pressure transducers weremounted inverted, minimizing the possibility of lubricantsettling on the diaphragm and affecting the reading. Differ-ential pressures across the gas cooler and evaporator weremeasured with Rosemount differential pressure transducers.Fig. 4 e Microchannel tube and water fin schematic.Table 1 e Seven-plate gas cooler dimensions.Overall DimensionsNumber of refrigerant tubes16Number of water passes7Gas cooler length191 mmGas cooler width54.0 mmGas cooler height84.0 mmRefrigerant-SideTotal refrigerant tube length516 mmRefrigerant pass tube length81.0 mmNumber of channels per tube4Channel diameter0.89 mmTube width6.35 mmTube height1.65 mmTube wall thickness0.38 mmTube web thickness0.64 mmRefrigerant-side heat transfer area92,292 mm2Water-SideFin height6.41 mmFin space2.23 mmFin thickness0.31 mmFin length3.18 mmFin pitch4.4 fins per cmWater-side heat transfer area385,140 mm2Fig. 5 e Test facility ernational journal of refrigeration 34 (2011) 7e1611The transducer across the gas cooler (P/N: 3051CD4) hada range from 0 to 20.7 bar with an uncertainty of ?0.005 bar inthe set span. The evaporator pressure transducer (P/N:3051CD5) had a range of 0e137.9 bar with an uncertainty of?0.0076 bar.TherefrigerantmassflowratewasmeasuredwithaMicromotionCoriolisflowmeter(P/N: CMF025H) andasinglevariable Micromotion Coriolis flow meter transmitter (P/N:1700).Theflowmeter/transmittercombinationhadanuncertainty of ?0.35% of the reading for gas flow. The flowmeter was mounted at the evaporator outlet. The refrigerantout of the evaporator was alwaysin a superheated vapor state,ensuring that only single-phase fluid flowed through themeter.Thegas coolerwas coupled to a closed waterloop as shownin Fig. 5. The closed water loop was then coupled to a buildingchilled water/glycol loop. Independent control of a) thebuilding chilled water flow rate, b) the building chilled watertemperature, and c) the closed water loop flow rate allowedprecise control of water temperature and flow rate at the inletof the gas cooler. The volumetric flow rate of the gas coolerclosed water loop was measured with a Rosemount magneticflow tube (P/N: 8705 TS) coupled to a Rosemount transmitter(P/N: 8732CT). The flow meter had a range of 0e106 L min?1and an uncertainty of ?0.5% of reading. Gas cooler waterpressure drop was not measured during the water heatingtests. However, pressure drop was measured using a Rose-mount pressure transducer (P/N: 3051CD4) with an uncer-tainty of ?0.005 bar during separate isothermal tests in whichthe refrigerant loop was turned off.The evaporator water loop was coupled to an adjustableelectric resistance heater as shown in Fig. 5. This setupallowed a controlled water flow and temperature at the inletof the evaporator. This water flow rate was measured using anOmega turbine flow meter (P/N: FTB- 902) with a range of2.84e18.9 L min?1and uncertainty of ?0.5% of reading.T-Typethermocouples(UP/N:TMQSS-062G-6)wereusedtomeasure all refrigerant and water temperatures at the variouslocationsindicated in Fig. 5. The thermocouple bodywas ratedfor temperatures up to 220?C with a standard error of ?0.5?C.Data were acquired with a National Instruments SCXIsystem. Two SCXI-1102 modules were connected to an SCXI-1303 terminal block. Each terminal block had the capability toread 32 channels of voltage, current and thermocouple data.An SCXI-1000 chassis housed the module/terminal blockcombinations and was connected to a Microsoft Windows-basedPCthroughanNIPCI-6280MSeries DAQcard.Data wereacquired, analyzed and displayed using LabVIEW version 7.1.Fifty-one data points were taken on each of the five andseven-plate heat exchangers, while 24 data points were takenonthesimulatedtwelve-plateheatexchanger,foratotalof126gas cooler performance data points. Data were taken atnominal refrigerant temperatures of 85, 100 and 115?C,refrigerant mass flows between 0.008 and 24 g s?1, or a massflux of 200 G 630 kg m?2s?1. As the density and viscosity ofthe supercritical CO2changed dramatically with temperature,the Reynolds numbers in the gas cooler varied from 3000 to25,000.Thenominalwaterinlettemperatureswere5and20?C,while the nominal water flow rates were 0.93, 2.38 and5.68Lmin?1.UsingtheReynoldsnumberdefinitionspecifiedbyManglik and Bergles (1995) for offset strip fin inserts, Re variedfrom115to875onthewater-side.Toachieveafixedmassflowand refrigerant inlet temperature, it was necessary to let thehigh-side pressure vary. For the data taken in this study, thehigh-side pressure was between 81 and 110 bar.Water and refrigerant heat duties were calculated based onmass flow rates and enthalpy changes of each fluid as shownin Eq. (1). A comparison of the two values was made to providean indication of energy balance. CO2enthalpy was calculatedas function of pressure and temperature using property dataavailable in Engineering Equation Solver (EES) (Klein, 2006),which are in turn based on the equations of state developedbySpan and Wagner (1996)._Q _ mhin? hout(1)The overall UA value of the gas cooler was determined fromthe average of the heat duties on each side and the counterflowformulation of log-mean-temperature difference (LMTD). TheconceptofLMTDreliesontheassumptionthatthefluidspecificheatsareapproximatelyconstant; however, refrigerantspecificheat varies greatly as a function of temperature in the super-critical region. Therefore, the global UA value calculated frommeasurements for the entire heat exchanger is only anapproximation, to be used in comparing the same heatexchanger geometry with the same fluids under similar condi-tions. A more detailed segmented analysis approach is pre-sented in companion paper Part II.Gas cooler approach temperature difference, shown in Eq.(2) is an important parameter for evaluating the effectivenessof a CO2gas cooler. It is defined as the difference between therefrigerant outlet and water inlet temperatures. As approachtemperature approaches zero, the effectiveness of the gascooler approaches one.DTapproach Tref;out? Twater;in(2)Calculated variables are subject to uncertainty propagationresulting from the combined uncertainties in the measuredquantities used in the calculations. This uncertainty propa-gation is accounted for using the root-mean-squares (RMS)approach as detailed in Taylor and Kuyatt (1994). Assumingthat all the measured variables are uncorrelated and random,the uncertainty of a calculated variable can be expressed asshown in Eq. (3).UyffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiXi?vyvxi?2U2xivuut(3)Uyis the uncertainty of the calculated variable, and Uxis theuncertainty of each measured variable. Uncertainties in themeasured variables were obtained from manufacturer/vendorspecifications.3.Results and discussionThe capacity, approach temperatures and UA values for eachof the 126 data points obtained in this study are presented anddiscussed in this section. The five-, seven- and ernational journal of refrigeration 34 (2011) 7e1612The average heat duty is defined as the average of thecalculated refrigerant and water-side heat duties. The abso-lute average deviation between the calculated refrigerant andwater heat duties was less than 5%. The average heat dutyranged from 1.0 to 6.5 kW, with an uncertainty between ?30and ?140 W, depending on the test conditions. At low waterflow rates, the uncertainty was dominated by uncertainty inthe water volumetric flow measurement. At high water flowrates, the uncertainty was dominated by the small water-sidetemperature difference. In general, the uncertainty in averageheat duty ranged from 1 to 9% over the range of refrigerantand water inlet conditions, with the highest relative uncer-tainty occurring at the lowest refrigerant mass flow rates andhighest water flow rate. Fig. 6 shows the average heat duties atrefrigerant inlet temperatures of 115?C at three differentrefrigerant mass flow rates and two different water inlettemperatures for the five- and seven-plate gas coolers.As can be seen in Fig. 6, increasing the refrigerant massflow rate increases the average heat duty for both heatexchangers. The increase in capacity is nearly linear for waterflow rates of 2.38 and 5.68 L min?1. For the low water flow rateof 0.93 L min?1, refrigerant flow rates greater than approxi-mately 12 g s?1do not yield significant increases in capacity.At the low water flow rates (0.93 L min?1), the average watertemperature lift is approximately 45?C at a refrigerant flowrate of 12 g s?1. At this same mass flow, the average lifts of the2.38 and 5.68 L min?1water flow rate points are 20 and 9?C,respectively. As refrigerant mass flow is increased, the highwater outlet temperature of the 0.93 L min?1water flowresults in a smaller driving temperature difference and leadsto the pinch effect observed in Fig. 6. This pinch effect is notobserved at the higher water flow rates for the refrigerantmass flow rates tested. Additionally, as seen in Fig. 6, fora given water flow rate, capacity is higher for the lower waterinlet temperature due to the higher driving temperaturedifference.In addition to heat duty, approach temperature differenceis an important indicator of the effectiveness of the heatexchangerunderdifferentconditions.Alowapproachtemperature difference in the gas cooler indicates the heatexchanger is properly sized for the application and willmaximize system efficiency by decreasing the inlet enthalpyat the evaporator. This effect is illustrated in a CO2P-hdiagram shown in Fig. 7. By reducing the gas cooler outlettemperature from 40 to 6?C, a significant increase in systemCOP can be realized.For the experimental conditions in the present study,approach temperature difference ranged from less than 1 K to45 K, with an uncertainty of ?0.7 K Fig. 8 shows approachtemperature difference versus refrigerant mass flow rate fora refrigerant inlet temperature of 115?C.Approach temperature difference is lowest for the high-est water and lowest refrigerant flow rates. Decreasing thewater flow rate at a constant refrigerant mass flow rateresults in an increased approach temperature difference.This is due to the lower thermal capacitance rate of thewater and the lower water-side heat transfer coefficientresulting from the reduced water velocity and Reynoldsnumber. Increasing refrigerant flow rate for a constantwater flow rate will also increase the approach temperaturedifferenceduetotheincreaseinrefrigerantthermalcapacitance rate with no change in the water-side rate. Anincrease in heat exchanger size from five to seven to twelveplates will yield a lower approach temperature difference forsimilar conditions due to the increased area for heattransfer.Fig. 6 e Average heat duty vs. refrigerant flow rate for 115?C refrigerant inlet.Fig. 7 e Effect of gas cooler outlet ernational journal of refrigeration 34 (2011) 7e1613The global calculated UA value allows for comparisonbetween similar gas cooler designs under varying inletconditions. In the present study, this UA ranged from 0.05 to0.35 kW K?1. The uncertainty in UA varied from ?4 to ?10%.Similar to average heat duty, the highest uncertainty was atthe highest water flow rate and lowest refrigerant flow rate.UA values versus refrigerant flow rate for the five and seven-plate heat exchangers at a refrigerant inlet temperature of100?C are shown in Fig. 9. For low water flow rates(0.93 L min?1), as refrigerant mass flow is increased from 12 to16 g s?1, the UA value remains approximately the same ordecreases slightly. As the refrigerant velocity and thus heattransfer coefficient increase, UA value is expected to generallyincrease rather than remain constant or decrease. However,two factors contribute to the actual observed trend. In Fig. 9,the water flow rate decreases by 0.04 L min?1for the five-plategas cooler and by 0.08 L min?1for the seven-plate gas coolerfrom the nominal as refrigerant mass flow rate increases from12 to 16 g s?1. Since the water-side resistance is more signifi-cant at low flow rates, this decrease in water flow rate morethan compensates for the increase in refrigerant-side heattransfer coefficient. Secondly, the higher mass fraction oflubricant with the dual compressor configuration has beenshown to have a deleterious effect on refrigerant-side heattransfer coefficient (Kuang et al., 2003; Cheng et al., 2008) andthe corresponding UA value.At the higher water flow rates, UA increases with higherrefrigerant mass flow rate, indicating that at these water flowrates, the refrigerant-side heat transfer coefficient is thelimiting factor. Increasing the heat exchanger size from five toseven to twelve platesincreases UA at a given condition due tothe increased heat transfer area.Refrigerant pressuredrop measurements were taken for alldata obtained using the dual compressor setup. Excessiverefrigerant line vibration/pulsing resulting from the operationof the non-muffled single compressor resulted in erroneousFig. 9 e Overall UA vs. refrigerant flow rate for 100?C refrigerant inlet.Fig. 8 e Approach temperature difference vs. refrigerant flow rate for 115?C refrigerant inlet.Fig. 10 e Measured refrigerant pressure ernational journal of refrigeration 34 (2011) 7e1614readings of refrigerant pressure drop for the lower mass flowtests. The entire set of pressure drop data is shown in a scatterplot in Fig. 10. It can be seen that pressure drop is notdependent on refrigerant mass flow alone. For a seven-plategas cooler with refrigerant mass flow rate of 21 g s?1and aninlet temperature of 85?C, the measured pressure drop was95 kPa for a gas cooler refrigerant temperature change frominlet to outlet of 45?C, and 64 kPa for a temperature change of76?C. The smaller temperature change results in a higheraverage refrigerant temperature, lower average density andhigher average velocity through the gas cooler, resulting ina higher pressure drop. For similar refrigerant mass flows andrefrigerant outlet temperatures, pressure drop increases forincreased refrigerant tube length. For a refrigerant inlettemperature of 100?C and mass flow rate of 16 g s?1with anapproximate temperature change of 70?C, the measuredpressure drops for the five, seven and twelve-plate gas coolersare 44, 48 and 54 kPa, respectively.Water-side pressure drop was measured for the five- andseven-plate heat exchangers in isothermal tests in which therefrigerant flow was turned off. The isothermal pressure dropsforwaterflowsfrom0.93to11.25Lmin?1areplottedonalogelogscale in Fig. 11. As expected, pressure drop increases withincreasing water velocity, with higher pressure drops measuredin the seven-plate gas cooler due to the longer flow path.4.ConclusionsA transcritical CO2heat pump facility was developed to allowfor “in situ” experimentation of a compact, water-coupledmicrochannel CO2gas cooler under varying operating condi-tions. Gas coolers of three different sizes were investigated. Bycoupling the closed gas cooler water loop to an externallychilled circuit, the water inlet temperature and flow rate couldbe precisely controlled. Refrigerant inlet conditions to the gascooler could be tuned by adjusting the expansion device, CO2charge level, and evaporation temperature.Refrigerantmassflowrateinthecurrentstudyrangedfrom8to24gs?1.High-sidepressurerangedfrom7930to11,030kPa.Data were obtained at points with refrigerant inlet tempera-turesof85?C,100?C and115?C.Thegascoolerinletwaterflowrate was varied from 0.93 to 5.68 L min?1at inlet temperaturesof 5?C and 20?C. Measured heating capacity for the threedifferent gas coolers ranged from less than 2.0 to 6.5 kW.In companion paper Part II, the results of this study areused to develop and validate computational models for opti-mizing gas cooler design for a particular system, as well asmore accurately predict the performance of a system oper-ating under various conditions. The effects of changingphysical heat exchanger parameters such as fin dimensions,microchannel size or number of water passes can be predictedusing this model without the need for expensive prototypedevelopment and testing.While the present study has resulted in a large bank of heatduty and pressure drop data for a water-coupled micro-channel gas cooler under various conditions, there are severalareas in which additional work is needed. Some of the keyareas that require attention are as follows:? Measurement of refrigerant pressure drop at low mass flowrates (12 g s?1) should be performed to provide a morecomplete bank of pressure drop data for the heat exchangergeometry under investigation.? The range of gas cooler inlet conditions could be expandedto include higher refrigerant flow rates and higher waterinlet temperatures, allowing the assessment of gas coolerperformance in a simulated “multi-pass” water heatingsystem.? Measurement of lubricant circulation rates in the test loopwould provide a more accurate estimate of lubricant flowrates in the gas cooler, which would assist in more accuratecomputation of heat duty and system UA.? The study conducted here should be extended to additionalport diameters other than the 0.89 mm case to betterunderstand the effect of port diameter on the gas cooler sizeneeded
- 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