ASHRAE HVAC SYSTEMS AND EQUIPMENT IP CH 39-2012 CONDENSERS.pdf

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1、39.1CHAPTER 39CONDENSERSWATER-COOLED CONDENSERS 39.1Heat Removal. 39.1Heat Transfer . 39.2Water Pressure Drop 39.4Liquid Subcooling 39.5Water Circuiting. 39.5Types 39.5Noncondensable Gases 39.6Testing and Rating . 39.7Operation and Maintenance 39.7AIR-COOLED CONDENSERS 39.8Types 39.8Fans and Air Req

2、uirements 39.9Heat Transfer and Pressure Drop 39.9Condensers Remote from Compressor. 39.10Condensers as Part of Condensing Unit 39.10Water-Cooled Versus Air-Cooled Condensing. 39.10Testing and Rating 39.11Control . 39.11Installation and Maintenance 39.13EVAPORATIVE CONDENSERS 39.14Heat Transfer . 39

3、.14Condenser Configuration. 39.15Condenser Location . 39.15Multiple-Condenser Installations 39.16Ratings . 39.16Desuperheating Coils. 39.17Refrigerant Liquid Subcoolers . 39.17Multicircuit Condensers and Coolers 39.17Water Treatment . 39.18Water Consumption 39.18Capacity Modulation 39.18Purging. 39.

4、18Maintenance. 39.18Testing and Rating 39.18HE CONDENSER in a refrigeration system is a heat ex-Tchanger that rejects all the heat from the system. This heat con-sists of heat absorbed by the evaporator plus the heat from theenergy input to the compressor. The compressor discharges hot,high-pressure

5、 refrigerant gas into the condenser, which rejects heatfrom the gas to some cooler medium. Thus, the cool refrigerant con-denses back to the liquid state and drains from the condenser to con-tinue in the refrigeration cycle.Condensers may be classified by their cooling medium as (l)water-cooled, (2)

6、 air-cooled, (3) evaporative (air- and water-cooled), and (4) refrigerant-cooled (cascade systems). The firstthree types are discussed in this chapter; see Chapter 48 in the 2010ASHRAE HandbookRefrigeration for a discussion of cascade-cooled condensers.WATER-COOLED CONDENSERSHEAT REMOVALThe heat rej

7、ection rate in a condenser for each unit of heatremoved by the evaporator may be estimated from the graph in Fig-ure 1. The theoretical values shown are based on Refrigerant 22 with10F suction superheat, 10F liquid subcooling, and 80% compres-sor efficiency. Depending on compressor efficiency, the h

8、eatremoved could be higher or lower than these values. Usually, the heatrejection requirement can be accurately determined by adding theknown evaporator load and the heat equivalent of the power requiredfor compression (obtained from the compressor manufacturers cat-alog). (Note that heat from the c

9、ompressor is reduced by an indepen-dent heat rejection processes such as oil cooling, motor cooling, etc.)The volumetric flow rate of condensing water required may becalculated as follows:Q = (1)whereQ = volumetric flow rate of water, ft3/h (multiply ft3/h by 0.125 to obtain gpm)qo= heat rejection r

10、ate, Btu/h = density of water, lb/ft3t1= temperature of water entering condenser, Ft2= temperature of water leaving condenser, Fcp= specific heat of water at constant pressure, Btu/lbFThe preparation of this chapter is assigned to TC 8.4, Air-to-RefrigerantHeat Transfer Equipment; TC 8.5, Liquid-to-

11、Refrigerant Heat Exchangers;and TC 8.6, Cooling Towers and Evaporative Condensers.qocpt2t1-Fig. 1 Heat Removed in Condenser39.2 2012 ASHRAE HandbookHVAC Systems and Equipment Example 1. Estimate volumetric flow rate of condensing water required forthe condenser of an R-22 water-cooled unit operating

12、 at a condensingtemperature of 105F, an evaporating temperature of 40F, 10F liquidsubcooling, and 10F suction superheat. Water enters the condenser at86F and leaves at 95F. The refrigeration load is 100 tons.Solution: From Figure 1, the heat rejection factor for these conditionsis about 1.19.qo= 100

13、 1.19 = 119 tons = 62.1 lb/ft3at 90.5Fcp= 1.0 Btu/(lbF)From Equation (1):Q = = 319 gpmNote: The value 1496 is a unit conversion factor.HEAT TRANSFERA water-cooled condenser transfers heat by sensible cooling inthe gas desuperheating and condensate subcooling stages and bytransfer of latent heat in t

14、he condensing stage. The condensing stageis by far the dominant process in normal refrigeration applications,accounting for approximately 83% of the heat rejection. Becausethe tube wall temperature is normally lower than the condensingtemperature at all locations in the condenser, condensation takes

15、place throughout the condenser.The effect of changes in the entering gas superheat is typicallyinsignificant because of an inverse proportional relationshipbetween temperature difference and heat transfer coefficient. As aresult, an average overall heat transfer coefficient and the mean tem-perature

16、 difference (calculated from the condensing temperaturecorresponding to the saturated condensing pressure and the enteringand leaving water temperatures) give reasonably accurate predic-tions of performance.Subcooling affects the average overall heat transfer coefficientwhen tubes are submerged in l

17、iquid. The heat rejection rate is thendetermined asq = UAtm(2)whereq = total heat transfer rate, Btu/hU = overall heat transfer coefficient, Btu/hft2FA = heat transfer surface area associated with U, ft2tm= mean temperature difference, FChapter 4 of the 2009 ASHRAE HandbookFundamentals de-scribes ho

18、w to calculate tm.Overall Heat Transfer CoefficientThe overall heat transfer coefficient Uoin a water-cooled con-denser with water inside the tubes may be computed from calcu-lated or test-derived heat transfer coefficients of the water andrefrigerant sides, from physical measurements of the condens

19、ertubes, and from a fouling factor on the water side, using the follow-ing equation:Uo= (3)whereUo= overall heat transfer coefficient, based on external surface and mean temperature difference between external and internal fluids, Btu/hft2FAo /Ai= ratio of external to internal surface areahw= intern

20、al or water-side film coefficient, Btu/hft2Frfw= fouling resistance on water side, ft2hF/Btut = thickness of tube wall, ftk = thermal conductivity of tube material, Btu/hftFAo /Am= ratio of external to mean heat transfer surface areas of metal wallhr= external or refrigerant-side coefficient, Btu/hf

21、t2Fs= surface fin efficiency (100% for bare tubes)For tube-in-tube condensers or other condensers where re-frigerant flows inside the tubes, the equation for Uo, in terms ofwater-side surface, becomesUo= (4)wherehr= internal or refrigerant-side coefficient, Btu/hft2Fhw= external or water-side coeffi

22、cient, Btu/hft2FFor brazed or plate-and-frame condensers A0= Ai; thereforethe equation for UoisUo= (5)where t is plate thickness.Water-Side Film CoefficientValues of the water-side film coefficient hwmay be calculatedfrom equations in Chapter 4 of the 2009 ASHRAE HandbookFun-damentals. For turbulent

23、 flow, at Reynolds numbers exceeding10,000 in horizontal tubes and using average water temperatures,the general equation (McAdams 1954) is(6)whereD = inside tube diameter, ftk = thermal conductivity of water, Btu/hftFG = mass velocity of water, lb/hft2 = viscosity of water, lb/fthcp= specific heat o

24、f water at constant pressure, Btu/lbFThe constant 0.023 in Equation (6) reflects plain inner diameter(ID) tubes. Bergles (1995) and Pate et al. (1991) discuss numerouswater-side enhancement methods that increase the value of this con-stant.Because of its strong influence on the value of hw, a high w

25、atervelocity should generally be maintained without initiating erosionor excessive pressure drop. Typical maximum velocities from 6 to10 fps are common with clean water. Experiments by Sturley (1975)at velocities up to approximately 26 fps showed no damage to cop-per tubes after long operation. Wate

26、r quality is the key factor affect-ing erosion potential (Ayub and Jones 1987). A minimum velocityof 3 fps is good practice when water quality is such that noticeablefouling or corrosion could result. With clean water, the velocity maybe lower if it must be conserved or has a low temperature. In som

27、ecases, the minimum flow may be determined by a lower Reynoldsnumber limit.For brazed or plate-and-frame condensers, the equation is similarto Equation (6). However, the diameter D is replaced by H, which isthe characteristic spacing between plates.Refrigerant-Side Film CoefficientFactors influencin

28、g the value of the refrigerant-side film coeffi-cient hrareType of refrigerant being condensed1496 11962.1 1.0 95 86-1AoAi-1hw-AoAi-rfwAoAm-tk-1hrs-+-1AoAi-1hr-rfwtk-1hw-+ +-11 hrrfwtk1 hw+ +-hwDk- 0.023DG-0.8cpk-0.4=Condensers 39.3Geometry of condensing surface plain tube outer diameter (OD);finned

29、-tube fin spacing, height, and cross-sectional profile; andplate geometryCondensing temperatureCondensing rate in terms of mass velocity or rate of heat transferArrangement of tubes in bundle and location of inlet and outletconnectionsVapor distribution and rate of flowCondensate drainageLiquid subc

30、oolingValues of refrigerant-side coefficients may be estimated fromcorrelations in Chapter 5 of the 2009 ASHRAE HandbookFunda-mentals. Information on the effects of refrigerant type, condensingtemperature, and loading (temperature drop across the condensatefilm) on the condensing film coefficient is

31、 in the section on Con-densing in the same chapter. Actual values of hrfor a given physicalcondenser design can be determined from test data using a Wilsonplot (Briggs and Young 1969; McAdams 1954).The type of condensing surface has a considerable effect on thecondensing coefficient. Most halocarbon

32、 refrigerant condensers usefinned tubes where the fins are integral with the tube. Water veloc-ities normally used are large enough for the resulting high water-side film coefficient to justify using an extended external surface tobalance the heat transfer resistances of the two surfaces. Pearsonand

33、 Withers (1969) compared refrigerant condensing performanceof integral finned tubes with different fin spacing. Some otherrefrigerant-side enhancements are described by Pate et al. (1991)and Webb (1984a). The effect of fin shape on the condensing coef-ficient is addressed by Kedzierski and Webb (199

34、0). Ghaderi et al.(1995) reviewed in-tube condensation heat transfer correlations forsmooth and augmented tubes.In the case of brazed-plate or plate-and-frame condensers inletnozzle size, chevron angle, pitch, and depth of the nozzles areimportant design parameters. For trouble-free operation, refri

35、gerantshould flow counter to the water flow. Little specific design infor-mation is available; however, film thickness is certainly a factor inplate condenser design because of the falling-film nature along thevertical surface. Kedzierski (1997) showed that placing a brazedcondenser in a horizontal

36、position improved Uoby 17 to 30%because of the shorter film distance.Huber et al. (1994a) determined condensing coefficients forR-134a, R-12, and R-11 condensing on conventional finned tubeswith a fin spacing of 26 fins per inch (fpi) and a commercially avail-able tube specifically developed for con

37、densing halocarbon refrig-erants (Huber et al. 1994b). This tube has a sawtooth-shaped outerenhancement. The data indicated that the condensing coefficientsfor the sawtoothed tube were approximately three times higher thanfor the conventional finned tube exchanger and two times higher forR-123.Furth

38、er, Huber et al. (1994c) found that for tubes with 26 fpiR-134a condensing coefficients are 20% larger than those for R-12at a given heat flux. However, on the sawtoothed tube, R-134a con-densing coefficients are nearly two times larger than those for R-12at the same heat flux. The R-123 condensing

39、coefficients were 10 to30% larger than the R-11 coefficients at a given heat flux, with thelargest differences occurring at the lowest heat fluxes tested. Thedifferences in magnitude between the R-123 and R-11 condensingcoefficients were the same for both the 26 fpi tube and the saw-toothed tube.Phy

40、sical aspects of a given condenser design (e.g., tube spacingand orientation, shell-side baffle arrangement, orientation of multi-ple water-pass arrangements, refrigerant connection locations, num-ber of tubes high in the bundle) affect the refrigerant-side coefficientby influencing vapor distributi

41、on and flow through the tube bundleand condensate drainage from the bundle. Butterworth (1977)reviewed correlations accounting for these variables in predicting theheat transfer coefficient for shell-side condensation. These effectsare also surveyed by Webb (1984b). Kistler et al. (1976) developedan

42、alytical procedures for design within these parameters.As refrigerant condenses on the tubes, it falls on the tubes inlower rows. Because of the added resistance of this liquid film, theeffective film coefficient for lower rows should be lower than thatfor upper rows. Therefore, the average overall

43、refrigerant film coef-ficient should decrease as the number of tube rows increases. Webband Murawski (1990) present row effect data for five tube geome-tries. However, the additional compensating effects of added filmturbulence and direct contact condensation on the subcooled liquidfilm make actual

44、row effect uncertain.Huber et al. (1994c, 1994d) determined that the row effect onfinned tubes is nearly negligible when condensing low-surface-tension refrigerants such as R-134a. However, the finned-tube filmcoefficient for higher-surface-tension refrigerants such as R-123can drop by as much as 20

45、% in lower bundle rows. The row effectfor the sawtoothed condensing tube is quite large for both R-134aand R-123, as the film coefficient drops by nearly 80% from top tobottom in a 30-row bundle.Randall and Eckels (2005a, 2005b) measured refrigerant-sidecoefficients for smooth, finned, and three-dim

46、ensionally enhancedtube surfaces with and without liquid inundation for R-134a. Theresults show that the three-dimensionally enhanced tube surfaces hadthe highest refrigerant-side coefficients at low inundation conditions,but their performance was most sensitive to inundation rates. Therefrigerant-s

47、ide coefficient of the finned tube was found to be lesssensitive to inundation rates that those of the three-dimensionallyenhanced tubes.Eckels (2007) studied the effect of two lubricating oils on therefrigerant-side coefficient during condensation of R-134a forsmooth, finned, and three-dimensionall

48、y enhanced tube surfaces atvarying oil concentrations. The reduction in refrigerant-side coeffi-cient caused by oil for all three tube types was found to be less than15% for all three tube types, and independent of oil concentration.Honda et al. (1994, 1995) demonstrated that row effects causedby condensate drainage and inundation are less for staggered tubebundles than fo