ASHRAE HVAC SYSTEMS AND EQUIPMENT SI 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 . 39.16Multiple-Condenser Installations 39.16Ratings . 39.16Desuperheating Coils. 39.17Refrigerant Liquid Subcoolers . 39.17Multicircuit Condensers and Coolers 39.17Water Treatment . 39.17Water 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 with5 K suction superheat, 5 K liquid subcooling, and 80% compressorefficiency. Depending on compressor efficiency, the hea

8、t removedcould be higher or lower than these values. Usually, the heat rejectionrequirement can be accurately determined by adding the knownevaporator load and the heat equivalent of the power required forcompression (obtained from the compressor manufacturers cata-log). (Note that heat from the com

9、pressor 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, m3/sqo= heat rejection rate, kW = density of water, kg/m3t1= temper

10、ature of water entering condenser, Ct2= temperature of water leaving condenser, Ccp= specific heat of water at constant pressure, kJ/(kgK)The preparation of this chapter is assigned to TC 8.4, Air-to-RefrigerantHeat Transfer Equipment; TC 8.5, Liquid-to-Refrigerant Heat Exchangers;and TC 8.6, Coolin

11、g Towers and Evaporative Condensers.qocpt2t1-Fig. 1 Heat Removed in Condenser39.2 2012 ASHRAE HandbookHVAC Systems and Equipment (SI)Example 1. Estimate volumetric flow rate of condensing water required forthe condenser of an R-22 water-cooled unit operating at a condensingtemperature of 40C, an eva

12、porating temperature of 4C, 5 K liquid sub-cooling, and 5 K suction superheat. Water enters the condenser at 30Cand leaves at 35C. The refrigeration load is 300 kW.Solution: From Figure 1, the heat rejection factor for these conditionsis about 1.19.qo=1.19 300 = 357 kW = 995 kg/m3at 33Ccp= 4.18 kJ/(

13、kgK)From Equation (1):Q = = 0.0172 m3/sHEAT TRANSFERA water-cooled condenser transfers heat by sensible cooling inthe gas desuperheating and condensate subcooling stages and bytransfer of latent heat in the condensing stage. The condensing stageis by far the dominant process in normal refrigeration

14、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 takesplace throughout the condenser.The effect of changes in the entering gas superheat is typicallyi

15、nsignificant 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 difference (calculated from the condensing temperaturecorresponding to the saturated condensing

16、 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 liquid. The heat rejection rate is thendetermined asq = UAtm(2)whereq = total heat transfer rate,

17、 WU = overall heat transfer coefficient, W/(m2K)A = heat transfer surface area associated with U, m2tm= mean temperature difference, KChapter 4 of the 2009 ASHRAE HandbookFundamentals de-scribes how to calculate tm.Overall Heat Transfer CoefficientThe overall heat transfer coefficient Uoin a water-c

18、ooled 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 condensertubes, and from a fouling factor on the water side, using the follow-ing equation:Uo= (3)whereUo= ove

19、rall heat transfer coefficient, based on external surface and mean temperature difference between external and internal fluids, W/(m2K)Ao /Ai= ratio of external to internal surface areahw= internal or water-side film coefficient, W/(m2K)rfw= fouling resistance on water side, (m2K)/Wt = thickness of

20、tube wall, mk = thermal conductivity of tube material, W/(mK)Ao /Am= ratio of external to mean heat transfer surface areas of metal wallhr= external or refrigerant-side coefficient, W/(m2K)s= surface fin efficiency (100% for bare tubes)For tube-in-tube condensers or other condensers where re-frigera

21、nt flows inside the tubes, the equation for Uo, in terms ofwater-side surface, becomesUo= (4)wherehr= internal or refrigerant-side coefficient, W/(m2K)hw= external or water-side coefficient, W/(m2K)For brazed or plate-and-frame condensers A0= Ai; thereforethe equation for UoisUo= (5)where t is plate

22、 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 flow, at Reynolds numbers exceeding10 000 in horizontal tubes and using average water temperatures,the general equati

23、on (McAdams 1954) is(6)whereD = inside tube diameter, mk = thermal conductivity of water, W/(mK)G = mass velocity of water, kg/(sm2) = viscosity of water, kg/(sm)cp= specific heat of water at constant pressure, kJ/(kgK)The constant 0.023 in Equation (6) reflects plain inner diameter(ID) tubes. Bergl

24、es (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 watervelocity should generally be maintained without initiating erosionor excessive pressure drop. Typical maximum velo

25、cities from 2 to3 m/ s are common with clean water. Experiments by Sturley (1975)at velocities up to approximately 8 m/ s showed no damage to cop-per tubes after long operation. Water quality is the key factor affect-ing erosion potential (Ayub and Jones 1987). A minimum velocityof 1 m/ s is good pr

26、actice 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 somecases, the minimum flow may be determined by a lower Reynoldsnumber limit.For brazed or plate-and-frame condensers,

27、the equation is similarto Equation (6). However, the diameter D is replaced by H, which isthe characteristic spacing between plates.Refrigerant-Side Film CoefficientFactors influencing the value of the refrigerant-side film coeffi-cient hrareType of refrigerant being condensed357995 4.18 35 30-1AoAi

28、-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-tube fin spacing, height, and cross-sectional profile; andplate geometryCondensing temperatureCondensing rate in terms of

29、 mass velocity or rate of heat transferArrangement of tubes in bundle and location of inlet and outletconnectionsVapor distribution and rate of flowCondensate drainageLiquid subcoolingValues of refrigerant-side coefficients may be estimated fromcorrelations in Chapter 5 of the 2009 ASHRAE HandbookFu

30、nda-mentals. Information on the effects of refrigerant type, condensingtemperature, and loading (temperature drop across the condensatefilm) on the condensing film coefficient is in the section on Con-densing in the same chapter. Actual values of hrfor a given physicalcondenser design can be determi

31、ned 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 refrigerant condensers usefinned tubes where the fins are integral with the tube. Water veloc-ities normally used are lar

32、ge 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 Withers (1969) compared refrigerant condensing performanceof integral finned tubes with different fin spacing. Some other

33、refrigerant-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 (1990). Ghaderi et al.(1995) reviewed in-tube condensation heat transfer correlations forsmooth and augmented tubes.In the cas

34、e 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, refrigerantshould flow counter to the water flow. Little specific design infor-mation is available; however, film thickness is

35、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 position improved Uoby 17 to 30%because of the shorter film distance.Huber et al. (1994a) determined condensing coefficien

36、ts forR-134a, R-12, and R-11 condensing on conventional finned tubeswith a fin spacing of 0.977 mm and a commercially available tubespecifically developed for condensing halocarbon refrigerants(Huber et al. 1994b). This tube has a sawtooth-shaped outerenhancement. The data indicated that the condens

37、ing coefficientsfor the sawtoothed tube were approximately three times higher thanfor the conventional finned tube exchanger and two times higher forR-123.Further, Huber et al. (1994c) found that for tubes with 0.977 mmfin spacing R-134a condensing coefficients are 20% larger thanthose for R-12 at a

38、 given heat flux. However, on the sawtoothedtube, R-134a condensing coefficients are nearly two times largerthan those for R-12 at the same heat flux. The R-123 condensingcoefficients were 10 to 30% larger than the R-11 coefficients at agiven heat flux, with the largest differences occurring at the

39、lowestheat fluxes tested. The differences in magnitude between the R-123and R-11 condensing coefficients were the same for both the0.977 mm fin spacing tube and the sawtoothed tube.Physical aspects of a given condenser design (e.g., tube spacingand orientation, shell-side baffle arrangement, orienta

40、tion of multi-ple water-pass arrangements, refrigerant connection locations, num-ber of tubes high in the bundle) affect the refrigerant-side coefficientby influencing vapor distribution and flow through the tube bundleand condensate drainage from the bundle. Butterworth (1977)reviewed correlations

41、accounting for these variables in predicting theheat transfer coefficient for shell-side condensation. These effectsare also surveyed by Webb (1984b). Kistler et al. (1976) developedanalytical procedures for design within these parameters.As refrigerant condenses on the tubes, it falls on the tubes

42、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 refrigerant film coef-ficient should decrease as the number of tube rows increases. Webband Murawski (1990) present

43、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 row effect uncertain.Huber et al. (1994c, 1994d) determined that the row effect onfinned tubes is nearly negligible

44、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% in lower bundle rows. The row effectfor the sawtoothed condensing tube is quite large for both R-134aand R-123, as

45、 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-dimensionally enhancedtube surfaces with and without liquid inundation for R-134a. Theresults show that the three-dimen

46、sionally enhanced tube surfaces hadthe highest refrigerant-side coefficients at low inundation conditions,but their performance was most sensitive to inundation rates. Therefrigerant-side coefficient of the finned tube was found to be lesssensitive to inundation rates that those of the three-dimensi

47、onallyenhanced tubes.Eckels (2007) studied the effect of two lubricating oils on therefrigerant-side coefficient during condensation of R-134a forsmooth, finned, and three-dimensionally enhanced tube surfaces atvarying oil concentrations. The reduction in refrigerant-side coeffi-cient caused by oil

48、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 for in-line tubes. In addition, performance improve-ments as high as 85% were reported for optimized two-dimensionalfin profiles compared to co