1、 International Journal of Heating, Ventilating, Air-conditioning and Refrigerating Research Editor John W. Mitchell, Ph.D., P.E. Professor of Mechanical Engineering, University of Wisconsin-Madison, USA Associate Editors Michael J. Brandemuehl, Ph.D., P.E., Professor, Joint Center for Energy Managem
2、ent, University of Colorado, Boulder, USA James E. Braun, Ph.D., P.E., Associate Professor, Ray W. Herrick Laboratories, School of Mechanical Engineering, Purdue University, West Lafayette, Indiana, USA Alberto Cavallini, Ph.D., Professor, Dipartmento di Fisicia Tecnica, University of Padova, Italy
3、Arthur L. Dexter, D.Phil., C.Eng., Reader in Engineering Science, Department of Engineering Science, University of Oxford, United Kingdom Leon R. Glicksman, Ph.D., Professor, Departments of Architecture and Mechanical Engineering, Massachusetts Institute of Technology, Cambridge, USA Richard R. Gonz
4、alez, Ph.D., Director, Biophysics and Biomedical Modeling Division, U.S. Army Research Institute of Environmental Medicine, Natick, Massachusetts, USA Anthony M. Jacobi, Ph.D., Professor and Associate Director ACRC, Department of Mechanical and Industrial Engineering, University of Illinois, Urbana-
5、Champaign, USA Reinhard Radermacher, Ph.D., Professor and Director, Center for Environmental Energy Engineering, Department of Mechanical Engineering, University of Maryland, College Park, USA Keith E. Starner, P.E., Engineering Consultant, York, Pennsylvania, USA Jean-Christophe Visier, Ph.D., Head
6、 Centre Scientifique et Technique du Btiment, Energy Management Automatic Controller Division, Mame La Valle, France Policy Committee Editorial Assistant Stephen W. Ivesdal, Chair, Member ASHRAE P. Ole Fanger, Fellow/Life Member ASHRAE Ken-Ichi Kimura, Fellow ASHRAE John W. Mitchell, Fellow ASHRAE
7、Frank M. Coda, Member ASHRAE W. Stephen Comstock, Associate Member ASHRAE Bany Kurian, Publishing Manager Jennifer A. Haukohl W. Stephen Comstock Mark S. Owen, Handbook Editor Heather E. Kennedy, Handbook Associate Editor Nancy F. Thysell, Typographer Publisher ASHRAE Staff 02002 by the American Soc
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9、ransmitted in any form or by any means-electronic, photocopying, recording, or other-without permission in writing from ASHRAE. Abstracts-Abstracted and indexed by ASHRAE Abstract Center; Ei (Engineering Information, Inc.) Ei Compendex and Engineering Index; IS1 (Institute for Scientific Information
10、) Web Science and Research Alert; and BSRIA (Building Services Research x 0.51, the ratio of mean EHD pressure to axial momentum flux is greater than 1, implying a significant effect of EHD force on the flow field, and therefore on heat transfer enhancements and pressure drop penalties. The stratifi
11、ed flow regime corresponds to such flow parametric values. The flow regimes with and without the EHD force are quite different for the same parametric conditions. The flow regime corresponds to the base case when no EHD field is applied. Figure 6A and B depicts the effect of mass flux on the heat tr
12、ansfer enhancements and pressure drop penalties for R-404a condensing inside a smooth tube. The trends are similar to those observed for R-134a in Singh (1995). The results for R-134a at an applied voltage of 15 kV are presented in Table 2 for comparison. The mass transfer resistance associated with
13、 near-azeotrope R-404a is negligible. An EHD enhancement of 10.6-fold with a pressure drop penalty of 13.4-fold is obtained at a mass flux of 50 kg/(m2.s), as compared to an enhancement and penalty of 5.4-fold and 2.3-fold, respectively, at a mass flux of 300 kg/(m2.s). At lower mass fluxes, the flo
14、w is stratified and less turbulent. Higher mass fluxes with semi- annular or annular flow regimes already have high levels of turbulence at the heat transfer surface because of the increased motion and thinner boundary layer. The liquid-extraction phenomenon associated with the EHD technique makes t
15、he condensate film thinner. Although it can barely affect the shear-dominated, thin, annular film associated with a highly turbulent flow, it has more room to affect the body-force-dominated, thick, stratified condensate film, characteristic of a lower mass flow. The heat transfer enhancement curve
16、for mass flux of 50 kg/(m2.s) overtakes 228 HVAC interfacial shear increases with increasing mass flux. The electric-body force cannot be quantified simply in such a situation with multicomponent refrigerant with local permittivity and fluid density gradients. The complex interplay of these three fo
17、rces can give rise to the behaviors shown in VOLUME 8, NUMBER 3, JULY 2002 R-407 IN SMOOTH TUBE X, = 30% 229 x, = 30% Q“ = 10 W/m +- G = 50 ks/(mz.s) +G=lOOkg/(I$.s) Tm=30C,Pm= 1250kPa -t G = 300 kg/(I$.S) 5 10 15 APPLIED VOLTAGE, kV (A) Figure 7. (A) Heat Transfer Enhancements and (B) Pressure Drop
18、 Penalty for R-407 in Smooth Tube (Mass Flux Variation) Table 2. Heat Transfer Enhancements for R-l34a, R-404a, and R407c (15 kV, T, = 3OOC) Smooth Tube Mass Flux, kg/(m2.s) R-134aX R-404a R-407 50 4.95 8.45 2.52 1 O0 4.90 7.35 2.69 200 3.48 4.33 2.28 300 2.63 3.67 1.83 *Data from earlier measuremen
19、ts in this laboratory (Singh 1995) Figure 6. A detailed theoretical and computational study is needed to give more insight into the complex interplay of these three forces. The results of mass flux variation for refrigerant R-407c condensing inside the smooth tube, however, show a different behavior
20、 (Figure 7A). The maximum EHD enhancement during in-tube condensation of R-407c does not necessarily take place at the highest applied voltage. For in-tube condensation, it is believed that the enhancement increases with the applied EHD 230 HVAC a flow meter, absolute and differential pressure trans
21、ducers, and a condenser. A gear pump circulated the ammonia through the loop, which ensured that there was no oil in the system while the pure ammonia tests were conducted. The flow from the pump was directed either through the main line to the boiledpreheater or through the bypass line to the conde
22、nser VOLUME 8, NUMBER 3, JULY 2002 241 rl CHILLER R-134a LOOP I I I EXCHANGER BLPSOEWACCUYUIATOR I I I BOILER LOOP HEATER m v BALLVALVE DIFFERENTIAL PRESSURE SG TRANSDUCER (ni, FLOWMETER PUMP 0 GLASS Figure 1. Schematic of Test Faciiity located at the outlet of the test section. Before the ammonia e
23、ntered the boiler, its mass flow rate was measured with a Coriolis flow meter with an operational range of O to 0.6 kgs (O to 80 lb/min) and an uncertainty of *O. 15% of reading. In the boiler, the ammonia was evaporated to the quality desired at the inlet of the test section. After leaving the boil
24、er, the ammonia was routed to the desired test section by a series of ball valves. Following evaporation in the test sec- tion, the two-phase ammonia mixture entered the inner tube of a coiled tube-in-tube condenser and was recondensed before returning to the pump. The bladder/accumulator assembly c
25、on- trolled the saturation pressure of ammonia by varying the mass of ammonia in the system. Test Sections The test sections were tube-in-tube counterflow heat exchangers with ammonia flowing in the inner tube and liquid R-134a flowing in the annulus. The dimensions of the smooth tube and the microf
26、in tube are summarized in Table 1. The test sections for both tubes were constructed with a 3.04 m (10 ft) long, 12.5 mm (0.5 in.) OD aluminum tube, which was placed inside a 19.0 mm (0.75 in.) OD outer tube. The outer tube was assembled in 0.61 m (2.0 ft) sections with compres- sion fittings. The s
27、aturation temperature of the ammonia in the test section was inferred from the saturation pressure measured with a O to 690 kPa (O to 100 psia) pressure transducer with an accuracy of kO.1 psi, located at the inlet of the test sections. The pressure drop across the test sections was measured with a
28、strain-gage-type differential pressure transducer. R-134a Loop The R-134a loop supplied the desired energy for the test sections to the ammonia. This loop consisted of a gear pump, a flow meter, a bladderaccumulator, a pressure transducer, and an 242 HVAC&R RESEARCH Table 1. Physical Dimensions of S
29、mooth and Microfin Tubes Smooth Microfin Outside diameter 12.5 mm (0.5 in.) 12.5 mm (0.5 in.) Inside diameter 10.9 mm (0.43 in.) 1 1.8 mm (0.466 in.) Number of fins O 60 Fin height - 0.254 mm (0.010 in.) Helix angle - 17 degrees Wall thickness 0.89 mm (0.035 in.) 0.43 mm (0.017 in.) immersion heater
30、 Measurements were required at low temperatures and low heat fluxes for the realistic boundary conditions obtained with a fluid-heated test section. These three constraints (low temperature, low heat flux, and fluid heating) were met by selecting R-134a as the annulus fluid. A water/glycol mixture
31、would have also been possible, but the higher viscosity would have made it more difficult to maintain turbulent flow in the annulus. The bladdedaccumulator and pressure transducer were used to control and monitor the pres- sure of the R-134a. Monitoring the pressure and temperature of the R-134a in
32、the loop ensured that the refrigerant remained liquid by maintaining the saturation temperature well above the R-134a temperatures. Thermocouples inserted into the fluid stream at the inlet and outlet of the test section were used to measure the R-134a temperature. A Coriolis flow meter with a range
33、 of O to 0.303 kgls (O to 40 lbmin) and an accuracy of *0.2% was used to record the ammonia mass flow rate. Boiler Loop The vapor quality of the ammonia entering the test section was set in the boiledpreheater. The loop consisted of a positive-displacement pump, a flow meter, and an immersion heater
34、 The boiler was a tube-in-tube heat exchanger with ammonia in the inner tube and a mixture of water and propylene glycol flowing counter to the ammonia in the annular gap. Thermocouples inserted into the fluid stream were used to measure the temperature of the water/glycol mixture entering and leav
35、ing the boiler. The flow rate of the watedglycol mixture was measured with a Coriolis flow meter that had an accuracy of *0.2%. Chiller Loop The chiller loop, which acted as the heat sink in the system, consisted of a refrigeration unit that provided 70.4 kW (20 tons) of refrigeration capacity, a 0.
36、946 m3 (250 gallon) reservoir con- taining a mixture of water and ethylene glycol, a centrifugal pump, and a heat exchanger. The two-phase ammonia mixture leaving the test section was condensed on the tube side of a coiled tube-in-tube heat exchanger located at the exit of the test section. The cent
37、rifugal pump circu- lated the watedglycol mixture from the fluid reservoir to the heat exchanger. Data Acquisition System The outputs from the various sensors in the system were recorded with a digital multimeter connected to two switchkontrol units, each having 50 channels. A personal computer cont
38、rolled the operation of the multimeter and the switcWcontro1 units. A program was developed that monitored the mass flow rates, temperatures, and pressures in the system. Once the system reached the desired steady-state conditions, the final data acquisition program was run, which converted signals
39、to temperatures, flow rates, and pressures in the loops. The system was considered in steady-state once the refrigerant flow rate and saturation pres- sure in the test section achieved a constant value for a period of 5 min. All variables were then VOLUME 8, NUMBER 3, JULY 2002 243 scanned sequentia
40、lly 10 times over approximately 25 min. The data were then averaged and inspected for anomalies, typically hl% on flow rate and saturation pressure. Due to the intermit- tent nature of the flow patterns observed in this study, a true steady-state value could not be achieved for all points. For examp
41、le, in the dryout region at high qualities the heat transfer coef- ficient varied dramatically as the tube wetted and dried, thus making it impossible to obtain completely stable temperature readings. The average values from the 10 scans were used to cal- culate the average and sectional-average hea
42、t transfer coefficients discussed in the next section. DATA ANALYSIS This section describes the types of heat transfer coefficients presented in this paper, as well as the method used to calculate these values. The heat transfer coefficients referred to as average heat transfer coefficients were mea
43、sured for an inlet vapor quality of 15% and an outlet vapor quality of 95%. Sectional-average heat transfer coefficients were determined for a given mass flux, saturation temperature, and heat flux, at nominal vapor qualities of 10, 25, 50, 75, and 95%. Both average and sectional-average heat transf
44、er coefficients were calculated in the fol- lowing manner for any data run with turbulent fluid flow in the annulus. The amount of energy transferred from the R-l34a, in the annulus of the test section, to the ammonia was determined fi-om an energy balance on the R- 134a, This energy transfer can al
45、so be defined in terms of the overall heat transfer coefficient U,: where the log-mean temperature difference is defined as Equations (i), (2), and (3) can then be combined and rearranged to solve for the overall heat transfer coefficient U,. The saturation temperatures at the inlet and outlet of th
46、e test section, as determined by the pressure readings, are used in Equation (3). The annulus-side heat transfer coefficient ho was calculated using the Gnielinski (1976) cor- relation, r 1 where the leading coefficient Co was determined using a modified Wilson plot technique (Briggs and Young 1969) and Dhyd,ann was the hydraulic diameter of the annular gap. The fric- tion factor was calculated with the Petukhov equation for in-tube flow (Incropera and DeWitt
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