Experimental Study of Water Cooled Condenser Made of Three

EXPERIMENTAL STUDY OF WATER COOLED CONDENSER MADE OF THREE DIMENSIONAL AND HIGH FIN DENSITY INTEGRAL-FINNED TUBES

Wen-Tao JI Chuang-Yao ZHAO

MOE Key Laboratory of Thermo-Fluid Science and Engineering, Xi’an Jiaotong University, Xi’an, China

Qi-Bin DAI Shu-Heng HAN Chongqing Midea General Refrigeration Equipment Co,

Ltd. Chongqing, China

Ding-Cai ZHANG Ya-Ling HE Wen-Quan TAO* MOE Key Laboratory of Thermo-Fluid Science and Engineering

Xi’an Jiaotong University, Xi’an, China * Corresponding author, Tel: +86-29-82669106, Email: [email protected]

ABSTRACT

The thermo-hydraulic performance of two shell and tube condensers was investigated with an experimental approach. The experiment is conducted in a water cooled centrifugal chiller test rig. The condensers are made of three-dimensional (3-D) and high fin density integral-finned (2-D) tubes. 2-D and 3-D tubes all have the diameter of 3/4 inch (19mm). The 2-D tube has external fin density of 56fpi (fins per inch), fin height 1.023mm and 48 internal ribs per circle. The 3-D enhanced tube has the external fin density of 45fpi, fin height of 0.981mm and 45 internal ribs per circle. The 3-D tube is widely used in the water cooled chillers. 2-D tube is a newly designed surface with enhanced external fin density. Condensing heat transfer coefficient of R134a outside single horizontal tube is firstly tested at saturate temperature of 40℃. At the internal water velocity of 2.2m/s, the overall heat transfer coefficients of 2-D tube is in the range of 10364.7 to 12420.9W/m2K, 4.2%～9.0% higher than 3-D tube. External condensing heat transfer coefficient is 16.3% ～ 25.2% higher than 3-D tube. The condensers are manufactured with these two types of tubes. Both condensers have the same geometric parameters except the tubes and tube bundle space. The length of tube in the condenser is 4000mm. The tube bundles are arranged in a staggered mode. For the integral-fin tube condenser, the longitudinal tube pitch of tube arrays is 23mm in rows and the transverse is 20mm. At the same power input and cooling water inlet temperature of 32℃, the cooling power of 2-D tube condenser are respectively of 1755.4kW and 1769.4kW; 3-D tube condenser is 1727.5kW and 1770.5kW. The pressure drop increased about 11.2%～15.9% for the 2-D tube condenser compared with 3-D tube condenser. Generally, the two condensers have the same heat transfer performance, while the

integral-fin tube condenser saves 15% of copper material consumption.

1. INTRODUCTION

Shell and tube water cooled condenser have been widely used in air-conditioning applications. Typically, refrigerant vapor enters the shell side as superheated state and condenses outside the tube bundle fixed in the tube-sheet. To enhance the condensation heat transfer, many types of enhanced surfaces were developed [1-4]. Enhanced surfaces can reduce the size of heat exchanger and increase the heat duty for a given exchanger. Two and three dimensional tube with very sharp-edged fins has a very high condensation heat transfer coefficient in comparison with a smooth surface. A number of commercially enhanced tubes are now available in the market, such as Thermoexcel-C, Turbo-C, GEWA-SC, Tred-D. The basic idea to enhance the film condensation process is to make the liquid film as thin as possible over the possibly widest heat transfer areas.

Integral-finned tube has been used in liquid-liquid heat exchangers since 1940s[5]. Not long thereafter, it was introduced to condensing applications and shown efficient heat transfer performance in condensing. At the beginning, the major concern was to increase the heat transfer surface area. The effect of surface tension on the film condensation outside the finned surface was firstly recognized by Gregorig[6]. Surface tension forces may play a major role in the condensation process. The effect of surface tension pulls the condensate into concave grooves.

According to the research of Kedzierski and Webb[7, 8], when designing the fin thickness, two factors to consider are condensate retention and number of fins per meter. The fin efficiency can be increased by reducing the fin thickness. The condensate acts as an insulating blanket on the tube. High fin

Proceedings of the ASME 2014 International Mechanical Engineering Congress and Exposition IMECE2014

November 14-20, 2014, Montreal, Quebec, Canada

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1 Copyright © 2014 by ASME

heights can benefit heat transfer by reducing the row effect caused by condensate inundation. Thus, a small fin thickness is one factor that can lead to high heat transfer performance for finned tubes.

For the 3-D tubes, by cutting or plowing the basic grooves, the density of very sharp and acute angle tips is increased. Hence, the actual vapor-condensing portions are increased in number. The small sharp tips aided not only in forming thin film of condensate, but also dividing the liquid films into pieces. It may expose effective condensing surfaces and allowing rapid introduction of condensed liquid into grooves. Compared with the integral-finned tubes, the 3-D enhanced tubes are most widely used in shell and tube condensers. However, according to the prior researches [3, 8-12], the bundle effect of three dimensional surfaces is relatively intense and severe. Surface tension may also have detrimental effects on the condensation process of tube bundles, especially the three dimensional tubes.

The effect of inundation for enhanced tubes is not very clearly established. Through literature survey, it is found that the following papers studied the inundation effect of three dimensional and integrally finned tubes. Webb and Murawski [8] tested the condensing heat transfer outside the horizontal tube bundles. The tubes included low fin tube with fin density of 26fpi and three 3-D commercially available enhanced tubes: Turbo-C, GEWA-SC and Tred-D. Honda et al [14] studied the film condensation of R113 outside six different fin geometries, two low fin tubes(with fin density of 27 and 51fpi) and four three dimensional tubes. Cheng and Wang[15] tested the condensing heat transfer of R134a on one plain tube, three low-fin tubes (26fpi, 32fpi and 41fpi) and three three-dimensional finned tubes. Honda et al [17] studied the condensation heat transfer of R134a in a staggered 3×15 (column × row) tube bundle of two kinds of integral-finned tubes and three kinds of three dimensional tubes. Gstoehl and Thome[12] experimentally tested the film condensation of four three-dimensional tubes and one integral finned tubes on three tube arrays. Through the research above, it is found that different ratios of bundle effect are observed for the three dimensional tube. The condensate from upper tube has less effect on the condensing heat transfer of integrally finned tubes. The heat transfer coefficient of 3-D enhanced tubes decreased more or less as the increment of condensate flow rate.

Three-dimensional enhanced tubes are now widely used in the condensers of air conditioning systems[13, 14]. Taking into account the high price of material, the penalty of high cost is expected to increase. The trend is to develop more compact and cheaper heat exchangers or to increase the systems’ thermodynamic savings.

As the efforts to produce more efficient heat exchangers, high fin density of integrally finned tube is manufactured; Heat transfer performance of condensers with integral-finned and 3-D enhanced tubes is reported in this paper.

The rest of the paper is organized as follows. In the second section, condensing heat transfer of two single tubes is introduced, including the test loop and the specific structure of

the tested enhanced tubes. Then the heat transfer performance of two condensers is tested in a centrifugal chillers test rig. Finally, some conclusions are summarized in Section 4.

2. CONDENSING HEAT TRANSFER OF SINGLE TUBE 2.1 Experimental apparatus

An experimental apparatus for condensing heat transfer outside single horizontal tube has been used for the test. It consists of one refrigerant circulating system (including a condenser and a boiler) and water cooling systems. The schematic figure of the apparatus is shown in Fig.1. Liquid refrigerant is firstly heated electrically and boiled in the boiler, and then the refrigerant vapor goes up to the condenser via the connecting tube. In the condenser, it is condensed and the liquid condensate returns to the boiler vessel by gravity. The test tube is fixed in the condenser. After the cooling water circulates through the tested tube, it flows through a weight-time flow meter to measure the flow rates of cooling water, and then gets back to the water storage tank by a centrifugal pump.

10 9

8

T

P

T

T

P

TT TT T

7

6

5

43 2

1

(1)Boiler;(2)Condenser; (3)Thermocouple; (4)Pressure gauge; (5)Condensate measuring container; (6)Exhausting valve; (7)Electric heater; (8)Weight-time flow meter of Condensate controlling water cycle; (9)Water pump; (10) Water storage tank;

Fig.1 Schematic diagram of the experimental apparatus

A pressure gauge is used to monitor the pressure of the condenser. The tested range is from 0 to 2.5MPa, which has the precision of ±0.00625MPa. The power of the heater can be adjusted from 0 to 15kW, which is measured by a dynamometer with the accuracy of ±0.1W. Five platinum resistance temperature transducers (PT100) with a precision of ±(0.15+0.002|t|)K are used to measure the temperatures of the refrigerant in different part of the system. Thermocouples are used to measure the circulating water’s inlet and outlet temperature. The difference between inlet and outlet water’s temperature of water cycle is measured by six-junction copper-constantan thermocouple piles. The thermocouples and thermocouple piles were calibrated against a temperature calibrator that has the precision of ±0.2K. A Keithley digital


voltmeter having the resolution of ±0.1μV is used to measure the electric potential.

The cross section of the 3D enhanced and integral-fin tubes are shown in Fig.2. Integral fin structure is obtained by single rolling process with high, dense and approximate trapezoidal cross-section. 3-D tube is manufactured with rolling, plowing and extrusion process with sparse fins. Specifications are given in Table 1(See in the ANNEX), where di is the root diameter of inner embryo tube. The integral finned tube has fin density of 56fpi.

(a) Integral-fin tube

(b) 3-D-enhanced tube

Fig.2 Geometries of enhanced tubes

2.2. Data reduction

Heat balance is examined by the heat transfer rate of cooling water and electric heating power.

The power output from cooling water: ( )c c p in outq m c t t= − (1)

Where, int and outt are the inlet and outlet temperatures

of cooling water(K), pc is the specific heat capacity of cooling water corresponding to the mean temperature of inlet and outlet water(J/kg.K), cm is the mass flow rate of cooling water(kg/s). The properties of water are taken from [15].

The maximum difference between the heat transfer rates of cooling and heating is within 3%. The average is used to determine the overall heat transfer coefficients of tubes.

o m

qUA t

=⋅∆

(2)

External condensing heat transfer coefficient ho is separated from the overall thermal resistance:

ow

i i i o

1 1 1+ +A RU A c h h

= (3)

cihi is the internal convective heat transfer coefficient. hi is the internal heat transfer coefficient of smooth tube with the same Re and physical properties as internal grooved tube, determined by Gnielinski equations[16], the application range of which is: Re=2300-106, Pr=0.6-105. ci is the enhanced ratio of internal grooved tube compared with smooth tube, determined by Wilson plot technique[17, 18].

According to [19-20], the measurement uncertainty is estimated. The estimated uncertainties of overall heat transfer coefficient U are all less than 5.7%. ho is not measured directly, the uncertainty of which is estimated using the method suggested in [21]. The estimated uncertainty of ho for all tubes at the test range is within 35.1%.

2.3. Validation of experimental apparatus

In order to verify the reliability of experimental apparatus, experimental condensing heat transfer coefficient outside plain tube is firstly compared with Nusselt analytical solution. In Fig.3, condensing heat transfer coefficient of R134a is plotted against heat flux.

10 15 20 25 30 35 40 45 50

1

2

3

4

Ps=1.01MPaTs=40℃R134a

q' / (kWm-2)

h o/ (k

Wm

-2K-1

)

Plain Tube

Experimental Result Nusselt Analytical Solution

-10%

Fig.3 Comparison of experimental result with Nusselt

analytical solution for plain tube

As can be seen from the figure, at the heat flux from 10 to 40kW/m2, saturate temperature of 40℃, the deviation of experimental result and Nusselt analytical solution is basically in the scope of ±10%. The comparison should validate the experimental apparatus and procedure.

2.4. Heat transfer coefficient

In this sub-section, the overall, internal coolant and condensation heat transfer coefficient for the tested tubes are presented and compared.

Figs.4(a)(b)(c) shows the relation between overall heat transfer coefficient of two tubes and internal water velocity. The saturate temperature is 40℃ and internal water velocity is in the range of 1.5 to 2.5m/s. The overall heat transfer coefficient of 3-D tube is a little bit larger than 2-D tube at lower water velocity of 1.5m/s. In general, at the internal water


velocity greater than 2m/s, the deviations of heat transfer coefficients of different tubes are within 3% at heat flux 10, 20 and 30kW/m2.

1.0 1.5 2.0 2.5 3.010

11

12

13

14

15

16

17

v / (ms-1)

U / (

kWm

-2K-1

)


q' = 10k/m2

Integral-Fin 3 Dimensional

(a)Heat flux=10kW/m2

1.0 1.5 2.0 2.5 3.09

10

11

12

13

14

15

v / (ms-1)

q' = 20k/m2


U / (

kWm

-2K-1

)


(b) Heat flux=20kW/m2

1.0 1.5 2.0 2.5 3.09

10

11

12

13

14

15

q' = 30k/m2


v / (ms-1)

U / (

kWm

-2K-1

)


(c) Heat flux=30kW/m2

Fig.4 Overall heat transfer coefficient versus velocity at heat flux 10, 20 and 30kW/m2

The internal convective heat transfer coefficient of water is obtained by Wilson plot. The plot of the two enhanced tubes is shown in Fig.5. Overall thermal resistance1/U versus i1/h is plotted. The slope of integral tube is higher than 3-D enhanced tube. The enhanced ratio of the integral tube is 2.2 and 3-D enhanced tube is 2.7. As the fin-height of 3-D tube is higher and the helix angle is generally the same, hence, the enhanced ratio of 3-D tube is higher than integral-finned tube.

0.5 1.0 1.5 2.0 2.5

0.6

0.8

1.0

1.2

1.4

1.6

hi-1x104/(m2·K·W-1)

U-1

x104 /(m

2 ·K·W

-1)


Fig.5 Wilson plot of two enhanced tubes

4.6 6.48.2 28 46 64 820.640.82

2.84.66.48.2

28466482

Nusselt Analytical Solution

ho=a2q

-0.23


q' / (kWm-2)

ho/ (

kWm

-2K-1

)

Ts=40℃R134a

ho=a1q

-0.27

Fig.6 Condensing heat transfer coefficient versus heat flux

Fig.6 shows the condensing heat transfer coefficient of R134a versus heat flux. The heat flux is in the range of 0.7 to 126.5kW/m2. The heat transfer coefficient is decreasing as the increasing of heat flux. It is between 20 to 40kW/m2

·K at the experimental test range. A linear regression is fitted. The slope of integral-finned tube and 3-D enhanced tube are respectively of -0.27 and -0.23. For the Nusselt analytical solution, it is -0.33. Slightly larger difference of heat transfer coefficient is observed at lower heat flux. The fin density of integral-fin tube is 56fpi and 3-D enhanced tube is 43fpi. Compared with 3-D enhanced tube, the heat transfer area increased about 27.7% for integral-finned tube. The heat transfer coefficients of integral-finned tube are 10.8% to 25.8% higher than the 3-D


enhanced tube. Nusselt analytical solution is also presented in Fig.6. Compared with Nusselt analytical solution, the enhanced ratios of integral-fin tube are 17.7 to 21.0 and 3-D enhanced tube is in the range of 14 .3 to 20.0.

In the experiment, it is observed that the condensate from the integral-fin tube drips stably from a certain position, in droplet or column mode. It is found that dripping condensate position is moving in a certain axial direction outside the 3-D enhanced tube(Fig.7). The high fins may act as dams, which will block axial flow of the condensate. Hence, it is assumed that integral fin tube yielded the less row effect as the prior research[3, 8, 10, 11]. However, it should be stressed here that for a specific fin density, 3-D tube have more efficient heat transfer rate for an optimum geometry without the effect of condensate.

Fig.7 Moving direction of condensate column

3.HEAT TRANSFER OF CONDENSERS 3.1. Experimental apparatus

Heat transfer performance of condensers is tested in a centrifugal chiller test rig. The schematic diagram is shown in Fig.8(See in the ANNEX). The system includes three circuits: cooling water, chilled water and refrigeration.

P

T

P

T

T

P

P

TP

T

T

P

1

2

3

4

56

7

T P

PT

PT

PT

Chilled Water In

Chilled Water Out

O i l

Cooling Water In

Cooling Water Out

1. Centrifugal compressors; 2.Condenser; 3. Evaporator; 4. Oil separator; 5. Dry filter; 6. Sight glass; 7.Expansion valve;

Fig.8 Schematic diagram of centrifugal chiller test rig

Water is circulated with pump and distributed in the heat exchangers. Refrigeration circulations include centrifugal

compressor, condenser, evaporator, expansion valve and other fittings. In the evaporator, refrigerant R134a absorbs heat from chilled water, evaporates and enters the centrifugal compressor at a saturate state. It is compressed to a higher pressure and flow into the condenser at a superheated state. In the condenser, superheated vapor is cooled and condensed by the cooling water. Condensed liquid is next routed through the expansion valve and flows back to the evaporator. Since the lubricating oil is possibly to discharged from the compressor, oil separator is configured in the outlet of compressor. In order to observe the level and flowing of liquid, sight glass is mounted in the refrigeration cycle, evaporator and condenser.

After the condenser is being fixed, the whole system is charged with high pressure nitrogen to test the tightness of system. Firstly, high pressure is charged into the test rig and kept for at least 8 hours to ensure the whole system is well sealed. Secondly, the system is evacuated by a vacuum pump. Then the system is insulated with plastic rubber materials. Finally, refrigerant R134a is charged into the system.

Inlet temperature of chilled and cooling water can be adjusted according to the conditions of experiment. The cooling power of the test rig could even reach 2200kW. The temperature is measured with platinum resistance thermometer. The accuracy is ±0.2℃. Dynamometer is used to measure the motor power, precision of which is in the range of ±1.0W. Flow rate of water is measured with electromagnetic flow meter, with an accuracy of ±2%. Accuracy of pressure gauge and differential pressure transmitter are respectively in the range of ±1.5kPa and ±0.1kPa.

3.2. Parameters of condensers

Performance of two condensers is tested in this experiment. The tubes used in the condensers are listed in Table 1.The enhanced length of the 3-D enhanced tube and integral-finned tube fixed in the condenser is both 4000mm. The longitudinal tube pitch of integral-finned condenser is 22 mm and the transverse is 24mm. The integral-finned tube condenser has the tube numbers less than 15% compared with 3-D tube condenser. Since the internal volumes of two condensers are the same, the longitudinal and transverse tube pitch is a little bit closer for the 3-D tube condenser. It saves 80 tubes for the integral-finned tube condenser. Sub-cooling section is included in the bottom of two condensers, where the same types of tube are used. Layouts of tube bundle are staggered. Since the precise design is confidential, it cannot be given here.

3.3. Heat transfer performance of condensers

In this section, the condensation heat transfer of two condensers will be presented and compared. Table 2(See in the ANNEX) shows the refrigeration capacity and pressure drop of 3-D(1#) and integral-fin tube condenser(2#). Flow rate, motor power input, temperatures and pressures in different section is also presented. Test procedure and experimental conditions of the two condensers are kept the same. Two experimental conditions are tested in the experiment. Cooling water inlet temperatures are respectively of 32 and 30℃. The saturation


temperatures of condensers are respectively of 37.2 and 35.2℃. At the same motor power input, refrigeration capacity of water inlet temperature 30℃ is 5% higher than 32℃.

Firstly, it is observed from Table 2 that the refrigeration capacity of 2# condenser is generally the same as the 1# condenser at the same experimental conditions. The difference is less than 30kW. At the cooling water inlet temperature of 32℃, the refrigeration capacity of 1# condenser is 1727.5 and 1770.5kW, and 2# condenser is 1755.4 and 1769.4kW. At the same motor power input, the deviations of Coefficient of Performance(COP) are within 2%. Through analysis, the overall heat transfer coefficient of the 2# condenser is 16.7% higher than 1#.

Second, it should be note that the total flow cross sectional area of integral-fin tube condenser is 15% less than 3-D tubes. Pressure drop of 2# condenser is 11.2%-15.6% higher than 1# condenser. At the same cooling water flow rate 432m3/h, the internal water velocity inside the tube of #1condenser is 2.1m/s, and 2# is 2.5m/s. Internal-finned tube has an internal fin height smaller than 3-D tube. But the internal fins per circle are 2 ribs more than 3-D enhanced tube. The pressure drop of the two condensers is equivalent for its heat transfer area difference. The pressure drop per specific length of integral-fin tube should be very close to that of 3-D tube.

For energy and material savings considerations, integral-fin tubes here reduce heat exchange surface area by approximately 15%. Integral-fin can possibly reduce the size of a heat exchanger for a specified heat duty, but the penalty is increment of coolant flowing resistance.

The possible reasons to explain why integral-fin tube has less copper material consumption are as follows. Firstly, the condensers with such refrigeration capacity have more than 18 rows of tube bundle. The condensate inundation effect is severely for the lower tube bundles. Prior researches [3, 8, 10, 11] have also indicated that integral-fin tube has less bundle effect than the 3-D enhanced tubes. Secondly, for the integral finned tubes, the fin gap channeled the drainage of condensate, as the condensate left the tube, it does not influence the neighbor fin regions. Hence, it may leaves free condensate regions in the upper tube fin surface. This region is even not influenced by the condensate. The condensate impinging effect will enlarge the heat transfer area of low-fin tubes. The external fin height is higher than 3-D fins. The impinge effect of condensate to 3-D fins are limited as the fin height is low and the surface of tube is spread with film.

It is worth noting that compared with other researches for condensation outside tube bundles, one major novelty in this study is the high density integral-fin tubes with 56fins per inch and its applications in the tube bundles more than 18 rows.

The defect of high fin density integrally finned tube condenser is easy squeeze of fins in transportation or in the installment of long tubes in the heat exchangers. The squeeze or even cut-down may occur in the tube supporting plates and ends of condensers. This will decrease the heat transfer area of the tubes. Scraps from the tubes may also clog the filters of the

refrigerant circulations. Hence, it is recommended that additional protection measures are adopted in the installment of tubes to prevent the squeeze of external fins.

4. CONCLUSIONS

In this paper, the integrally finned tubes with higher fin density and fin height are designed and studied. The overall, internal and condensing heat transfer coefficients are compared with the 3-D enhanced tube. The thermo-hydraulic performance of two shell and tube condensers with integral-fin and 3-D enhanced tube is also investigated with an experimental approach. The major findings are as follows:

(1) Integral fin tubes with high integrally-finned density (fpi of 56) presents higher overall and condensing heat transfer coefficient than the 3-D enhanced tubes.

(2) At the same motor power input, the deviations of the refrigeration capacity for integral-fin tube condenser and 3-D enhanced tube condenserare within 1.6%.

(3) Copper material consumption of integral-fin tube is only 85% of 3-D condenser, while coolant pressure drop increases 11.2%-15.6%.

NOMENCLATURE a1 Fitted coefficient of integral-fin tube a2 Fitted coefficient of 3-D tube A Area，m2

ci Enhanced ratio of inside heat transfer coefficient

cp Specific heat capacity，J⋅kg-1⋅K-1 d Diameter of tube，mm e Height of outside fin，mm f Drag coefficient

h Heat transfer coefficients，W⋅m-2⋅K-1

U Overall heat transfer coefficients，W⋅m-2⋅K-1

L Tube’s tested length，m m Mass flow rate，kg⋅s-1 P Pressure, MPa q Heat transfer rate，W q’ Heat flux, kW/m2 r Latent heat of refrigerant, kJ/kg Re Reynolds number

Rw Thermal resistance of tube wall

t Temperature,℃; height of inside fin/ mm

Greek alphabet λ Thermal conductivity，W⋅m-1⋅K-1

Δtm Logarithmic mean temperature difference

Subscript


c Cooling i Inside of tube in Inlet of tube o Outside of tube out Outlet of tube s Saturation w Wall

ACKNOWLEDGMENT The supports from the National Key Projects of Fundamental R/D of China (973)(No.2013CB228304), Specialized Research Fund for the Doctoral Program of Higher Education (SRFDP) (No.20130201120057), Postdoctoral Science Foundation of China(No.2013M542344) and Chongqing Midea General Refrigeration Equipment Co, Ltd. of China are greatly ack-nowledged.

REFERENCES [1] Nakayama, W., Daikoku, T., Kuwahara, H., and Kakizaki, K., 1975, "High-Flux Heat Transfer Surface "THERMO-EXCEL"," Hitachi Review, 24, pp. 329-331. [2] Ali, H. M., and Briggs, A., 2012, "Condensation heat transfer on pin-fin tubes: Effect of thermal conductivity and pin height" Applied Thermal Engineering, 60(1), pp. 465-471. [3] [3]Gstoehl, D., and Thome, J. R., 2006, "Film condensation of R-134a on tube arrays with plain and enhanced surfaces: Part II---empirical prediction of Inundation effects," Journal of Heat Hransfer, 128(1), pp. 33-43. [4] Jung, D., Chae, S., Bae, D., and Yoo, G., 2005, "Condensation heat transfer coefficients of binary HFC mixtures on low fin and Turbo-C tubes," International Journal of Refrigeration, 28(2), pp. 212-217. [5] Marto, P., 1988, "An evaluation of film condensation on horizontal integral-fin tubes," Journal of Heat Transfer, 110(4b), pp. 1287-1305. [6] Gregorig, R., 1954, "Film condensation on finely rippled surfaces with consideration of surface tension," Z. Angew. Math. Phys, 5, pp. 36-49. [7] Kedzierski, M., and Webb, R., 1990, "Practical fin shapes for surface-tension-drained condensation," Journal of Heat Transfer, 112(2), pp. 479-485. [8] Webb, R. L., and Murawski, C. G., 1990, "Row effect for R-11 condensation on enhanced tubes," Journal of Heat Transfer, 112(3), pp. 768-776. [9] Honda, H., Uchima, B., Nozu, S., Torigoe, E., and Imai, S., 1992, "Film condensation of R-113 on staggered bundles of horizontal finned tubes," Journal of Heat Transfer 114(2), pp. 442-449. [10] Cheng, W. Y., and Wang, C. C., 1994, "Condensation of R134a on enhanced tubes," ASHRAE Transactions, 100(part 2), pp. 809-817. [11] Gstoehl, D., and Thome, J. R., 2006, "Film condensation of R-134a on tube arrays with plain and enhanced surfaces: Part I---experimental heat transfer coefficients," Journal of Heat Transfer, 128(1), pp. 21-32.

[12] Ji, W.-T., Zhao, C.-Y., Zhang, D.-C., He, Y.-L., and Tao, W.-Q., 2012, "Influence of condensate inundation on heat transfer of R134a condensing on three dimensional enhanced tubes and integral-fin tubes with high fin density," Applied Thermal Engineering, 38, pp. 151-159. [13] Marner, W., Bergles, A., and Chenoweth, J., 1983, "On the presentation of performance data for enhanced tubes used in shell-and-tube heat exchangers," Journal of Heat Transfer, 105(2), pp. 358-365. [14] Webb, R. L., and Kim, N. H., 2005, Principle of enhanced heat transfer, Taylor & Francis, Boca Raton. [15] Yang, S. M., and TAO, W. Q., 2006, Heat Transfer, Higher Education Press, Beijing. [16] Gnielinski, V., 1976, "New equations for heat and mass transfer in turbulent pipe and channel flows," Int. Chem. Eng., 16, pp. 359-368. [17] Rose, J. W., 2004, "Heat-transfer coefficients, Wilson plots and accuracy of thermal measurements," Experimental Thermal and Fluid Science, 28(2-3), pp. 77-86. [18] Cheng, B., and Tao, W. Q., 1994, "Experimental study of R-152a film condensation on single horizontal smooth tube and enhanced tubes," Journal of Heat Transfer, 116(1), pp. 266-270. [19] Kline, S. J., and Mcclintock, F. A., 1953, "Describing uncertainties in single-sample experiments," Mechanical Engineering, 75(7), pp. 3-9. [20] Moffat, R. J., 1988, "Describing the uncertainties in experimental results," Experimental Thermal and Fluid Science(1), pp. 3-17. [21] Cheng, B., and Tao, W. Q., 1994, "Experimental study of R-152a film condensation on single horizontal smooth tube and enhanced tubes," Journal of heat transfer, 116(1), pp. 266-270.


ANNEX A

TABLE 1. SPECIFICATIONS OF TUBES

Tubes Outside diameter do(mm)

Inside diameter

di(mm)

Height of outside Fin e(mm)

Outside fin numbers per inch

Height of inside fin

t(mm)

Length of test section L(mm)

Plain 19.17 16.40 - - - 1463 3-Dimensional 19.04 16.66 0.981 43 0.373 1369

Integral-fin 19.07 16.59 1.023 56 0.284 1464

TABLE 2. HEAT TRANSFER RATE AND PRESSURE DROP OF 3-D(1#) AND INTEGRAL-FIN TUBE CONDENSER(2#)

Item 3-D enhanced tube(1#) Integral-fin(2#) Exp. conditions 1 2 3 1 2 3

Chilled water inlet temperature /℃ 11.1 11.2 11.4 11.1 11.2 11.4 Chilled water outlet temperature /℃ 7.0 7.0 7.0 7.0 7.0 7.0

Chilled water flow rate /m3h-1 362.1 361.9 362.3 361.8 362.3 362 Cooling water inlet temperature/℃ 32.0 32.0 30.3 32.0 32.0 30.0

Cooling water outlet temperature/℃ 36.2 36.3 34.5 36.2 36.2 34.5 Cooling water flow rate/m3h-1 432.4 433.0 432.9 433.3 433.2 433.1

Cooling water pressure drop/kPa 95.1 97.9 96.0 110.7 110.0 111.1 Refrigeration capacity /kW 1727.5 1770.5 1842.4 1755.4 1769.4 1866.1

Suction temperature/℃ 5.6 5.9 5.6 5.6 6.0 6.8 Discharge temperature/℃ 48 49.2 47 49.4 49.2 47.2

Condensing saturate temperature /℃ 36.9 37.2 35.2 37.1 37.1 35.2

Condenser output temperature/℃ 36 33 34.6 34.4 34.6 34.4

Pressure before expansion/MPa 0.931 0.927 0.886 0.929 - 0.888

Pressure after expansion/MPa 0.437 0.435 0.435 0.432 - 0.437 Evaporating saturate temperature /℃ 6 6.1 5.8 6.1 5.8 5.7

Motor power/kW 402.4 407.7 404.2 402.4 402.4 412.6 Coefficient of performance 4.29 4.34 4.56 4.36 4.39 4.52


Documents

Experimental Study of Water Cooled Condenser Made of Three