The influence of surface structure and thermal conductivity of the tube on the condensation heat transfer of R134a and R404A over single horizontal enhanced tubes

Accepted Manuscript The influence of surface structure and thermal conductivity of the tube on the condensation heat transfer of R134a and R404A over single horizontal enhanced tubes Chuang-Yao Zhao, Wen-Tao Ji, Pu-Hang Jin, Ying-Jie Zhong, Wen-Quan Tao PII: DOI: Reference:

S1359-4311(17)33263-5 http://dx.doi.org/10.1016/j.applthermaleng.2017.06.133 ATE 10653

To appear in:

Applied Thermal Engineering

Received Date: Revised Date: Accepted Date:

11 May 2017 25 June 2017 28 June 2017

Please cite this article as: C-Y. Zhao, W-T. Ji, P-H. Jin, Y-J. Zhong, W-Q. Tao, The influence of surface structure and thermal conductivity of the tube on the condensation heat transfer of R134a and R404A over single horizontal enhanced tubes, Applied Thermal Engineering (2017), doi: http://dx.doi.org/10.1016/j.applthermaleng.2017.06.133

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The influence of surface structure and thermal conductivity of the tube on the condensation heat transfer of R134a and R404A over single horizontal enhanced tubes Chuang-Yao Zhaoa, Wen-Tao Jib, Pu-Hang Jinb, Ying-Jie Zhonga,*, Wen-Quan Taob,* a College of Mechanical Engineering, Zhejiang University of Technology, Hangzhou 310014, P. R. China b Key Laboratory of Thermal-Fluid Science and Engineering, MOE, Xi’an Jiaotong University, Xi’an 710049, P. R. China * Correspondent author: [email protected] (Ying-Jie Zhong) * Correspondent author: Tel./fax: +86 29 8266 9106, [email protected] (Wen-Quan Tao)

Abstract The influences of tube’s surface structure and thermal conductivity on the filmwise condensation of R134a and R404A are investigated on four single horizontal enhanced tubes with 2D and 3D finned structures. The test tubes are made of iron cupronickel and aluminum brass. The results indicate that the condensation heat transfer of R404A is more sensitive to surface structure and thermal conductivity than R134a; the 3D finned surface is more suitable than the 2D finned one to enhance the condensation heat transfer for the tubes with low thermal conductivity; the different material thermal conductivities produce different subcooling temperature distributions of the fins, and therefore provide different condensation heat transfer performances. Besides, the influencing mechanisms of the tube’s thermal conductivity on condensation heat transfer are elucidated. Keywords: Condensation; R134a; R404A; Thermal conductivity; Refrigeration

Nomenclature A

Area, m2

1

cp

Specific heat capacity, Jkg−1K−1

d

Diameter of tube, m

EF

Enhancement factor

g

Gravity acceleration, ms−2

h

Heat transfer coefficients, Wm−2K−1

k

Overall heat transfer coefficients, Wm−2K−1

L

Tested length of tube, m

m


Mass flow rate, kgs−1

P

Pressure, Pa

q

Heat flux, Wm−2

R

Thermal resistance, m2KW−1

r

Latent heat, Jkg−1

T

Temperature, °C

u

Velocity, ms−1

Γ

The mass flux of the refrigerant, kgm −1s−1

φ

Heat transfer rate, W

η

Efficiency index

λ

Thermal conductivity, Wm−1K−1

µ

Dynamic viscosity, kgm−1 s−1

ρ

Density, kgm−3

Greek

ΔT

Temperature difference, °C

Subscript

2

c

Condensing

e

Evaporating

flank

The fin flank

flood

The area flooded by the condensate

l LMTD

Liquid refrigerant Logarithmic mean temperature difference

i

Inside of tube

o

Outside of tube

s

Saturation

w

Wall

1. Introduction The filmwise condensation heat transfer outside a horizontal tube is widely encountered in the shell-and-tube condensers of refrigeration and air conditioning units. Previous studies have paid massive attention to enhance condensation heat transfer, in which the finned tubes play significant roles by providing higher heat transfer coefficients due to the change in the condensate surface profile under the effect of the surface tension [1-3], such as the 2D and 3D finned tubes [4, 5], pin finned tubes [6] and petal-shaped finned tubes [7], etc. The enhanced tubes are generally made of copper with high thermal conductivity to reduce the tube wall thermal resistance [8-10]. However, in some engineering applications with serious corrosion, erosion or impingement of high speed coolant, the tubes made of titanium [11], stainless steel [2] and alloy of copper containing nickel or manganese [12-14] with low thermal conductivities are employed. Many studies on 3

the condensation outside the copper tubes are available in literature, while the results about the low thermal conductivity tubes are less widely reported. The following is a brief review on the relevant investigations in recent years. The thermal conductivity of the tube has significant effects on the condensation heat transfer over the enhanced surfaces. Briggs and Rose [12] analyzed condensation heat transfer on horizontal integral finned tubes utilizing a semi-empirical model to account for fin efficiency effects. They found that the thermal conductivity has a weak effect on the optimum fin spacing but has a strong effect on the optimum fin thickness. Das et al. [15] experimentally studied the condensation heat transfer of steam outside eight single horizontal tubes, and the test tubes were made of stainless steel with rectangular fins. In their study, the heat transfer coefficients were found significantly affected by the surface tension of the condensate film and the thermal conductivity of the tubes. Zhang et al. [13] evaluated the condensation heat transfer performances of R134a and R12 on three enhanced tubes and the results suggested that the cupreous integral finned tube provides more than twice heat transfer coefficient than the cupronickel Thermoexcel-C tube. Yun et al. [16] measured the condensation heat transfer coefficient of R134a over stainless steel integral finned tubes. And the results indicated that the 19 fpi tube provides much higher heat transfer coefficients than 26 fpi tube due to the effects of condensate profile on the heat transfer coefficient. Ji et al. [2] experimentally studied filmwise condensation of R134a on eight enhanced tubes made of titanium, cupronickel (B10 and B30) and stainless steel, and found that the enhanced copper tube yields about 1.6 – 2.1 times of heat transfer coefficients than

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the low thermal conductivity tubes with the same surface structures, and that the mean enhanced ratio of titanium, B10, B30 and stainless steel tubes are 8.48, 8.31, 8.22 and 7.52, respectively. The effects of the tube’s thermal conductivity on condensation heat transfer vary with working fluid even though for the same surface structures. Briggs and Rose [12] reported that the thermal conductivity has a weak effect on the condensation heat transfer of R113 but has a strong effect on that of steam. Zhang et al. [13] found that R134a provides around 32.6% higher heat transfer coefficients than R12 for the Thermoexcel-C tube. The group of Fernández-Seara [11, 17, 18] studied the condensation heat transfer of R22 and mixture refrigerants (R417A, R422A and R422D) on the cupronickel Turbo-C (40 fpi) tube. They found that the increasing in the wall subcooling and the condensate Reynolds number decreases the heat transfer coefficients of R22 while increases the ones of the mixture refrigerants, and provides the rank of the condensation heat transfer coefficients of the four refrigerants: R22 > R422A > R422D > R417A. They also tested the condensation heat transfer of R134a and ammonia on the titanium integral finned (32 fpi) tube, and obtained the lower enhancement factors (from 0.77 to 1.22) utilizing ammonia while higher enhancement ratios (from 3.09 to 4.10) utilizing R134a. They pointed out that the different condensate retention fractions on the integral finned tube perimeter due to different surface tensions between ammonia and R134a are responsible for the different enhancement ratios, such that the flooded fraction of the tube perimeter ranges from 62.9% to 73.2% for ammonia (with higher surface tension) at saturation temperature

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from 30 °C to 45°C, respectively, while varies from 25% to 20% for R134a (with lower surface tension) at saturation temperature from 30 °C to 50°C, respectively. Ali and Briggs [14] conducted condensation tests of ethylene glycol and R113 on three identical pairs of pin finned tubes made of copper, brass and bronze, and found that the thermal conductivity has a weak influence on the condensation heat transfer for R113 while has a strong effect for ethylene glycol. Nowadays, R134a and R404A (R125/R143a/R134a = 44/52/4 mass %) are the most popular hydrofluorocarbons fluids used in the refrigeration systems mounted for commercial applications [19]. The study of R134a condensation outside the enhanced tubes has been widely reported, while the one of R404A is very few in the literature. R404A, a near zeotropic refrigerant, is widely used in marine condenser. According to the authors’ awareness, there are only several studies related to the condensation heat transfer of R404A inside tubes [20-22] while no information is available for filmwise condensation outside the tube. Besides, the existing publications have not been compared for the condensation heat transfer performances of R404A and R134a on the tubes with lower thermal conductivities. The condensation of R404A differs from that of R134a probably by following two features: a decrease in local heat transfer due to the apparent gliding temperature difference, and an extra mass transfer resistance caused by the vapor film between the bulk vapor and the condensate surface [7], both of which are related to the different thermal physical properties of the two refrigerants, as listed in Table 1. The present study is aiming to elucidate the influence of the thermal conductivity

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on filmwise condensation outside the enhanced tubes and to compare the condensation heat transfer coefficients of R134a and R404A over the tubes with low thermal conductivities. The test tubes are made of iron cupronickel and aluminum brass titanium, and are machined with 2D or 3D finned structures. In the following sections, the experimental apparatus and procedure, the data reduction process, as well as the results and discussion are stated in order, and finally the conclusions are drawn.

2. Experimental apparatus All experimental data in this study are obtained using the facility of phase change heat transfer outside the doubly-enhanced tube in Key Laboratory of Thermo-Fluid Science and Engineering of MOE in Xi’an Jiaotong University, as schematically shown in Fig. 1. From this figure, we can see three loops of refrigerant, heating water and cooling water. A brief introduction of the test rig is presented as follows, and it can be seen in details in our previous studies [1, 2, 13, 23]. The refrigerant loop contains a boiler, a condenser, a vapor upward pipe and a condensate downward pipe, all of which are made of stainless steel (AISI 304). The boiler vessel has an inner diameter of 257 mm and a total length of 1100 mm. And the condenser vessel has an inner diameter of 147 mm and a total length of 1500 mm. To guarantee the thermal insulation the whole apparatus is well covered with rubber plastic material of thickness 40 mm enwrapped by a layer of aluminum foil. During run, the liquid refrigerant in the boiler is heated to evaporate by the heating tubes flowing through hot water, and then the vapor flows into the condenser via the

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associate pipe, where the vapor is distributed over the test tube through a perforated plate to ensure uniform distribution and then condensates over the test tube, and finally the condensate flows back into the boiler by gravity via the downward pipe. The cooling water circulates through the test tube and returns to the cooling water tank driven by one submersible pump. And similarly, the heating water flows through the heating tube and returns to the hot water tank driven by another submersible pump. The temperatures of cooling and heating water are adjusted by their own electronic heating and cooling devices, respectively. The flow rates of cooling and heating water are measured by their own mass-time flow meter. A pressure gauge with a range of 2.5 MPa and an accuracy of 0.00625 MPa is used to measure the pressure of the condenser. The temperatures of refrigerant in the system (including the vapor and liquid phase) are measured by platinum resistance temperature transducers (Pt100) with accuracy of ± (0.15 ± 0.002|T|) K within the measurement range (T denotes the real temperature in K). The temperature difference between the water inlet and outlet is measured by a six-junction copper-constantan thermocouple pile. The thermocouples and thermocouple pile are calibrated against a temperature calibrator with an accuracy of ± 0.2 K. A Keithley digital voltmeter of 0.1 µV resolution is used to measure the electric potential. Table 2 shows the specification of test tubes. The materials of tubes are aluminum brass (copper/zinc: 78/22%) and iron cupronickel (copper/nickel: 70/30%), of which the conductivities are obtained in [24]. The values of outside and inside diameter listed in this table are based on the plain part of the enhanced tubes. In addition, for

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the convenience of comparison, all test tubes have plain inner surface.

3. Experimental procedure The condensing tube and heating tube are first cleaned and are then fixed in the condenser and boiler. And then the leakage hunting, airproof and pressure maintenance test for the entire system are conducted, during which the boiler and condenser are charged with nitrogen to a pressure of 1.2 MPa for R134a and 2.0 MPa for R404A, respectively, and all valves are closed. This pressure is maintained for 24 hours, until it remains constant except for small fluctuations in the surrounding condition. The nitrogen is then discharged, and the entire system is vacuumed to an absolute pressure of at least 800 Pa and is maintained for 24 hours. After completing the steps above, small quantity of refrigerant is charged into the system and then vacuumed and this procedure is repeated three times. Finally, the refrigerant is charged into the boiler with suitable quantity. In general, the effect of the non-condensable gas can be judged by the difference (± 0.2 K is allowable) between the measured temperature of the condenser and the saturation temperature obtained from the REFPROP database [25] corresponding to the measured pressure of the condenser, otherwise the system is heated to higher pressure and the exhausting valve is opened for some time till the temperature difference is within ± 0.2 K [26]. During the experiments, all results are recorded after the system reaches a stable state, which is characterized by the variation of the required saturation temperature of refrigerant and the inlet temperature of cooling water should be within ± 0.05 K.

4. Data reduction and uncertainty analysis 9

For the system, the input and output power by heating and cooling water is: φe = m
e c p ,e (Te,1 − Te,2 )  φc = m
c c p ,c (Tc,2 − Tc,1 )

(1)

where Te,1 and Te,2 denote the temperatures of the inlet and outlet heating water (K), respectively, Tc,1 and Tc,2 are the temperatures of inlet and outlet cooling water (K), respectively, m
e and m
c represent the mass flow rates of the two water circulations (kgm−1), and cp is the specific heat capacity (Jkg−1K−1) of water based on the mean temperature of inlet and outlet water. The properties of water are obtained in reference [26]. Considering the heat loss of the experimental system to the environment, the acceptable difference between the two heat transfer rates should be within 3% for each recorded data. The overall heat transfer coefficient of the condensation tube is determined by the following equation: k=

φ Ao ∆TLMTD

(2)

where the mean value of the two heat transfer rates, φ = (φe + φc)/2, is used to determine the overall heat transfer coefficient of the test tube; Ao is the nominal area based on the outside diameter of the plain embryo tube d o, and ∆TLMTD is the logarithmic mean temperature difference, which is defined as follows: ∆TLMTD =

Tc,2 − Tc,1

ln  (Ts − Tc,1 ) (Ts − Tc,2 ) 

(3)

where Ts denotes the saturation temperature. The overall thermal resistance can be expressed as the sum of the fouling resistance, and the resistances of inner single phase convection, wall conduction, and outside 10

condensation, as follows:

1 1 do 1 = + + Rw k hi di ho

(4)

where Rw is the thermal resistance of the wall, Ao and Ai are the outside and inside heat transfer area of the test tube based on the plain embryonic tube, ho and hi are the outside condensation heat transfer coefficient and inside convection heat transfer coefficient, respectively. The test tubes have plain inner surface, so the interior side average heat transfer coefficient is calculated by Gnielinski equation [27]. To study the influence of the condensate on the condensation heat transfer, the mass flux of the refrigerant, kg per meter of tube per second at one side of tube, can be determined by the condensation heat transfer rate, tube length and the latent heat of the refrigerant, as follows:

Γ =

φ 2rL

(5)

According to references [28, 29], uncertainty analysis of experimental data and the reduced results is performed. The confidence level for all measurement uncertainties are 95%. The estimated uncertainty of k is less than 5% for all the test points. So if the uncertainty of the internal coefficient predicted by Gnielinski equation is estimated as 10% [30], the uncertainty of ho, derived from the measured variables, is less than 20%.

5. Results and discussion 5.1 Validation by comparison with theoretical solutions

Validation experiment is executed to check the reliability of the test rig. The condensation heat transfer coefficients outside a horizontal plain titanium tube 11

according to the Nusselt analytical solution [31,32] can be obtained from Eq. (6). 14

 gr ρ l2λl3  ho = 0.729    µl d o (Ts − Tw ) 

(6)

In Fig. 3, experimental results are compared with the Nusselt analytical solution [31,32] for R134a condensation outside a horizontal plain copper-nickel tube at the saturation temperature of 40 °C. The deviations of the present study from Nusselt analytical solution are within ± 10%. So the experimental data in this study are reliable. 5.2 Overall heat transfer performance

Figures 4 and 5 display the dependences of overall heat transfer coefficients of four test tubes on the water velocity, in which the mass fluxes of the refrigerants based on Eq. (5) are also displayed. The water velocity varies from 0.8 to 2.3 ms‒1 and the saturation temperature and inlet water temperature are 40 °C and 32 °C, respectively, for both refrigerants. All tubes has plain internal surface, and therefore the differences in the overall heat transfer coefficients are caused by thermal conductivities and outside enhanced structures under the same range of water velocities. As shown in Figs. 4 and 5, the 3D finned aluminum brass tube (No. 3) exhibits the highest overall heat transfer coefficients, the 2D finned iron cupronickel tube (No. 2) yields the lowest overall heat transfer coefficients, and the tubes Nos. 1 and 4 have moderate performance. Regardless of R134a and R404A, for the tubes with the same heat transfer conductivity, like No. 1 vs. 2, as well as No. 3 vs. 4, the 3D finned structures provide superior overall heat transfer performance. It is because the longitudinal channels (or 12

circumferential cuts among fins) prevent condensate retention among fins while the condensate always retains in the long channels among the integral fins of 2D finned tubes, which makes 3D fins having smaller thermal resistance than 2D fins during condensation. For the tubes with the same enhanced structure, like No. 1 vs. 3, as well as No. 2 vs. 4, the tube higher thermal conductivity is favorable to enhance the overall heat transfer due to the higher fin efficiency, the mechanism of which will be provided in the later section. The different slopes of the curves indicate different dependences of the overall heat transfer coefficients on the coolant velocity, which determines the inside thermal resistance. Table 3 lists the thermal resistance ratios of each part at the water velocity of 2.0 ms‒1 and saturation temperature of 40 °C for both refrigerants, from which it can be seen that the thermal resistance ratios vary with the tube thermal conductivity and surface structure. Taking the cases of R404A for example, tube No. 3 has 3D finned surface and a larger thermal conductivity, namely it has the smallest ratio of thermal resistance of tube wall and outside condensation heat transfer (33.06% + 9.56%), so the overall heat transfer coefficients are appreciably dependent on the water velocity. While tube No. 2 has the largest ratio of thermal resistance of tube wall and outside condensation heat transfer (47.21% + 20.23%), so its overall heat transfer coefficients are weakly dependent on the water velocity. From comparison between Fig. 4 and Fig. 5, we can see that the results of R404A are more scattered than R134a among four tubes, which implies that the surface structure and thermal conductivity have more prominent effects on the condensation

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heat transfer of R404A than that of R134a. Taking the cases of R134a for instance, the overall heat transfer coefficient, k, of tube No. 3 is on average 30% higher than that of tube No. 4, and k of tube No. 3 is on average 32% higher than that of tube No. 1. While for R404A, k of tube No. 3 is on average 38% higher than that of tube No. 4, and k of tube No. 3 is on average 40% higher than that of tube No. 1. These findings are reasonable considering the different surface tensions of the two refrigerants. The drainage of the condensate is much easier for R404A due to smaller surface tension, namely, the flooded area by the condensate of R404A is smaller than that of R134a, which is favorable to condensation heat transfer. 5.3 Condensation heat transfer over plain tubes

Figure 6 compares the experimental results of the condensation heat transfer coefficients of R134a on several plain tubes (all tubes have 19.0 mm external diameter) made of different materials at saturation temperature of 40 ºC (some results are cited from Ji et al. [2] and Ji [33]), where the range of thermal conductivity spans from 10 to 398 Wm−1K−1. As shown in Fig. 6, the deviations of the heat transfer coefficients from the Nusselt analytical solutions [31] are all within ±10%. To further confirm the independence of condensation heat transfer outside the plain tube on the thermal conductivity, the comparisons of the experimental results of condensation heat transfer coefficients of water steam outside the plain tubes with different thermal conductivities are shown in Fig. 7, where the results are collected from literatures [15, 33-35] and the tubes made of titanium, stainless steel and copper. As seen in Fig. 7, the deviations of these results from the Nusselt theoretical solutions

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are all within 10% at saturation temperatures of 100 ºC and 54-69ºC. Thus we can conclude that the condensation heat transfer coefficients outside plain tubes are independent on the tubes’ materials, and are only related to the thermal physical property and saturation temperature and can be predicted by Nusselt theoretical solutions [31,32]. It is worth noting that for the conditions with a smaller heat transfer rate, for example, the wall subcooling is smaller than 1 ºC, the condensation heat transfer coefficients of the plain tube are much lower than those predicted by the Nusselt theoretical solutions [31], such as lower than 18% for R134a on copper tube [16], lower than 10% for R134a on copper tube [37], and lower than 13% for R12 [38]. These phenomena can be explained by the numerical results of Koch et al [39]: the tube material has significant effects on the condensation heat transfer over the plain tube if the condensate liquid film is very thin when the thermal resistance of the surface roughness is comparative to that of the condensate film. 5.4 Condensation heat transfer over enhanced tubes

Figs. 8 and 9 demonstrate the variations of the condensation heat transfer coefficients outside the tubes with different enhanced types and materials. It can be seen that the condensation heat transfer coefficients increase obviously with increasing the heat flux for the tube No. 3 of R134a and tubes Nos. 1 and 3 of R404A. The similar trends on 2D and 3D finned tubes have been reported by many researchers [7, 18, 39]. This variation trend is obviously different from that of pure working medium for which condensation heat transfer coeffcient decreases with the

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increae of heat flux. According to [7] this abnormal variation trend may be attributed to the fact that for the condensation of zeotropic refrigerant mixture mass transfer mainly occurring in a vapor diffusion layer formed between the condensate film and the vapor bulk, in which the thermal resistance reduces appreciably due to the vapor diffusion as the heat flux (or wall subcooling) increases [7]. Specifically, Fig. 8 reveals that for the tubes with the same enhanced structure, the one with larger thermal conductivity provides higher condensation heat transfer coefficient, such as h o of tube No. 3 is on average 34% higher than that of tube No. 1 and h o of tube No. 4 is on average 20% higher than that of tube No. 2; For the tubes made of the same material, the h o of tube No. 1 is on average 30% higher than that of tube No. 2 and h o of tube No. 3 is on average 49% higher than that of tube No. 4. In Fig. 9, we can observe that the similar trends as displayed in Fig. 8. Through comparison between Fig. 8 and Fig. 9, it can be seen that for the 3D finned tubes, the thermal conductivity has more significant effect on ho for R404A than on that for R134a (average gap of 75% vs. 34%), while for the 2D finned tubes, the thermal conductivity effect is almost negligible between R404A and R134a (average gap of 15% vs. 20%); for aluminum brass tube, the fin type has greater influence on ho of R404A than on that of R134a (average gap of 125% vs. 49%), but fin type effect for iron cupronickel is not so strong between R404A and R134a (average gap of 50% vs. 30%). In summary, the enhanced structure and thermal conductivity have more significant impacts on the condensation heat transfer performance for R404A than for R134a.

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The comparison of the two refrigerants at different heat flux are plotted together in Fig. 10, from which we can find that: R134a provides higher condensation heat transfer coefficient, h o, than R404A over the 2D finned tubes (Nos. 2 and No. 4), and the difference between the two refrigerants on tube No. 4 is more prominent than on tube No. 2; The difference of ho between R134a and R404A on tube No. 1 is insignificant except at the heat flux smaller than 40 kWm‒2; The condensation heat transfer performance of R404A over tube No. 3 is obviously superior to R134a. These comparison results indicate that the thermal physical properties of the refrigerants have significant influence on condensation heat transfer. For the two refrigerants the most different one is the liquid surface tension which is related to the condensate retention between fins (the surface tension of R134a is about 47% higher than that of R404A); The second important difference is the liquid viscosity (R404A’s value is about 35% lower than that of R134) and a lower value is helpful to the condensate moving downward. Thus there is an appropriate fin spacing depending strongly on fluid properties as indicated in [40] , notably surface tension and viscosity. 5.5 Influence mechanism of tube’s thermal conductivity on condensation heat transfer

According to the above analysis, we can find that the thermal conductivity of the tube has significant effects on the condensation heat transfer when the tube has exterior enhanced structures, which may be an easily accepted fact, but the mechanisms need to be further understood. The following analysis coupling the temperature variation of the tube wall could provide a quantitative understanding of

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the tube conductivity’s influence. Rose [41] proposed a semi-empirical model of condensation heat transfer on a horizontal tube with integral rectangular fins (2D fins), and Briggs and Rose further considered the fin efficiency effect [12]. This model classified the tube surface during condensation into three regions: flooded region by the condensate, fin flank, and the unflooded inter-fin region, and the heat transfer rates of each part are called φflooded,

φfinflank and φintfin, respectively. And therefore the enhancement factor of condensation heat transfer for enhanced tubes is expressed as:

EF = (φflooded + φfinflank + φintfin ) φplain

(7)

where φplain is a reference heat transfer rate on the plain tube with length equal to one fin pitch of finned tube, which is independent of the tube’s material. According to the expressions of all heat transfer rates (see in [12] for detail), there exist positive relationships between (1) φflooded and the temperature difference ∆Tflooded (= Ts–(Ttip –Troot)1/3) in flooded region, (2) between φfinflank and average vapor side temperature

differences ∆Tfinflank (= (Ts–Tflank)3), and temperature difference ∆Ttipave (= (Ts–Ttip)3) at fin tip and fin flank, and (3) between φintfin and temperature difference ∆Tintfin (= (Ts –Tintfin)3) at inter-fin space. In this sense, an enhanced condensing tube made of

material with low thermal conductivity produces higher Ttip, Tflank and Tintfin, namely provides smaller ∆Tfloofed, ∆Tfinflank, ∆Ttipave and ∆Tintfin, and undoubtedly yields smaller φflooded + φfinflank + φintfin, namely a smaller condensation enhancement factor. And therefore, for the tubes with the identical enhanced structure, the one with lower thermal conductivity provides a smaller heat transfer coefficient. This explanation can

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be also used to describe the effect of thermal conductivity on the condensation heat transfer outside the enhanced tubes with complicated fins, such as 3D fin, which can be viewed as the composition of primary, secondary even tertiary fins. While for the plain tubes made of different materials, there exists Ttip = Tflank = Tintfin = Troot, and all tubes have the uniform outside surface temperature, so all enhancement factors equal to 1, and the heat transfer coefficients can be predicted by Nusselt theoretical solutions [31], as seen in Figs. 6 and 7. According to the above analysis, we can understand that the temperatures of the fin surface (Ttip, Tflank and Tintfin) are the key parameters determining the condensation heat transfer performance of an enhanced tube, which are influenced by the thermal conductivity of the tube. In other words, the different thermal conductivities produce various subcooling distributions along the fins with the same geometries, and then provide different thermal resistances and finally lead to different condensation heat transfer performances. The above discussion may explain why the two refrigerants R134a and R404A show different responses to the tube wall thermal resistance, i.e, in Figs.8 and 9 R404A is more sensitive to the enhanced structure than that of R134a. Both the different thermophysical properties of the two tested refrigerants and the different temperature difference between saturation and tube wall surface caused by different material thermal conductivity eventually leads to different condensation performances of the two tested tubes. 5.6 Comparison with previous predicted correlations of pure refrigerant

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The existed filmwise condensation heat transfer correlations outside a horizontal tube are generally based on the pure refrigerants, while the condensation heat transfer prediction of the near zeotropic refrigerant is rarely studied. However, it is convenient and practical to use the existed condensation correlations of the pure refrigerants to reveal whether they have some approximate feasibility . Motivated by this consideration, the following comparison is carried out. The heat transfer coefficients of the present two 2D finned tubes are compared with the previous correlations for tubes with low integrated fin developed by Beatty-Katz [43], Briggs-Rose [12], Owen [44], Rose [42] and Webb [45], and the results are displayed in Fig. 11. From the figure following features may be noted . (1) With the increase of heat flux the ratio of ho,exp/ho,pre increases from less than 1 to larger than 1, indicating that for low heat flux the predicted values are larger than the test data while at high heat flux the predicted ones are less than the test ones. (2) The general agreement between predicted values and test data for R134a is better than that of R404A, which is obvious since R134a is a pure substance. Within ±20% deviation correlations of [42-44] can be applied for tube No. 2, and correlations of [42, 43] can be applied for tube No. 4. (3) Even for R404A within the same deviation correlations of [42, 43] can be applied for both tube Nos. 2 and 4. Such comparison is meaningful in that before enough test data are accumulated such that specific correlations can be developed for R404A, for the time being correlations of [42,43] can be used for condensation of R404A outside tube Nos. 2 and 4.

6. Conclusions 20

The condensations of R134a and R404A have been experimentally investigated on four single horizontal tubes with 2D and 3D finned structures and different tube thermal conductivities. Based on the experimental results, the following conclusions are obtained: (1) The thermophysical properties of R404A are more favorable to the moving of condensate between fins (low liquid surface tension and viscosity), hence the condensation heat transfer of R404A is more sensitive to the enhanced structure than that of R134a. (2) The tube with higher thermal conductivity can provide higher condensation heat transfer coefficient, and the 3D finned surface is more suitable than the 2D finned one to enhance the condensation heat transfer for the lower thermal conductivity condensing tube. (3) For tubes Nos. 1, 2 and 4, R134a provides higher condensation heat transfer coefficients than R404A; for tube No. 3, R134a provides lower condensation heat transfer coefficients than R404A. (4) The diferent material thermal conductivities produce different subcooling temperature distributions of the fins, and therefore provide different condensation heat transfer performances. The lower the material thermal conductivity, the smaller the subcooling temperature distribution, and the lower the outside condensation heat transfer coefficient.

Acknowledgment This work was supported by the National Key Basic Research Program of China

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(973 Program) (2013CB228304)

References [1] W.T. Ji, C.Y. Zhao, D.C. Zhang, Y.L. He, W.Q. Tao, Influence of condensate inundation on heat transfer of R134a condensing on three dimensional enhanced tubes and integral-fin tubes with high fin density, Appl. Therm. Eng., 38 (2012) 151-159. [2] W.T. Ji, C.Y. Zhao, D.C. Zhang, Z.Y. Li, Y.L. He, W.Q. Tao, Condensation of R134a outside single horizontal titanium, cupronickel (B10 and B30), stainless steel and copper tubes, Int. J. Heat. Mass. Transf., 77 (2014) 194-201. [3] A.R. Al-Badri, A. Bär, A. Gotterbarm, M.H. Rausch, A.P. Fröba, The influence of fin structure and fin density on the condensation heat transfer of R134a on single finned tubes and in tube bundles, Int. J. Heat. Mass. Transf., 100 (2016) 582-589. [4] S.K. Sajjan, R. Kumar, A. Gupta, Experimental investigation during condensation of R-600a vapor over single horizontal integral-fin tubes, Int. J. Heat. Mass. Transf., 88 (2015) 247-255. [5] D.J. Kukulka, R. Smith, W. Li, Comparison of condensation and evaporation heat transfer on the outside of smooth and enhanced 1EHT tubes, Appl. Therm. Eng., 105 (2016) 913-922. [6] H.M. Ali, M. Abubaker, Effect of circumferential pin thickness on condensate retention as a function of vapor velocity on horizontal pin-fin tubes, Appl. Therm. Eng., 91 (2015) 245-251. [7] Z. Zhang, Q. Li, T. Xu, X. Fang, X. Gao, Condensation heat transfer characteristics of zeotropic refrigerant mixture R407C on single, three-row petal-shaped finned tubes and helically baffled condenser, Appl. Therm. Eng., 39 (2012) 63-69. [8] C.Y. Zhao, W.T. Ji, P.H. Jin, W.Q. Tao, Cross vapor stream effect on falling film evaporation in horizontal tube bundle using R134a, Heat. Transfer. Eng., In press (2017). [9] C.Y. Zhao, P.H. Jin, W.T. Ji, L.H. Ya, W.Q. Tao, Experimental investigations of R134a and R123 falling film evaporation on enhanced horizontal tube, Int. J. Refrig., 75 (2017) 190-203. [10] W.T. Ji, A.M. Jacobi, Y.L. He, W.Q. Tao, Summary and evaluation on single-phase heat transfer enhancement techniques of liquid laminar and turbulent pipe flow, Int. J. Heat. Mass. Transf., 88 (2015) 735-754. [11] J. Fernández-Seara, F. Uhía, J., D. Rubén, D. Alberto, Condensation of R-134a on horizontal integral-fin titanium tubes, Appl. Therm. Eng., 30 (2010) 295-301. [12] A. Briggs, J.W. Rose, Effect of fin efficiency on a model for condensation heat transfer on a horizontal, integral-fin tube, Int. J. Heat. Mass. Transf., 37 (1994) 457-463. [13] D.C. Zhang, W.T. Ji, W.Q. Tao, Condensation heat transfer of HFC134a on horizontal low thermal conductivity tubes, Int. Commun. Heat Mass, 34 (2007) 917-923. [14] H.M. Ali, A. Briggs, Condensation heat transfer on pin-fin tubes: Effect of thermal conductivity and pin height, Appl. Therm. Eng., 60 (2013) 465-471. [15] A. Das, G.A. Incheck, P.J. Marto, The effect of fin height during steam condensation on a horizontal stainless steel integral-fin tube, J. Enhanc. Heat. Transf., 6 (1999). [16] R. Yun, J. Heo, Y. Kim, Film condensation heat transfer characteristics of R134a on horizontal stainless steel integral-fin tubes at low heat transfer rate, Int. J. Refrig., 32 (2009) 865-873. [17] F.S. José, J.U. Francisco, R. Diz, A. Depazo, Experimental analysis of ammonia condensation on smooth and integral-fin titanium tubes, Int. J. Refrig., 32 (2009) 1140-1148. [18] J. Fernández-Seara, F.J. Uhía, R. Diz, J.A. Dopazo, Vapour condensation of R22 retrofit

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substitutes R417A, R422A and R422D on CuNi turbo C tubes, Int. J. Refrig., 33 (2010) 148-157. [19] K. Michael, Trends and Perspectives in Supermarket Refrigeration, in: International Technical Meeting on HCFC phase-out, 05/04/2008. The Directorate-General for Climate Action, 2015. [20] B.M. Fronk, S. Garimella, In-Tube Condensation of Zeotropic Fluid Mixtures: A Review, Int. J. Refrig., 36 (2013) 534-561. [21] H. Charun, Thermal and flow characteristics of the condensation of R404A refrigerant in pipe minichannels, Int. J. Heat. Mass. Transf., 55 (2012) 2692-2701. [22] P.A. Patil, S.N. Sapali, Condensation pressure drop of HFC-134a and R-404A in a smooth and micro-fin U-tube, Exp. Therm. Fluid. Sci., 35 (2011) 234-242. [23] W.T. Ji, D.C. Zhang, Y.L. He, W.Q. Tao, Prediction of fully developed turbulent heat transfer of internal helically ribbed tubes – An extension of Gnielinski equation, Int. J. Heat. Mass. Transf., 55 (2012) 1375-1384. [24] T.Y. S, P.R. W, C.C. Y, Thermophysical properties of matter. Vol.7 Thermal conductivity: metallic elements and alloys, in, New York:IFI/Plenum Press, 1970. [25] E.W. Lemmon, M.O. McLinden, M.L. Huber, Reference Fluid Thermodynamic and Transport Properties (REFPROP), Version 8.0, in, National Institute of Standards and Technology (NIST), 2008. [26] S.M. Yang, W.Q. TAO, Heat Transfer, 4th Edition, Higher Education Press, Beijing, 2006. [27] V. Gnielinski, New equations for heat and mass transfer in the turbulent flow in pipes and channels, International Chemical Engineering, 16 (1976) 359-368. [28] B. Cheng, W.Q. Tao, Experimental study of R-152a film condensation on single horizontal smooth tube and enhanced tubes, ASME J. Heat Transfer, 116 (1994) 266-270. [29] S.J. Kline, The purposes of uncertainty analysis, ASME J. Fluids. Eng., 107 (1985) 153-160. [30] Y.A. Cengel, A.J. Ghajar, Chapter eight: Internal forced convection, in: Heat and Mass Transfer, Fundamentals and Applications, 4th Edition, McGraw-Hill, New York, 2011, p. 489. [31] W. Nusselt, Die Oberflachencondensation Des Wasserdampfes, VDI, 60 (1916) 541-569. [32] V.K.Dhir, J.H.Linhard, Laminar film condensation on plane and axisymmertic bodies in non-uniform gravity. ASME J. Heat Transfer, 93(1971) 97-105 [33] W.T. Ji, Experimental and Application Study of Enhanced Refrigerant Phase Change Heat Transfer and Water-cooling Condenser, (Ph.D Thesis), Xi'an Jiaotong University, 2012. [34] M.H. Jaber, R.L. Webb, Steam condensation on horizontal integral-fin tubes of low thermal conductivity, J. Enhanc. Heat. Transf., 3 (1996) 55-71. [35] K. Hwang, J. Jeong, S. Hyun, K. Saito, S. Kawai, K. Inagaki, R. Ozawa, Heat transfer and pressure drop characteristics of enhanced titanium tubes, Desalination, 159 (2003) 33-41. [36] M.H. Jaber, A theoretical and experimental study of steam condensation on horizontal enhanced tubes, (Ph.D Thesis), The Pennsylvania State University, 1991. [37] R. Kumar, H.K. Varma, B. Mohanty, K.N. Agrawal, Prediction of heat transfer coefficient during condensation of water and R-134a on single horizontal integral-fin tubes, Int. J. Refrig., 25 (2002) 111-126. [38] R.E. White, Condensation of refrigerant vapors: apparatus and film coefficients for R-12, Refrig. Eng,, 55 (1948) 375–379. [39] G. Koch, K. Kraft, A. Leipertz, Parameter study on the performance of dropwise condensation, Revue Générale de Thermique, 37 (1998) 539-548. [40] D. Jung, S. Chae, D. Bae, G. Yoo, Condensation heat transfer coefficients of binary HFC mixtures on low fin and Turbo-C tubes, Int. J. Refrig., 28 (2005) 212-217.

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[41] J.W. Rose, Surface Tension Effects and Enhancement of Condensation Heat Transfer, Chemical Engineering Research and Design, 82 (2004) 419-429. [42] J.W. Rose, An approximate equation for the vapour-side heat-transfer coefficient for condensation on low-finned tubes, Int. J. Heat. Mass. Transf., 37 (1994) 865-875. [43] K.O. Beatty, D.L. Katz, Condensation of vapors on outside of finned tubes, Chem. Eng. Prog., 44 (1948) 55-70. [44] R. Owen, R. Sardesai, R. Smith, W. Lee, Gravity controlled condensation and horizontal low-fin tube, in: Condensers: theory and practice. Symposium, 1983, 415-428. [45] T. Rudy, R. Webb, An analytical model to predict condensate retention on horizontal integral-fin tubes, ASME J. Heat. Transfer., 107 (1985) 361-368.

Figure captions Fig. 1 Schematic diagram of experimental apparatus Fig. 2 The photos of the test tubes, (a) 2D finned, (b) 3D finned Fig. 3 Comparisons of experimental condensation heat transfer coefficients with the Nusselt predictions for plain titanium tube Fig. 4 Variation of overall heat transfer coefficients of R134a versus water velocity Fig. 5 Variation of overall heat transfer coefficients of R404A versus water velocity Fig. 6 Condensation heat transfer coefficients of R134a versus heat flux outside plain tubes Fig. 7 Condensation heat transfer coefficients of water steam versus heat flux outside plain tubes Fig. 8 Condensation heat transfer coefficients of R134a versus heat flux outside enhanced tubes Fig. 9 Condensation heat transfer coefficients of R404A versus heat flux outside enhanced tubes Fig. 10 Comparisons of condensation heat transfer coefficients of R134a and R404A versus heat fluxes on four enhanced tubes Fig. 11 Comparison of experimental results with predicted correlations for 2D finned tubes

24

(1) Boiler; (2) Condenser; (3) Thermal couple; (4) Pressure gauge; (5) Condensate measuring container; (6) Exhausting valve; (7) Hot water flow meter; (8) Cooling water pump; (9) Cooling water tank; (10) Cooling water flow meter; (11) Hot water pump; (12) Hot water tank; (13) Condensation tube; (14) Heating tube; (15) Cooling water heater; (16) Hot water heater; (17) Cooling water condenser; (18) Hot water condenser.

Fig. 1 Schematic diagram of experimental apparatus

25

Fig. 2 The photos of the test tubes, (a) 2D finned, (b) 3D finned

26

3.5

Nusselt prediction [31, 32] Experimental Result

3.0 2.5 2.0 -2

ho/ kWm K

-1

+10%

1.5

1.0

-10%

R134a Ts=40°C Ps=1.01Mpa

0.5 10

15 20 -2 q/ kWm

25

30 35 40 45 50

Fig. 3 Comparisons of experimental condensation heat transfer coefficients with the Nusselt predictions for plain titanium tube

27

5.5 5.0 4.5

No.1 No.2 No.3 No.4

R134a Ts=40° C Ps=1.014Mpa −1 −1

-2

k/ kWm K

-1

Γ=4.23−6.29gm s

4.0 30%

−1 −1

Γ=3.15−4.24gm s

32%

3.5

−1 −1

Γ=3.51−4.40gm s

3.0

−1 −1

Γ=2.86−4.02gm s

2.5 2.0 0.8

1.0

1.2

1.4

1.6 1.8 -1 u/ ms

2.0

2.2

2.4

Fig. 4 Variation of overall heat transfer coefficients of R134a versus water velocity

28

5.5 5.0

R404A Ts=40° C Ps=1.829Mpa

−1 −1

Γ=5.40−8.81gm s

-2

k/ kWm K

-1

4.5

No.1 No.2 No.3 No.4

4.0 38%

3.5

40%

−1 −1

Γ=4.55−6.29gm s

3.0

Γ=4.18−5.85gm s

2.5

Γ=3.81−5.15gm s

2.0 0.8

−1 −1

−1 −1

1.0

1.2

1.4

1.6 1.8 -1 u/ ms

2.0

2.2

2.4

Fig. 5 Variation of overall heat transfer coefficients of R404A versus water velocity

29

3.5 2.5

R134a Ts=40° C

2.0

Ps=1.01Mpa

3.0

-2

ho/ kWm K

-1

+10%

1.5

Nusselt prediction [31] Titanium, Present study Aluminum brass, Ji et al. [2] 2014 Stainless steel, Ji et al. [2] 2014 Iron cupronickle, Ji et al. [1] 2012 Copper, Ji et al. [1] 2012

1.0

0.5 5

10

15 20 -2 q/ kWm

25 30 35 40 4550

Fig. 6 Condensation heat transfer coefficients of R134a versus heat flux outside plain tubes

30

18 16

+10%

Steam

14 +10%

-2

ho/ kWm K

-1

12 10 8

6

T =5 4-69 s

T =1 s 00°C °C

Nusselt prediction [31] Copper, Jaber [35]1991 Stainless steel, Jaber and Webb [33] 1996 Titanium, Hwang et al. [34] 2003 Stainless steel, Das et al. [15] 1999 10

20

30

40

50 60 70

∆T/ K

Fig. 7 Condensation heat transfer coefficients of water steam versus heat flux outside plain tubes

31

-1 -2

ho/ kWm K

14 13 12 11 10 9 8

34%

20%

7

49%

30%

6 5

No.1 No.2 No.3 No.4

R134a Ts=30°C

4

Ps=0.771Mpa 3

20

30

40 50 -2 q/ kWm

60

70 80 90 100

Fig. 8 Condensation heat transfer coefficients of R134a versus heat flux outside enhanced tubes

32

16 14 12 10

75%

-2

ho/ kWm K

-1

125%

8 50%

6

15%

No.1 No.2 No.3 No.4

R404A Ts=30°C

4

Ps=1.432Mpa 20

30

40

50

60

70

80 90 100

-2

q/ kWm

Fig. 9 Condensation heat transfer coefficients of R404A versus heat flux outside enhanced tubes

33

20

No.1 R134a No.2 R134a No.3 R134a No.4 R134a

18

14

-2

ho/ kWm K

-1

16

No.1 R404A Ps-R134a=0.77Mpa No.2 R404A No.3 R404A Ps-R404A=1.43Mpa No.4 R404A

12 10 8 6 4

20

30

40

50 60 -2 q/ kWm

70

80

90

100

Fig. 10 Comparisons of condensation heat transfer coefficients of R134a and R404A versus heat fluxes on four enhanced tubes

34

1.4

No.2 R134a Ts=30°C

Beatty-Katz [42] Briggs-Rose [12] Owen [43] Rose [41] Webb [44]

1.3 1.2 1.1

No.2 R404A Ts=30°C

1.0 0.9 0.8 0.7 ho-exp/ ho-pre

0.6 0.5 1.4 1.3 1.2

No.4 R134a Ts=30°C

No.4 R404A Ts=30°C

1.1 1.0 0.9 0.8 0.7 0.6 0.5 20

30

40

50

60

70

80

90

20

30

40

50

60

70

80

90

-2

q / kWm

Fig. 11 Comparison of experimental results with predicted correlations for 2D finned tubes

35

Table 1 Properties of R134a and R404A and their relative variation at Ts = 30 °C Fluid properties ‒3

‒1

Surface tension (10 Nm ) Latent heat (103Jkg‒1) Liquid density (kgm‒3) Vapor density (kgm‒3) Liquid specific heat (Jkg‒1K‒1) Liquid thermal conductivity (10‒3Wm‒1K‒1) Liquid viscosity (10‒ 6Pas)

R134a

R404A

Relative variation, %

7.42 173.10 1187.5 37.54 1446.5 78.99 183.13

3.97 134.11 1019.4 75.61 1590.3 61.75 119.44

46.50 22.52 14.16 ‒101.41 ‒9.94 21.83 34.78

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Table 2 Specifications of test tubes Test tube 3D fin No.1 2D fin No.2 3D fin No.3 2D fin No.4

Material

Outside tube diameter, (mm)

Inside tube diamet er (mm)

Fin height, (mm)

Fin spacing (mm)

Fin thickness (mm)

Length of test section (mm)

Material conductivity (Wm‒1 K‒1)

Iron cupronickel

18.93

16.48

‒

‒

‒

1500

28.9

Iron cupronickel

19.25

16.59

1.29

1.26

0.30

1464

28.9

Aluminum brass

18.90

15.16

‒

‒

‒

1450

104.7

Aluminum brass

19.27

16.58

1.27

1.24

0.29

1471

104.7

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Table 3 Thermal resistance ratio at water velocity of 2.0 ms‒ 1 and saturation temperature of 40 °C Tube No.1 No.2 No.3 No.4

R134a

R404A

Ro%

Rw%

Ri%

Ro%

Rw%

Ri%

37.78% 40.10% 34.56% 49.67%

23.50% 22.89% 10.36% 7.95%

38.73% 37.01% 55.08% 42.38%

40.24% 47.21% 33.06% 54.15%

22.89% 20.23% 9.56% 6.61%

36.87% 32.56% 57.38% 39.24%

38