Pool boiling heat transfer on tandem tubes in vertical alignment

International Journal of Heat and Mass Transfer 87 (2015) 138–144

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Pool boiling heat transfer on tandem tubes in vertical alignment Myeong-Gie Kang Department of Mechanical Engineering Education, Andong National University, 388 Songchun-dong, Andong-city, Kyungbuk 760-749, South Korea

a r t i c l e

i n f o

Article history: Received 22 December 2014 Received in revised form 3 April 2015 Accepted 5 April 2015

Keywords: Pool boiling Tube pitch Lower tube heat flux Heat transfer Advanced nuclear reactors

a b s t r a c t The combined effects of a tube pitch and the heat flux of a lower tube on saturated pool boiling heat transfer of tandem tubes were investigated experimentally. For the test, two smooth stainless steel tubes of 19 mm diameter and the water at atmospheric pressure were used. The pitch was varied from 28.5 to 114 mm and the heat flux of the lower tube was changed from 0 to 110 kW/m2. The bundle effect was clearly observed when the heat flux of the lower tube was greater than that of the upper tube and the heat flux of the upper tube is less than 60 kW/m2. The maximum bundle effect was 3.08 as the tube pitch was 95 mm when the heat fluxes of the upper and the lower tubes were 10 and 90 kW/m2, respectively. In general, the bundle effect was decreased with increasing the tube pitch. The major heat transfer mechanisms are considered as the convective flow of bubbles, rising from the lower tube, and liquid agitation caused by the bubbles. As a way of quantifying the bundle effects a simple empirical correlation was suggested. Ó 2015 Elsevier Ltd. All rights reserved.

1. Introduction Pool boiling is closely related to the design of passive type heat exchangers, which have been investigated in nuclear power plants to meet safety functions in case of no power supply [1,2]. Since the space for a heat exchanger install is usually limited, the exact estimation of the heat transfer is very important to keep up a reactor integrity. One of the major issues in the design of a heat exchanger is the heat transfer in a tube bundle. The passive condensers adopted in SWR1000 and APR+ has U-type tubes [1,2]. The heat exchanging tubes are in vertical alignment. For the cases, the upper tube is affected by the lower tube. Therefore, the results for a single tube are not applicable for the tube bundles. Since heat transfer is closely related to the conditions of tube surface, bundle geometry, and liquid, lots of studies have been carried out for the several decades to investigate the combined effects of those factors on pool boiling heat transfer [3,4]. One of the most important parameters in the analysis of a tube array is the pitch (P) between tubes. Many researchers have been investigating its effect on heat transfer enhancement for the tube bundles [5–7] and the tandem tubes [6,8,9]. The effect of tube array on heat transfer enhancement was also studied for application to the flooded evaporators [10–13]. The upper tube within a tube bundle can significantly increase nucleate boiling heat transfer compared to the lower tubes at moderate heat fluxes. At high heat fluxes these influences disappear and the data merge onto the pool boiling curve of a single tube [13]. http://dx.doi.org/10.1016/j.ijheatmasstransfer.2015.04.015 0017-9310/Ó 2015 Elsevier Ltd. All rights reserved.

Cornwell and Schuller [14] studied the sliding bubbles by high speed photography to account the enhancement of heat transfer observed at the upper tubes of bundles. The study by Memory et al. [15] shows the effects of the enhanced surface and oil adds to the heat transfer of tube bundles. They identified that, for the structured and porous bundles, oil addition leads to a steady decrease in performance. The flow boiling of n-pentane across a horizontal tube bundle was investigated experimentally by Roser et al. [16]. They identified that convective evaporation played a significant part of the total heat transfer. The fouling of the tube bundle under pool boiling was also studied by Malayeri et al. [17]. They identified that the mechanisms of fouling on the middle and top heater substantially differ from those at the bottom heater. The heat transfer on the upper tube is enhanced compared with the single tube [10]. The enhancement of the heat transfer on the upper tube is estimated by the bundle effect (hr ). It is defined as the ratio of the heat transfer coefficient (hb ) for an upper tube in a bundle with lower tubes activated to that for the same tube activated alone in the bundle [18]. The upper tube within a tube bundle can significantly increase nucleated boiling heat transfer compared to the lower tubes at moderate heat fluxes. At high heat fluxes these influences disappear and the data merge the pool boiling curve of a single tube [13]. It was explained that the major influencing factor is the convective effects due to the fluid velocity and the rising bubbles [4]. For tandem tubes having an equal heat flux, the greatest heat transfer coefficient of the upper tube

M.-G. Kang / International Journal of Heat and Mass Transfer 87 (2015) 138–144

139

Nomenclature AT D hb hr I L P PT

data acquisition error, °C diameter of the heating tube, m boiling heat transfer coefficient, W/m2-°C bundle effect (=hb =hb;q00L ¼0 ) supplied current, A heated tube length, m pitch distance, m precision limit, °C

decreases [8], increases [9], or negligible [6] with increasing tube pitch in pool boiling. Some possible explanations for the discrepancy are due to the liquid and the difference in geometric conditions of the test section and the pool. Ribatski et al. [6] performed an experiment with R123 and smooth brass tubes of 19.05 mm outer diameter (D). Through the investigation, the effects of reduced pressure (pr ) and tube pitch were studied. They identified that the effect of tube spacing on the local heat transfer coefficient along the tube array was negligible. At low heat fluxes, the tube positioning shows a remarkable effect on the heat transfer because of the free convection. However, at high heat fluxes the effect of natural convection is negligible and the effects of bubbles become dominant. According to Ribatski et al. [6] the spacing effects on the heat transfer became relevant as the tubes come closer to each other due to bubble confinement between consecutive tubes. Gupta et al. [8] investigated the effect of tube pitch on pool boiling of two tubes placed one above the other. They used a stainless steel tube of commercial finish having 19.05 mm outer diameter and the distilled water at 1 bar. They found that the heat transfer coefficients of the upper tube were increased as the pitch distance decreased due to the larger number of bubbles intercepted by the upper tube when the pitch was lowered. The authors also investigated three horizontal tubes, which were stacked one above another at a constant pitch distance and got the similar results. Hahne et al. [10] used finned type copper tubes submerged under R11 at 1 bar. The tube pitch was varied between P=D = 1.05 and 3.0. They found that the largest heat transfer of the upper tube increases with increasing tube pitch. This is expected as the convective flow of bubbles and liquid, rising from the lower tube, enhances the heat transfer on the upper tube. However, as the heat flux increases the heat transfer of the upper tube decreases with increasing tube pitch. Since the source of the convective flow in pool boiling is the lower heated tube, the heat flux of the lower tube (q00L ) is of interest. Kumar et al. [19] carried an experimental study using the combination of distilled water and two horizontal reentrant cavity copper tubes. They used a fixed spacing and developed a model to predict the heat transfer coefficient of individual tube in a multitube row and the bundle heat transfer coefficient. Ustinov et al. [20] investigated effects of the heat flux of the lower tube on pool boiling of the upper tube for the fixed tube pitch. They used microstructure-R134a or FC-3184 combinations and identified that the increase in the heat flux of the lower tube decreased the superheating (DT sat ) of the upper tube. Summarizing the previous results it can be stated that heat transfer coefficients are highly dependent on the tube geometry and the heat flux of the lower tube. As shown in Table 1 the aim of the published results for tandem tubes were to investigate the effect of pitch or lower tube heat flux individually. In general, there are many tube arrays where the analysis of the combined effects is necessary. Therefore, the focus of the present study is an

q00L q00T Ra T sat TW V DT sat

heat flux on the lower tube, W/m2 heat flux on the upper tube, W/m2 surface roughness, lm saturation temperature, °C tube wall temperature, °C supplied voltage, V tube wall superheat (¼ T W  T sat ), °C

identification of the effects of tube pitch as well as the heat flux of the lower tube on the heat transfer.

2. Experiments A schematic view of the present experimental apparatus is shown in Fig. 1. The water tank (Fig. 1(a)) is made of stainless steel and has a rectangular cross section (950  1300 mm) and a height of 1400 mm. The sizes of the inner tank are 800  1000  1100 mm (depth  width  height). Four auxiliary heaters (5 kW/ heater) are installed in the space between the inside and the outside tank bottoms. The heat exchanging tubes are resistance heaters (Fig. 1(b)) made of very smooth stainless steel tubes of 19 mm outside diameter and 400 mm heated length (L). The surface of the tube is finished through a buffing process to have a smooth surface. The value of the surface roughness was measured by a stylus type profiler. The arithmetic mean of all deviations from the center line over the sampling path has the value of Ra = 0.15 lm. Electric power of 220 V AC was supplied through the bottom side of the tube. For the tests, the assembled test section was in the water tank as shown in Fig. 1(a). The pitch was regulated from 28.5 to 114 mm by adjusting the space between the tubes, which were positioned one above the other and were assembled using bolts and nuts to the supporter. The values of the tube pitches and the heat fluxes of the lower tube are listed in Table 2. The heat flux of the lower tube was (1) set fixed values of 0, 30, 60, and 90 kW/m2 or (2) varied equal to the heat flux of the upper tube (q00T ). The water tank was filled with water until the initial water level reached 1.1 m; the water was then heated using four pre-heaters at constant power. When the water temperature was reached the saturation value (100 °C since all tests were done at atmospheric pressure), the water was then boiled for 30 min to remove the dissolved air. The temperatures of the tube surfaces (T W ) were measured when they were at steady state while controlling the heat flux on the tube surface with the input power. The tube outside was instrumented with six T-type sheathed thermocouples (diameter is 1.5 mm). The thermocouple tip (about 10 mm) was brazed on the sides of the tube wall. The brazing metal is a kind of brass and the averaged brazing thickness is less than 0.1 mm. The temperature decrease along the brazing metal was calibrated by the one dimensional conduction equation. Since the thermal conductivity of the brass is nearby 130 W/m-°C at 110 °C [21], the maximum temperature decrease through the brazing metal is 0.08 °C at 110 kW/m2. The value was calculated by the product of the heat transfer rate and the thermal resistance. The measured temperatures were calibrated considering the above error. The water temperatures were measured with six sheathed Ttype thermocouples attached to a stainless steel tube that placed vertically in a corner of the inside tank. To measure and/or control the supplied voltage and current, two power supply systems was used.

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Table 1 Summary of Published Results. Author

Liquid

Ribatski et al. [6] Gupta et al. [8] Hahne et al. [10] Ustinov et al. [20] *

Tube

P/D

Type

Pressure

Surface

Material

Diameter

R123 Distilled water R11 R134a FC-3284

0.023* 0.063* 1 bar 1 bar 1,5–9 bar

Smooth Smooth Finned type (19 fpi, 26 fpi) Microstructure

Brass Stainless steel Copper Copper

19.05 mm 19.05 mm 18.74 mm 18.89 mm 18 mm

1.32, 1.53, 2.0 1.5, 3.0, 4.5, 6.0 1.05, 1.3, 3.0 1.5

Reduced pressure.

Unit: mm Outer tank Inner tank

95

1300

0

Water supply line Flow holes

100

100

100

1400

Power supply T/C lines

Heaters

Water drain

Reinforced glass

Test section and supporter

Filter

Fig. 1. Schematic diagram of experimental apparatus.

Table 2 Test Matrix. P; mm 28.5 38 47.5 57 76 95 114

q00L ; kW/m2

P=D 1.5 2 2.5 3 4 5 6

0,30,60,90, 0,30,60,90, 0,30,60,90, 0,30,60,90, 0,30,60,90, 0,30,60,90, 0,30,60,90,

q00T ; kW/m2 q00T q00T q00T q00T q00T q00T q00T

10–110 10–110 10–110 10–110 10–110 10–110 10–110

The heat flux from the electrically heated tube surface is calculated from the measured values of the input power as follows:

q00T ¼

VI

pDL

¼ hb DT sat ¼ hb ðT W  T sat Þ

ð1Þ

where V and I are the supplied voltage and current, and D and L are the outside diameter and the length of the heated tube, respectively. T W and T sat represent the measured temperatures of the tube surface and the saturated water, respectively. Every temperature used in Eq. (1) is the arithmetic average value of the temperatures measured by the thermocouples. The uncertainties of the experimental data were calculated from the law of error propagation [22]. The data acquisition error (AT , ±0.05 °C) and the precision limit (P T , ±0.1 °C) were counted for the uncertainty analysis of the temperature. The 95 percent confidence uncertainty of the measured temperature was calcu1=2

lated from ðA2T þ P2T Þ and had the value of ±0.11 °C. The error bound of the voltage and current meters used for the test was

±0.5% of the measured value. Therefore, the uncertainty of the calculated power (voltage  current) was obtained as ±0.7%. Since the heat flux had the same error bound as the power, the uncertainty in the heat flux was estimated to be ±0.7%. When evaluating the uncertainty of the heat flux, the error of the heat transfer area was not counted since the uncertainty of the tube diameter and the length was ±0.1 mm and its effect on the area was negligible. To determine the uncertainty of the heat transfer coefficient, the uncertainty propagation equation was applied on Eq. (1). Since the values of the heat transfer coefficient were the results of the calculation of q00 =DT sat , a statistical analysis on the results was performed. After calculating and taking the mean of the uncertainties of the propagation errors, the uncertainty of the heat transfer coefficient was determined to be ±6%. 3. Results and discussion To identify the validity of the present investigation, the experimental data for q00L ¼ q00T were compared with the published results as shown in Fig. 2. All results show that the value of the bundle effect is the highest at the lower heat fluxes and merges hr = 1 as the heat flux increases. Although the conditions of the surface and the liquid adopted in Ribatski et al. [6] and Hahne et al. [10] are different from the present investigation, the general tendency and the value of hr is very similar to each other. There exists some discrepancy between the present data and the results of Gupta et al. [8]. The combination of tube surface and liquid used by Gupta et al. [8] is almost same to the present study. One of the causes is the difference in the water tank size. The tank tested by Gupta et al. has the size of 200  300  425 mm. If the tank size

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120

2.0 Present, P/D=1.5 Present, P/D=2.0 Ribatski, pr=0.023, P/D=1.53 Ribatski, pr=0.023, P/D=2.0 Ribatski, pr=0.063, P/D=1.53 Ribatski, pr=0.063, P/D=2.0

1.8

2

1.4

100

qT'', kW/m

hr

1.6

1.2 1.0 0.8

80

60

40

0

10

20

30

40

30 kW/m

50

2

qT'' , kW/m

2

20

(a) Ribatski et al. [6] 0

2.4 P/D 1.5 3.0 6.0

2.2 2.0

hr

1.8

Present

Gupta

1

2

3

4

5

6

7

8

9

o

ΔTsat , C Fig. 3. Plots of q00T versus DT sat for P=D = 6.

1.6 1.4 1.2 1.0 0.8

0

10

20

30

40

50

2

qT'' , kW/m (b) Gupta el al. [8] 2.0 Present, P/D=1.5 Present, P/D=3.0 Hahne, 19fpi, P/D=1.3 Hahne, 19fpi, P/D=3.0 Hahne, 26fpi, P/D=1.3 Hahne, 26fpi, P/D=3.0

1.8 1.6

hr

6.2

4.3

0

1.4 1.2 1.0 0.8

0

10

20

30

40

50

60

2

qT'' , kW/m

(c) Hahne et al. [10] Fig. 2. Comparison of present data with published results.

were small the circulating flow in the inside tank became increasing. This results in additional convective effect on the surface and the bundle effects can be enhanced at low heat fluxes. Another possible reason is due to the surface condition. The tube surface of the present study was manufactured through the buffing process while Gupta et al. used a commercial finish. However, two results show a similar tendency as the heat flux increases. Fig. 3 shows plots of q00T versus DT sat data obtained on the upper tube surface. The heat flux on the lower tube was kept constant while the upper tube experienced the whole range of heat flux variation at P=D = 6. As shown in the figure the activation of the lower tube enhances the heat transfer on the upper tube compared with the results for the single tube (i.e., q00L = 0 kW/m2). The heat transfer on the upper tube is increased as the heat flux of the lower tube is increased. The change of q00L from 90 to 0 kW/m2 results in 44.2% (from 4.3 to 6.2 °C) increase of DT sat when q00T = 30 kW/m2. The gradual increase in q00L results in the decrease in DT sat for the

given q00T . Throughout the heat fluxes tested the enhancement in heat transfer is much clearly observed at low or moderate heat fluxes. When q00T > 80 kW/m2 the curve for q00L – 0 kW/m2 converges to the curve of the single tube. The curve for q00L ¼ q00T is very unique compared to the other results and shows a kind of transition from enhanced to deteriorated heat transfer as the heat flux decreases. When the heat fluxes are lower than 30 kW/m2, the tube wall superheat is higher than the curve for q00L = 30 kW/m2. As the heat flux increases, the curve of q00T versus DT sat shifts left side and the heat transfer is enhanced. When q00T > 80 kW/m2 the curve converges to the curve for q00L = 90 kW/m2. The bundle effect is expected as the convective inflow of bubbles and liquid, rising from the lower tube, enhances the heat transfer on the upper tube [10]. The intensity of the convective flow is increased as q00L increases. The heat transfer on the upper tube is associated with (1) the bulk movement of bubbles and liquid coming from the lower side and (2) the bubble nucleation and growth on the tube surface [8]. The possible mechanisms affecting on heat transfer are convective flow, liquid agitation, and the nucleation site density. The increase in the heat flux also increases the nucleation sites which enhancing heat transfer because of the latent heat required for the phase change. The convective flow generated by the bulk movement increases heat transfer and is important for the heat transfer analysis at low heat fluxes. At low heat flux regions most of the bulk flow is liquid and the major mechanism is related with the single phase natural convection. The liquid agitation also enhances heat transfer. The intensity of the liquid agitation depends on the amount of bubbles and the active movement of the bubbles. The stronger intensity accelerates the access of the liquid to the heated surface and departure of the bubbles from the surface. This enhances the heat transfer rate. When the upper tube is at low heat flux a convection-controlled regime prevails. Therefore, the turbulent flow generated by the departed bubbles from the lower tube enhances heat transfer much. However, as the heat flux of the upper tube increases, the portion of the liquid convection gets decreased and, accordingly, the enhancement in heat transfer gets decreased. When q00L ¼ q00T , the heat flux of the lower tube is increasing as the heat flux of the upper tube increases. When q00L is less than 30 kW/m2, the convective flow generated by the lower side tube is weak. Therefore, the enhancement of the heat transfer is not observed around these heat fluxes. As the heat flux increases the intensity of the liquid agitation and the convective flow get increased and, as a result, enhanced heat transfer is observed at higher heat fluxes.

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(a) qL′′ =60 kW/m2

(b) qL′′ =90 kW/m2

(c) qL′′ =90 kW/m2

qT′′ =50 kW/m2

qT′′=50 kW/m2

qT′′=100 kW/m2

Fig. 4. Photos of pool boiling on tandem tubes of P=D=2.

3.5

3.5 qT''=10kW/m

3.0

qT''=30kW/m

qT''=90kW/m 2.5

2.0

2.0

1.5

1.5

1.0

1.0

1

2

qT''=30kW/m

2

2.5

0.5

qT''=10kW/m

3.0

2

hr

hr

qT''=90kW/m

2

3

4

5

0.5

6

1

2

2

4

5

6

5

6

P/D

(a) qL''=30kW/m

2

(b) qL''=60kW/m

3.5

2

3.5 2

qT''=10kW/m

2

2

3.0

qT''=10kW/m

3.0

qT''=30kW/m

2

qT''=30kW/m

2

qT''=90kW/m

2

qT''=90kW/m

2.5

2.5

2.0

2.0

hr

hr

2

3

P/D

1.5

1.5

1.0

1.0

0.5

2

1

2

3

4

5

6

0.5

1

P/D

2

3

4

P/D

(c) qL''=90kW/m

2

Fig. 5. Variation of bundle effect with P=D.

(d) qL''=qT''

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M.-G. Kang / International Journal of Heat and Mass Transfer 87 (2015) 138–144

Fig. 4 shows some photos of pool boiling on the tandem tubes of P=D = 2. Those photos were taken around the upper part of the tube. As shown in the photos relatively bigger bubbles are observed on the surface of the upper tube. When the heat fluxes are low (see Fig. 4(a)) the effects of convective flow and the liquid agitation are dominant and relatively small bubbles are observed at a small area of the surface. However, the size of the bubbles becomes increasing through coalescing with the nearby bubbles as the heat flux increases as shown in Fig. 4(c). In the case, the effect of bubbles on the heat transfer becomes predominant. The variation of bundle effect against P=D is shown in Fig. 5. The heat transfer on the upper tube of the twin tubes is enhanced compared with the single tube. In general, the increase of P=D decreases hr . The bundle effect is maximum at P=D = 1.5 and gets decreased and merges to hr = 1 as P=D increases when q00T P 30 kW/m2. This is because the effective area of the convective flow gets broadened and, then, the intensity becomes weaker as P=D

hr

4

3

qL''=30kW/m

2

qL''=90kW/m qL''=qT''

qL''=60kW/m

2 2 2

1

0

0

20

40

60

80

100

120

2

qT'' , kW/m

(a) P/D=2 4 qL''=30kW/m

hr

3

qL''=60kW/m qL''=90kW/m qL''=qT''

2

2 2 2

1

0

0

20

40

60

80

100

increases. However, the tendency is changing when q00T = 10 kW/ m2, where the effects of convective flow and liquid agitation are dominant. The value of hr gets decreased, and then, increased, and, finally, decreased as P=D increases. This tendency is clearly observed when q00L is high. The higher heat flux of the lower tube generates stronger convective flow and liquid agitation which enhancing heat transfer. The maximum hr is 3.08 at P=D = 5 when q00L = 90 kW/m2 and q00T = 10 kW/m2. As 26 P=D 65, the increase in P=D increases hr . The increase of P=D from 2 to 5 results in 44.6% increase of hr (from 2.13 to 3.08). One of the possible explanations for the tendency is the development of turbulence. The mixture of the liquid and bubbles coming from the lower tube needs some distance to generate enough turbulent effect. Another related cause is the static pressure of the liquid. The size of the departed bubbles gets increased while moving up due to the decrease of the static pressure [4]. The big size bubbles generate additional liquid agitation which enhances the heat transfer. The gradual increase in P=D more than 5 results in hr decrease. Through the tube pitch, two competing heat transfer mechanisms can be considered. One of them is the dispersion of the convective flow and the other one is the effects of turbulence and static pressure. The former mechanism, which deteriorates heat transfer, gets dominant at P=D > 5. Fig. 6 shows variations in the bundle effect against the heat flux on the upper tube. Throughout the heat fluxes tested, the increases in q00L increases the bundle effect regardless of the tube pitch. As q00L increases, the bundle effect increases significantly. The value of hr increases 63.6% (from 1.76 to 2.88) as q00L increases from 30 to 90 kW/m2. The bundle effect is clearly observed when q00L > q00T and q00T is less than 60 kW/m2. However, the bundle effect converges to unity at higher heat fluxes regardless of the value of q00L . The bundle effect has the minimum value when q00L ¼ q00T . The bundle effect is expected as the convective inflow of bubbles and liquid, rising from the lower tube, enhances the heat transfer on the upper tube [10]. When the heat flux of the upper tube is low, the major heat transfer mechanism is convective flow. Therefore, the turbulent flow generated by the departed bubbles from the lower tube enhances heat transfer much. However, as the heat flux of the upper tube increases the effect of convective flow decreases and the enhancement in heat transfer decreases. A total of 308 data points has been obtained for the heat flux versus superheat for various tube pitches and the heat flux of the lower tube. Although it is not realistic to obtain any general theoretical correlation for heat transfer coefficients in nucleate boiling since

120

2

qT'' , kW/m

5

(b) P/D=4

4 +10%

4

3

qL''=30kW/m

2

qL''=90kW/m qL''=qT''

qL''=60kW/m

2 2

-10%

2

hr,cal

hr

3

2

1 1

0

0

20

40

60

80

100

2

qT'' , kW/m

(c) P/D=6 Fig. 6. Variation of bundle effect with heat flux.

120 0.7 0.7

1

2

3

4

5

hr,exp Fig. 7. Comparison of experimental data with calculated values.

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M.-G. Kang / International Journal of Heat and Mass Transfer 87 (2015) 138–144

it contains inherent unidentified uncertain parameters, we continue the development of the correlation nevertheless. This is because the quantification of the experimental results could broaden its applicability to the thermal designs. Through the statistical analysis of the experimental data with the help of a computer program (which uses the least square method as a regression technique) a simple equation was determined as follows:

hr ¼

0:847 þ

0:397 ðP=DÞ2

!

001:279lnð2:424þP=DÞ=q00T

qL

ð2Þ

In the above equation, the dimension of q00T and q00L is kW/m2. The applicable ranges of the correlation are P=D = 1.56, q00L = 3090 kW/m2, and q00T = 10110 kW/m2. To confirm the validity of the correlation the statistical analyses on the ratios of the calculated and the measured heat transfer coefficients (i.e., hr;cor =hr;exp ) have been performed. The mean and the standard deviation are 1.003 and 0.076, respectively. A comparison between the bundle effect from the tests (hr;exp ) and the calculated value (hr;cal ) by Eq. (2) is shown in Fig. 7. The newly developed correlation predicts the present experimental data within ±10%, with some exceptions. The scatter of the present data is of similar size to that found in other existing pool boiling data. 4. Conclusions An experimental study was performed to investigate the combined effects of a tube pitch and the heat flux of the lower tube on pool boiling heat transfer of tandem tubes. For the test two smooth stainless steel tubes of 19 mm outside diameter and the water at atmospheric pressure were used. The major conclusions of the present study are as follows: (1) The increase of P=D decreases hr at q00T P 30 kW/m2. However, the tendency is changing when q00T = 10 kW/m2. As 2 6 P=D 6 5, the increase in P=D results in hr increase. The increase of P=D from 2 to 5 results in 44.6% hr enhancement. (2) Throughout the heat fluxes tested, the increases in q00L increases the bundle effect. The bundle effect is clearly observed when q00L > q00T and q00T is smaller than 60 kW/m2. (3) An experimental correlation was suggested to predict the heat transfer coefficients on the upper tube surface of the tandem tubes. The correlation can predict the experimental data within ±10% error bound. Conflict of interest None declared.

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