International Journal of Heat and Mass Transfer 135 (2019) 124–130
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Experimental study of pool boiling on a novel reentrant cavity tube surface with R134a Yonghui Wang, Jili Zhang, Zhixian Ma ⇑ Faculty of Infrastructure Engineering, Dalian University of Technology, Dalian 116024, PR China
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Article history: Received 21 March 2018 Received in revised form 13 January 2019 Accepted 28 January 2019
Keywords: Reentrant cavity tube Pool boiling Heat transfer Experiment
a b s t r a c t This study experimentally investigated pool boiling heat transfer performance on horizontal copper tubes with refrigerant R134a. Experiments were conducted at saturation temperature of 5.6 °C, while heat flux varied from 5 to 65 kW/m2 with an interval of 5 kW/m2. A plain tube and a reentrant cavity tube (enhanced tube) were tested. The nominal outside diameter of test tube was 25.14 mm and the length in test section was 1000 mm. Tubes were water heated by 15% volume faction water-ethylene glycol mixture with an insert bar in the fluid passage. The experimental data of the plain tube were compared against the Cooper correlation with deviation of 3% to 10%. The boiling heat transfer coefficient of the enhanced tube increased with the increasing heat flux in the experimental range. The heat transfer enhancement factor, the ratio of boiling Nusselt number for enhanced tube to that for plain tube, varied from 1.66 to 2.64 and reached the maximum at Reo = 0.38(heat flux q = 20 kW/m2). The boiling Nusselt number was correlated with boiling Reynolds number with predicting the experimental data within ±4%. Ó 2019 Elsevier Ltd. All rights reserved.
1. Introduction Most of the flooded evaporators used in HVAC system were the shell-and-tube heat exchangers. Heat transfer coefficients (HTCs) of boiling on tube bundle were the necessary knowledge to design the heat exchangers. The HTCs of pool boiling on a single tube essential to the heat exchanger design. To reduce the bulk and improve the heat capacity of the shelland-tube heat exchanger, plain tubes are processed into enhanced tubes which show higher heat transfer coefficients than plain tubes. According to former investigations on heat transfer enhancement, the basic principle for enhancing pool boiling is to increase the number of stable nucleation sites over a long period. With the development of manufacture technique, the plain tubes were gradually replaced by the high-efficiency enhanced tubes. The reentrant cavity tube (structured tube) was commonly used commercially. Various reentrant cavity tubes were invented and studied in the open literatures [1,2]. Tube and refrigerant were the two indispensable parts of the pool boiling heat transfer process. Tubes published in the open literatures [3–6] included Turbo-B, Gewa-B, Thermoexcel-E, TurboBⅡ HP, Turbo-BⅡ LP, Turbo-EDE2, Gewa-B4, Turbo-B5, Gewa-B5 and some reentrant cavity tubes listed in literatures. With the
⇑ Corresponding author. E-mail address:
[email protected] (Z. Ma). https://doi.org/10.1016/j.ijheatmasstransfer.2019.01.128 0017-9310/Ó 2019 Elsevier Ltd. All rights reserved.
development of the refrigerant industry, many investigations have been conducted on R11, R12, R22, R123, R114, R245a, R134a, R1234ze, R1234yf, R1233ze and so on[7–11]. Despite of the property of the refrigerant, some meaningful results have been reached. As well known, the heat transfer enhancement ratio (ho,en/ho,p), defined as HTC of an enhanced tube divided by that of a plain tube at the same heat flux, is one of the most important characteristics of enhanced tubes. Numerous researchers found that the pool boiling heat transfer enhancement ratio was greater than one and decreased with increasing heat flux in the experimental range [12–16]. In the study of Jung et al.[14], in the heat flux of 10– 80 kW/m2, the heat transfer enhancement ratios of Turbo-B and Thermoexcel-E were 5.41–1.77, 8.77–1.64 respectively. Ji et al. [15] aimed at a specific reentrant cavity tube and got an enhancement ratio of 4.9–1.8 in the heat flux range of 10–75 kW/m2. Ribatski and Thome [16] conducted experiments at saturation temperatures of 5, 10 and 20 °C and heat flux from 20 to 70 kW/m2. The heat transfer enhancement ratios of the Gewa-B, and TurboBII HP tubes were 5.2–2.4 and 7.0–1.8, respectively. Reentrant cavity tubes were indeed effective boiling heat transfer enhanced tubes. However, whether the heat transfer enhancement ratio decreased with heat flux is valid for novel reentrant cavity tubes couldn’t be concluded yet. With the same refrigerant, the pool boiling heat transfer coefficients of different tubes showed different variation trends with heat flux. An investigation of Memory et al. [13] showed that the
Y. Wang et al. / International Journal of Heat and Mass Transfer 135 (2019) 124–130
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Nomenclature A Ci Co cp D do f g hi hi,Gni hlv ho K k L M Nu p pc Pr pr Q q Re Rwal Rfoul T
heat transfer area(m2) coefficient for internal heat transfer correlation (-) constant for internal heat transfer correlation (W/m2°C) specific heat capacity(J/kg°C) diameter(mm) characteristic length of bubble (m) friction factor(-) acceleration due to gravity (m/s2) internal heat transfer coefficient(W/m2°C) internal heat transfer coefficient calculated by Gnielinski equation(W/m2°C) latent heat capacity of refrigerant(J/kg) outside heat transfer coefficient(W/m2°C) overall heat transfer coefficient(W/m2°C) thermal conductivity(W/m°C) test section length (m) molar mass (g/mol) Nusselt number (-) pressure (Pa) critical pressure(Pa) Prandtl number (-) reduced pressure (p/pc) heat transfer rate (W) heat flux(W/m2) Reynolds number (-) thermal resistance of tube wall(m2°C /W) thermal resistance of fouling(m2°C /W) temperature(K or °C)
HTCs of Thermoexcel-HE tube were lower than that of Turbo-B tube with the same refrigerant R114 at the same saturation temperature and the heat flux of 3–100 kW/m2. The HTCs of TurboB5 were nearly constant and that of Gewa-B5 decreased as the heat flux increased in the experimental range[17,18]. The HTCs of Turbo-EDE2 and Gewa-B4 with R134a showed similar slowly decreasing trend with increasing heat flux in spite of the values for Gewa-B4 were greater [19]. In addition, HTC data for Turbo-B, Thermoexcel-B and test tubes in [15,20] exhibited increasing tendency. Pool boiling heat transfer on outside surfaces of numerous tubes with various refrigerants should generate diverse results that were hard to be predicted accurately. Even for pool boiling on the same tube with the same refrigerant, heat transfer characteristics gained by different investigators were sometimes different, for example Turbo-BⅡ HP with R134a. Tatara and Payvar [21] got an increasing trend with increasing heat flux at the saturation temperature of 4.4 °Cand heat flux from 8 to 40 kW/m2. Gorgy and Eckels[22] also found the similar increasing trend at the heat flux from 4 to 45 kW/m2 and HTCs decreased and eventually kept constant from 45 to 135 kW/m2. Data from Ribatski and Thome [16] exhibited a decreasing trend at the saturation temperature of 5, 10, 20 °C and the heat flux from 20 to 70 kW/m2. Besides, Robinson and Thome [23] found HTCs increased and then decreased in the heat flux range of 16 to 40 kW/m2 at 4.4 °C. The above confusing phenomenon was perhaps due to boiling hysteresis induced by the surface structure and refrigerant property. From the above literature review, it is clear that the reentrant cavity incurs a significant contribution to the pool boiling heat transfer enhancement. However, there are likewise some unsolved problems on the application of reentrant cavity tubes. For a novel reentrant cavity tube, obtaining the pool boiling HTCs accurately was still largely dependent on experiment. This study conducted experiments for obtaining the pool boiling HTCs on a novel reen-
u V
flow velocity(m/s) volumetric flow rate(m3/s)
Greek symbols r surface tension (N/m) l viscosity (Pas) q fluid density(kg/m3) DTm logarithmic mean temperature difference(K) Subscripts ave averaged b bar inserted in the tube c cold medium exp experiment data f fluid h hot medium i inside in inlet l liquid refrigerant o boiling outside out outlet p plain tube pre predicted by equations sat saturation R134a v vapor refrigerant wal wall
trant cavity tube with R134a. The heat transfer enhancement factor was determined and correlations were established for predicting the experimental data. 2. Experiment 2.1. Experimental system The experimental system showed in Fig. 1 was established for obtaining the pool boiling heat transfer coefficients of R134a on the enhanced tube. The experimental system consisted of a refrigerant loop, a hot medium circuit, a cold medium loop and a computer based control and data acquisition system. The refrigerant loop included an upper condenser and a lower evaporator. The condenser and evaporator made of steel were well sealed. The refrigerant liquid was heated to the vapor on the surface of the test tube in the evaporator and then the vapor went up into the condenser. The refrigerant vapor was cooled to the liquid in the condenser and then the liquid flowed down to the bottom of the evaporator driven by gravity. The cold medium flowed through the condenser and got a temperature rise for obtaining heat from the vapor refrigerant. Then the cold medium flowed through the evaporator of a 5 kW cooling capacity chiller unit to make the temperature of the cold medium at the tank inlet lower than that at the tank outlet. The hot medium flowed through the test tube of the evaporator and got a temperature drop for transferring heat to the liquid refrigerant. The temperature at the tank inlet was also lower than that at the tank outlet. The temperatures of cold medium and hot medium were stabilized by controlling the power of the heaters inserted in both tanks with two PID control systems. The control system and the data acquisition system were based on a computer. The turbine flowmeters (0.6–6 m3/h, 0.5%Reading) were used to measure the flow rates of hot medium and clod medium. An absolute pressure transducer (0–0.7 MPa, 0.1%Full-Scale)
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Y. Wang et al. / International Journal of Heat and Mass Transfer 135 (2019) 124–130
Fig. 1. Schematic diagram of experimental system.
was used to measure the pressure of the test section i.e. the evaporator. Several PT-100 temperature sensors (calibrated accuracy ±0.05 °C) were used to measure the temperatures of the working fluid. The accuracy of measuring the temperature difference between the inlet and outlet of the condenser and the evaporator was within ± 0.03 °C. All the flow rates, pressure and temperatures were recorded every 20 s and more than 16 groups of values were arranging for getting the average values after the condition was stable. Fig. 2 shows the test section, a shell and tube heat exchanger. The hot medium flowed through the test tube and the refrigerant was in the shell side. The inside diameter of the cylinder is 200 mm and the wall is 10 mm thick. The total length of the test section is 1060 mm and the length of the tube immersed in the refrigerant is 1000 mm. A stainless bar with 16 mm outside diameter was placed in the center of the test tube. This would help reduce measured uncertainty for obtaining a larger temperature difference between the inlet and outlet temperatures with the same flow velocity and inlet temperature of the working fluid. The test tube was sealed with the test section depending on the
structure displayed in Fig. 2. Teflon ring was tightly squeezed with tube and shell wall by the steel ring and plate screwed with the shell wall. The test section was well isolated to reduce the probable heat absorbing from the surroundings. Test tubes in this experiment included a plain tube and an enhanced tube, both of which are made of copper. The enhanced tube is illustrated in Fig. 3. The enhanced tube was a commercial tube with cold machining on the integral-fin tube. The surface of the integral-fin tube was twice knurled to form the reentrant cavity surface. Two knurls were 90° crossing and both had a 45° angle with the tube axis. The inside diameter based on the bottom of the internal helical fins is 22.48 mm. The outside diameter based on the top of the outside structure is 25.14 mm. The number of the fins is 1890 per meter and the fin height is 0.60 mm. The refrigerant was R-134a for the experiment. The refrigerant was qualified for the difference between the measured temperature in the test section and the saturation temperature calculated with absolute pressure was
