Stratified Flow Condensation of CO 2 in a Tube at Low Temperatures Pei-hua Li 1, a, Joe Deans 2,b and Stuart Norris 3,c

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1 Applied Mechanics and Materials Submitted: ISSN: , Vols , pp Accepted: doi: / Online: Trans Tech Publications, Switzerland Stratified Flow Condensation of CO 2 in a Tube at Low Temperatures Pei-hua Li 1, a, Joe Deans 2,b and Stuart Norris 3,c 1,2,3 Department of Mechanical Engineering University of Auckland, Auckland 1142, New Zealand a lpei263@aucklanduni.ac.nz, b j.deans@auckland.ac.nz, c s.norris@auckland.ac.nz Keywords: Carbon dioxide, stratified flow, condensation Abstract. This study presents an experimental investigation of CO 2 flowing condensation at the saturation temperature of -10 C, mass flux in the range from 40 to 60kgm -2 s -1 and vapour quality ranging from 0.2 to 0.8, in a 6.52mm inside diameter horizontal tube. Previous research on refrigerant condensation has shown that under these conditions, CO 2 two phases are expected to develop as a stratified flow. The significant feature of the stratified flow heat transfer is vapour film condensation in the upper region which dominates the overall heat transfer process. Test series in this study confirm that the saturation-to-tube wall temperature difference has a significant influence on the condensing heat transfer coefficient when the temperature difference is within 3K. Comparisons between the experimental results and the predictions by the Dobson, Cavallini and Thome models show that CO 2 stratified flow condensation heat transfer coefficients are overpredicted by these models with mean deviations of 104%, 81% and 127%, respectively. Introduction It is generally believed that during condensation the flow distribution of a two-phase mixture in a tube influences the local heat transfer coefficient. If the two phases are completely separated by a horizontal interface to form a stratified two-phase mixture, this flow pattern is recognized as the stratified flow regime. In the absence of any high vapour shear stress, the condensate is formed in the upper portion of the tube and drains to the interface by gravitational force; meanwhile liquid accumulated beneath the interface flows along the axial direction. The overall heat transfer coefficients can be determined from the falling film condensation on the upper part of the tube and the convective heat transfer component in the lower part. Heat transfer through the falling condensate film in the upper tube surface accounts for majority of the heat transfer to the tube wall by the fluids. The film condensation model proposed by Nusselt[1] shows the film thickness to be a function of the tube wall-to-saturation temperature difference. The Nusselt model also demonstrates that both the thermal resistance and the local heat transfer coefficient change with the varying wallto-saturation temperature difference. From this it is possible that the overall heat transfer coefficient of stratified flow would change with different wall-to-saturation temperature difference. A specially designed test rig is used to measure carbon dioxide heat transfer data to verify this assertion. CO 2 is considered to be a promising secondary refrigerant in a cascade system, but limited information has been published on its condensation phase change at low temperatures. Frequently cited correlations for predicting condensation rates are those developed by Dobson[2], Cavallini[3] and Thome[4]. However, the accuracy of these correlations for the prediction of the condensation of CO 2 requires examination. The fundamental reason for this need is that CO 2 has unique heat transfer characteristics compared with common refrigerants. For example, under a saturation temperature of -15ºC, the specific heat of CO 2, is 247%, 186%, 179%, 195%, 133%, and 153% to that of R12, R125, R134a, R22, R32 and R410a, respectively. The thermal conductivity of CO 2 at - 15ºC is 152%, 166%, 120%, 124%, 81%, and 112% to that of R12, R125, R134a, R22, R32 and R410a, respectively. The higher specific heat and thermal conductivity make CO 2 a favorable refrigerant which can perform well under smaller temperature differences inside condensers. The experimental results are obtained when stratified flow conditions are assumed to exist in the test section. This assumption is based on the flow regime transition values used within the All rights reserved. No part of contents of this paper may be reproduced or transmitted in any form or by any means without the written permission of Trans Tech Publications, (ID: , Pennsylvania State University, University Park, USA-06/03/16,22:50:10)

2 Applied Mechanics and Materials Vols aforementioned correlations. The results obtained from the tests are compared with the heat transfer coefficients predicted using these correlations in this paper. Literature Review The overall heat transfer from a stratified condensing flow has two contributions, film condensation with the horizontal vapour shear stress being exerted on condensate in the upper portion of the tube and the convective heat transfer in the lower part of the tube. These heat transfer mechanisms are schematically illustrated in Fig. 1. At cross section A, the overall velocity of condensate on the inner tube surface has been divided to two contributions. Velocity components are expressed in cylindrical coordinate system; u x represents the horizontal velocity components caused by the vapour shear stress along the x direction while u θ relates to the velocity which the condensate film flows down the concave tube surface. The latter velocity is referred to as the peripheral velocity. Therefore, the flowing direction of the condensate under stratified regime is not totally perpendicular to the horizontal direction but has an axial angle. The actual condensate flowing direction is shown at position B. Figure 1. Schematic drawing for the stratified flowing condensation For most stratified flows, the heat transfer rate for the falling film greatly out-weights that of convection, and earlier researchers believed that the bottom heat transfer component could be neglected.[5] However, this proposal is doubtful if there is a high mass flow rate with low vapour quality condition. One common feature of the stratified heat transfer correlations[2-4] is their identifications of the two contributions, falling film condensation and convection heat transfer at the lower part of the tube, combined by local void fraction proportions. As the main heat transfer mechanism under stratified flow patterns is the falling film condensation, the total heat transfer coefficient obtained in falling film condensation is a function of the saturation-to-tube wall temperature difference. This point was also shown in Nusselt s the classical analysis. Mathematical Model of the Stratified Flow. In terms of fluid mechanics, the formation of condensation stratified flow is caused by the interaction of the condensate gravity force and the horizontal vapour shear stress. As shown in Fig. 2, this mechanism implies that the local mass flow rate of the condensate also has two components. These components are the condensate liquid entering into the control volume from both the axial direction and the azimuthal direction. The analysis of this flow structure can be simplified by making the flowing assumptions, 1. The condensate inside flowing liquid film has constant properties; 2. The condensate-film surface is smooth and condensate film flows in the axial and azimuthal directions as shown in Fig. 2;

3 186 Manufacturing Science and Technology VI 3. The inner tube wall temperature and the liquid surface temperature are constant and equal to T w, T s ; 4. Temperature profile inside liquid film is linear. Figure 2. Mass and energy analysis on a control volume inside the liquid condensate Most of these assumptions can be readily integrated into the model except the difficulties related to the interfacial waves and the effects of the vapour drag between the liquid and vapour on the falling film. Sparrow [6] presented a falling film condensation analysis that included both the inertia terms and convection terms in the momentum and energy equations by utilizing a multiplier along with original Nusselt integral analysis. Koh [7] investigated the effects of vapour drag and developed an extension of the similar solution of Sparrow [6]. This solution kept all parameters of original Nusselt s correlation except the latent heat term is enhanced by use of a multiplier. =.. (1) h =h (2) The overall local condensate mass flow rate can be expressed, = +. (3) and are the increased values to the azimuthal and axial mass flow rates. Expanding these two mass flow rate contributions, we can get the overall expression of local mass flow rate: = +. (4) From conservation of energy, the local heat flux that can be expressed by convection heat transfer equal to the increased latent heat of the mass flow rate within the control volume: =h =h. (5) h local is the local heat transfer coefficient, T s and T w are the saturation temperature and tube wall temperature, respectively, and h fg is the latent heat for saturated vapour. Substituting the locally increased mass flow rate into the above equation, the local heat transfer coefficient can be expressed as: h = +. (6)

4 Applied Mechanics and Materials Vols The peripheral velocity increases with larger values of θ and the axial velocity of the condensate film increases along the flow direction, while the condensate film thickness increases along both of these directions. If the assumptions were changed, the revised latent heat h can replace h to show the inertia force and interfacial vapour drag s effects. It can be shown that, there exists an inverse relation between the local heat transfer coefficients and the saturation-to-tube wall temperature differences. The primary objective of this paper is the confirmation of this point. Flow Pattern Map. During the design of a horizontal tube condenser, it is necessary to predict the internal flow distribution and then to calculate the overall heat transfer coefficient. The flow pattern distribution depends on the inlet mass velocity, vapour quality, tube inner diameter and refrigerant properties. The predictions of the flow distribution from earlier studies and representative maps used to predict two-phase condensation flow in horizontal tubes are those by the Dobson, Cavallini and Thome models. These researchers summarized previous accepted results and also modified the flow transition values based on their own experimental results. Dobson and Chato[2] correlated the results from their studies that examined the flowing condensation regimes of refrigerants. The fluids examined in their correlations included R-12, R-22, R-134a, and near-azeotropic blends of R-32/R-125 using both 50 percent/50 percent and 60 percent/40 percent compositions. Modified Froude number (Fr so ) was used as the key parameter in their transition criteria. They concluded that if the mass flux was greater than 500kgm -2 s -1 or the Froude number was greater than 20, the refrigerant flows can be expected to develop an annular flow regime. For a mass flux less than 500kgm -2 s -1 and a Froude number less than 20, a wavy flow correlation can be used to calculate the heat transfer process. The Froude number of 20 is therefore used as a transition value from wavy to annular flow. The modified Froude number is given by = for for >1250. (7) Cavallini et al.[3] compared several of the most frequently used heat transfer models and their associated transition criteria for various flow regimes. In their analysis they used a dimensionless vapour mass velocity to differentiate the transition criteria between slug, stratified and annular flows. They concluded that if > 2.5, the flow patterns could be regarded as annular flow. If <2.5 and 1.6, the flow patterns are in stratified and stratified-annular transition range and if <2.5 and > 1.6, the flow regimes will be slug and slug-stratified flows. El Hajal et al.[8]developed flow maps for flowing condensation in a horizontal tube from the flowing boiling map originally proposed by Kattan et al.[9]. Instead of using the void fraction model of Zivi [10], they defined a logarithmic mean void fraction method, that combined the homogenous void fraction and the Steiner [11] horizontal tube void fraction to predict the void fractions under a wide range of pressures. Finally, the transition values of Kattan et al. [9]for evaporation were modified to calculate the different transition regimes for condensation. The flow pattern transition values developed by the Dobson, Cavallini and Thome models as shown in Fig. 3 are converted into the mass flux-vapour quality coordinate system. The CO 2 experiemnt in this was at the mass flux values of 40kgm -2 s -1, 50kgm -2 s -1 and 60kgm -2 s -1 in a 6.52mm internal diameter horizontal tube under a saturation temperature of -10 C, and these are superimposed on each of the flow pattern maps. All of these sets of experimental data are shown to be within the stratified or stratified-wavy flow regime. Under this regime it is expected that the heat transfer coefficients will vary with the differences between the tube wall and the saturation temperatures if the CO 2 working conditions are maintained constant.

5 188 Manufacturing Science and Technology VI a b c Figure 3. Predicted CO 2 condensation flow patterns based on a(dobson s), b(cavallini s) and c(thome s) criterion Test Rig and Procedures A schematic of open circuit CO 2 condensation test rig is shown in Fig. 4.[12] CO 2 is stored in highpressure gas bottles at room temperature and the pressure drives the fluid through the rig. The CO 2 pipes leaving the bottles are immersed in a sub-cooling bath that contains ethyl alcohol and dry ice and is maintained at a temperature of -30 o C. The sub-cooled high-pressure CO 2 then flows through a mass flow meter, a throttling valve and an additional sub-cooler that ensures that liquid CO 2 leaves this tank with a measured level of sub-cooling. A preheater with an accurately measured energy input is then used to heat the liquid CO 2 to the desired quality. The entry pipe before the test section forms a 0.8 meter long adiabatic calming section. When the CO 2 leaves the test section, it is routed through another sub-cooled bath containing a glycol-water mixture. The coiled pipe-work in this bath effectively works as an accumulator of CO 2. The temperature of the bath is maintained equal to the system temperature by the addition of dry ice to avoid the possible pressure fluctuations caused by the regulators. The pressure regulator of this open loop test rig is composed of two separate backpressure regulators and these are housed in a hot water bath. The hot water bath delays the formation of dry ice blockage within the regulators. The operating range for the combined regulators is between 0 and 68.9bar and the maximum operating pressure for the test is 45bar. After leaving the backpressure regulators, CO 2 is directly discharged to the environment. The glycol-water circuit was designed to investigate the influence that temperature differences have on the local heat transfer coefficients. In the initial phase of each test the glycol-water mixture is cooled by the refrigeration system to the required temperature. The three-way valve was used to isolate the refrigeration loop, and the mass flow rate of the glycolwater mixture is routed though the test section. During the operation, the glycol-water temperature slowly increases in the operation window when the heat transfer data was recorded. CO 2 flowing through the open loop was discharged to the environment while the system pressure was stabilized and any residual air was purged from the system. When the operation conditions became steady, dry ice was added into the sub-cooler bath to maintain the sub-cooled condition of the CO 2 before it flowed into preheater. The mass flow rate and system pressure were then changed to the required working conditions by adjusting the throttle valve and backpressure regulator. The preheater was switched on and adjusted to provide the required vapour quality at entry to the test section. The refrigeration system was started at the beginning of each test and the temperature of the

6 Applied Mechanics and Materials Vols glycol-water mixture was set to the required values. The brine pump was then started and the glycol-water mixture was directed to the test section. CO 2 pressure and mass flow rate were then manually stabilized while the glycol-water temperature gradually increased. In a typical test the wall-to-refrigerant temperature differences fall from 5K to 2K in a 15-minute period. Data Reduction Figure 4. Schematic diagram of the test rig The test section used in this study is a smooth double pipe heat exchanger with a low-temperature glycol water mixture flowing in the annulus and the two-phase CO 2 mixture flowing through the inner tube. The local heat flux of the CO 2 partial condensation across the test section is determined by using the rate of heat gained by the glycol water mixture flowing over the external surface area of the inner tube, = =,, =, =,,. (9) The glycol water side convection thermal resistance, the tube conduction thermal resistance and CO 2 condensation thermal resistance are expressed by, 2 =,, = ln,,, =,. (10) The initial test rig did not allow for the measurement of the tube wall temperature. The quasilocal heat transfer coefficient was calculated indirectly by using the overall thermal resistances and the average water-glycol-water mixture temperature,. In Eq. 10, h is the glycol water convection heat transfer coefficient determined by Shah and London [13], while h is the experimental heat transfer coefficients of this study. The inner tube wall temperature, which affects the local heat transfer coefficient of stratified flow is determined by the above Eq. 9. = h +, =. (11) To get the desired vapour quality at the inlet of test section, the energy input for the preheater can be determined from Eq. 11, the voltage and current to the preheater were accurately controlled by use of a variac from an AC power supply.

7 190 Manufacturing Science and Technology VI Experimental Results Heat transfer coefficient T(K) = T(K) = T(K) = x=0.2 x=0.6 x=0.4 x=0.8 Figure5. Heat transfer coefficient of Carbon Dioxide varying with saturation-to-wall temperature difference ( T) The test results presented in Fig. 5 show that the heat transfer coefficients found at different mass fluxes when the tube wall surface temperature difference was reduced from 5K to 2K. These results were obtained when the quality of the condensing carbon dioxide was maintained at four representative values. The main identified trends shown by these results are that the internal flowing condensation heat transfer coefficient increases with the decreasing saturation-to-tube wall temperature difference. The experimental heat transfer coefficients given in the Fig. 5 are generally constant with an approximate value of 1500Wm -2 K -1 when the temperature differences are greater than 3K. When the temperature difference is lower than 3K, the experimental results confirm the assertion that the heat transfer coefficient is inversely proportion to the temperature difference. Consequently, when carbon dioxide is flowing in a gravity-controlled condensation regime, the heat transfer coefficients start to develop to higher values when the temperature difference is decreasing to 3K. A comparison of the test data presented in Fig. 5 shows that the heat transfer coefficient increases with higher mass flux. The results are compared with predictions of stratified heat transfer model of Dobson[2], Cavallni[3] and Thome[4]. The average and absolute deviations of these comparisons are listed in Table 1. These three models over-predict the experimental heat transfer coefficients by 103.7%, 80.77% and 126.8%, respectively. Developed for high pressure refrigerants, Cavallini s model has the closer predictions than Dobson s and Thome s. The over-predictions of Dobson s model for high saturation pressure refrigerants were confirmed by Nellis and Klein [14]. A possible reason for the difference is that these correlations were developed from refrigerants other than CO 2. When these models are applied to predict CO 2 condensation, the unique thermo-physical properties, such as high specific heat and thermal conductivity of CO 2 lead to prediction error. In Fig. 5, under the mass flux of 40kgm -2 s -1 and 50kgm -2 s -1, there are clear distinctions between the heat transfer coefficients found at low vapour qualities (x=0.2) and those found at higher vapour qualities. That is mainly caused by the low mass flow rate used in these tests and the probability that the carbon dioxide

8 Applied Mechanics and Materials Vols became fully condensed within the test section. Consequently, the heat transfer under these conditions was similar to the values obtained for single-phase forced convection. Table 1. Comparison with heat transfer models Heat transfer model Average deviation * Absolute mean deviation * Dobson[2] 104% 104% Cavallini[3] 81% 81% Thome[4] 127% 127% *Average deviation: 100% Absolute mean deviation: 100% Conclusion CO 2 condensation experimental tests are conducted under saturation temperature of -10 C with mass flux of 40 to 60kgm -2 s -1. Under these conditions, flow maps of Dobson, Cavallini and Thome s models predict that CO 2 two-phase mixtures develop as stratified flows, where the film condensation heat transfer dominates the overall heat transfer process. Experimental stratified flow heat transfer coefficients increase with decreasing saturation-to-tube wall temperature differences significantly when the temperature differences are controlled within 3K. The heat transfer coefficients are over-predicted by models of Dobson[2], Cavallni[3] and Thome[4], with the absolute deviations of 104%, 81% and 127%, respectively. Nomenclature A i internal area of horizontal tube (m 2 ) A o outside area of horizontal tube (m 2 ) cp specific heat (kjkg -1 K -1 ) D tube diameter (m) h heat transfer coefficient (Wm -2 K -1 ) h fg latent heat for saturated vapour (kj/kg) J G dimensionless vapour mass velocity k thermal conductivity (Wm -1 K -1 ) m g mass flow rate of glycol-water (kgs -1 ) q local heat flux Q the amount of transferred heat rate (W) r radius T temperature ( o C) u the velocity component x vapour quality Greek Symbols θ angle coordinate δ the liquid condensate thickness ρ the density Subscripts c e f Convection experimental falling film g glycol-water v vapour L Liquid pre predicted s saturation strat stratified sat saturation subc sub-cooled w tube wall µ dynamic viscosity Dimensional Groups Nu Nusselt number Fr Froude number, / Fr so Soliman s modified Froude number Eq. 7 &, Eq. 8 Ga Galileo number, g Ja Jakob number, /h Re l superficial liquid Reynolds number, 1 X tt turbulent-turbulent Lockhart Martinelli parameter,.. 1.

9 192 Manufacturing Science and Technology VI References [1] Nusselt, W., Die Oberflachencondensation des Wasserdampfes. Z.Vereins desutscher Ininuere, : p [2] Dobson, M.K. and J.C. Chato, Condensation in smooth horizontal tubes. Journal of Heat Transfer-Transactions of the Asme, (1): p [3] Cavallini, A., et al., In-tube Condensation of Halogenated Refrigerants. ASHRAE Transactions, : p [4] Thome, J.R., J. El Hajal, and A. Cavallini, Condensation in horizontal tubes, part 2: New heat transfer model based on flow regimes. International Journal of Heat and Mass Transfer, (18): p [5] Chato, J.C., Laminar Condensation Inside Horizontal and Inclined Tubes. ASHRAE Journal, : p [6] Sparrow, W.M.G., J.L., A boundary-layer treatment of laminar film condensation. Journal of Heat Transfer, : p [7] Koh, J.C.Y., Sparrow, E.M., Hartnett, J.P., The two-phase boundary layer in laminar film condensation. International Journal of Heat and Mass Transfer, : p [8] El Hajal, J., J.R. Thome, and A. Cavallini, Condensation in horizontal tubes, part 1: Two-phase flow pattern map. International Journal of Heat and Mass Transfer, (18): p [9] Kattan, N., J.R. Thome, and D. Favrat, Flow boiling in horizontal tubes: Part 1 - Development of a diabatic two-phase flow pattern map. Journal of Heat Transfer, (1): p [10] Zivi, S.M., Estimation of Steady-State Steam Void-Fraction by Means of the Principle of Minimum Entropy Production. Journal of Heat Transfer, (2): p [11] Steiner, D., Heat transfer to boiling saturated liquids. VDI-Wär Meatlas (VDI Heat Atlas), [12] Iqbal, O. and P. Bansal, In-tube condensation heat transfer of CO2 at low temperatures in a horizontal smooth tube. International Journal of Refrigeration, (2): p [13] Shah, R.K. and A.L. London, Laminar flow forced convection in ducts : a source book for compact heat exchanger analytical data. 1978, New York: Academic Press. [14] Nellis, G. and S. Klein, Heat Transfer. 2012: Cambridge University Press.

10 Manufacturing Science and Technology VI / Stratified Flow Condensation of CO 2 in a Tube at Low Temperatures / DOI References [2] Dobson, M.K. and J.C. Chato, Condensation in smooth horizontal tubes. Journal of Heat Transfer- Transactions of the Asme, (1): pp / [4] Thome, J.R., J. El Hajal, and A. Cavallini, Condensation in horizontal tubes, part 2: New heat transfer model based on flow regimes. International Journal of Heat and Mass Transfer, (18): pp /s (03) [7] Koh, J.C.Y., Sparrow, E.M., Hartnett, J.P., The two-phase boundary layer in laminar film condensation. International Journal of Heat and Mass Transfer, : pp / (61) [8] El Hajal, J., J.R. Thome, and A. Cavallini, Condensation in horizontal tubes, part 1: Two-phase flow pattern map. International Journal of Heat and Mass Transfer, (18): pp /s (03)00139-x [9] Kattan, N., J.R. Thome, and D. Favrat, Flow boiling in horizontal tubes: Part 1 - Development of a diabatic two-phase flow pattern map. Journal of Heat Transfer, (1): pp / [10] Zivi, S.M., Estimation of Steady-State Steam Void-Fraction by Means of the Principle of Minimum Entropy Production. Journal of Heat Transfer, (2): pp / [12] Iqbal, O. and P. Bansal, In-tube condensation heat transfer of CO2 at low temperatures in a horizontal smooth tube. International Journal of Refrigeration, (2): pp /j.ijrefrig [14] Nellis, G. and S. Klein, Heat Transfer. 2012: Cambridge University Press /cbo