R-134a和R-407C冷媒在水平多排小圓管內蒸發熱傳及壓降特性之實驗研究

全文

(1)國立交通大學機械工程學系碩士論文. R-134a 和 R-407C 冷媒在水平多排小圓管內蒸發熱傳及壓降特性之實驗研究 Experimental Study of Evaporation Heat Transfer and Pressure Drop Characteristics of R-134a and R-407C in Horizontal Small Tubes. 研究生：蘇峯慶指導教授：林清發博士. 中華民國九十三年六月 1.

(2) R-134a 和 R-407C 冷媒在水平多排小圓管內蒸發熱傳及壓降特性之實驗研究. Experimental Study of Evaporation Heat Transfer and Pressure Drop Characteristics of R-134a and R-407C in Horizontal Small Tubes. 研究生：蘇峯慶. Student：Feng-Qing Su. 指導教授：林清發. Advisor：Tsing-Fa Lin 國立交通大學機械工程學系碩士論文 A Thesis. Submitted to Institute of Mechanical Engineering College of Engineering National Chiao Tung University In Partial Fulfillment of the Requirements For the degree of Master of Science In Mechanical Engineering June 2004 Hsinchu, Taiwan, Republic of China 中華民國九十三年六月 2.

(3) 誌謝時光飛逝，想當初才剛進交大，對於一切總是那麼地陌生。由於學長與學弟的關心與幫助，讓我在這兩年的研究所生活過得非常充實與開心。在這兩年的研究過程中，林清發老師的教學態度與研究精神都是值得我學習的對象，相信對於往後不管是工作或是學習上都是很有幫助的。然而，從一開始對於冷凍空調方面的知識一竅不通，直到現在本論文的完成，要感謝的人很多，若不是有祐民學長和瑞琳學長的細心教導，以及召漢學弟與家銘學弟的從旁協助，可能我要花兩倍甚至更多的時間才能完成。在此，我由衷的感謝老師、學長與學弟。當然在研究室的兩年中，實驗室的好夥伴是不可少的，佳鴻、榮源(小白)與信介(介哥)不管在修課或做實驗的過程都是亦師亦友。其次對研究室成員瑞青學長、威伸學長、尚緯學長、汎鈞學長、文賢、及立傑等人感謝他們分享在實驗室的低潮與喜悅。還要感謝清華大學潘欽主任、核能所李堅雄教授與工研院能資所顏貽乙學長在口試中提出各項建設性的建議，使得本論文更趨於完備。最後要感謝家人與所有好朋友在背後的的默默關心與支持，對於每一個人的好，我一直心存感謝著。蘇峯慶. 3. 2004, 6 於新竹交大.

(4) R-134a 和 R-407C 冷媒在水平多排小圓管內蒸發熱傳及壓降特性之實驗研究研究生: 蘇峯慶. 指導教授: 林清發博士. 國立交通大學機械工程學系摘要本論文是分別針對 R-134a 和 R-407C 兩種冷媒在 0.83 mm 和 2.0 mm 兩種內徑的 28 根水平小圓管內蒸發熱傳及壓降特性之實驗研究。實驗的目的是探討改變冷媒的飽和溫度、冷媒的質通量、測試段的熱通量及測試段的進口蒸氣乾度對 R-134a 和 R-407C 兩種冷媒在此兩種小管熱交換器的影響。在實驗參數的範圍上，冷媒的飽和溫度 Tsat 被設定從 5 到 15℃，測試段熱通量 q 從 5 到 15 kW/m2，進口蒸氣乾度 xin 大約從 0.2 到 0.8。在冷媒的質通量 G 方面，由於管徑從 2.0 mm 縮小到 0.83 mm，因此在系統迴路的穩定性考量上，我們設定 G 在 2.0-mm 小管是從 200 到 400 kg/m2s，在 0.83-mm 小管是從 800 到 1500 kg/m2s。經由實驗的結果可知，R-134a 和 R-407C 兩種冷媒在 0.83-mm 和 2.0-mm 兩種管徑中，我們發現除了在較低的質通量和較高的熱通量，蒸發熱傳係數與摩擦壓降係數皆會隨著進口乾度的增加而明顯地上升。另外，在熱通量的影響方面，熱傳係數也會隨著熱量的增加而. 4.

(5) 上升。但是相較摩擦壓降係數的影響則是很微弱的。而在飽和溫度的影響上，蒸發熱傳係數會隨著系統的飽和溫度上升而增加；但是對於摩擦壓降係數則會隨著飽和溫度的上昇而減少。此外，R-134a 和 R-407C 的蒸發熱傳與摩擦壓降均隨著冷媒質通量的增加而增加。值得一提的是，R-134a 冷媒在 0.83-mm 小管中，若供給較低的冷媒質通量和較高的熱通量，蒸發熱傳係數可能會隨著乾度的增加而下降。這是由於冷媒在小管內流動時，產生部分乾化的現象，因此影響蒸發熱傳係數的下降。在相同的實驗條件與管徑的比較下，我們發現 R-407C 冷媒其整體的蒸發熱傳係數會比 R-134a 冷媒來得高。然而，在摩擦壓降係數方面，則較 R-134a 冷媒低。最後，我們將蒸發熱傳係數與摩擦壓降係數在 0.83-mm 和 2.0-mm 小管中的實驗資料作分析，並求出經驗公式。. 5.

(6) Experimental Study of Evaporation Heat Transfer and Pressure Drop Characteristics of R-134a and R-407C in Horizontal Small Tubes Student：Feng-Qing Su. Advisor：Prof. Tsing-Fa Lin. Institute of Mechanical Engineering National Chiao Tung University. ABSTRACT An experiment is carried out in the present study to investigate the characteristics of the evaporation heat transfer and frictional pressure drop for refrigerants R-134a and R-407C flowing in horizontal small tubes having the same inside diameter of 0.83 mm or 2.0 mm. In the experiment for the 2.0-mm small tubes, the refrigerant mass flux G is varied from 200 to 400 kg/m2s, imposed heat flux q from 5 to 15 kW/m2, inlet vapor quality xin from 0.2 to 0.8 and refrigerant saturation temperature Tsat from 5 to 15℃. While for the 0.83-mm small tubes, G is varied from 800 to 1500 kg/m2s with the other parameters varied in the same ranges as those for Di=2.0 mm. In the study the effects of the refrigerant vapor quality, mass flux, saturation temperature and imposed heat flux on the measured evaporation heat transfer coefficients and frictional pressure drops are examined in detail. For R-134a and R-407C in the 2.0-mm or 0.83-mm small tubes, the experimental data clearly show that both the R-134a and R-407C evaporation heat transfer coefficients and frictional pressure drops increase almost linearly and significantly with the vapor quality of the refrigerant, except at low mass flux and high heat flux. Besides, the evaporation heat transfer coefficients also increase substantially with the rise in the imposed heat flux. But the effect of q on the frictional pressure drop is rather weak. Moreover, a significant increase in the evaporation heat transfer i.

(7) coefficients results for a rise in Tsat, but an opposite trend is noted for the frictional pressure drop. Furthermore, both the R-134a and R-407C evaporation heat transfer coefficients and frictional pressure drops increase substantially with the refrigerant mass flux. At low R-134a mass flux and high imposed heat flux the evaporation heat transfer coefficient in the small tubes (Di=0.83 mm) may decline at increasing vapor quality when the quality is high. This decline of hr at rising x is attributed to the partial dryout of the refrigerant flow in the small tubes at these conditions. We also note that under the same xin, Tsat, G, q and Di, refrigerant R-407C has a higher evaporation heat transfer coefficient and a lower frictional pressure drop when compared with that for R-134a. Finally, the empirical correlations for the R-134a and R-407C evaporation heat transfer coefficient and friction factor in 0.83-mm and 2.0-mm small tubes are proposed.. ii.

(8) TABLE OF CONTENTS ABSTRACT. i. TABLE OF CONTENTS. iii. LIST OF FIGURES. v. LIST OF TABLE. xv. NOMENCLATURE. xvi. CHAPTER 1 INTRODUCTION. 1. 1.1 Motivation of the Present Study. 1. 1.2 Literature Review - Boiling Heat Transfer and Pressure Drop in Small Channels. 2. 1.2.1 Single Phase heat transfer in small tubes. 2. 1.2.2 Flow boiling heat transfer in small tubes. 3. 1.2.3 Flow Boiling Heat Transfer in Conventional Tubes. 7. 1.2.4 Pressure Drop in Small Tubes. 8. 1.3 Review of Correlation Equations for Two Phase Flow Boiling Heat Transfer. 8. 1.4 Objective of This Study CHAPTER 2 EXPERIMENTAL APPARATUS AND PROCEDURES. 10 17. 2.1 Refrigerant Loop. 17. 2.2 Test Section. 18. 2.3 Water Loop for Pre-heater. 19. 2.4 Water-Glycol Loop. 20. 2.5 DC Power Supply. 20. 2.6 Data Acquisition. 20. 2.7 Experimental Procedures. 21. CHAPTER 3 DATA REDUCTION. 28. 3.1 Single Phase Heat Transfer. 28. 3.2 Two Phase Heat Transfer. 29. 3.3 Friction Factor. 30 iii.

(9) 3.4 Uncertainty Analysis. 31. CHAPTER 4 EVAPORATION HEAT TRANSFER. 34. 4.1 Single Phase Heat Transfer. 34. 4.2 Evaporation Heat Transfer in 2.0-mm Tubes. 35. 4.3 Evaporation Heat Transfer in the Smaller Tubes (Di=0.83 mm). 38. 4.4 Correlation equation for Evaporation Heat Transfer Coefficients. 40. 4.5 Concluding Remarks. 41. CHAPTER 5 TWO PHASE FRICTIONAL PRESSURE DROP. 83. 5.1 Friction Pressure Drop in 2.0-mm Tubes. 83. 5.2 Friction Pressure Drop in the Smaller Tubes (Di=0.83 mm). 85. 5.3 Correlation equation for Frictional Pressure Drops. 85. 5.4 Concluding Remarks. 86. CHAPTER 6 CONCLUDING REMARKS. 125. REFERENCES. 127. iv.

(10) LIST OF FIGURES Experiment Apparatus Fig. 2.1 Schematiclayout of the experimental system ------------------------------------------- 24 Fig. 2.2 Schematic diagram of test section along with the inlet and exit sections ----------- 25 Fig. 2.3 (a) Test section of the small tubes and (b) locations of the thermocouples --------- 26 Fig. 2.4 The detail of test section ------------------------------------------------------------------- 27 R-407C & R-134a Single-Phase Heat Transfer in 2.0 & 0.83-mm Small tubes Fig. 4.1 Comparison of the present data in 2.0-mm small tubes for the R-407C & R-134a liquid phase heat transfer coefficient with the Dittu-Boelter and Gnielinski correlations --------------------------------------------------------------------- 44 Fig. 4.2 Comparison of the present data in 0.83-mm small tubes for the R-407c & R-134a liquid phase heat transfer coefficient with the Dittu-Boelter and Gnielinski correlations --------------------------------------------------------------------- 45 R-134a & R-407C Evaporation Heat Transfer in 2.0 & 0.83-mm Small tubes R-134a in 2.0-mm small tubes For Tsat effects Fig. 4.3 Variations of R-134a evaporation heat transfer coefficient with inlet vapor quality in 2.0-mm small tubes at q=5 kW/m2 for various Tsat for (a) G=200 kg/m2s, (b) G=300 kg/m2s, and (c) G=400 kg/m2s. ------------------------------------ 46 Fig. 4.4 Variations of R-134a evaporation heat transfer coefficient with inlet vapor quality in 2.0-mm small tubes at q=10 kW/m2 for various Tsat for (a) G=200 kg/m2s, (b) G=300 kg/m2s, and (c) G=400 kg/m2s. ------------------------------------ 47 Fig. 4.5 Variations of R-134a evaporation heat transfer coefficient with inlet vapor quality in 2.0-mm small tubes at q=15 kW/m2 for various Tsat for (a) G=200 kg/m2s, (b) G=300 kg/m2s, and (c) G=400 kg/m2s. ------------------------------------ 48 For G effects Fig. 4.6 Variations of R-134a evaporation heat transfer coefficient with inlet vapor quality in 2.0-mm small tubes at q=5 kW/m2 for various G for (a) Tsat=5℃, (b) Tsat=10℃, and (c) Tsat=15℃. --------------------------------------------------------- 49 v.

(11) Fig. 4.7 Variations of R-134a evaporation heat transfer coefficient with inlet vapor quality in 2.0-mm small tubes at q=10 kW/m2 for various G for (a) Tsat=5℃, (b) Tsat=10℃, and (c) Tsat=15℃. --------------------------------------------------------- 50 Fig. 4.8 Variations of R-134a evaporation heat transfer coefficient with inlet vapor quality in 2.0-mm small tubes at q=15 kW/m2 for various G for (a) Tsat=5℃, (b) Tsat=10℃, and (c) Tsat=15℃. --------------------------------------------------------- 51 For q effects Fig. 4.9 Variations of R-134a evaporation heat transfer coefficient with inlet vapor quality in 2.0-mm small tubes at Tsat=5℃ for various q for (a) G=200 kg/m2s, (b) G=300 kg/m2s, and (c) G=400 kg/m2s. ------------------------------------ 52 Fig. 4.10 Variations of R-134a evaporation heat transfer coefficient with inlet vapor quality in 2.0-mm small tubes at Tsat=10℃ for various q for (a) G=200 kg/m2s, (b) G=300 kg/m2s, and (c) G=400 kg/m2s. ------------------------------------ 53 Fig. 4.11 Variations of R-134a evaporation heat transfer coefficient with inlet vapor quality in 2.0-mm small tubes at Tsat=15℃ for various q for (a) G=200 kg/m2s, (b) G=300 kg/m2s, and (c) G=400 kg/m2s. ------------------------------------ 54 R-407C in 2.0-mm small tubes For Tsat effects Fig. 4.12 Variations of R-407C evaporation heat transfer coefficient with inlet vapor quality in 2.0-mm small tubes at q=5 kW/m2 for various Tsat for (a) G=200 kg/m2s, (b) G=300 kg/m2s, and (c) G=400 kg/m2s. ------------------------------------ 55 Fig. 4.13 Variations of R-407C evaporation heat transfer coefficient with inlet vapor quality in 2.0-mm small tubes at q=10 kW/m2 for various Tsat for (a) G=200 kg/m2s, (b) G=300 kg/m2s, and (c) G=400 kg/m2s. ------------------------------------ 56 Fig. 4.14 Variations of R-407C evaporation heat transfer coefficient with inlet vapor quality in 2.0-mm small tubes at q=15 kW/m2 for various Tsat for (a) G=200 kg/m2s, (b) G=300 kg/m2s, and (c) G=400 kg/m2s. ------------------------------------ 57 For G effects Fig. 4.15 Variations of R-407C evaporation heat transfer coefficient with inlet vapor vi.

(12) quality in 2.0-mm small tubes at q=5 kW/m2 for various G for (a) Tsat=5℃, (b) Tsat=10℃, and (c) Tsat=15℃. --------------------------------------------------------- 58 Fig. 4.16 Variations of R-407C evaporation heat transfer coefficient with inlet vapor quality in 2.0-mm small tubes at q=10 kW/m2 for various G for (a) Tsat=5℃, (b) Tsat=10℃, and (c) Tsat=15℃. --------------------------------------------------------- 59 Fig. 4.17 Variations of R-407C evaporation heat transfer coefficient with inlet vapor quality in 2.0-mm small tubes at q=15 kW/m2 for various G for (a) Tsat=5℃, (b) Tsat=10℃, and (c) Tsat=15℃. --------------------------------------------------------- 60 For q effects Fig. 4.18 Variations of R-407C evaporation heat transfer coefficient with inlet vapor quality in 2.0-mm small tubes at Tsat=5℃ for various q for (a) G=200 kg/m2s, (b) G=300 kg/m2s, and (c) G=400 kg/m2s. ------------------------------------ 61 Fig. 4.19 Variations of R-407C evaporation heat transfer coefficient with inlet vapor quality in 2.0-mm small tubes at Tsat=10℃ for various q for (a) G=200 kg/m2s, (b) G=300 kg/m2s, and (c) G=400 kg/m2s. ------------------------------------ 62 Fig. 4.20 Variations of R-407C evaporation heat transfer coefficient with inlet vapor quality in 2.0-mm small tubes at Tsat=15℃ for various q for (a) G=200 kg/m2s, (b) G=300 kg/m2s, and (c) G=400 kg/m2s. ------------------------------------ 63 R-134a in 0.83-mm small tubes For Tsat effects Fig. 4.21 Variations of R-134a evaporation heat transfer coefficient with inlet vapor quality in 0.83-mm small tubes at q=5 kW/m2 for various Tsat for (a) G=800 kg/m2s, (b) G=1150 kg/m2s, and (c) G=1500 kg/m2s. --------------------------------- 64 Fig. 4.22 Variations of R-134a evaporation heat transfer coefficient with inlet vapor quality in 0.83-mm small tubes at q=10 kW/m2 for various Tsat for (a) G=800 kg/m2s, (b) G=1150 kg/m2s, and (c) G=1500 kg/m2s. ------------------------ 65 Fig. 4.23 Variations of R-134a evaporation heat transfer coefficient with inlet vapor quality in 0.83-mm small tubes at q=15 kW/m2 for various Tsat for (a) G=800 kg/m2s, (b) G=1150 kg/m2s, and (c) G=1500 kg/m2s. ------------------------ 66 vii.

(13) For G effects Fig. 4.24 Variations of R-134a evaporation heat transfer coefficient with inlet vapor quality in 0.83-mm small tubes at q=5 kW/m2 for various G for (a) Tsat=5℃, (b) Tsat=10℃, and (c) Tsat=15℃. --------------------------------------------------------- 67 Fig. 4.25 Variations of R-134a evaporation heat transfer coefficient with inlet vapor quality in 0.83-mm small tubes at q=10 kW/m2 for various G for (a) Tsat=5 ℃, (b) Tsat=10℃, and (c) Tsat=15℃. ----------------------------------------------------- 68 Fig. 4.26 Variations of R-134a evaporation heat transfer coefficient with inlet vapor quality in 0.83-mm small tubes at q=15 kW/m2 for various G for (a) Tsat=5 ℃, (b) Tsat=10℃, and (c) Tsat=15℃. ----------------------------------------------------- 69 For q effects Fig. 4.27 Variations of R-134a evaporation heat transfer coefficient with inlet vapor quality in 0.83-mm small tubes at Tsat=5℃ for various q for (a) G=800 kg/m2s, (b) G=1150 kg/m2s, and (c) G=1500 kg/m2s. --------------------------------- 70 Fig. 4.28 Variations of R-134a evaporation heat transfer coefficient with inlet vapor quality in 0.83-mm small tubes at Tsat=10℃ for various q for (a) G=800 kg/m2s, (b) G=1150 kg/m2s, and (c) G=1500 kg/m2s. --------------------------------- 71 Fig. 4.29 Variations of R-134a evaporation heat transfer coefficient with inlet vapor quality in 0.83-mm small tubes at Tsat=15℃ for various q for (a) G=800 kg/m2s, (b) G=1150 kg/m2s, and (c) G=1500 kg/m2s. --------------------------------- 72 R-407C in 0.83-mm small tubes For Tsat effects Fig. 4.30 Variations of R-407C evaporation heat transfer coefficient with inlet vapor quality in 0.83-mm small tubes at q=5 kW/m2 for various Tsat for (a) G=800 kg/m2s, (b) G=1150 kg/m2s, and (c) G=1500 kg/m2s. --------------------------------- 73 Fig. 4.31 Variations of R-407C evaporation heat transfer coefficient with inlet vapor quality in 0.83-mm small tubes at q=10 kW/m2 for various Tsat for (a) G=800 kg/m2s, (b) G=1150 kg/m2s, and (c) G=1500 kg/m2s. ------------------------ 74 Fig. 4.32 Variations of R-407C evaporation heat transfer coefficient with inlet vapor viii.

(14) quality in 0.83-mm small tubes at q=15 kW/m2 for various Tsat for (a) G=800 kg/m2s, (b) G=1150 kg/m2s, and (c) G=1500 kg/m2s. ------------------------ 75 For G effects Fig. 4.33 Variations of R-407C evaporation heat transfer coefficient with inlet vapor quality in 0.83-mm small tubes at q=5 kW/m2 for various G for (a) Tsat=5℃, (b) Tsat=10℃, and (c) Tsat=15℃. --------------------------------------------------------- 76 Fig. 4.34 Variations of R-407C evaporation heat transfer coefficient with inlet vapor quality in 0.83-mm small tubes at q=10 kW/m2 for various G for (a) Tsat=5 ℃, (b) Tsat=10℃, and (c) Tsat=15℃. ----------------------------------------------------- 77 Fig. 4.35 Variations of R-407C evaporation heat transfer coefficient with inlet vapor quality in 0.83-mm small tubes at q=15 kW/m2 for various G for (a) Tsat=5 ℃, (b) Tsat=10℃, and (c) Tsat=15℃. ----------------------------------------------------- 78 For q effects Fig. 4.36 Variations of R-407C evaporation heat transfer coefficient with inlet vapor quality in 0.83-mm small tubes at Tsat=5℃ for various q for (a) G=800 kg/m2s, (b) G=1150 kg/m2s, and (c) G=1500 kg/m2s. --------------------------------- 79 Fig. 4.37 Variations of R-407C evaporation heat transfer coefficient with inlet vapor quality in 0.83-mm small tubes at Tsat=10℃ for various q for (a) G=800 kg/m2s, (b) G=1150 kg/m2s, and (c) G=1500 kg/m2s. --------------------------------- 80 Fig. 4.38 Variations of R-407C evaporation heat transfer coefficient with inlet vapor quality in 0.83-mm small tubes at Tsat=15℃ for various q for (a) G=800 kg/m2s, (b) G=1150 kg/m2s, and (c) G=1500 kg/m2s. --------------------------------- 81 Fig. 4.39 Comparison of the measured data for heat transfer coefficient in the evaporation of R-134a and R-407 in 0.83-mm and 2mm small tubes with the proposed correlation. ------------------------------------------------------------------ 82 R-134a & R-407C Two Phase Pressure Drop in 2.0 & 0.83-mm Small tubes R-134a in 2.0-mm small tubes For Tsat effects Fig. 5.1 Variations of R-134a frictional pressure drop with inlet vapor quality in ix.

(15) 2.0-mm small tubes at q=5kW/m2 for various Tsat for (a) G=200 kg/m2s, (b) G=300 kg/m2s, and (c) G=400 kg/m2s. -------------------------------------------------- 88 Fig. 5.2 Variations of R-134a frictional pressure drop with inlet vapor quality in 2.0-mm small tubes at q=10kW/m2 for various Tsat for (a) G=200 kg/m2s, (b) G=300 kg/m2s, and (c) G=400 kg/m2s. -------------------------------------------------- 89 Fig. 5.3 Variations of R-134a frictional pressure drop with inlet vapor quality in 2.0-mm small tubes at q=15kW/m2 for various Tsat for (a) G=200 kg/m2s, (b) G=300 kg/m2s, and (c) G=400 kg/m2s. -------------------------------------------------- 90 For G effects Fig. 5.4 Variations of R-134a frictional pressure drop with inlet vapor quality in 2.0-mm small tubes at q=5 kW/m2 for various G for (a) Tsat=5℃, (b) Tsat=10℃, and (c) Tsat=15℃.-------------------------------------------------------------- 91 Fig. 5.5 Variations of R-134a frictional pressure drop with inlet vapor quality in 2.0-mm small tubes at q=10 kW/m2 for various G for (a) Tsat=5℃, (b) Tsat=10℃, and (c) Tsat=15℃.-------------------------------------------------------------- 92 Fig. 5.6 Variations of R-134a frictional pressure drop with inlet vapor quality in 2.0-mm small tubes at q=15 kW/m2 for various G for (a) Tsat=5℃, (b) Tsat=10℃, and (c) Tsat=15℃.-------------------------------------------------------------- 93 For q effects Fig. 5.7 Variations of R-134a frictional pressure drop with inlet vapor quality in 2.0-mm small tubes at Tsat=5℃ for various q for (a) G=200 kg/m2s, (b) G=300 kg/m2s, and (c) G=400 kg/m2s. -------------------------------------------------- 94 Fig. 5.8 Variations of R-134a frictional pressure drop with inlet vapor quality in 2.0-mm small tubes at Tsat=10℃ for various q for (a) G=200 kg/m2s, (b) G=300 kg/m2s, and (c) G=400 kg/m2s. -------------------------------------------------- 95 Fig. 5.9 Variations of R-134a frictional pressure drop with inlet vapor quality in 2.0-mm small tubes at Tsat=15℃ for various q for (a) G=200 kg/m2s, (b) G=300 kg/m2s, and (c) G=400 kg/m2s. -------------------------------------------------- 96 R-407C in 2.0-mm small tubes For Tsat effects x.

(16) Fig. 5.10 Variations of R-407C frictional pressure drop with inlet vapor quality in 2.0-mm small tubes at q=5kW/m2 for various Tsat for (a) G=200 kg/m2s, (b) G=300 kg/m2s, and (c) G=400 kg/m2s. -------------------------------------------------- 97 Fig. 5.11 Variations of R-407C frictional pressure drop with inlet vapor quality in 2.0-mm small tubes at q=10kW/m2 for various Tsat for (a) G=200 kg/m2s, (b) G=300 kg/m2s, and (c) G=400 kg/m2s. -------------------------------------------------- 98 Fig. 5.12 Variations of R-407C frictional pressure drop with inlet vapor quality in 2.0-mm small tubes at q=15kW/m2 for various Tsat for (a) G=200 kg/m2s, (b) G=300 kg/m2s, and (c) G=400 kg/m2s. -------------------------------------------------- 99 For G effects Fig. 5.13 Variations of R-407C frictional pressure drop with inlet vapor quality in 2.0-mm small tubes at q=5 kW/m2 for various G for (a) Tsat=5℃, (b) Tsat=10℃, and (c) Tsat=15℃.-------------------------------------------------------------- 100 Fig. 5.14 Variations of R-407C frictional pressure drop with inlet vapor quality in 2.0-mm small tubes at q=10 kW/m2 for various G for (a) Tsat=5℃, (b) Tsat=10℃, and (c) Tsat=15℃.-------------------------------------------------------------- 101 Fig. 5.15 Variations of R-407C frictional pressure drop with inlet vapor quality in 2.0-mm small tubes at q=15 kW/m2 for various G for (a) Tsat=5℃, (b) Tsat=10℃, and (c) Tsat=15℃.-------------------------------------------------------------- 102 For q effects Fig. 5.16 Variations of R-407C frictional pressure drop with inlet vapor quality in 2.0-mm small tubes at Tsat=5℃ for various q for (a) G=200 kg/m2s, (b) G=300 kg/m2s, and (c) G=400 kg/m2s. -------------------------------------------------- 103 Fig. 5.17 Variations of R-407C frictional pressure drop with inlet vapor quality in 2.0-mm small tubes at Tsat=10℃ for various q for (a) G=200 kg/m2s, (b) G=300 kg/m2s, and (c) G=400 kg/m2s. -------------------------------------------------- 104 Fig. 5.18 Variations of R-407C frictional pressure drop with inlet vapor quality in 2.0-mm small tubes at Tsat=15℃ for various q for (a) G=200 kg/m2s, (b) G=300 kg/m2s, and (c) G=400 kg/m2s. -------------------------------------------------- 105 R-134a in 0.83-mm small tubes xi.

(17) For Tsat effects Fig. 5.19 Variations of R-134a frictional pressure drop with inlet vapor quality in 0.83-mm small tubes at q=5kW/m2 for various Tsat for (a) G=800 kg/m2s, (b) G=1150 kg/m2s, and (c) G=1500 kg/m2s. ----------------------------------------------- 106 Fig. 5.20 Variations of R-134a frictional pressure drop with inlet vapor quality in 0.83-mm small tubes at q=10kW/m2 for various Tsat for (a) G=800 kg/m2s, (b) G=1150 kg/m2s, and (c) G=1500 kg/m2s. ------------------------------------------- 107 Fig. 5.21 Variations of R-134a frictional pressure drop with inlet vapor quality in 0.83-mm small tubes at q=15kW/m2 for various Tsat for (a) G=800 kg/m2s, (b) G=1150 kg/m2s, and (c) G=1500 kg/m2s. ------------------------------------------- 108 For G effects Fig. 5.22 Variations of R-134a frictional pressure drop with inlet vapor quality in 0.83-mm small tubes at q=5 kW/m2 for various G for (a) Tsat=5℃, (b) Tsat=10℃, and (c) Tsat=15℃.-------------------------------------------------------------- 109 Fig. 5.23 Variations of R-134a frictional pressure drop with inlet vapor quality in 0.83-mm small tubes at q=10 kW/m2 for various G for (a) Tsat=5℃, (b) Tsat=10℃, and (c) Tsat=15℃.-------------------------------------------------------------- 110 Fig. 5.24 Variations of R-134a frictional pressure drop with inlet vapor quality in 0.83-mm small tubes at q=15 kW/m2 for various G for (a) Tsat=5℃, (b) Tsat=10℃, and (c) Tsat=15℃.-------------------------------------------------------------- 111 For q effects Fig. 5.25 Variations of R-134a frictional pressure drop with inlet vapor quality in 0.83-mm small tubes at Tsat=5℃ for various q for (a) G=800 kg/m2s, (b) G=1150 kg/m2s, and (c) G=1500 kg/m2s. ----------------------------------------------- 112 Fig. 5.26 Variations of R-134a frictional pressure drop with inlet vapor quality in 0.83-mm small tubes at Tsat=10℃ for various q for (a) G=800 kg/m2s, (b) G=1150 kg/m2s, and (c) G=1500 kg/m2s. ----------------------------------------------- 113 Fig. 5.27 Variations of R-134a frictional pressure drop with inlet vapor quality in 0.83-mm small tubes at Tsat=15℃ for various q for (a) G=800 kg/m2s, (b) G=1150 kg/m2s, and (c) G=1500 kg/m2s. ----------------------------------------------- 114 xii.

(18) R-407C in 0.83-mm small tubes For Tsat effects Fig. 5.28 Variations of R-407C frictional pressure drop with inlet vapor quality in 0.83-mm small tubes at q=5kW/m2 for various Tsat for (a) G=800 kg/m2s, (b) G=1150 kg/m2s, and (c) G=1500 kg/m2s. ----------------------------------------------- 115 Fig. 5.29 Variations of R-407C frictional pressure drop with inlet vapor quality in 0.83-mm small tubes at q=10kW/m2 for various Tsat for (a) G=800 kg/m2s, (b) G=1150 kg/m2s, and (c) G=1500 kg/m2s. ------------------------------------------- 116 Fig. 5.30 Variations of R-407C frictional pressure drop with inlet vapor quality in 0.83-mm small tubes at q=15kW/m2 for various Tsat for (a) G=800 kg/m2s, (b) G=1150 kg/m2s, and (c) G=1500 kg/m2s. ------------------------------------------- 117 For G effects Fig. 5.31 Variations of R-407C frictional pressure drop with inlet vapor quality in 0.83-mm small tubes at q=5 kW/m2 for various G for (a) Tsat=5℃, (b) Tsat=10℃, and (c) Tsat=15℃.-------------------------------------------------------------- 118 Fig. 5.32 Variations of R-407C frictional pressure drop with inlet vapor quality in 0.83-mm small tubes at q=10 kW/m2 for various G for (a) Tsat=5℃, (b) Tsat=10℃, and (c) Tsat=15℃.-------------------------------------------------------------- 119 Fig. 5.33 Variations of R-407C frictional pressure drop with inlet vapor quality in 0.83-mm small tubes at q=15 kW/m2 for various G for (a) Tsat=5℃, (b) Tsat=10℃, and (c) Tsat=15℃.-------------------------------------------------------------- 120 For q effects Fig. 5.34 Variations of R-407C frictional pressure drop with inlet vapor quality in 0.83-mm small tubes at Tsat=5℃ for various q for (a) G=800 kg/m2s, (b) G=1150 kg/m2s, and (c) G=1500 kg/m2s. ----------------------------------------------- 121 Fig. 5.35 Variations of R-407C frictional pressure drop with inlet vapor quality in 0.83-mm small tubes at Tsat=10℃ for various q for (a) G=800 kg/m2s, (b) G=1150 kg/m2s, and (c) G=1500 kg/m2s. ----------------------------------------------- 122 Fig. 5.36 Variations of R-407C frictional pressure drop with inlet vapor quality in 0.83-mm small tubes at Tsat=15℃ for various q for (a) G=800 kg/m2s, (b) xiii.

(19) G=1150 kg/m2s, and (c) G=1500 kg/m2s. ----------------------------------------------- 123 Fig. 5.37 Comparison of the measured data for frictional pressure drop coefficient in the evaporation of R-134a and R-407 in 0.83-mm and 2.0-mm small tubes with the proposed correlation. ------------------------------------------------------------ 124. xiv.

(20) LIST OF TABLE Table 1.1. Comparison of Properties of three HFC refrigerants for air conditioning and refrigeration applications --------------------------------------------------------- 11. Table 1.2. Classification of Two-phase heat transfer coefficient and friction factor correlations ------------------------------------------------------------------------------ 12. Table 2.1. Comparison of thermophysical properties of R-134a and R-407C--------------- 23. Table 3.1. Summary of the uncertainty analysis------------------------------------------------- 33. Table 4.1. List of conditions of the experimental parameters for R-134a and R-407C refrigerants -------------------------------------------------------------------- 43. xv.

(21) NOMENCLATURE A. heat transfer area of the small tube, m2. Bo. Boiling number, Bo=. Cc. coefficient of contraction. Co. Convective number, Co = (. cp. specific heat, J/kg℃. D. diameter of small tube, mm. E. enhancement factor. f. friction factor. G. mass flux, kg/m2s. h. heat transfer coefficient, W/m2℃. i fg. enthalpy of vaporization, J/kg℃. I. measured current from DC power supply, A. L. length of small tube, m. q , dimensionless G ⋅ i fg. ρ g 0.5 (1 − x) 0.8 ) ( ) , dimensionless x ρl. [. σ g( ρl − ρ g ). Nconf. Confinement number, N conf =. Nu. Nusselt number, Nu =. P. system pressure, kPa. Pr. Prandtl number, Pr =. Pr. reduced pressure. q. average imposed heat flux, W/m2. Q. heat transfer rate, W. Re. Reynolds number, Re =. ]0.5. Di. , dimensionless. hD , dimensionless k. µ ⋅ cp k. , dimensionless. G ⋅ Di. µ. , dimensionless. xvi.

(22) S. suppression factor. T. temperature,℃. Tr,sat. bubble point temperature of the refrigerant R-134a or R-407C, ℃. V. measured voltage from DC power supply, V. W. mass flow rate. Wel. Weber number based on liquid, Wel =. x. vapor quality. X. Martinelli parameter. Greek Symbols ∆P. pressure drop. ∆T. temperature difference, ℃. ∆x. total quality change in the small tube. ρ. density, kg/m3. ν. specific volume, m3/kg. σ. surface tension. σc. contraction ratio. α. void fraction, dimensionless. ε. relative heat loss. Subscripts a. acceleration. ave. average. conv. convection boiling. cs. cross-section of small tube. eq. equilibrium. evp. evaporation heat transfer xvii. G2 ⋅ D , dimensionless ρl ⋅ σ.

(23) exp. experiment. f. friction or liquid phase. fg. difference between liquid phase and vapor phase. g. vapor phase. hom. homogeneous. i. inside diameter of the tube. i,o. at inlet and exit of the test section. l. all-liquid nonboiling heat transfer. lat,sen. latent and sensible heats. m. average value for the two phase mixture or between the inlet and exit. n. net power input to the refrigerant R-134a or R-407C. nb. nucleate boiling. p. preheater. pool. pool boiling. r. refrigerant. s. heater surface for area or single phase for heat transfer rate. sat. saturation. t. total. tp. two phase. tt. both liquid and vapor phases in turbulent flow. w. water. wall. the inner wall of small tube. xviii.

(24) CHAPTER 1 INTRODUCTION. 1.1 Motivation of the Present Study The Montreal Protocol signed in 1987 allowed the use of the chlorofluorocarbons refrigerants (CFCs) such as CFC-11, CFC-12, CFC-113, CFC-114 and CFC-115 up to 1996, due to the presence of chlorine and carbon in these refrigerants which were found to destruct the ozone layer in the outer atmosphere around the earth and increase the Total Equivalent Warming Impact (TEWI) [1]. Moreover, refrigerant R-22 which is widely used in medium- and small-sized air conditioning system is scheduled to be phased out at the year 2030. Thus, the search for the replacement of CFCs and R-22 becomes very urgent since 1987. Several new refrigerants have been developed for this purpose. For examples the hydrofluorocarbons refrigerants (HFCs) such as R-134a, R-410A, R-407C, R-409A, R-401A and R-290 have been considered to be suitable and some are currently in use. It is well known that R-134a is a single-component HFC refrigerant and has similar thermophysical properties to R-12. However, R-407C is a multicomponent HFC mixture consisting of R-32/R-125/R-134a (23/25/52 wt%) with its thermophysical properties close to R-22. The operating conditions of refrigerants R-134a, R-410A and R-407C are given in Table 1.1. In air conditioning system design, the use of compact heat exchangers to promote the performance of the air conditioning is beneficial. Ultra-compact heat exchangers also called microchannel heat exchangers or compact high intensity coolers. Such heat exchangers are distinguished by a very high ratio of surface area to volume, low thermal resistances, small volumes, lower total mass and low inventory of working 1.

(25) fluids [2]. The understanding of heat transfer and flow characteristics in the small channels consisted in compact heat exchangers are therefore very important. In the present study we plan to investigate the heat transfer and pressure drop characteristics for the evaporation of refrigerants R-134a and R-407C in small pipes.. 1.2 Literature Review - Boiling Heat Transfer and Pressure Drop in Small Channels The size of the channels in a compact heat exchanger can significantly affect the performance of the exchanger when it is small enough. Mehendale et al. [2] suggested to use the hydraulic diameter Dh of the channels to classify the compact heat exchangers: (1) micro heat exchangers for Dh =1 ~ 100 µm, (2) meso heat exchangers for Dh =100 µm ~ 1 mm, (3) compact heat exchangers for Dh =1 ~ 6 mm, and (4) conventional heat exchangers for Dh >6 mm. Another classification from Kandlikar [3] proposed a different criterion: Dh >3 mm for the convectional channels, 0.2 mm< Dh <3 mm for the mini-channels, and 0.01 mm< Dh <0.2 mm for the micro channels.. 1.2.1 Single Phase heat transfer in small tubes Before reviewing the works on the flow boiling heat transfer, several studies on the single-phase heat transfer in small tubes are reviewed first. A decade age, Peng and Wang [4] investigated single-phase heat transfer for liquid methanol in a duct made of small rectangular channels with Dh=0.31, 0.51 and 0.646 mm. Their experimental data showed that the liquid velocity, liquid subcooling, liquid properties and geometry of the channels all exhibited significant effects on the heat transfer performance. Enhancement in heat transfer was noted at increasing liquid velocity, subcooling and channel number. In a continuing study for water [5] they also found 2.

(26) that the geometric configuration of channels had a significant effect on the single-phase convective heat transfer and flow characteristics. Specifically, the heat transfer was noted to depend on the aspect ratio and the ratio of the hydraulic diameter to the center-to-center distance of the channels. Furthermore, the turbulent heat transfer depends on a new dimensionless variable, Z, defined as. min(H, W ) , H and max(H, W ). W are the height and width of the channel, respectively. Recently in an experiment for R-134a liquid flow in mini-channels, Agostini et al. [6] compared their heat transfer data with some existing correlations. They concluded that the correlation of Shah and London [7] was good for Re <500 and the Gnielinsky correlation [8] was appropriate for 2,300< Re <106 and 0.6< Pr <105. Moreover, the Dittus-Boelter correlation [9] was suitable for Re >105 and 0.7< Pr <16,700.. 1.2.2 Flow boiling heat transfer in small tubes Heat transfer associated with flow boiling in a conventional channel is normally considered to result from two mechanisms: nucleate boiling and convective boiling. In the situation with a dominant nucleate boiling mechanism, the boiling heat transfer coefficient is mainly dependent on the imposed heat flux and the refrigerant saturated temperature or pressure. While for the cases with a predominated convective boiling mechanism, heat transfer coefficient is mainly dependent on the refrigerant mass flux and vapor quality. Therefore in the study of flow boiling heat transfer in mini- and micro- channels, which mechanism is prevalent is the main issue of many investigations. Yu et al. [10] recently examined flow boiling heat transfer for water in a 2.98 mm diameter channel. They found that the boiling heat transfer of water in the small channel (vapor quality above 0.5) was dependent on the heat flux but was independent of the mass flux. They concluded that the nucleate boiling was dominant over the 3.

(27) convective boiling in small channels. The results are significantly different from these for the conventional channels, where the mass flux effect can be substantial. Similarly, Sumith et al. [11] measured the saturated flow boiling heat transfer and pressure drop of water in the test section which was made of a stainless steel tube with an inner diameter of 1.45 mm. They indicated that the dominant flow pattern in the tube was a slug-annular or an annular flow, and liquid film evaporation dominated the heat transfer. An experimental measurement carried out by Yan and Lin [12, 13] to study the evaporation heat transfer and pressure drop of R-134a in a tube bank forming by 28 small pipes (Di=2.0 mm) revealed that both the nucleate and convective boiling mechanisms were important and the evaporation heat transfer in the small pipes was significantly higher than that in large tubes. A visualization investigation was presented by Nino et al. [14] to examine R-134a refrigerant in a multiport microchannel tube with Dh=1.5 mm. They proposed a method to describe the fraction of time or the probability that a flow pattern existed in a particular flow condition. A recent study from Fujita et al. [15] for R-123 boiling in a horizontal small tube with an inside diameter of 1.12 mm suggested that heat transfer in the flow was dominated by the nucleate boiling and the effects of the refrigerant mass flux and vapor quality to the boiling heat transfer were very weak. A similar study for R-113 boiling conducted by Lazarek and Black [16] noted the negligible variation of the boiling heat transfer coefficient with the local vapor quality, which implied that the wall heat transfer process was again controlled by nucleate boiling. In a vertical small tube with an inside diameter of 1 mm with R-141b refrigerant flowing in it, Lin et al. [17] found that at low quality, nucleate boiling dominated. But at higher quality, convective boiling dominated. In a further study [18], they examined the same refrigerant in four tubes with diameters of 1.1, 1.8, 2.8, 3.6 mm and one square tube of cross section 2×2 4.

(28) mm2. Their results indicate that the mean heat transfer coefficient in a tube or channel is independent of the mass flux and tube diameter but is a function of the imposed heat flux. Cornwell and Kew [19] observed boiling in refrigerant R-113 flow and measured the heat transfer for two geometries: one had 75 channels with Dh=1.03 mm and other had 36 channels with Dh=1.64 mm. Their experimental work suggested the presence of three two-phase flow patterns in the channels: isolated bubble, confined bubble and annular-slug flow. In a continuing study [20], they investigated refrigerant R-141b boiling in a horizontal tube with its inner diameter ranging from 1.39 to 3.69 mm and proposed that flow boiling in a narrow channel might be through one of four mechanisms: nucleate boiling, confined bubble boiling, convective boiling and partial dry-out. They further indicated that, except at very low heat flux, the boiling showed a strong dependence on the heat flux, a weak dependence on the mass flux, and independence of the quality. Besides, they introduced a new dimensionless group named as the confinement number, N conf ≡. [σ /(g( ρl − ρ g ))]1/ 2 Dh. , which represented the. importance of the restriction of the flow by the small size of the channel. The dimensionless number Nconf can be used to find the transition from the isolated to confined bubble regimes. To a first approximation, they showed that the confined boiling occurred when Nconf >0.5. The effects of the tube confinement were found to be significant for micro- and mini-channels. Examining the boiling of refrigerants in a small circular tube (D=2.46 mm) and a rectangular duct (Dh=2.40 mm) with nearly the same hydraulic diameters, Tran et al. [21] showed that there was no significant geometry effect for the two channels tested. Furthermore, their results implied that the nucleation mechanism dominated over the convection mechanism in small-channel evaporators over the full range of quality (0.2 ~ 0.8), which was contrary to the situations in larger channels where the convection 5.

(29) mechanism dominates at qualities typically above 0.2. Bao et al. [22] also found that the boiling heat transfer coefficient was a strong function of the heat flux and system pressure in a 1.95 mm diameter tube with R-11 and R-123, while the effects of the mass flux and vapor quality were very small, suggesting that the heat transfer was mainly through the nucleate boiling. Wambsganss et al. [23] studied boiling heat transfer of refrigerant R-113 in a small diameter (2.92 mm) tube and evaluated 10 different heat transfer correlations. They found that the high boiling number and slug flow pattern led to the domination by the nucleation mechanism and the two-phase correlations based on this dominance also predicted the data best. Moreover, Warrier et al. [24] used FC-84 in five parallel channels with each channel having a hydraulic diameter of 0.75 mm and compared their results with five widely used correlations. They then proposed two new correlations, one for subcooled flow boiling heat transfer and the other for saturated flow boiling heat transfer. Oh et al. [25] examined R-134a in capillary tubes of 500 mm long and 2, 1 and 0.75 mm inside diameters. They concluded that the heat transfer in the forced convective boiling was more influenced by the refrigerant mass flux than by the boiling number and the heat transfer coefficient was controlled by the Reynolds number. Besides, their results also showed that the dry-out point moved to the lower quality with decreasing size of the tubes. Vaporization of CO2 in 25 flowchannels of 0.8 mm ID was recently examined by Pettersen [26]. He also observed the two-phase flow pattern with another separate test rig with 0.98-mm heated glass tube. The results showed that nucleate boiling dominated at low/moderate vapor fractions, where the boiling heat transfer coefficient increased with the heat flux and refrigerant temperature but was less affected by the mass flux and vapour fraction. Moreover, the dryout effects became very important at higher mass flux and temperature, where the boiling heat transfer coefficient dropped rapidly at increasing x. And the two-phase flow was in intermittent and annular flow 6.

(30) regimes, the latter becoming more important at high mass flux. More complete information on the two phase flow boiling heat transfer in compact evaporators or microchannels is available from the recent critical review conducted by Ghiaasiaan and Abdel-khalik [27], Thome [28], Sobhan and Garimella [29], Kandlikar [30, 31] and Watel [32].. 1.2.3 Flow Boiling Heat Transfer in Conventional Tubes The experimental data taken from boiling of refrigerants R-22 and R-407C in a conventional tube (Di=6.5 mm) by Wang and Chiang [33] suggested that the major heat transfer mechanism at low mass flux (G= 100 kg/m2s) was the nucleate boiling. The convective evaporation becomes dominant as the mass flux is increased to 400 kg/m2s. A similar study from Shin et al. [34] showed that the boiling heat transfer coefficients depended strongly on the heat flux in a low quality region. Choi et al [35] examined boiling of several refrigerants in a horizontal tube (Di=7.75 mm) and found that the boiling heat transfer coefficients increased with the refrigerant temperature and mass flux. Yu et al. [36] examined the relations between the flow patterns and local heat transfer coefficient for the flow boiling of R-134a in a 10.7 mm diameter smooth tube and a micro-fin tube. The results showed that the flow pattern changed from wavy, intermittent, semi-annular to annular patterns at increasing mass flux and quality, and the heat transfer could be enhanced with the microfin surface up to 200% at low mass flux and high quality where the annular flow dominated in the microfin tube while the wavy flow occurred in the smooth tube. Yun et al. [37] investigated the evaporation heat transfer characteristics of CO2 in a horizontal tube with an inner diameter of 6.0 mm and found that the heat transfer coefficient decreased with an increase in the. 7.

(31) quality due to both a lower dryout quality and dominance of nucleate boiling as compared to the conventional refrigerants.. 1.2.4 Pressure Drop in Small Tubes The single phase pressure drop for flow in micro-channels was examined by Peng and Peterson [5]. Their results showed that the laminar friction factor or flow resistance reached a minimum value as the parameter Z approached 0.5. In the turbulent flow, the smaller the resistance, the higher the flow transition Reynolds number. Lazarek and Black [16] developed empirical correlations for frictional, spatial acceleration and bend pressure drops for saturated boiling of R-113 in a vertical U-tubes with Dh=3 mm. Yan and Lin [12, 13] found that the frictional pressure drop for R-134a evaporation increased with the refrigerant mass flux and imposed wall heat flux in a tube bank with Dh=2 mm for each tube. Similar results were obtained by Fujita et al. [15], Warrier et al. [24] and Tran et al. [38]. Recently Qu and Mudawar [39] investigated the pressure drop in multi-port parallel micro-channels (Dh=0.348 mm). They identified two types of two-phase hydrodynamic instability: severe pressure drop oscillation and mild parallel channel instability. They further noted that the pressure drop increased appreciably upon commencement of boiling in micro-channels.. 1.3 Review of Correlation Equations for Two Phase Flow Boiling Heat Transfer Based on the available experimental data from open literatures, some correlation equations for two phase flow boiling heat transfer and pressure drop were developed.. 8.

(32) An early general correlation model for the flow boiling in channels was proposed by Chen [40]. He divided the boiling heat transfer coefficient into two parts: a microconvective (nucleate boiling) contribution which could be estimated by the pool boiling correlations and a macroconvective (non-boiling force convection) contribution which could be estimated by the single-phase correlation such as the Dittus-Boelter equation [9]. A new correlation for boiling heat transfer in small diameter channels was proposed by Kew and Cornwell [19]. They introduced a confinement number in the correlation. Besides, Tran et al. [21] modified the heat transfer correlation of Lazarek and Black [16] with the Reynolds number of the flow replaced by the Weber number to eliminate viscous effects in favor of the surface tension. Similar correlations were proposed by Fujita et al. [41] and Xia et al. [42]. A flow boiling correlation based on an explicit nucleate boiling term rather than on an empirical boiling number dependence for vertical and horizontal flows in tubes and annuli was proposed by Gungor and Winterton [43]. They modified the Chen’s correlation to include the enhanced (E) and suppression factors (S). A new correlation from Liu and Winterton [44] introduced an asymptotic function to predict the heat transfer coefficient for vertical and horizontal flows in tubes and annuli. Moreover, Kandlikar [45] gave a general correlation for saturated flow boiling heat transfer inside horizontal and vertical tubes. The correlation was also based on a model utilizing the contributions due to nucleate boiling and convective mechanisms. It incorporated a fluid-dependent parameter in the nucleate boiling term. In a following study [46 ~ 48], he developed a new flow boiling map to depict the relationships among the heat transfer coefficient, vapor quality, heat flux and mass flux for subcooled and saturated flow boiling of different fluids inside circular tubes. Chen et al. [49] proposed an empirical correlation for two-phase frictional performance in small diameter tubes. The Chisholm correlation was noted to have a 9.

(33) poor predictive ability for boiling in small diameter tubes. However, the homogeneous model gave good predictions for the refrigerant and air-water data. Selected correlations from the literature for the boiling heat transfer coefficient and pressure drops are given in Table 1.2.. 1.4 Objective of This Study The above literature review clearly indicates that the experimental data for the evaporation heat transfer and pressure drop of HFC refrigerants in small tubes are still rare. In this study we intend to measure the evaporation heat transfer coefficient and pressure drop of refrigerants R-134a and R-407C in horizontal small tubes. The effects of the vapor quality, refrigerant mass flux, imposed heat flux and system pressure on the evaporation heat transfer and pressure drop in the small tubes will be examined in detail.. 10.

(34) Table 1.1. Comparison of Properties of three HFC refrigerants for air conditioning and refrigeration applications. Refrigerant. R-134a. R-410A. R-407C. Component. HFC-134a. HFC-32/125. HFC-32/125/134a. Wt %. 100 %. 50/50%. 23/25/52%. 1. near-azeotropic refrigerant. 2. the working pressure is five times than R-22. 3. the frication pressure drop is smaller.. 1. zeotropic refrigerant, and the components charge easy. 2. the working pressure is same with R-22.. Comparison with R-22. 1. the lower working pressure. 2. the frication pressure drop is larger in the same capability of freezing.. The energy efficiency ratio relative to R-22. 72~90. 94~100. 90~97. Molecule quality. 102.3. 72.6. 85.62. Remark. 1. the volume of operating system becomes larger. 2. the air-out volume of compress is larger.. Green-house effect (100 years). 1300. 1725. 1526. Toxicity limit (kg/m3). 0.25. 0.44. 0.31. Boiling point (℃). -26.2. -52.7 R32 (-51.8℃) / R125 (-48.5℃). -43.6 R32 (-51.8℃) / R125 (-48.5℃) / R134a (-26.2℃). Temperature glides. ─. ＜1 ℉. 10 ℉. 1. the design of system must to consider the strong and optimum elements.. 11. 1. the solutions of variation of R-407C components..

(35) Table 1.2 Classification of Two-phase heat transfer coefficient and friction factor correlations Reference W. Yu et al. [10]. Fluid. Heat Transfer Coefficient Correlations. Pressure Drop Correlations. －. ρ g 0.5 1 − x Re g 0.1 ) X tt = 18.65( ) ( x Rel 0.5 ρl. water. φlo 2 = X tt −1.9 B. Sumith et al.. water. [11]. Conditions. D = 2.98mm G = 50 ~ 200kg / m 2 s P = 2bar Din = 1.45mm. hz = 2.83 ⋅ (1/ X tt + 0.213)0.736 hLO. G = 23.4 ~ 152.7 kg / m 2 s. －. q = 10 ~ 715kW / m 2 xm = 0 ~ 0.8. Yan and Lin [12, 13]. R-134a. For X m ≤ 0.7. ftp = 0.127 Reeq. k 13 htp = 4.36 l Prl (1 − X m ) −0.5 (C1 ⋅ Reeq + C2 ) Di. Reeq =. ⋅ (C3 ⋅ Bo + C4 ) For X m > 0.7 htp = 4.36. kl 1 3 Prl (1 − X m ) −0.5 (C1 ⋅ Reeq + C2 ) Di. C1 = −0.0124G −0.368 C2 = 1.49G 0.514 C3 = −1166 X m + 1028 C4 = 0.53e0.931 X m. 12. GD. µl. D = 2mm. −0.1925. ((1 − x) + x(. ρl 0.5 ) ) ρg. Tsat = 5 ~ 31 oC G = 50 ~ 200kg / m 2 s x = 0.2 ~ 0.8.

(36) Table 1.2 Continued (1) Reference. Fluid. Y. Fujita et al. [15]. R-123. Heat Transfer Coefficient Correlations. Pressure Drop Correlations. htp = 0.884G 0.143 q 0.714. ∆Pf = ftp Retp =. Lazarek and Black [16]. R-113. htp = 30 Rel. 0.857. Bo 0.714 (kl / D). ∆Ptp ∆P. = 1+. ∆P = f. K. Cornwell and P.. R-113. L G2 D 2ρ. 1. ρ. GD. 1. µ. µ. 30 1 + 2 X tt X tt. L G2 (1 − x ) 2 D 2 ρl. =. +. ρg. +. 1− x. G : 50 − 400 kg m 2 s. D :1.12mm. ρl. q : 5 − 20 kW m2. 1− x. Re :135 − 1070. P :1.1 − 1.2bar. µl. f = 0.316[. x : −0.2 − 0.9. Bo : 3 × 10−4 − 8.9 ×10−4. G (1 − x) D. µl. −1. ]. 4. G :125 − 750 kg m 2 s Re : 860 − 5500. −4. Bo : 2.3 × 10 − 76 × 10. −4. P :1.3 − 4.1bar. DH = 1.03mm, G = 124 ~ 627 kg / m 2 s. Nu = C1 ⋅ Bo0.7 ⋅ Nulo. 1.64 mm, G = 211 ~ 317 kg / m 2 s. Confined Bubble Regime :. －. Nu = C2 ⋅ Bo ⋅ Co ⋅ Nulo m. µg. x. 1 − x 0.9 ρ g 0.5 µl 0.1 D : 3.1mm ] [ ] [ ] X tt = [ x ρl µg q :14 − 380 kW m 2. Isolated Bubble Regime :. A. Kew [19]. x. =. Conditions. n. L = 320 mm q = 4.3 ~ 20 kW / m 2. Annular − Slug Flow : Nu = C3 ⋅ F ⋅ Nulo T. N. Tran et al.. R-12,. [21]. R113. for Z. Y. Bao et al. [22]. R-11, R-123. D : 2.46mm, 2.92mm ; Dh = 2.4mm. ρ h = (8.4 × 10−5 )( Bo 2Wel )0.3 ( l ) −0.4 ρg. G : 44 − 832 kg m2 s. －. ∆T > 2.75 oC. 1.12( x /(1 − x)). Bo : 2 ×10−4 − 23 ×10−4. ∆Tsat : 2.8 − 18.2 oC. ( ρl ρ g ). G : 50 − 1800 kg m 2 s. D :1.95mm. htp hl = 1 + 3000 Bo0.86 + 0.75. Pr : 0.045 − 0.2. q : 7.5 − 129 kW m 2. 0.41. －. q : 50 − 200 kW m2 P : 2 − 5bar. x : −0.3 − 0.9. ∆Tsat : 5 − 15 C o. 13. Re : 860 − 5500.

(37) Table 1.2 Continued (2) Reference. Fluid. G. R. Warrier et al.. FC-84. [24]. Heat Transfer Coefficient Correlations. htp hl. = 1 + 6 Bo. 1 16. + f 2 ( Bo)( x). Pressure Drop Correlations. ∆Pf = ∆Pflo {1 + (4.3Γ 2 − 1). 0.65. R-134a. hTP 240 ⎛ 1 ⎞ = ⎜ ⎟ hLZ X tt ⎝ ReTP ⎠. D : 0.75mm. ⋅ [ N conf x 0.875 (1 − x)0.875 + x1.75 ]}. f 2 ( Bo) = −5.3 ⋅ [1 − 855Bo] H. K. Oh et al. [25]. Conditions. φlo 2 =. ∆Pf ∆Pflo. G : 557 − 1600 kg m 2 s. q : 0 − 59.9 kW m 2. Re : 418 − 2015. −1. α = [1 + 0.28 X tt 0.71 ]. DH = 2, 1 and 0.75mm. 0.6. L = 500mm －. G = 240 ~ 720 kg / m2 s q = 10 ~ 20 kW / m2 x = 0.1 ~ 1. C. C. Wang and C. S. Chiang [33]. R-22. f tp = −2.714 × 10−3 + 2.3 × 10−5 ⋅ Re0.5. R-407C. − 4.6 × 10−14 Re 2. －. Din = 6.5mm G = 100 and 400kg/m 2s G=100 ~ 700kg/m 2s q=2.5~ 10kW/m 2 x=0.11~0.93. T. N. Tran et al. [38]. ∆Pf = ∆PfLo {1 + (4.3Γ 2 − 1). R-134a R-12 R-113. ⋅ [ N conf x 0.875 (1 − x)0.875 + x1.75 ]} －. N conf = Γ2 =. 14. [σ g ( ρl − ρ g )]0.5 D. (dP dz ) fgo (dP dz ) flo. D = 2.46 and 2.92mm G = 33 ~ 832kg / m 2 s q = 2.2 ~ 129kW / m 2 xout = 0.02 ~ 0.95 Pr = 0.04 ~ 0.23.

(38) Table 1.2 Continued (3) Reference W. Qu and I.. Fluid. Heat Transfer Coefficient Correlations. water. Pressure Drop Correlations. ∆Ptp , f =. Mudawar [39]. φ f2 = 1 + －. Ltp xe ,out. ∫. xe ,out. 2 f f G 2 (1 − xe ) 2 v f. 0. dh. Conditions. φ f2 dxe. Dh = 0.348mm Tin = 30 ~ 60 oC G = 134.9 ~ 400.1kg / m 2 s. 1 C + 2 X tt X vv. Pout = 1.17bar. C = 21[1 − exp(−0.319 ×10 d h )] 3. ⋅ (0.00418G + 0.0613) ⎛ µf X vv = ⎜ ⎜ µg ⎝ Chen [40]. ⎞ ⎛ 1 − xe ⎟⎟ ⎜ ⎠ ⎝ xe. ⎞⎛ vf ⎟ ⎜⎜ ⎠ ⎝ vg. ⎞ ⎟⎟ ⎠. 0.5. htp = hconv + hnb ⇔ htp = Fhl + Shpool hpool = 0.00122 ⋅ ((kl (σ 0.5 µl. 0.29. i fg. 0.24. 0.79. c p ,l. 0.45. ρl 0.49 ). ρ g 0.24 )) ⋅ ∆Tsat 0.24 ∆Psat 0.75 ;. －. F = 2.35(1 X tt + 0.213)0.736 , forX tt > 0.1. －. 1.14 −1. S = (1 + 0.12 Retp. ) , for Retp < 32.5. Retp = G (1 − x) D / µl ( F 1.25 )(10−4 ) K. E. Gungor et al. [43]. D = 2.95 ~ 32mm. htp = hl ⋅ E + hpool ⋅ S. G = 12.4 ~ 61518kg / m 2 s. hl = 0.023 ⋅ Rel0.8 ⋅ Prl0.4 (kl / D) Pr −0.55 ) hpool = 55 ⋅ Pr 0.12 ⋅ (− log10 ⋅ M −0.5 ⋅ q 0.67. －. E = 1 + 24000 ⋅ Bo1.16 + 1.34 ⋅ (1/ X tt )0.86 ) −1 S = (1 + 1.15 ×10−6 ⋅ E 2 ⋅ Re1.17 l 15. q = 0.35 ~ 91534kW / m 2 x = 0 ~ 1.46.

(39) Table 1.2 Continued (4) Reference. Fluid. Heat Transfer Coefficient Correlations. T. Y. Choi et al.. R-32. For pure refrigerants. Din = 7.75mm. R-134a. htp = E ⋅ hDB + S ⋅ hSA. Tsat = −12 ~ 17 oC. For mixture refrigerants htp = E ⋅ hDB + FM S ⋅ hSA. G = 240 ~ 1060kg / m 2 s. [35]. R-32/134a R-407C. hDB = 0.023Rel. 0.8. Prl. Conditions. q′′ = 4.1 ~ 28.6kW / m 2. 0.4. k ⎛ q′′bd ⎞ hSA = 207 l ⎜ ⎟ bd ⎝ klTs ⎠ FM =. Pressure Drop Correlations. 0.674. ⎛ ρv ⎞ ⎜ ⎟ ⎝ ρl ⎠. 0.581. Prl. 0.533. 1. －. 1 + cM {[C pl (Tdew − Tbub )] / h fg }. bd = 0.0146 β [2σ / g ( ρl − ρv )] with β =35o E = c1 Boc2 X tt. c3. S = c4Coc5 c1 = 49.971 c4 = 0.909 Z. Liu et al. [44]. c2 = 0.383. c3 = −0.758. c5 = 0.039. 2. htp = ( Fhl ) 2 + ( Shpool ) 2 hl = 0.023(kl D ) Rel hpool = 55Pr. 0.12. 0.8. D : 2.95 − 32mm. Prl. q : 0.35 − 2620 kW m 2 Re : 568.9 − 87500 Pr : 0.0023 − 0.895 x : 0 − 0.948. 0.4. (− ln Pr ) −0.55 M −0.5 q 0.67. F = (1 + x Prl ( ρl ρ g − 1)) S = (1 + 0.55F 0.1 Rel. G :12.4 − 8179.3 kg m 2 s. －. 0.35. 0.16 −1. ). 16. Fr : 2.66 ×10−4 − 2240. Prl = 0.83 − 9.1.

(40) CHAPTER 2 EXPERIMENTAL APPARATUS AND PROCEDURES. The experimental system which is modified slightly from that used in the previous study [12] is employed here to investigate the evaporation heat transfer and pressure drop of the HFC refrigerants in small tubes. It is schematically depicted in Fig. 2.1. The modified test section along with the entry and exit sections are schematically shown in Fig 2.2. The experimental apparatus consists of three main loops, namely, a refrigerant loop, a water-glycol loop, and a hot-water loop.. 2.1 Refrigerant Loop The refrigerant loop consists of an oil-free variable–speed refrigerant pump, an accumulator (low pressure), a receiver (high pressure), a mass flow meter, the test section and the inlet and exit sections, a condenser, a sub-cooler, a filter/dryer and four sight glasses. The refrigerant mass flow rate is mainly controlled by an AC motor through the change of its inverter frequency. In addition, the refrigerant flow rate can also be further adjusted by regulating the bypass valve in the flow path from the refrigerant pump. The refrigerant R-134a or R-407C at the outlet of the refrigerant pump must be kept subcooled to avoid any vapor flow through the mass flow meter, which has an estimated measurement error of ± 1%. The preheater is used to heat the subcooled refrigerant R-134a or R-407C to a specified vapor quality at the test section inlet by receiving heat from the hot water in the hot-water loop. Then, a vapor-liquid refrigerant mixture is generated and moves through the small tubes in the test section. The vapor flow leaving the test section is re-liquefied in an oversized 17.

(41) condenser/subcooler in the cold water-glycol loop. After leaving the subcooler, the liquid R-134a or R-407C flows back to the receiver at the bottom of the system. An accumulator is connected to a high-pressure nitrogen tank to dampen the fluctuations of the refrigerant flow rate and pressure. The filter/dryer is used to filter the impurities and noncondensable gas possibly existing in the loop. Varying the temperature and flow rate of the water-glycol mixture flowing through the condenser and subcooler allows us to control the pressure of the refrigerant loop. Two absolute pressure transducers are installed at the inlet and exit of the test section with a resolution up to ± 2kPa. All the refrigerant and water temperatures are measured by copper-constantan thermocouples (T-type) with a calibrated accuracy of ± 0.2℃.. 2.2 Test Section Due to the tubes to be tested being relatively small, the refrigerant flow rates in them are very low and direct measurement of boiling heat transfer coefficient and pressure drop in the tubes is difficult and can be subject to large error. Thus 28 small tubes all made of copper, each having the same diameter and length, are put together to form the test section, as schematically shown in Fig. 2.2. Each small tube has the same diameter of 0.83 or 2.0 mm, outside diameter of 1.83 or 3.0 mm and length of 150 mm. Specifically, these 28 tubes are placed together side by side forming a plane tube bundle. In order to allow the refrigerant to flow smoothly into the small tubes, a section including divergent, convergent and straight portions is connected to the inlets of the tubes. Besides, another section including straight and convergent parts is attached to the exits of the tubes. Both the inlet and exit sections are formed by the stainless steel plates. At the middle axial location of the small tubes 14 thermocouples. 18.

(42) are soldered onto the outer surface of the tubes. Specifically, these thermocouples are soldered onto 14 selected tubes at the circumferential position of 45° from the top of the tube or from the bottom of the tube, as shown in Fig. 2.3. Two copper plates of 5-mm thick are respectively soldered on the upper and lower sides of the tube bundle shown in Fig. 2.4. Obviously small crevices exist between the tube outside surface and copper plates. Due to the good thermal contact of the copper plates and the tubes, there is no need to fill the crevices with conducting grease. Instead, the crevices provide the space for the thermocouple wires leading to the data logger. The copper plates are heated directly by an electric-resistance heater of 2.6-mm wide, 0.5-mm thick and 2.5-m long. The heater is connected to a 500W DC power supply. Mica sheet is placed in the narrow space between the heater and copper plates to prevent leak of electric current to the copper plates. The power input to the heater is measured by a power meter with an accuracy of ±0.5﹪. In order to reduce the heat loss from the heaters, the whole test section is wrapped with a 10-cm thick polyethylene layer.. 2.3 Water Loop for Pre-heater In order to maintain the refrigerant R-134a or R-407C at the preset temperature at the test section inlet, a water loop is used to preheat the refrigerant before it arrives at the test section inlet. The water loop for the preheater includes a double-pipe heat exchanger having a heat transfer area of 0.12 m2, a 125-liter hot water container with three 2.0 kW heaters in it, and a 0.5 hp water pump which can drive the hot water at a specified flow rate to the preheater. In the preheater the hot water passes through the outer pipe while the liquid refrigerant flows in the inner pipe. The water flow rate is controlled by an AC motor through the change of its inverter frequency and by the. 19.

(43) by-pass valve. The connecting pipe between the preheater and test section is thermally insulated with a 5-cm think polyethylene layer to reduce the heat loss from the pipe.. 2.4 Water-Glycol Loop The water-glycol loop is designed for condensing the R-134a or R-407C vapor. The water-glycol loop is cooled by a water cooled R-22 refrigeration system. The cooling capacity is 3.5 kW for the water-glycol mixture at -20℃. The cold water-glycol mixture at a specified flow rate is driven by a 0.5 hp pump to the condenser as well as to the subcooler. A by-pass loop is provided to adjust the flow rate. By adjusting the mixture temperature and flow rate, the bulk temperature of the R-134a or R-407C in the subcooler can be controlled at a preset level.. 2.5 DC Power Supply As described above, the small tubes in the test section are heated by a resistance heater. A 50V-25A DC power supply delivers the required electric current to the heater. A Yokogawa DC meter is used to measure the DC current through the heater with an accuracy of ±1%. Then the voltage drop across the heater is measured by a Yokogawa multimeter. Thus the power input to the heater can be calculated.. 2.6 Data Acquisition The data acquisition system includes a recorder, a 24V-3A power supply and a controller. The recorder is used to record the temperature and voltage data. The water flowmeter and differential pressure transducer need the power supply as a driver to output an electric current of 4 to 20 mA. The IEEE488 interface is used to connect the controller and the recorder, allowing the measured data to be transmitted from the. 20.

(44) recorder to the controller and the data are analyzed by the computer immediately. The system automatically monitors 22 copper-constantan thermocouples, two pressure transducers, and three mass flowmeters. The thermodynamic and transport properties of the refrigerants R-134a and R-407C are obtained by the manual from AlliedSignal Co. Ltd. and by the ASHRAE handbook.. 2.7 Experimental Procedures Before a test is started, the system temperature is compared with the saturation temperature of refrigerant R-134a or R-407C corresponding to the measured saturation pressure of the refrigerant and the allowable difference is kept in the range of 0.2-0.3K. Otherwise, the system is re-evacuated and then re-charged to remove some noncondensible gases possibly existing in the refrigerant loop. In each test the liquid refrigerant leaving the subcooler is first maintained at a specified temperature by adjusting the water-glycol temperature and flow rate. In addition, we adjust the thermostat in the water loop to stabilize the refrigerant temperature at the test section inlet. Next, the temperature and flow rate of the hot water loop for the preheater are adjusted to keep the vapor quality of R-134a or R-407C at the test section inlet at the desired value. Then, we regulate the refrigerant pressure at the test section inlet by adjusting the gate valve locating right after the exit of the test section. Meanwhile, by changing the current of the DC motor connecting to the refrigerant pump, the refrigerant flow rate can be varied. The imposed heat flux from the heater to the refrigerant is adjusted by varying the electric current delivered from the DC power supply. By measuring the current delivered to and voltage drop across the heater, we can calculate the heat transfer rate to the refrigerant. All tests are run when the experimental system has reached statistically steady state. Finally, all the data. 21.

(45) channels are scanned every 5 seconds for a period of 50 seconds. The thermophysical properties of the refrigerants R-134a and R-407C are given in Table 2.1.. 22.

(46) Table 2.1. Comparison of thermophysical properties of R-134a and R-407C. Thermal properties. R-407C. R-134a. Temperature (℃) (Saturated Pressure). 4.895 (665 kPa). 10.04 (776 kPa). 15.05 (900 kPa). 5 (349.8 kPa). 10 (414.6 kPa). 15 (488.6 kPa). Liquid µl - µg Vapor Liquid Kinematic Viscosity vl - vg (µm2/s) Vapor Liquid Density ρl - ρg (kg/m3) Vapor Liquid Enthalpy Latent heat of (kJ/kg) vaporization Vapor Liquid Conductivity (W/m·K) Vapor. 196.9 184.96 11.94 0.1616 0.1554 0.4219 1218 1189.7 28.3 206.9. 184.6 172.4 12.2 0.1539 0.1478 0.3689 1199 1165.93 33.07 214.3. 173.4 160.95 12.45 0.147 0.1411 0.3236 1179 1140.53 38.47 221.7. 254.4 243.46 10.94 0.199 0.193 0.6382 1278 1260.86 17.14 206.75. 238.8 227.65 11.15 0.1893 0.1834 0.5511 1261 1240.77 20.23 213.6. 224.3 212.94 11.36 0.1803 0.1745 0.478 1243.5 1219.735 23.765 220.5. 207.1. 201.7. 196.3. 194.75. 190.7. 186.55. 414 0.0983. 416 0.09588. 418 0.09585. 401.5 0.0898. 404.3 0.0876. 407.05 0.08545. 0.0122. 0.01264. 0.013113. 0.01195. 0.0124. 0.01286. 0.0099. 0.00914. 0.00839. 0.010705. 0.01014. 0.00944. Viscosity (µN·s/m2). Surface Tension (N/m). Liquid. 23.

(47) REFRIGERANT LOOP. T THERMOCOUPLE. PRESSURE TAP. WATER-GLYCOL LOOP. TEST SECTION. DIFFERENTIAL. DP PRESSURE. P T. T P. DEGASED VALVE. SIGHT GLASS VALVE FLOW DIRECTION. COOLING TOWER. +. -. PUMP DC Power Supper 50V - 30A. CONDENSER. WATER LOOP. T T. PREHEATER. N2. BY - PASS FLOW. T. WATER THERMOSTAT. PUMP. SUBCOOLER. RECEIVER. T. RELEASE VALVE. BY-PASS FLOW. FILTER/DRYER. WATER FLOWMETER. ACCUMULATOR. P. WATER-GLYCOL THERMOSTAT PUMP. MASS FLOWMETER. Fig. 2.1 Schematic layout24of the experimental system. 24.

(48) FLOW. Test Section. Inlet Section. Exit Section. Units：mm. Fig. 2.2 Schematic diagram of test section along with the inlet and exit sections. 25.

(49) R E T A E H. A. ) a (. m m 0 5 1 R E T A E H T E E H S A C I M E L P U O C O M R E H T S E B U T mm mm 33 88 . . 1 0 == o i DD. 26. m m. m m 0 0 1 W E I V L A N O I T C E S S S O R C A A. Fig. 2.3 (a) Test section of the small tubes and (b) locations of the thermocouples.. = o D. mm mm 15. A. ) b ( Di=0.83 or 2.0 mm Do=1.83 or 3.0 mm. Do.

(50) 28 x SMALL TUBES. HEATER PLATE. COPPER PLATE. Fig. 2.4 The detail of test section. 27.

(51) CHAPTER 3 DATA REDUCTION. In the present measurement of the evaporation heat transfer coefficient and pressure drop of the refrigerant flow in the horizontal small tubes, a data reduction analysis is needed to calculate the heat transfer coefficient and friction factor from the measured raw data. The data reduction process is described in the following.. 3.1 Single Phase Heat Transfer Before the two-phase experiments, the total heat loss from the test section is evaluated by comparing the total power input from the power supply Qt with that calculated from the energy balance in the single phase refrigerant flow, which can be expressed as Q s = G ⋅ A cs ⋅ C p ,r ⋅ ∆Ti ,o. (3.1). where A cs is the total cross-sectional area of the small tubes and ∆Ti,o is the difference in the refrigerant temperature at the exit and inlet of the test section. The total power input is computed from the product of the measured voltage drop across the resistance heater and the electric current passing through it, Q t =VI . The imposed heat flux at the inside surface of the small tubes is then evaluated from the relation q=. Qt As. (3.2). where V and I respectively represent the measured voltage drop and current, and A s denotes the total inside surface area of the small tubes. The relative heat loss from the test section is defined as. ε = (Q t -Qs )/Q t. (3.3). 28.

(52) The measured results indicated that for all runs in the energy balance test ε was within 2%. The average single phase convection heat transfer coefficient in the small tubes is defined as hl =. Qn A s ⋅ (Twall -Tr,ave ). (3.4). Here Q n = Q t − ε ⋅ Q t = Qs is the net power input to the liquid refrigerant R134a and R-407C refrigerants, Twall is the average of the measured tube wall temperature at all detected locations, and Tr,ave is the average refrigerant temperature in the tubes which in turn is estimated from the measured refrigerant temperatures at the inlet and exit of the test section as (Tr,i+Tr,o)/2.. 3.2 Two Phase Heat Transfer The vapor quality of the R-134a and R-407C refrigerants entering the test section inlet is evaluated from the energy balance for the preheater. Based on the temperature drop on the water side the heat transfer in the preheater is calculated from the relation,. Q w,p = Ww,p c p,w (Tw,p,i -Tw,p,o ). (3.5). While the heat transfer to the refrigerant in the preheater is the summation of the sensible heat transfer (for the temperature rise of the refrigerant to the saturated value) and latent heat transfer (for the evaporation of the refrigerant),. Q w,p = Q sens +Q lat. (3.6). here. Qsens = Wr c p,r (Tr,sat -Tr,p,i ). (3.7). Qlat = Wr ⋅ i fg ⋅ xp,o. (3.8). The above equations can be combined to evaluate the refrigerant quality at the exit of the preheater that is considered to be the same as the vapor quality of the refrigerant 29.

(53) entering the test section. Specifically, xin = xp,o =. 1 i fg. ⎡ Q w,p ⎤ -c p,r (Tr,sat -Tr,p,i ) ⎥ ⎢ ⎣ Wr ⎦. (3.9). The total change of the refrigerant vapor quality in the test section is then deduced from the net heat transfer rate from the heater to the refrigerant in the test section, Q n . Thus ∆x =. Qn Wr i fg. (3.10). Finally, the average heat transfer coefficient for the evaporation of R-134a or R-407C refrigerants in the test section is determined from the definition hr ≡. Qn A s (Twall -Tr,sat ). (3.11). 3.3 Friction Factor Note that in the evaporation of R-134a or R-407C refrigerant in the tubes the flow accelerates, causing the pressure to drop, as it moves downstream. Besides, the refrigerant pressure also drops due to the contraction at the test section inlet and rises due to the expansion at the exits of the small tubes. Thus, in the refrigerant flow the two phase frictional pressure drop ∆Pf associated with the refrigerant evaporation in the small tubes is calculated by subtracting the pressure drop due to flow acceleration ∆Pa and the pressure drop at the test section inlet ∆Pi and by adding the pressure rise at the test section exit ∆Po from the measured total pressure drop ∆Pexp . The frictional pressure drop is hence given as ∆Pf = ∆Pexp -∆Pa -∆Pi +∆Po. (3.12). Note that the acceleration pressure drop is estimated by the homogeneous model for two phase flow [50] as 30.

(54) ∆Pa = G 2 ν fg ∆x. (3.13). Moreover, the pressure drops associated with the sudden contraction and the pressure rises associated with the sudden expansion ∆Po and ∆Pi for two phase flow moving through the inlet and exit ports estimated by Collier [50] based on a separated flow model are chosen here. They can be expressed as 3. 2. 2. 2. (1+Cc )[xin ν g /α 2 +(1-xin )3 ν l /(1-α) 2 ] x ν (1-xin ) 2 ν g G ∆Pi =( ) 2 (1-Cc ){ -Cc [ in g + ]} Cc 2[xin ν g +(1-xin )ν l ] α (1-α) (3.14) and ∆Po =G 2 σ(1-σ)ν l [. 2 (1-xin ) 2 ν g xin +( ) ] (1-α) νl α. (3.15). where Cc in equation (3.14) is the coefficient of contraction and it is a function of the contraction ratio σ . The void fraction α in the above equations is calculated from the correlation given by Zivi [51] as α=. 1. (3.16). ρ 2 1-x 1+( in )( g ) 3 xin ρl. Finally, for the evaporation of R-134a or R-407C in the small tubes the two phase friction factor is expressed as. f tp =. ∆Pf Di 2G 2 ν m L. (3.17). where L is the length of the small tubes and ν m is the mean specific volume of the vapor-liquid mixture in the small tubes when they are homogeneously mixed and can be expressed as ν m =[xm ν g +(1-xm )ν l ]=(ν l +xm ν fg ). (3.18). 3.4 Uncertainty Analysis 31.

(55) As the tubes are miniaturized, it becomes harder to maintain the geometric precision of the tube cross section and to insure the smooth transition from the inlet section to the heat transfer section [52]. Uncertainty of a similar nature arises from the micro-roughness of the heat transfer surface and the presence of dust particles in the flow. Besides, the heat conduction from the test heater to the unheated part of the tubes increases as the test section is miniaturized. These factors result in higher uncertainties of the measured data. Uncertainties of the heat transfer coefficients are estimated according to the procedures proposed by Kline et al. [53]. The detailed results from this uncertainty analysis are summarized in Table 3.1.. 32.