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Heat Rejection in Condensers: Desuperheating, Condensation in Superheated Region and Two Phase Zone

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Heat Rejection in Condensers: Desuperheating, Condensation in Superheated Region and Two Phase Zone View metadata, citation and similar papers at core.ac.uk brought to you by CORE provided by Purdue E-Pubs 2503 , Page 1 Heat Rejection in Condensers: Desuperheating, Condensation in Superheated Region and Two Phase Zone Chieko KONDOU 1,2 , Pega HRNJAK 1,3 *, 1University of Illinois, Department of Mechanical Science and Engineering ACRC, Urbana, IL, USA also 3Creative Thermal Solutions, Inc. Urbana, IL, USA E-mail; [email protected] 2Kyushu University, Interdisciplinary Graduate School of Engineering Science, Fukuoka, Japan E-mail; [email protected] ABSTRACT Conventional modeling of condensers always assumes three zones: desuperheating, condensation and subcooling even condensation occurs in desuperheating zone at some conditions and that subcooling occurs during condensation. This paper discusses the actual situation and provides experimental validation of the hypothesis. The -2 -1 experimental results show heat transfer coefficients of CO 2 and R410A at mass fluxes 100 to 240 kg m s and reduced pressures from 68% to 100% in a horizontal smooth tube of 6.1 mm inner diameter. Data are compared to correlations proposed for other working fluids or other conditions. Results show much higher values of HTC than correlation proposed for single-phase turbulent flow in superheated zone. The occurrence of condensation in superheat zone is evident when tube wall temperature is below saturation temperature. The results suggest that simplified calculations of heat rejection in superheated zone significantly oversize the condenser. The semi- empirical correlation, which is here proposed as the combination of existing correlations for single-phase turbulent and saturated condensation, satisfactorily predicts the heat transfer coefficient of the superheat zone condensation. 1. INTRODUCTION Balekjian and Katz [1] experimentally investigated condensation from superheated vapors of R114 and steam on outer surface of a horizontal tube. Their experimental HTC suggested the lowering cooling surface temperature, below saturation point, generates condensate from superheated vapor. Altman et al. [2] provided six points of experimentally determined HTC, which are averaged from various superheated inlet to saturated outlet in a test section of 1.22 m length 8.71 mm ID. With those six data having various superheat degree at test section inlet, approximately 30 to 70 % decrease in HTC from a correlation valid for two-phase zone condensation was confirmed. Bell [3] accepted the criterion of condensation occurrence in desuperheating zone is the tube wall temperature below saturation point and cautioned that the simple use of LMTD method to calculate the overall coefficient of condensers could be invalid. Miropolskiy et al. [4] provided experimental data for the quasi-local HTC of superheated steam vapor flowing downward in a cooled vertical smooth tube. The results experimentally presented the criterion of start of condensation which is when the tube wall temperature is below saturation point. Further, their data showed the behavior of superheat zone condensation for various reduced pressure until 0.82. Fujii et al. [5] experimentally investigated condensation of R11 and R113 flow in horizontal smooth tubes. From the temperature distribution in the radial direction of the horizontal middle plane of the tube, they proved coexistence of superheated vapor and subcooled liquid in condensation flow. They varied the vapor mass quality, which indicates actual vapor and liquid mass flow rate in non-equilibrium, to analyze the mass transport process. Lee et al. [6] experimentally investigated condensation in superheated R22 vapor and proposed a physical model accounting the sensible heat on the condensation heat transfer. Their model defines HTC bulk temperature as the reference temperature. In response, Webb [7] reported that sensible heat is negligible and simplified model defines HTC with saturation temperature gives same results to Lee’s model. Various systems and application drove operating conditions closer to the critical point, very often rejected heat in superheat zone is greater than in two-phase zone. Conditions just below the critical point are increasingly common for both R410A water heaters and CO2 commercial refrigeration systems in most frequent operations from autumn to spring. Kondou and Hrnjak [8] presented experimental results for only CO 2 heat rejection (cooling/condensation) International Refrigeration and Air Conditioning Conference at Purdue, July 16-19, 2012 2503 , Page 2 flow in 6.1 mm ID smooth tube at the pressure from 5.0 to 7.5MPa. The experimental results identified the condensation in superheated vapor. In order to strengthen experimental verification one additional refrigerant R410A was explored. It was chosen because of its importance in applications and the effect of the refrigerant properties on condensation heat transfer. 2. CONDENSATION IN SUPERHEAT ZONE Figure 1 illustrates the heat flow and the temperature profile in condensing superheat zone, where bulk mean refrigerant temperature is above saturation point. According to Soliman’s flow regime [9], condensation begins as mist flow and then changes into annular flow. According to Altman [2], thin ridges or droplets flow on the interior tube surface. Figure 1 explains the heat exchange with annular flow model for simplification. From continuity, the total mass flow rate of vapor and liquid refrigerant is, mɺ total mɺtotal= m ɺ V, i + m ɺ L, i = m ɺ V, o + m ɺ L, o (1) The amount of condensate ∆mɺ L generated through a segment is expressed from the continuity as, ∆mmmɺɺɺɺɺL= L,o − L,i = mm V,i − V,o = ∆ m ɺ V (2) The average enthalpies in superheated vapor and subcooled liquid are represented with heat capacities and degree of superheat and subcool ∆TSH ∆TSC as, hhV =+Vsat CpT V∆ SH =+ h Vsat ∆ h SH, hhL =− Lsat CpTh L ∆ SC =−− Vsat ∆∆ h LV h SC (3) Total inlet heat at the entrance of a segment is, ɺɺɺ ɺ ɺ hmb, i total= hmhm V, i V, i + L, i L, i =+() h Vsat∆ hm SH,i V,i +−−() h Vsat ∆ h LV ∆ hm SC,i L,i (4) Similarly, the total outlet heat at the exit of a segment is, ɺɺɺ ɺɺ ɺɺ hmhmhmb, o total= V, o V, o + L, o L, o =+() h Vsat∆ hm SH,o() V,i −+−− ∆ mh L() Vsat ∆∆ h LV hm SC,o() L,i + ∆ m L (5) The subtraction from Eq. (5) to Eq. (4) gives the total heat exchange through a segment. ɺ ɺɺɺ ɺɺ ()hhmb, i−=− b, o total()∆∆ h SH,i hm SH,o V,i + ∆∆∆∆∆∆ hmhmh SH,o L+ LV +−+() SC,o hm SC,i L,i ∆∆ hm SC,o L (6) latent SH SC In the right of Eq. (6), the first, second, and third terms shows heat transfer rate caused by de-superheating of vapor flow, latent heat rejection to generate condensate, and subcooling of condensate. ɺ ɺɺ ɺ Qtotal= Q SH + Q latent + Q SC (7) qtotal= q SH +() q latent + q SC (8) The driving temperature difference for the heat flux caused by de-superheating qSH is probably the difference between bulk temperature to saturation temperature on the liquid film surface ( Trb - Tsat ). The remaining summation III: TP zone II: condensing SH zone I: single-phase SH zone xb = 1, Twi < Tsat ≈ Trb Twi < Tsat < Trb Twi = Tsat < Trb Outlet Inlet Tube wall Condensate ∆mL h hV, o V, i mV, o mV, i qSH T ∆TSH T ≈T rb Trb rb sat qlatent Tsat hL, o hL, i Tsat T =T qSC wi sat mL, o m L, i ∆T SC Twi Twi qtotal Twi Figure 1: Illustration of heat flow and temperature distribution in the condensing superheat zone International Refrigeration and Air Conditioning Conference at Purdue, July 16-19, 2012 2503 , Page 3 of heat fluxes ( qlatent + qSC ) is manipulated as a heat flux by saturated condensation. Because condensation requires degree of subcooling of cooling surface, the HTC of saturated condensation αTP always includes both heat fluxes. That driving temperature difference is normally taken as a difference from saturation temperature to wall temperature ( Tsat - Twi ). Thus, Eq. (9) is converted with those HTCs driving temperature differences to, α()()()TTrb−= wi α SH TT rb −+ sat α TP TT sat − wi (9) 3. EXPERIMENTAL SETUP AND METHOD ΔP Mixer 3.1 Test Section and Test Tube Test section TTT Pre-cooler PPP Figure 2 shows the schematic diagram of experimental TTT apparatus. The loop for CO 2 and R410A consists TTT TTT Pre-heater TTT TTT mainly of a variable speed gear pump, a coriolis-type TTT mass flow meter, an electric pre-heater, a mixer, a pre- TTT Mass cooler, a test section, two after-coolers, and receiver flow tank. In a mixer, placed at entrance of the pre-heater, meter Charge pressure and bulk-mean temperature of superheated port CO 2 or R410A Water vapor are measured. System pressure is adjusted Chilled water (7 oC) roughly by refrigerant charge amount and precisely by TTT Thermocouple inlet temperature and flow rate of cooling water flow ∆∆∆ PPP PPP Pressure transducers through the after-coolers. After-cooler Receiver tank Gear pump Figure 3 (a) and (b) show a structure of the test section and dimensions of the test tube. The test tube, which is a smooth copper tube of 6.1 mm ID and 9.53 mm OD, Figure 2: Schematic diagram of the experimental apparatus is placed horizontally and covered with a thick brass jacket. The halved brass jacket is pressed over the test tube and the small gap between them is filled with a thermal paste. On the outside of the brass jacket, copper tubes are attached with solder allowing cooling water to flow through. This structure yields cooling conditions with an almost uniform temperature. Twelve thermocouples are embedded into the top, bottom, right, and left of the test tube wall at three positions in an axial direction. The active cooling length by the brass jacket is 150 mm, which is relatively short for measuring quasi-local HTC in axial direction.
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