Thermal Design of Condenser Using Ecofriendly Refrigerants R404A-R508B for Cascade Refrigeration System A D Parekh, and P R Tailor Abstract Because of damaging effect of CFC refrigerants on stratospheric ozone layer, international agreement (Montreal Protocal) has been signed by many countries to curtailing the use of CFCs on a global scale. Here an attempt is made to design condenser for cascade refrigeration using ecofriendly refrigerant pair R404A- R508B. These refrigerants have zero ozone depletion potential and minimum global warming potential. The design results have been obtained by developing simulation program in Engineering Equation Solver (EES). The heat transfer results of two geometries have been compared at various operating conditions of cascade system. It is revealed from the results that the required heat transfer area for tube in tube type condenser is less than shell and coil type under same operating conditions. The required heat transfer area of the both types of condensers decreases as condensing temperature and evaporating temperature increases. Keywords cascade refrigeration system, condenser, design, heat transfer area, shell and coil, tube in tube I. INTRODUCTION N low temperature refrigeration application such as storage Iof frozen food, liquefaction of petroleum vapour, manufacturing of dry ice etc. where the required evaporating temperature between -40 0 C and -90 0 C, the cascade system is often a convenient option, in which two suitable refrigerants are chosen. As per Montreal protocols chlorofluorocarbons (CFCs) refrigerants are completely phased out due to its effect on stratospheric ozone layer. Thus, alternative refrigerants must be found to replace the CFC refrigerants. But, replacement of the alternate refrigerant may affect the performance of the system. So, in order to improve the performance, there is a need to re-design each component of the cascade refrigeration system. Also, refrigerant vapor discharged from a compressor in refrigeration system is generally cooled and condensed in a condenser via heat transfer to a secondary fluid. If the condenser does not dissipate the heat well, the discharge pressure would build up resulting in an increase in compressor power. So, the required size of condenser is large to increase the condenser effectiveness. This becomes impractical from the viewpoint of system maintenance and initial cost since A.D Parekh, Mechanical Engineering Department, S.V National Institute of Technology, Surat, Gujarat, India, e-mail : adp@med.svnit.ac.in. P.R Tailor, Mechanical Engineering Department, S.V National Institute of Technology, Surat, Gujarat, India, e-mail : prt@med.svnit.ac.in. more refrigerant is to be charged. Therefore, the best way to design effective condensers would be to keep the size as small as possible with special heat transfer enhancement mechanisms. For this it is important to know the flow condensation heat transfer coefficients (HTCs) of refrigerants. To investigate the overall performance of the condenser, heat transfer behaviors for refrigerant side as well as water side must be considered simultaneously. Several investigators have performed studies on heat transfer during condensation. Schlager et al. [1] investigated evaporation and condensation heat transfer and pressure drop characteristics in three horizontal 12.7 mm microfin tubes with R22. They found evaporation and condensation heat transfer coefficients in the microfin tubes were 1.6 2.2 and 1.5 2.0 times, respectively, larger than those in the smooth tube. Chamra et al. [2] presented R22 condensation heat transfer and pressure drop data for new microfin geometries applied to the inner surface of 15.88 mm O.D. tubes. They reported that the condensation heat transfer performance for cross grooved tubes was higher than those for single-grooved tubes. They found also that a cross-grooved tube with the helix angle of 278 had the highest heat transfer performance. Dobson and Chato [3] performed an experimental study on the condensation of R12, R22, R134a, R410A and R32/125 (60/40%) in horizontal smooth tubes. They presented two correlations, which can predict the heat transfer coefficient in the stratifying flow regime and in the annular flow regime, respectively. Rennie and Raghavan [4] performed an experimental study of a double-pipe helical heat exchanger using two differently sized heat exchangers. Both parallel flow and counter flow configurations were investigated. Hot water and cold water were used as working fluids. Overall heat transfer coefficients were determined and heat transfer coefficients in the inner tube and the annulus were calculated using Wilson plots. Nusselt number obtained for the inner tube and the annulus were compared to the data published in literature. Kwon et al. [5] measured flow condensation HTCs of R22 and R410A in plain and microfin tubes at 31 0 C in the mass flux range of 97 202 kgm -2 s -1 and showed that HTCs of R410A are slightly higher than those of R22 for a plain tube while those of R410A are almost same as those of R22 for a microfin tube. In the present study, two geometries of condenser are studied viz. shell and coil and tube in tube type; in which 32
refrigerant flows inside tube and water over tube. The objectives of the present study are to: (i) To find the heat transfer coefficients for both geometries of condensers under study, (ii) compare the thermal performances of two geometries using refrigerant pair viz. R404A/R508B and (iii) To investigate the influence of the various operating conditions on the heat transfer area of condensers. subcooling (DSC) of 4 0 C in both cycles, isentropic efficiency of compressor of 80% II. CASCADE REFRIGERATION SYSTEM A simplified sketch of cascade refrigeration system showing the main components is shown in Fig. 1. The refrigeration system comprises two separate refrigeration cycles the high temperature cycle (HTC) and low temperature cycle (LTC). Both cycles are thermally connected with each other through cascade condenser which acts as evaporator for HTC and condenser for LTC. Water acts as a secondary fluid for both types of condensers considered in the study. Fig. 2 shows the actual T-s diagrams for cascade system for R404A/R508B. Fig.2 Actual T-S plot of R404A-R508B cascade system for both cycles.resistance due to wall and fouling factor is neglected for the evaluation of heat transfer coefficient. The refrigerant mass flow rate through the low temperature cycle is given by the following equation. The energy balance across the cascade condenser allows the evaluation of the mass flow rate of refrigerant in HTC. It is specified through the following equation. (1) Where, ε cc is cascade condenser effectiveness. Heat transfer through the all condenser, evaporator and cascade condenser is given by (2) (3) Fig. 1 Schematic diagram of cascade refrigeration system III. MODELING OF THE CONDENSER To simplify the analysis, the following assumptions were considered in the simulation. 1. The changes in kinetic and potential energy of the components are negligible. 2. Heat transfer with the ambient is negligible. 3. Compression process is adiabatic but not isentropic. 4. Pressure drop on water side and connecting pipes is negligible. 5. Only single phase heat transfer occurs for secondary fluid. Design of cascade system has been carried out for evaporating temperature (T EL ) of -82 0 C, condensing temperature (T CH ) of 36 0 C, condensing temperature of LTC (T CL ) of -30 0 C, temperature difference in cascade condenser (DT) of 2.5 0 C, refrigerating effect (Q evap ) of 0.5kW, degree of superheating (DSH) of 2 0 C in both cycles, degree of The overall heat transfer coefficient is defined by following equation, The log mean temperature difference for fluids flowing through the condenser is correlated by A. Shell and coil type Fig. 3 shows the shell and coil type condenser in which refrigerant flows through the coil and water from the shell side or over the coil. (4) (5) 33
Fig. 4 Tube in tube type condenser Fig. 3 Shell and coil type condenser The two phase heat transfer coefficient for the refrigerant side is calculated by Traviss [7] for in-tube condensation. where the liquid Reynold s and Prandl number are calculated as (8) The non dimensional parameter F 1 and F 2 are calculated by the following equations [8]. (9) Where (1/X tt ) is the Lockhart-Martinelli factor for turbulent flow is given by [7] If (6) (7) (10) Kern [7] suggested the correlation for shell side heat transfer coefficient which is given below (11) Where h w is the shell side heat transfer coefficient, d h is the equivalent diameter on the shell side and G W is the shell side mass velocity. B. Tube in tube type Fig. 4 displays the tube in tube type condenser in which refrigerant flows through the inner tube and water through the outer tube or over the inner tube. Heat transfer coefficient of refrigerant side for tube in tube type is calculated by the same procedure as for the shell and coil type condenser. Heat transfer coefficient for water side is calculated by Prandl s correlation [7], Friction factor is calculated by the following equation [7] Reynolds number is calculated as (12) (13) (14) Table I shows the condenser base data which is common for the simulation of both types of condenser considered in the present study. Based on the above mentioned mathematical model, a code has been developed in Engineering Equation Solver (EES) for the detail numerical simulation of both shell and coil and tube in tube type condensers [8]. Table II shows the physical properties of refrigerants R404A and R508B used in HTC and LTC of cascade system. TABLE I BASE DATA FOR CONDENSER Shell and coil type Outside diameter of coil Thickness of coil Shell inside diameter Tube in tube type Inner tube diameter Outer tube diameter Tube thickness 9.5 mm 0.7 mm 281 mm 9.5 mm 18.82 mm 0.7 mm TABLE II PROPERTIES OF REFRIGERANTS Properties R508B R404A Molecular weight (kg/kmol) 95.39 97.6 Critical temperature ( 0 C) 13.7 72.2 Critical pressure (bar) 39.35 36.68 Bubble point ( 0 C) -88.27-46.2 IV. RESULTS AND DISCUSSION Table III compares the heat transfer area for two different geometries of condenser at specified operating condition. It is seen that heat transfer area required for tube in tube type condenser is lower than the shell and coil condenser. 34
\TABLE III DESIGN DATA FOR CONDENSERS Heat transfer area in m 2 for R404A-R508B cascade system Type of condenser (T EL = -82 0 C, T CH = 36 0 C, T CL = -30 0 C, DT = 2.5 0 C) Shell and coil 0.3616 Tube in tube 0.3169 Fig.5 shows the effect of varying evaporating temperature on the heat transfer area of condenser keeping the other parameter constant. It is seen that when evaporating temperature decreased from -85 0 C to -58 0 C the required heat transfer area of condenser decreased linearly. Heat transfer area of R404A/R508B cascade system decreased by 7.53% for shell and coil type condenser and 7.53% for tube in tube type condenser. The heat transfer area for tube in tube type condenser is lower than the shell and coil at all evaporating temperature. Heat transfer area of condenser (m 2 ) 0.38 0.37 0.36 0.35 0.34 0.33 0.32 0.31 R404A/R508B Shell and coil 0.3-85 -82-79 -76-73 -70-67 -64-61 -58 Evaporating temperature ( 0 C) Tube in tube Fig. 5 Effect of evaporating temperature on condenser heat transfer area Fig. 6 shows the effect of condensing temperature on the heat transfer area of condenser for shell and coil type and tube in tube type geometries. As the condensing temperature increases from 32 0 C to 41 0 C, heat transfer area of condenser decreased by 79%. There is significant decrease in area when temperature reduces from 32 0 C to 35 0 C but after 35 0 C the decrease in area is minimal. Heat transfer area of condenser (m 2 ) 1 0.9 0.8 0.7 0.6 0.5 0.4 0.3 0.2 R404A/R508B Shell and coil Tube in tube 32 33 34 35 36 37 38 39 40 41 Condensing temperature ( 0 C) Fig. 6 Effect of condensing temperature on condenser heat transfer area V. CONCLUSION To predict the required heat transfer area for the water cooled condenser of R404A-R508B cascade refrigeration system simulation model has been developed in Engineering equation solver for two geometries viz. tube in tube and shell and coil. It was found that tube in tube type condenser requires less heat transfer area compared to shell and coil condenser under same operating conditions. The effect of evaporating and condensing temperatures and temperature difference between cascade condenser on heat transfer area of both types of condenser viz. shell and coil and tube in tube types are studied. It is found that the required heat transfer area of the both types of condensers decreases as condensing temperature and evaporating temperature increases but increases as temperature difference in cascade condenser increases. NOMENCLATURE LTC Low temperature cycle HTC High temperature cycle DSH Degree of superheating ( 0 C) DSC Degree of subcooling ( 0 C) Q Heat duty (kw) m Mass flow rate (kg/s) h Enthalpy (kj/kg) P Pressure (bar) T Temperature ( 0 C) d Diameter (m) r Radius (m) A Area (m 2 ) G Mass flux (kg/m 2 s) υ Specific volume (m 3 /kg) c p Specific heat capacity (kj/kgk) k Thermal conductivity (kw/k) Nu Nusselt number Pr Prandl number Re Reynolds number LMTD Logarithmic mean temperature difference ( 0 C) U Overall heat transfer coefficient (kw/m 2 K) x Dryness fraction 1/X tt Lockhart-Martinelli factor h Heat transfer coefficient (kw/m 2 K) f Friction factor g Acceleration due to gravity (m/s 2 ) Greek symbols ε Effectiveness or relative roughness Subscripts L Low temperature refrigerant side H High temperature refrigerant side evap Evaporation cond Condensation cascond Cascade condenser EL Evaporation of low temperature cycle CL Condensation of low temperature cycle 35
EH CH 1.8 in out i o ref w cs f g h Evaporation of high temperature cycle Condensation of high temperature cycle State points Inner tube Outer tube Inside diameter or inlet Outside diameter or outlet Refrigerant side Water side or wall Cross sectional Liquid phase Gas phase Hydraulic REFERENCES [1] L. M. Schlager, M. B. Pate, A. E. Bergles, Evaporation and condensation heat transfer and pressure drop in horizontal, 12.7-mm microfin tubes with refrigerant 22, International Journal of Heat Transfer 112, 1990, pp. 1041 1047. [2] L. Chamra, R. Webb, M. Randlett, Advanced microfin tubes for condensation, Int J Heat Mass Transfer 39 (9),1996, pp. 1839 1846. [3] M. K. Dobson, J. C. Chato, Condensation in smooth horizontal tubes, J Heat Transfer 120 (1998) 193 213. [4] T. J. Rennie, G. S.V. Ranghavan, Experimental studies of a double-pipe helical heat exchanger, Exp. Thermal Fluid Sci. 29, 2005, pp. 919 924. [5] Kwon J. T., Park S. K., Kim M. H., Enhanced effect of a horizontal microfin tube for condensation heat transfer with R22 and R410A, Enhanced Heat Transfer 7, 2000, pp. 97 107. [6] C. P. Arora, Refrigeration and air conditioning, 3rd edition, McGraw-Hill publication, New Delhi, 2009. [7] Sadik Kakac and Hongtan Liu, Heat exchangers: Selection, rating and thermal design, 2nd ed., CRC Press, Boca Raton, New York, 2002. [8] S. A. Klein, Engineering Equation Solver, commercial V7.027. 36