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Prediction of condensation of refrigerant mixtures under magnetic field inside enhanced surface tubing
International Communications in Heat and Mass Transfer 34 (2007) 261 – 268 www.elsevier.com/locate/ichmt
Prediction of condensation of refrigerant mixtures under magnetic field inside enhanced surface tubing ☆ S.M. Sami ⁎, J. Comeau Research Centre for Energy Conversion, Mechanical Engineering, School of Engineering, University of Moncton, Moncton, NB, Canada E1A 3E9 Available online 27 October 2006
Abstract Two phase flow heat transfer characteristics observed under magnetic field during condensation of refrigerant mixtures R-404A, R407C and R507 as well as R-410A were observed, analyzed and presented in this paper. Experiments showed that magnetic field tends to enhance the condensation characteristics at lower Reynolds numbers. The data clearly indicated that values of the condensation heat transfer coefficient of refrigerant mixtures under investigation were significantly influenced by the Magnetohyrdodynamic (MHD) effects depending upon the type of refrigerant mixture. The proposed correlation appears to predict the heat transfer coefficient with an average deviation of ±10. © 2006 Elsevier Ltd. All rights reserved. Keywords: Heat transfer; MHD; Condensation; Refrigerant mixtures; Numerical study
1. Introduction The growing popularity of enhanced surface tubing combined with the use of multi-component refrigerant mixtures as new alternatives prompted a lot of heat transfer studies in recent years. Furthermore, most refrigeration and air conditioning systems experience load variation. High efficiency and high performance are greatly in demand. Among techniques employed for improvement, capacity control, and optimization of vapour compression systems are the refrigerant liquid and vapour injection as well as Magnetohydrodynamic. The effect of magnetism and magnetic field on fluids is still considered as not a well known subject. However, it is well established that there are major changes caused by the passage of fluid through a magnetic field. The magnetic measurements to evaluate the thermodynamic behaviour of magnetic material have been presented by Foldeaki et al. [1]. As reported in this reference depending on the thermodynamic cycle selected, the isothermal magnetic entropy change temperature or the adiabatic temperature change upon the field application should be reselected as a function of temperature. This paper presented classical magnetic measurements, when evaluated within the framework of the Landau theory. Furthermore, most refrigeration and air conditioning systems experience load variation. High efficiency and high performance are greatly in demand. Among techniques employed for improvement, capacity control, and optimization of vapour compression systems are the refrigerant liquid and vapour injection. However, it is believed that the magnetic ☆
Communicated by W.J. Minkowycz. ⁎ Corresponding author. E-mail address: [email protected] (S.M. Sami).
0735-1933/$ - see front matter © 2006 Elsevier Ltd. All rights reserved. doi:10.1016/j.icheatmasstransfer.2006.09.002
262
S.M. Sami, J. Comeau / International Communications in Heat and Mass Transfer 34 (2007) 261–268
Fig. 1. Schematic diagram of the air/air heat pump test facility.
field can be employed as an enhancing technique. To the authors' knowledge, several studies have been reported in the literature on condensation of refrigerant mixtures by Wijaya, and Spatz [2], Sami et al. [3–6] as well as Wang et al. [7] and others Sami and Maltais [6] and Dobson and Chato [8]. The issue of condensation for different refrigerant mixtures such as R-404A, R-407C, and R-410A were addressed by these authors, inside internally and externally enhanced surfaces, such as double fluted tubes and air-fined tubes; Sami and Poirier [11]. This research work has been undertaken to further enhance our understanding of the subject and to provide new information on the condensational heat transfer of refrigerant mixtures R-404A, R410A, R507, and R507C under Magnetohydrodynamic forces (MHD) since none was published on the subject. 2. Experimental apparatus and measurements Fig. 1 shows a schematic diagram of an air/refrigerant vapor compression heat pump. The experimental set up was composed mainly of a 3 kW compressor, oil separator, condenser, pre-condenser, pre-evaporator, adjustable expansion Table 1 Air coils specifications Tube outer diameter Rows deep Fins per inch Fin depth Fin height Fin length Fin thickness Rif led tubes
3/8″ (0.09525 m) 4 12 3.46″ (0.087 m) 20″ (0.508 m) 30″ (0.762 m) 0.0045″ (0.01143 m) Micro-fins
S.M. Sami, J. Comeau / International Communications in Heat and Mass Transfer 34 (2007) 261–268
263
Table 2 Geometry of micro-fin tubes Outside diameter Root diameter Tip diameter Fin height Pitch
0.375″ (0.09525 m) 0.344″ (0.08376 m) 0.331″ (0.084 m) 0.0074″ (0.018 m) 0.016″ (0.04065 m)
device, and condenser/evaporator test sections. The oil content in the refrigerant loop was estimated by gas chromatography to be about ±1%. The geometry of the coils is presented in Tables 1 and 2. Table 2 outlines the geometrical parameters of the micro-fin tubes. Furthermore, the test conditions as well as key parameters of the set up are presented in Table 3. Pressure, temperature and flow rate as well as pressure drop measuring stations are shown in Fig. 1. All pressures were measured using calibrated pressure transducers (0–3400 kPa). The accuracy of pressure transducers was ±2.5%. Differential pressure transducers were employed to measure the refrigerant/coolant pressure drop and their accuracy is ±0.25% full scale. RTD sensors were used to measure temperatures with accuracy of ±0.5% full scale. Various permanent magnetic elements with mono pole configurations at gauss level of 4000 each have been employed in this study. These magnets were intended for flow applications lines of 1/4 in. diameter; they were clamped at the refrigerant line of same diameter. The units were single-type with two brackets strapped around the outside of the refrigerant pipes at three locations of the outlet of post condenser. They were clamped on the refrigerant liquid line at the post condenser outlet at various distances, before the capillary tube/thermal expansion valve used as flow control device. During the course of this experimental study, a series of permanent magnets were used totalling 14000 gauss levels. A calibrated orifice installed in the liquid line after a liquid receiver was used to measure the refrigerant mass flow rate. Also, airflow rates were measured through a calibrated Pitot tube based air-measuring stations. The water flow rates are needed in the pre-condenser and pre-evaporator to control the flow quality. These flows were measured with calibrated orifices and both orifice pressure taps were connected to a differential pressure transducer (0–70 kPa). The accuracy of the mass flow measurements was ±3% of the nominal flow. Power supplied to the compressor and fans was measured in order to complete heat balance. An AC/DC clamp-on device was calibrated for power measurements with accuracy of ±2.5%. Data collection was carried out using a lab view acquisition system equipped with a capacity of 112 channels. This enabled us to record with a single scan, local properties such as pressure drops, pressures, temperatures, refrigerant as well as coolant flow rates, heat flux and power. All tests were performed under steady state conditions. The channels were scanned every second and stored every 10 s. All recorded measurements were obtained according to the ARI and ASHRAE standards. The flow qualities of the refrigerant mixture entering the condenser test section were kept between saturated vapour conditions. Furthermore, the use of the multi-capillary tubes system enables us to perform the twophase pressure testing at different entry conditions. In order to determine the heat transfer flow condensation and boiling characteristics, the thermodynamic properties as well as transport properties of pure and/or isotropic refrigerant mixtures should be known. The REFPROP version 6.01 (McLinden et al. [10]) was employed with caution, in selecting the interaction parameters, to determine the transport and thermodynamic properties of the mixed refrigerants. This is quite important since the selection of the interactive parameters can have a significant impact on the thermodynamic and thermophysical properties (McLinden et al. [10], and Sami and Poirier [11]). 3. Results and discussions In the following sections, the heat transfer characteristics results at different conditions will be presented and discussed. Table 3 Test conditions Temperature of air of the condenser inlet Temperature of air at the evaporator inlet Air flow rate Refrigerant mass flow rate Condenser pressure Evaporator pressure Standard relative humidity at condenser
21 °C − 15 °C to +8 °C 7.07 * 10− 2 m3/s to 9.4 * 10− 2 m3/s 8 to 40 g/s 600 to 1800 kPa 170 to 450 kPa 45%
264
S.M. Sami, J. Comeau / International Communications in Heat and Mass Transfer 34 (2007) 261–268
Fig. 2. Condensation data for R407C.
3.1. Heat transfer modelling The following equations have been employed to calculate the heat transfer coefficients from the data stored during each particular test at equilibrium conditions. The heat transfer rate in the condenser test section can be determined from the heat balance of the airflow; Qa ¼ m a ðHain −Haout Þ
ð1Þ
where; Haout and Hain are the enthalpies of airflow leaving and the evaporator, respectively. The vapour quality at the exit of the test section was calculated from an energy balance of the system and the refrigerant properties were determined at saturation conditions in the test section. Qa xout ¼ xin þ ð2Þ mr hfg The total heat transferred to the refrigerant is; Qrt ¼ m r hfg Dx
ð3Þ
where Δx is the quality change in the test section. The overall heat transfer coefficient based on the outside surface area of the test section is; U¼
Qrt A0 LMTD
ð4Þ
where LMTD is the logarithmic mean temperature difference based on the inlet/outlet of air/refrigerant flows. The calculation of the LMTD was based on the change of non-linear enthalpy with temperature. Furthermore, the LMTD was calculated for one straight
Fig. 3. Heat transfer condensation data for R-507.
S.M. Sami, J. Comeau / International Communications in Heat and Mass Transfer 34 (2007) 261–268
265
Fig. 4. Condensation data for R-404A.
run of the test section since it is representative of the whole heat exchanger test section. No integration of LMTD values was needed, since only one section of the heat exchanger was considered in this study. Assuming no fouling and Rwal is the thermal resistance in the copper wall of the tube; the refrigerant heat transfer coefficient hr can be calculated as follows; 1 1 1 ð5Þ ¼ − þ Rwal hr Ai U A0 ha A0 Where ha is the air heat transfer coefficient and is calculated using the Wilson plot technique as developed by Khartabil and described by Sami and Poirier [11]. The Wilson plot analysis was conducted for each refrigerant mixture at each condensation condition. Rwal is the wall resistance evaluated using the actual thickness and the outside diameter of the tube. During the course of this study for data resolution purposes, the enhanced surface tube has been treated as a plain tube with an equivalent diameter. Using the equivalent diameter is consistent with the approach suggested by references (Sami and Fontaine [3], Sami and Poirier [11]) for the enhanced surface tubing heat exchangers. 3.2. Condensation data Samples of the experimental condensation data have been presented in Figs. 2–4 where the condensation heat transfer coefficient is plotted versus Reynolds number under various MHD conditions. As expected, the data showed that the condensation heat transfer coefficient increased with the increase of the Reynolds number. The data clearly indicated that the condensation heat transfer coefficient of the refrigerant mixtures under investigation was significantly influenced by the MHD conditions depending upon the type of refrigerant mixture and its boiling point. The heat transfer coefficient was enhanced under magnetic field. This was quite evident for flows with Reynolds number higher than 7 × 106 as shown in Fig. 3 where heat transfer coefficient was plotted against either refrigerant Reynolds number. The results
Fig. 5. Condensation correlation for R-407C.
266
S.M. Sami, J. Comeau / International Communications in Heat and Mass Transfer 34 (2007) 261–268
Fig. 6. Condensation data for R-407C.
also showed that the magnetic field impact on the heat transfer coefficient varies depending upon the thermophysical properties as well as Reynolds numbers as demonstrated in the above mentioned figures. It also appears from the data presented in Figs. 2–8 that the use of magnetic field is beneficial to the condensation process and has a positive effect on the heat transfer coefficient for higher Reynolds numbers. In our opinion, these results showed that the heat transfer coefficient in this region was also affected by the changes in the thermophysical properties due to the impact of the magnetic field on the thermophysical properties. After detailed analysis of the convective condensation two phase flow data, the following form proposed by Sámi and Fontaine [3] is considered. Figs. 5–7 show the proposed correlation; Nue ¼ AðRe2 dKf Þ0:3 þ B
ð6Þ
Where Nue ¼
h Db G Db ; Re ¼ k l
And Kf ¼
D X hfg L*g
The parameters A and B are determined experimentally. Therefore, Eq. (6) takes the following linear form; Nue ¼ AX þ B
ð7Þ
Figs. 5–8 have been constructed in an attempt to validate the proposed correlation, where values of measured heat transfer coefficients and non-dimensional heat transfer coefficient predicted by the proposed correlation were compared with experimental
Fig. 7. Condensation data for R-507.
S.M. Sami, J. Comeau / International Communications in Heat and Mass Transfer 34 (2007) 261–268
267
Fig. 8. Prediction of condensation data for R-404A.
ones. The data plotted in these figures showed, that the correlation is applicable to the entire heat and mass flux range, investigated in the present study for the proposed blends under question. The average deviation between the experimental and predicted values was less than ±10%, for the majority of data. Similar deviations have been observed with the other refrigerant mixtures under investigation. No attempts have been made to compare the proposed condensation correlations to existing ones, since to the authors' knowledge; none were reported in the literature under liquid injection test conditions. Eq. (6) clearly shows that there is a functional dependence of condensation heat transfer on the thermophysical and transport properties, and particularly thermal conductivity and viscosity of the refrigerant mixtures as well as other parameters. However, a sensitivity analysis demonstrated that the thermal conductivity and viscosity are the most crucial parameters to the prediction of the condensation characteristics (Dobson and Chato [8] and Ebisu and Torikoshi [9]), which are influenced by the magnetic treatment. This is consistent to what has been reported by Sami and Kita [12]. Furthermore, the data showed the inherent dependence of the proposed correlation in Eq. (7) on the thermophysical properties. The data displayed in these figures showed that the correlation is applicable to the entire heat and mass flux range for the MHD investigated in the present study for the proposed blends under question.
4. Conclusions During the course of this experimental study, the condensation heat transfer characteristics of some refrigerant mixtures have been presented inside enhanced surface tubing. It was evident from the condensation data that magnetic field has a significant impact on the heat transfer coefficient. Condensation data also showed an increase of heat transfer coefficient with the increase Reynolds number under magnetic treatment. Proposed correlations predicted the heat transfer coefficient with an average deviation of ±10%. Nomenclature A Heat transfer area (m2) D Diameter of tube (m) Dbo Fin height (m) G Mass flux (kg m− 2 s− 1) h Heat transfer coefficient (kW m− 2 K− 1) hfg Latent heat of vaporisation (kJ kg− 1) H Total air enthalpy (kJ kg− 1) K Thermal conductivity of liquid (kW m− 1 K− 1) L Length (m) m Mass flow rate (kg s− 1) Q Heat input (kW) Rwal Wall thermal resistance (K kW− 1) T Temperature (°C or K) U Overall heat transfer coefficient (kW m− 2 K− 1) x Quality based on mass (–)
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Greek Letters μ Viscosity of liquid (Pa s) ρ Density (kg m− 3) Dimensionless Numbers Ph Phase change number (hfg / Cp(Tv − Twa)) Pr Prandtl number of liquid (Cp μ / K) Re Reynolds number (GDe / μ) Nu Nusselt number (hDe / K) ReL Reynolds number for condensation (ρl uvin L / μ) Nuc Nusselt number for condensation (hL / K) Subscripts a Air en entrance L Liquid m Measured o Outside r Refrigerant rt Transferred to refrigerant tp Two-phase v Vapour wal Wall Acknowledgement The research work presented in this paper was made possible through grants from NSERC. The authors wish to acknowledge the continuous support of the University of Moncton. References [1] M. Foldeaki, R. Chahine, T.K. Bose, Magnetic measurements: a powerful tool in magnetic refrigerator design, Journal of Applied Physics 77 (7) (April 1 1995) 3528–3537. [2] H. Wijaya, M.W. Spatz, Two phase flow heat transfer and pressure drop characteristics of R-22 and R32/R125, ASHRAE transactions 101 (Pt.1) (1995). [3] S.M. Sami, M. Fontaine, Two-phase flow condensation characteristics of alternatives to R 502 inside enhanced surface tubing, ASHRAE transactions 105 (Pt.2) (1999). [4] S.M. Sami, D.E.J. Desjardins, Prediction of convective boiling characteristics of alternatives to R 502 inside air/refrigerant enhanced surface tubing, Applied Thermal Engineering 20 (2000) 579. [5] S.M. Sami, J. Grell, Heat transfer prediction of two-phase flow boiling of alternatives to R-22, International Journal of Energy Research 24 (2000) 349. [6] S.M. Sami, H.J. Maltais, Experimental investigation of two-phase flow condensation of alternatives to HCFC-22 inside enhanced surface tubing, Applied Thermal Engineering 20 (2000) 1113. [7] C. Wang, C.S. Kuo, Y. Chang, D.C. Lu, Nucleate condensation performance of R-22, R123, R134a, R410A and R407C on smooth and enhanced tubes, ASHRAE transactions 104 (Part 1) (1998). [8] M.K. Dobson, J.C. Chato, Condensation in smooth horizontal tubes, Journal of Heat Transfer 120 (1998) 193–213. [9] T. Ebisu, K. Torikoshi, Heat transfer characteristics and correlations for R-410A flowing inside a horizontal smooth tube, ASHRAE transactions 104 (Pt. 2) (1998) 556–561. [10] M.O. McLinden, (1998) NIST Thermodynamic Properties of Refrigerants and Refrigerant Mixtures Data Base, Version 6.01, NIST, Gaithersburg, ND. [11] S.M. Sami, B. Poirier, Two phase heat transfer of binary mixtures inside enhanced surface tubing, International Communication in Heat and Mass Transfer Journal 25 (6) (1998). [12] S.M. Sami, R.J. Kita, Behaviour of new refrigerant mixtures under magnetic field, International Journal of Energy Research 29 (2003) 1205–1213.
Prediction of condensation of refrigerant mixtures under magnetic field inside enhanced surface tubing ☆ S.M. Sami ⁎, J. Comeau Research Centre for Energy Conversion, Mechanical Engineering, School of Engineering, University of Moncton, Moncton, NB, Canada E1A 3E9 Available online 27 October 2006
Abstract Two phase flow heat transfer characteristics observed under magnetic field during condensation of refrigerant mixtures R-404A, R407C and R507 as well as R-410A were observed, analyzed and presented in this paper. Experiments showed that magnetic field tends to enhance the condensation characteristics at lower Reynolds numbers. The data clearly indicated that values of the condensation heat transfer coefficient of refrigerant mixtures under investigation were significantly influenced by the Magnetohyrdodynamic (MHD) effects depending upon the type of refrigerant mixture. The proposed correlation appears to predict the heat transfer coefficient with an average deviation of ±10. © 2006 Elsevier Ltd. All rights reserved. Keywords: Heat transfer; MHD; Condensation; Refrigerant mixtures; Numerical study
1. Introduction The growing popularity of enhanced surface tubing combined with the use of multi-component refrigerant mixtures as new alternatives prompted a lot of heat transfer studies in recent years. Furthermore, most refrigeration and air conditioning systems experience load variation. High efficiency and high performance are greatly in demand. Among techniques employed for improvement, capacity control, and optimization of vapour compression systems are the refrigerant liquid and vapour injection as well as Magnetohydrodynamic. The effect of magnetism and magnetic field on fluids is still considered as not a well known subject. However, it is well established that there are major changes caused by the passage of fluid through a magnetic field. The magnetic measurements to evaluate the thermodynamic behaviour of magnetic material have been presented by Foldeaki et al. [1]. As reported in this reference depending on the thermodynamic cycle selected, the isothermal magnetic entropy change temperature or the adiabatic temperature change upon the field application should be reselected as a function of temperature. This paper presented classical magnetic measurements, when evaluated within the framework of the Landau theory. Furthermore, most refrigeration and air conditioning systems experience load variation. High efficiency and high performance are greatly in demand. Among techniques employed for improvement, capacity control, and optimization of vapour compression systems are the refrigerant liquid and vapour injection. However, it is believed that the magnetic ☆
Communicated by W.J. Minkowycz. ⁎ Corresponding author. E-mail address: [email protected] (S.M. Sami).
0735-1933/$ - see front matter © 2006 Elsevier Ltd. All rights reserved. doi:10.1016/j.icheatmasstransfer.2006.09.002
262
S.M. Sami, J. Comeau / International Communications in Heat and Mass Transfer 34 (2007) 261–268
Fig. 1. Schematic diagram of the air/air heat pump test facility.
field can be employed as an enhancing technique. To the authors' knowledge, several studies have been reported in the literature on condensation of refrigerant mixtures by Wijaya, and Spatz [2], Sami et al. [3–6] as well as Wang et al. [7] and others Sami and Maltais [6] and Dobson and Chato [8]. The issue of condensation for different refrigerant mixtures such as R-404A, R-407C, and R-410A were addressed by these authors, inside internally and externally enhanced surfaces, such as double fluted tubes and air-fined tubes; Sami and Poirier [11]. This research work has been undertaken to further enhance our understanding of the subject and to provide new information on the condensational heat transfer of refrigerant mixtures R-404A, R410A, R507, and R507C under Magnetohydrodynamic forces (MHD) since none was published on the subject. 2. Experimental apparatus and measurements Fig. 1 shows a schematic diagram of an air/refrigerant vapor compression heat pump. The experimental set up was composed mainly of a 3 kW compressor, oil separator, condenser, pre-condenser, pre-evaporator, adjustable expansion Table 1 Air coils specifications Tube outer diameter Rows deep Fins per inch Fin depth Fin height Fin length Fin thickness Rif led tubes
3/8″ (0.09525 m) 4 12 3.46″ (0.087 m) 20″ (0.508 m) 30″ (0.762 m) 0.0045″ (0.01143 m) Micro-fins
S.M. Sami, J. Comeau / International Communications in Heat and Mass Transfer 34 (2007) 261–268
263
Table 2 Geometry of micro-fin tubes Outside diameter Root diameter Tip diameter Fin height Pitch
0.375″ (0.09525 m) 0.344″ (0.08376 m) 0.331″ (0.084 m) 0.0074″ (0.018 m) 0.016″ (0.04065 m)
device, and condenser/evaporator test sections. The oil content in the refrigerant loop was estimated by gas chromatography to be about ±1%. The geometry of the coils is presented in Tables 1 and 2. Table 2 outlines the geometrical parameters of the micro-fin tubes. Furthermore, the test conditions as well as key parameters of the set up are presented in Table 3. Pressure, temperature and flow rate as well as pressure drop measuring stations are shown in Fig. 1. All pressures were measured using calibrated pressure transducers (0–3400 kPa). The accuracy of pressure transducers was ±2.5%. Differential pressure transducers were employed to measure the refrigerant/coolant pressure drop and their accuracy is ±0.25% full scale. RTD sensors were used to measure temperatures with accuracy of ±0.5% full scale. Various permanent magnetic elements with mono pole configurations at gauss level of 4000 each have been employed in this study. These magnets were intended for flow applications lines of 1/4 in. diameter; they were clamped at the refrigerant line of same diameter. The units were single-type with two brackets strapped around the outside of the refrigerant pipes at three locations of the outlet of post condenser. They were clamped on the refrigerant liquid line at the post condenser outlet at various distances, before the capillary tube/thermal expansion valve used as flow control device. During the course of this experimental study, a series of permanent magnets were used totalling 14000 gauss levels. A calibrated orifice installed in the liquid line after a liquid receiver was used to measure the refrigerant mass flow rate. Also, airflow rates were measured through a calibrated Pitot tube based air-measuring stations. The water flow rates are needed in the pre-condenser and pre-evaporator to control the flow quality. These flows were measured with calibrated orifices and both orifice pressure taps were connected to a differential pressure transducer (0–70 kPa). The accuracy of the mass flow measurements was ±3% of the nominal flow. Power supplied to the compressor and fans was measured in order to complete heat balance. An AC/DC clamp-on device was calibrated for power measurements with accuracy of ±2.5%. Data collection was carried out using a lab view acquisition system equipped with a capacity of 112 channels. This enabled us to record with a single scan, local properties such as pressure drops, pressures, temperatures, refrigerant as well as coolant flow rates, heat flux and power. All tests were performed under steady state conditions. The channels were scanned every second and stored every 10 s. All recorded measurements were obtained according to the ARI and ASHRAE standards. The flow qualities of the refrigerant mixture entering the condenser test section were kept between saturated vapour conditions. Furthermore, the use of the multi-capillary tubes system enables us to perform the twophase pressure testing at different entry conditions. In order to determine the heat transfer flow condensation and boiling characteristics, the thermodynamic properties as well as transport properties of pure and/or isotropic refrigerant mixtures should be known. The REFPROP version 6.01 (McLinden et al. [10]) was employed with caution, in selecting the interaction parameters, to determine the transport and thermodynamic properties of the mixed refrigerants. This is quite important since the selection of the interactive parameters can have a significant impact on the thermodynamic and thermophysical properties (McLinden et al. [10], and Sami and Poirier [11]). 3. Results and discussions In the following sections, the heat transfer characteristics results at different conditions will be presented and discussed. Table 3 Test conditions Temperature of air of the condenser inlet Temperature of air at the evaporator inlet Air flow rate Refrigerant mass flow rate Condenser pressure Evaporator pressure Standard relative humidity at condenser
21 °C − 15 °C to +8 °C 7.07 * 10− 2 m3/s to 9.4 * 10− 2 m3/s 8 to 40 g/s 600 to 1800 kPa 170 to 450 kPa 45%
264
S.M. Sami, J. Comeau / International Communications in Heat and Mass Transfer 34 (2007) 261–268
Fig. 2. Condensation data for R407C.
3.1. Heat transfer modelling The following equations have been employed to calculate the heat transfer coefficients from the data stored during each particular test at equilibrium conditions. The heat transfer rate in the condenser test section can be determined from the heat balance of the airflow; Qa ¼ m a ðHain −Haout Þ
ð1Þ
where; Haout and Hain are the enthalpies of airflow leaving and the evaporator, respectively. The vapour quality at the exit of the test section was calculated from an energy balance of the system and the refrigerant properties were determined at saturation conditions in the test section. Qa xout ¼ xin þ ð2Þ mr hfg The total heat transferred to the refrigerant is; Qrt ¼ m r hfg Dx
ð3Þ
where Δx is the quality change in the test section. The overall heat transfer coefficient based on the outside surface area of the test section is; U¼
Qrt A0 LMTD
ð4Þ
where LMTD is the logarithmic mean temperature difference based on the inlet/outlet of air/refrigerant flows. The calculation of the LMTD was based on the change of non-linear enthalpy with temperature. Furthermore, the LMTD was calculated for one straight
Fig. 3. Heat transfer condensation data for R-507.
S.M. Sami, J. Comeau / International Communications in Heat and Mass Transfer 34 (2007) 261–268
265
Fig. 4. Condensation data for R-404A.
run of the test section since it is representative of the whole heat exchanger test section. No integration of LMTD values was needed, since only one section of the heat exchanger was considered in this study. Assuming no fouling and Rwal is the thermal resistance in the copper wall of the tube; the refrigerant heat transfer coefficient hr can be calculated as follows; 1 1 1 ð5Þ ¼ − þ Rwal hr Ai U A0 ha A0 Where ha is the air heat transfer coefficient and is calculated using the Wilson plot technique as developed by Khartabil and described by Sami and Poirier [11]. The Wilson plot analysis was conducted for each refrigerant mixture at each condensation condition. Rwal is the wall resistance evaluated using the actual thickness and the outside diameter of the tube. During the course of this study for data resolution purposes, the enhanced surface tube has been treated as a plain tube with an equivalent diameter. Using the equivalent diameter is consistent with the approach suggested by references (Sami and Fontaine [3], Sami and Poirier [11]) for the enhanced surface tubing heat exchangers. 3.2. Condensation data Samples of the experimental condensation data have been presented in Figs. 2–4 where the condensation heat transfer coefficient is plotted versus Reynolds number under various MHD conditions. As expected, the data showed that the condensation heat transfer coefficient increased with the increase of the Reynolds number. The data clearly indicated that the condensation heat transfer coefficient of the refrigerant mixtures under investigation was significantly influenced by the MHD conditions depending upon the type of refrigerant mixture and its boiling point. The heat transfer coefficient was enhanced under magnetic field. This was quite evident for flows with Reynolds number higher than 7 × 106 as shown in Fig. 3 where heat transfer coefficient was plotted against either refrigerant Reynolds number. The results
Fig. 5. Condensation correlation for R-407C.
266
S.M. Sami, J. Comeau / International Communications in Heat and Mass Transfer 34 (2007) 261–268
Fig. 6. Condensation data for R-407C.
also showed that the magnetic field impact on the heat transfer coefficient varies depending upon the thermophysical properties as well as Reynolds numbers as demonstrated in the above mentioned figures. It also appears from the data presented in Figs. 2–8 that the use of magnetic field is beneficial to the condensation process and has a positive effect on the heat transfer coefficient for higher Reynolds numbers. In our opinion, these results showed that the heat transfer coefficient in this region was also affected by the changes in the thermophysical properties due to the impact of the magnetic field on the thermophysical properties. After detailed analysis of the convective condensation two phase flow data, the following form proposed by Sámi and Fontaine [3] is considered. Figs. 5–7 show the proposed correlation; Nue ¼ AðRe2 dKf Þ0:3 þ B
ð6Þ
Where Nue ¼
h Db G Db ; Re ¼ k l
And Kf ¼
D X hfg L*g
The parameters A and B are determined experimentally. Therefore, Eq. (6) takes the following linear form; Nue ¼ AX þ B
ð7Þ
Figs. 5–8 have been constructed in an attempt to validate the proposed correlation, where values of measured heat transfer coefficients and non-dimensional heat transfer coefficient predicted by the proposed correlation were compared with experimental
Fig. 7. Condensation data for R-507.
S.M. Sami, J. Comeau / International Communications in Heat and Mass Transfer 34 (2007) 261–268
267
Fig. 8. Prediction of condensation data for R-404A.
ones. The data plotted in these figures showed, that the correlation is applicable to the entire heat and mass flux range, investigated in the present study for the proposed blends under question. The average deviation between the experimental and predicted values was less than ±10%, for the majority of data. Similar deviations have been observed with the other refrigerant mixtures under investigation. No attempts have been made to compare the proposed condensation correlations to existing ones, since to the authors' knowledge; none were reported in the literature under liquid injection test conditions. Eq. (6) clearly shows that there is a functional dependence of condensation heat transfer on the thermophysical and transport properties, and particularly thermal conductivity and viscosity of the refrigerant mixtures as well as other parameters. However, a sensitivity analysis demonstrated that the thermal conductivity and viscosity are the most crucial parameters to the prediction of the condensation characteristics (Dobson and Chato [8] and Ebisu and Torikoshi [9]), which are influenced by the magnetic treatment. This is consistent to what has been reported by Sami and Kita [12]. Furthermore, the data showed the inherent dependence of the proposed correlation in Eq. (7) on the thermophysical properties. The data displayed in these figures showed that the correlation is applicable to the entire heat and mass flux range for the MHD investigated in the present study for the proposed blends under question.
4. Conclusions During the course of this experimental study, the condensation heat transfer characteristics of some refrigerant mixtures have been presented inside enhanced surface tubing. It was evident from the condensation data that magnetic field has a significant impact on the heat transfer coefficient. Condensation data also showed an increase of heat transfer coefficient with the increase Reynolds number under magnetic treatment. Proposed correlations predicted the heat transfer coefficient with an average deviation of ±10%. Nomenclature A Heat transfer area (m2) D Diameter of tube (m) Dbo Fin height (m) G Mass flux (kg m− 2 s− 1) h Heat transfer coefficient (kW m− 2 K− 1) hfg Latent heat of vaporisation (kJ kg− 1) H Total air enthalpy (kJ kg− 1) K Thermal conductivity of liquid (kW m− 1 K− 1) L Length (m) m Mass flow rate (kg s− 1) Q Heat input (kW) Rwal Wall thermal resistance (K kW− 1) T Temperature (°C or K) U Overall heat transfer coefficient (kW m− 2 K− 1) x Quality based on mass (–)
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Greek Letters μ Viscosity of liquid (Pa s) ρ Density (kg m− 3) Dimensionless Numbers Ph Phase change number (hfg / Cp(Tv − Twa)) Pr Prandtl number of liquid (Cp μ / K) Re Reynolds number (GDe / μ) Nu Nusselt number (hDe / K) ReL Reynolds number for condensation (ρl uvin L / μ) Nuc Nusselt number for condensation (hL / K) Subscripts a Air en entrance L Liquid m Measured o Outside r Refrigerant rt Transferred to refrigerant tp Two-phase v Vapour wal Wall Acknowledgement The research work presented in this paper was made possible through grants from NSERC. The authors wish to acknowledge the continuous support of the University of Moncton. References [1] M. Foldeaki, R. Chahine, T.K. Bose, Magnetic measurements: a powerful tool in magnetic refrigerator design, Journal of Applied Physics 77 (7) (April 1 1995) 3528–3537. [2] H. Wijaya, M.W. Spatz, Two phase flow heat transfer and pressure drop characteristics of R-22 and R32/R125, ASHRAE transactions 101 (Pt.1) (1995). [3] S.M. Sami, M. Fontaine, Two-phase flow condensation characteristics of alternatives to R 502 inside enhanced surface tubing, ASHRAE transactions 105 (Pt.2) (1999). [4] S.M. Sami, D.E.J. Desjardins, Prediction of convective boiling characteristics of alternatives to R 502 inside air/refrigerant enhanced surface tubing, Applied Thermal Engineering 20 (2000) 579. [5] S.M. Sami, J. Grell, Heat transfer prediction of two-phase flow boiling of alternatives to R-22, International Journal of Energy Research 24 (2000) 349. [6] S.M. Sami, H.J. Maltais, Experimental investigation of two-phase flow condensation of alternatives to HCFC-22 inside enhanced surface tubing, Applied Thermal Engineering 20 (2000) 1113. [7] C. Wang, C.S. Kuo, Y. Chang, D.C. Lu, Nucleate condensation performance of R-22, R123, R134a, R410A and R407C on smooth and enhanced tubes, ASHRAE transactions 104 (Part 1) (1998). [8] M.K. Dobson, J.C. Chato, Condensation in smooth horizontal tubes, Journal of Heat Transfer 120 (1998) 193–213. [9] T. Ebisu, K. Torikoshi, Heat transfer characteristics and correlations for R-410A flowing inside a horizontal smooth tube, ASHRAE transactions 104 (Pt. 2) (1998) 556–561. [10] M.O. McLinden, (1998) NIST Thermodynamic Properties of Refrigerants and Refrigerant Mixtures Data Base, Version 6.01, NIST, Gaithersburg, ND. [11] S.M. Sami, B. Poirier, Two phase heat transfer of binary mixtures inside enhanced surface tubing, International Communication in Heat and Mass Transfer Journal 25 (6) (1998). [12] S.M. Sami, R.J. Kita, Behaviour of new refrigerant mixtures under magnetic field, International Journal of Energy Research 29 (2003) 1205–1213.