- Email: [email protected]
Heat transfer and pressure drop characteristics of forced convective evaporation in horizontal tubes with coiled wire inserts
International Communications in Heat and Mass Transfer 36 (2009) 1089–1095
Contents lists available at ScienceDirect
International Communications in Heat and Mass Transfer j o u r n a l h o m e p a g e : w w w. e l s ev i e r. c o m / l o c a t e / i c h m t
Heat transfer and pressure drop characteristics of forced convective evaporation in horizontal tubes with coiled wire inserts☆ M.A. Akhavan-Behabadi ⁎, S.G. Mohseni, H. Najafi, H. Ramazanzadeh School of Mechanical Engineering, University College of Engineering, University of Tehran, Tehran, Iran
a r t i c l e
i n f o
Available online 12 August 2009 Keywords: Heat transfer Evaporation Pressure drop R-134a Coiled wire
a b s t r a c t Heat transfer enhancement and pressure drop increasing during evaporation of R-134a due to the presence of coiled wire insert inside a horizontal evaporator was studied experimentally. The test evaporator was an electrically heated copper tube of 1200 mm length and 7.5 mm inside diameter. Helically wire coils with different wire diameters of 0.5, 0.7, 1.0 and 1.5 mm and different coil pitches of 5, 8, 10 and 13 mm were made and used in full length of the test evaporator. For each inserted tube and also the plain tube, several test runs were carried out with different mass velocities and heat fluxes. From analysis of acquired data, it was found that the coiled wire inserts enhance the heat transfer coefficient but with a higher penalty due to the increasing of pressure drop, in comparison to that for the plain flow. An empirical correlation has been developed to predict the heat transfer coefficient during evaporation inside a horizontal tube in the presence of a coiled wire insert. © 2009 Elsevier Ltd. All rights reserved.
1. Introduction The limited availability of the world's conventional energy resources and the ever increasing cost of energy over the last several years has accelerated research in the field of energy conservation with the aim of possible reduction in its consumption in a given process. In the heat transfer field, extensive efforts are made to develop compact and more efficient heat exchangers. Evaporator is an important and widely used heat exchanger in air conditioning and refrigeration industries. Different methods have been used by investigators to increase the heat transfer rates in evaporators [1–3]. Webb [4] summarized various kinds of heat transfer enhancement techniques and has classified them into two main categories, viz., active technique and passive technique. The use of various types of insertions, such as, wire mesh or brush, twisted tape and coil or spiral spring, is considered as the passive techniques of enhancing the heat transfer, which does not require any external energy. The internal fin tubes are also used to enhance the heat transfer coefficient but they are expensive in comparison to the in-tube inserts. Moreover, inserts are easy to install inside the tubes. It should be also noted that, the pressure drop is also usually increased by using these techniques. The main objective of all the research work in this field is to obtain the highest heat transfer coefficient with the least pressure drop. A review of the existing literature, revealed that the use of coiled wire inserts to enhance the heat transfer rates during evaporation of
☆ Communicated by W.J. Minkowycz. ⁎ Corresponding author. E-mail address: [email protected] (M.A. Akhavan-Behabadi). 0735-1933/$ – see front matter © 2009 Elsevier Ltd. All rights reserved. doi:10.1016/j.icheatmasstransfer.2009.07.009
refrigerants flowing inside a horizontal tube, has been studied for quite some time [5]. Furthermore, the refrigerant R-134a, nowadays, is seen to be the eco-friendly substitute of R-12. It has zero ozone depletion potential (ODP) and 10% global warming potential (GWP) in comparison to R-12. Therefore, an experimental investigation was carried out to study the effects of wire coil inserts on evaporation heat transfer and pressure drop characteristics of R-134a in a horizontal tube. Finally, the performance evaluation of the wire coil inserts from the point view of heat transfer enhancement and pressure drop increasing should be done. 2. Experimental apparatus and procedure The test apparatus was a well instrumented vapor compression refrigeration system. The schematic diagram of experimental apparatus is shown in Fig. 1. It consisted of a compressor, a condenser, an expansion valve, a pre-evaporator, a test evaporator or test section, an after-evaporator, and necessary instruments for measurements and controls. The test evaporator was a copper tube of 7.5 mm inside diameter, di, 9.54 mm outside diameter, do, and 1200 mm length, L. Refrigerant R-134a flowing inside the test-evaporator tube was heated by a flexible electrical heating tape of 2 kW capacity, wrapped around it. During the experiments, a short range of vapor quality could be achieved in a single test run. Therefore, to cover the entire ranges of vapor qualities a pre-evaporator was installed before the test evaporator. By regulating the voltage to the electrical coil around the pre-evaporator, the vapor quality at the inlet of the test evaporator was controlled. It was ensured that the refrigerant coming from the test evaporator was superheated before it entered the compressor. This was attained by installing an after-evaporator
1090
M.A. Akhavan-Behabadi et al. / International Communications in Heat and Mass Transfer 36 (2009) 1089–1095
Nomenclature Bo C1,C2 cp d e g G h h̅ L ṁ P p q Q Rh RΔp T k x
boiling number, q/Gλ constants of Eq. (11) specific heat, kJ/kg K diameter, m wire diameter, m gravitational acceleration, m/s2 mass velocity, kg/s m2 heat transfer coefficient, W/m2 K average heat transfer coefficient, W/m2 K tube length, m mass flow rate, kg/s pressure, kPa coil pitch, m heat flux, kW/m2 electrical work on evaporator, W ratio of heat transfer coefficient of rough tube to plain tube ratio of pressure loss of rough tube to plain tube temperature, K thermal conductivity, W/m K vapor quality
Greek symbols α helix angle of coiled wire, degree λ latent heat, kJ/kg Δp pressure drop, kPa ΔT temperature difference, K ρ density, kg/m3
Different types of full length wire coils were inserted, one by one, having different wire diameters of 0.5, 0.7, 1.0 and 1.5 mm and different coil pitches of 5, 8, 10 and 13 mm. Table 1 represents the specification of the used coiled wire inserted tubes. After collecting experimental data for plain flow (tube set H), the data were acquired for the all the roughed tubes (coiled wire inserted tubes) given in Table 1. The data were acquired at refrigerant mass velocities of 54, 85, 114 and 136 kg/s m2. All together 194 test runs were performed. The ranges of operating parameters are as follows: Refrigerant mass velocity: 54–136 kg/s m2 Heat flux: 1.8–5.3 kW/m2 Boiling temperature: − 3 to − 19 °C Inlet vapor quality: 0.2–0.9 Exit vapor quality: 0.3–1.0 Liquid Reynolds number: 1250–3500 Twist ratio (p/di): 0.66–1.73. The thermo physical properties of R-134a were taken from [6]. The vapor quality at the inlet of pre-evaporator was computed by considering the iso-enthalpic expansion of refrigerant in the needle valve. An energy balance technique was used to get the vapor quality of the inlet and the exit of the test evaporator. The mean vapor quality was taken as the average of inlet and outlet vapor qualities of the test section. The heat transfer coefficient of the test section was determined using the heat gain from the electrical heater and the temperature difference between the evaporator inside wall surface and the boiling refrigerant. The average outside tube-wall temperature of the test evaporator at a particular station, Tws, was calculated by using the following equation: Tws =
Subscripts b bottom c wire coiled inserted tube e equivalent f liquid g vapor i inside tube o outside tube p plain tube s saturated, side t top w tube wall wi inside wall wo outside wall
ð1Þ
The average outside tube-wall temperature of the test evaporator, Two, was computed by taking the average temperatures of six axial stations as: 6
∑Tws Two =
1
6
:
ð2Þ
The radial heat flux, q, for the test evaporator was calculated by: q = Q =ðπdi LÞ:
ð3Þ
The temperature drop across the tube wall, ΔTw, was evaluated by: ΔTw =
downstream of the test evaporator. All the three evaporators were thermally insulated. An accumulator was also provided in the suction line of the compressor. The refrigerant mass flow rate was measured by a pre-calibrated rotameter installed downstream of the condenser. The outside wall temperatures of the test-evaporator tube were measured at six axial locations. At each location the temperature of the tube was measured at the top, both sides in the middle of the tube and bottom positions. The refrigerant temperatures at the inlet and outlet of the test evaporator were also measured. Fig. 2 shows the schematic diagram of the test section. All the temperature measurements were done by T-type thermocouples with a calibrated accuracy of 0.1 °C. To measure the pressure drop across the test evaporator, a differential pressure transducer apparatus was utilized. This apparatus was calibrated to measure the pressure loss in the ranges of 0 to 150 kPa. All other necessity parameters were also measured by accurate instruments.
Tt + 2Ts + Tb : 4
qdi lnðdo = di Þ : 2kw
ð4Þ
The average inside tube-wall temperature, Twi, was calculated by subtracting the temperature drop across the wall, ΔTw, from measured average outside temperature, Two. Twi = Two −ΔTw :
ð5Þ
The average static pressure in the test evaporator was taken to be the mean of the inlet and outlet pressure. The vapor saturation temperature in the test evaporator, Ts, was taken as the saturation temperature corresponding to this average static pressure. The heat transfer coefficient of the test evaporator was calculated by the following equation: h=
q : ðTwi −Ts Þ
ð6Þ
M.A. Akhavan-Behabadi et al. / International Communications in Heat and Mass Transfer 36 (2009) 1089–1095
1091
Fig. 1. Schematic diagram of experimental set-up.
3. Results and discussion 3.1. Heat transfer The Gungor and Winterton correlation for estimation of heat transfer coefficient during flow boiling inside the horizontal plain tube is as follows [7]: " 0:86
hp = hfp 1 + 3000Bo
x 0:75 ρ f + 1:12 ρg 1−x
!0:41 # ð7Þ
where, Gð1−xÞdi 0:8 0:4 kf : Prf hfp = 0:023 μf di
ð8Þ
The above mentioned correlation predicts the plain flow experimental data within an error band of −14% to +9%. This establishes the integrity of experimental apparatus. The results of plain flow are also used in comparing the performance of different coiled wire inserts. The variation of heat transfer coefficient with vapor quality for plain tube and the tubes with coiled wire inserts has been shown in Figs. 3–6. The effects of coiled wire inserts having different coil pitches and a constant wire thickness of 1 mm are shown in Figs. 3 and 4 for the mass velocities of 54 kg/m2 s and 136 kg/s m2, while the effects of coiled wire inserts having different wire thicknesses and a constant coil pitch of 10 mm on the enhancement of heat transfer coefficients are shown in Figs. 5 and 6 for mass velocities of 85 kg/s m2 and 114 kg/s m2. From these figures it is observed that, the insertion of a coiled wire inside the test-evaporator tube has produced higher heat transfer coefficient compared to plain tube value. The magnitude of the augmentation is not consistent for a special coiled wire and it is a complex function of mass velocity, vapor quality, and the geometry of a coil. For all the flow rates the best performing coiled wire insert is that having a coil pitch of 10 mm and wire thickness of 1.5 mm (the thickest coiled wire insert). In the high vapor quality region the insert
Table 1 Characteristic parameters of the coiled wire inserted tubes.
Fig. 2. Details of the test section.
Tube set
di (mm)
de (mm)
e (mm)
p (mm)
α (degree)
A B C D E F G H
7.5 7.5 7.5 7.5 7.5 7.5 7.5 7.5
6.44 6.09 5.61 4.95 4.48 5.29 5.94 7.5
0.5 0.7 1.0 1.5 1.0 1.0 1.0
10 10 10 10 5 8 13 Plain tube
73 73 73 73 81.5 76.6 69
1092
M.A. Akhavan-Behabadi et al. / International Communications in Heat and Mass Transfer 36 (2009) 1089–1095
Fig. 3. Heat transfer coefficient versus vapor quality for the plain tube and the roughed tubes with constant wire diameter of 1 mm at mass velocity of 54 kg/s m2.
Fig. 5. Heat transfer coefficient versus vapor quality for the plain tube and the roughed tubes with constant coil pitch of 10 mm at mass velocity of 85 kg/s m2.
enhances the heat transfer coefficient in a range of 72 to 98% in comparison to that for the plain flow. For the same insert at vapor quality, x, of 0.2 the enhancement in heat transfer coefficient is in a range of 0 to 49% and at the vapor quality, x, of 0.8 the enhancement in heat transfer coefficient is in a range of 74 to 88%. Akhavan-Behabadi et al. [8] investigated the effect of coiled wire inserts during forced convective condensation of R-134a inside a 10.7-mm I.D. horizontal copper tube. They reported a maximum of 85% heat transfer enhancement for coiled wire having a coil pitch of 10 mm and wire thickness of 1.5 mm. Since, the present investigation has been carried out for the evaporation of R-134a, the highest enhancement in heat transfer coefficient is 98%. From Figs. 3 and 4, it is seen that, the heat transfer coefficient increases with the decrease of coil pitch, so that the highest heat transfer coefficient is achieved by the coil with the pitch of 8 mm (tube set ‘F’); further reduction in the coil pitch reduces the heat transfer coefficient. This trend had been observed previously and was explained by this fact that in coil pitches smaller than a particular value, coil resists to heat transfer like a solid shell [9]. Also, it is revealed from Figs. 5 and 6 that in the high vapor quality region, heat transfer coefficient increases with the increase of the wire thickness but in the low vapor quality region,
although the maximum heat transfer coefficient is related to the thickest coiled wire this trend is not general. Yun et al. [10] also reported the same observations in their study of flow boiling heat transfer characteristics of nitrogen in coiled wire inserted tubes. In Fig. 7, the variation of heat transfer coefficient with vapor quality for different mass velocities of R-134a is shown for the tube having a coiled wire insert of 0.5 mm diameter and 10 mm pitch. It is observed that for a given vapor quality, the heat transfer coefficient is increased with the increase of vapor mass velocity. This phenomenon is simply contributed to the fact that at higher mass velocities, the Reynolds number is also higher, which in turn, increases the Nusselt number and, hence, the heat transfer coefficient. In addition, as the mass velocity increased, the swirl flow became much stronger and the dryout vapor quality increased. The same observation has also been made by Yun et al. [10]. The coiled wire insert of 0.5-mm diameter wire and 10 mm pitch has enhanced the average heat-transfer coefficient, h ,̅ in a range of 11– 20% in comparison to that for the plain flow. Literature survey showed that there is no correlation to predict the heat transfer coefficient during forced convective boiling of R-134a in spring inserted tubes. Therefore, a new correlation was developed to predict the heat transfer coefficient in coiled wire inserted tubes. For
Fig. 4. Heat transfer coefficient versus vapor quality for the plain tube and the roughed tubes with constant wire diameter of 1 mm at mass velocity of 136 kg/s m2.
Fig. 6. Heat transfer coefficient versus vapor quality for the plain tube and the roughed tubes with constant coil pitch of 10 mm at mass velocity of 114 kg/s m2.
M.A. Akhavan-Behabadi et al. / International Communications in Heat and Mass Transfer 36 (2009) 1089–1095
Fig. 7. Heat transfer coefficient versus vapor quality for tube set “A” at different mass velocities.
this purpose, Eq. (7) was used as a base equation and three additional parameters are added to it, i.e., ratio of the wire diameter to the tube inside diameter (e/di), the ratio of the coil pitch to tube inside diameter (p/di), and equivalent diameter (de). The equivalent diameter was calculated by using Eq. (9): 2
de = ðdi −eγÞ = ðdi + γÞ
ð9Þ
where, γ = πeðdi −eÞ = ðp sin αÞ:
ð10Þ
These variables have also been used by Akhavan-Behabadi et al. [8]. The following form of the functional relationship was developed as follows: 2
hc = C1
e pdi
!C
2
0:86
hfc 1 + 3000Bo
+ 1:12
x 0:75 ρ f ρg 1−x
!0:41 ! ð11Þ
where, Gð1−xÞde 0:8 0:4 kf : Prf hfc = 0:023 μL de
ð12Þ
1093
Fig. 8. Comparison of the experimental heat transfer coefficient with that predicted by Eq. (13).
3.2. Pressure drop Since insertion of coiled wire in the plain tube also increases the pressure drop across the test evaporator, the effect of wire geometry on pressure drop rates is also required to be investigated. The variation of pressure drop with vapor quality for the plain tube and the tubes with coiled wire inserts at different mass velocities has been shown in Figs. 9–12. It is found that, in general, by insertion of coiled wire inside horizontal tubes, the pressure drop is increased in comparison to that of plain flow. It is also observed that, the pressure drop is increased with the rise of vapor quality. This is due to the fact that, as the quality of vapor inside the tube grows; the velocity of fluid increases which in its own turn leads in more shear stress and more pressure loss. Furthermore, it is also observed that for a given vapor quality, the pressure drop is reduced with the decrease in mass velocity. The effects of the coiled wire inserts having different pitches and a constant wire thickness of 1.0 mm on the pressure drop are shown in Figs. 9 and 10 for mass velocities of 54 kg/s m2 and 136 kg/s m2. It is seen that, the lower pressure drop is achieved for the tube set with larger coil pitch; this can be explained as for larger coil pitches, the frictional surface per length is reduced, results in, lower frictional
In which, hc is the coiled wire inserted tube heat transfer coefficient. Using the present data and with the help of least squares regression analysis, the following correlation was obtained: hc = 1:6
e2 pdi
!0:065 0:86
hfc 1 + 3000Bo
+ 1:12
x 0:75 ρ f ρg 1−x
!0:41 ! : ð13Þ
The mean deviation and standard deviation of heat transfer coefficients calculated by Eq. (13) from the experimental data are 2.8% and 12.5%, respectively. The comparison between the computed heat transfer coefficient from developed correlation and the present experimental data has been shown in Fig. 8. As it is observed, most of the estimated values are within an error band of ±25% of the experimental data. So, the above mentioned correlation is in good agreement with the present experimental data. Also, uncertainty analysis done by the method described by Schultz and Cole [11] showed that the uncertainty in the determination of heat transfer coefficient is less than 10% for all test runs.
Fig. 9. Pressure gradient versus vapor quality for smooth tube and the roughed tubes with constant wire diameter of 1 mm at mass velocity of 54 kg/s m2.
1094
M.A. Akhavan-Behabadi et al. / International Communications in Heat and Mass Transfer 36 (2009) 1089–1095
Fig. 10. Pressure gradient versus vapor quality for the plain tube and the roughed tubes with constant wire diameter of 1 mm at mass velocity of 136 kg/s m2.
Fig. 12. Pressure gradient versus vapor quality for the plain tube and the roughed tubes with constant coil pitch of 10 mm at mass velocity of 114 kg/s m2.
pressure loss. The tube set ‘E’ which has the coil with the lowest pitch, increases the pressure drop by 1000% in comparison to that for the plain tube, at high vapor quality region and the mass velocities of 54 and 85 kg/s m2. Varma et al. [9] reported an increase of 1000–1200% in pressure drop compared to plain tube value in the boiling of R-22 inside coiled wire inserted tubes. Figs. 11 and 12 have also been drawn to illustrate the effect of the coiled wire inserts having different wire thicknesses and a constant coil pitch of 10 mm on the pressure drop for mass velocities of 85 kg/s m2 and 114 kg/s m2. From these figures, it is observed that the increase of wire diameter causes more pressure loss. Therefore, tube set ‘D’ which has the thickest coiled wire insert has the most pressure loss. This can be contributed to the fact that the rise in wire thickness increases the frictional surface which results in more frictional pressure loss. Also, thicker coils promote turbulence of the vapor core and liquid film which causes increased frictional pressure loss. For these coils, the momentum pressure loss is also increased because of the increased velocity resultant from reduced flow cross section area. The minimum value of pressure drop enhancement is due to tube set ‘A’ which has the lowest wire thickness. This coiled wire increases the pressure drop by 270% in comparison with the plain tube values at mass velocities of 54 and 85 kg/s m2.
3.3. The performance evaluation of coiled wire inserted tubes Analysis of the present experimental data showed that the use of coiled wire inserts inside the horizontal tubes increases the heat transfer coefficient with a pressure drop increasing penalty. The heat transfer coefficient and pressure drop are two independent parameters which do not relate by an equation, so comparison of two different cases such as two different coil pitches is difficult. Therefore, a third parameter is to be considered which relates to both of them and can produce the conditions for the comparison of these two parameters. For this purpose, Agrawal and Varma [12] proposed the ratio of pumping power to enhanced heat transfer rate or enhanced heat transfer coefficient as an alternative to be used as the criterion of performance evaluation. They calculated the increased pumping power due to the presence of twisted tape in the test evaporator by multiplying the volumetric flow rate and produced pressure drop in the test evaporator: ˙ W = vΔP:
In the present work, firstly, the ratio of the power consumption of compressor to heat transfer coefficient of plain flow (W/h)p is calculated, then the same ratio for coiled wire inserted tubes (W/h)c is computed. The compressor power is calculated from Eq. (14). Finally, the ratio from plain flow is divided by that for the flow with insert:
˙ vΔP h p
Fig. 11. Pressure gradient versus vapor quality for the plain tube and the roughed tubes with constant coil pitch of 10 mm at mass velocity of 85 kg/s m2.
ð14Þ
=
˙ vΔP h c
hc hp
ðΔPÞc ðΔPÞp
=
Rh RΔp
ð15Þ
where, Rh is the ratio of rough tube heat transfer coefficient to plain tube heat transfer coefficient and RΔp is the ratio of rough tube pressure loss to plain tube pressure loss. If Rh/RΔp is greater than one, using in-tube insert is beneficial, otherwise it is not recommended to utilize the inserts unless under specific circumstances. Fig. 13 illustrates this ratio for different mass flow rates and different coiled wire inserted tubes. From this figure, it is clear that for all the conditions the performance ratio, Rh/RΔp, is lower than one viz. the rough tubes are unfavorable. Therefore, the coiled wire inserted tubes should only been used in special cases when the compact heat exchangers are required e.g. for the scarcity of space, and more pumping power is justifiable. Akhavan-Behabadi et al. [13] also reported the same observations in their study of forced convective condensation in coiled wire inserted tubes. Comparing the different
M.A. Akhavan-Behabadi et al. / International Communications in Heat and Mass Transfer 36 (2009) 1089–1095
1095
2. The coiled wire inserts also increase the pressure drop. The evaporation pressure drop increases by as much as 1000% above that for plain flow. 3. Based on the present experimental data, a correlation was developed to predict the heat transfer coefficient during evaporation of R-134a inside a coiled wire inserted tube. 4. By considering the heat transfer and pressure drop, the plain tube has the best performance. The coiled wire inserts are only recommended under special conditions by considering the appropriate consistence between the heat transfer performance and the rate of pumping power increasing. References
Fig. 13. Ratio of Rh to RΔp for different mass velocities and various coiled wire inserted tubes.
coiled wire inserts of Fig. 13, it is revealed that at low mass velocities of 54 and 85 kg/s m2 the wire with coil pitch of 13 mm and wire thickness of 1.0 mm has a better performance relative to other wires. The reason for better performance of this wire at low mass velocities is that, at low mass flow rates the flow pattern in plain tube is separatedwavy and by insertion of coiled wire inside it, the flow regime changes to annular. This flow pattern changing is one of the most important reasons to enhance the heat transfer rates and all of the coiled wires can produce this flow pattern changing. On the other hand, the wire with coil pitch of 13 mm and wire thickness of 1.0 mm produces the least pressure drop among all other coiled wires. Thus, at low mass velocities, the tube set ‘G’ shows a better performance. From Fig. 13, it is also seen that at mass velocities of 114 and 136 kg/s m2, the best performing coiled wire insert is that having a coil pitch of 8 mm and wire thickness of 1.0 mm (tube set ‘F’) and the tube set ‘E’ which has the lowest coil pitch has the worst performance at all mass velocities. 4. Conclusions The following conclusions have been drawn from the present study: 1. The use of coiled wire inside horizontal tubes increases the evaporation heat transfer coefficient. This enhancement depends upon the test conditions and the geometry of a coiled wire insert.
[1] M.A. Akhavan-Behabadi, R. Kumar, M. Jamali, Investigation on heat transfer and pressure drop during swirl flow boiling of R-134a in a horizontal tube, Int. J. Heat Mass Transfer 52 (2009) 1918–1927. [2] R.S. Reid, M.B. Pate, A.E. Bergles, A comparison of augmentation techniques during in-tube evaporation of R-113, J. Heat Transfer 113 (1991) 451–458. [3] Waldemar Targanski, Janusz T. Cieslinski, Evaporation of R407C/oil mixtures inside corrugated and micro-fin tubes, Appl. Therm. Eng. 27 (2007) 2226–2232. [4] R.L. Webb, Principles of Enhanced Heat Transfer, Wiley, New York, 1994, pp. 3–11. [5] J. Lan, P.J. Disimile, J. Weisman, Two-phase flow patterns and boiling heat transfer in tubes containing helical wire inserts – part I – flow patterns and boiling heat transfer coefficients, J. Enhanc. Heat Transf. 4 (1997) 269–282. [6] American Society of Heating Refrigerating and Air-conditioning Engineers Handbook, Fundamentals, 2005, pp. 20–17. [7] K.E. Gungor, R.H. Winterton, Simplified general correlation for saturated flow boiling and comparison of correlations to data, Chem. Eng. Res. Des. 65 (1987) 148–156. [8] M.A. Akhavan-Behabadi, M.R. Salimpoor, R. Kumar, K.N. Agrawal, Augmentation of forced convection condensation heat transfer inside a horizontal tube using spiral spring inserts, J. Enhanc. Heat Transf. 12 (2005) 373–384. [9] H.K. Varma, K.N. Agrawal, M.L. Bansal, Turbulence promoters in horizontal R22 evaporators, AIRAH J. (Aust.) 46 (1992) 21–30. [10] R. Yun, J. Hwang, J.T. Chung, Y. Kim, Flow boiling heat transfer characteristics of nitrogen in plain and wire coil inserted tubes, Int. J. Heat Mass Transfer 50 (2007) 2339–2345. [11] R.R. Schultz, R. Cole, Uncertainty analysis in boiling nucleation, AICHE Symp. Ser. 75 (189) (1979) 32–38. [12] K.N. Agrawal, H.K. Varma, Experimental study of heat transfer augmentation versus pumping power in a horizontal R-12 evaporator, Int. J. Refrig. 14 (1991) 273–281. [13] M.A. Akhavan-Behabadi, M.R. Salimpour, V.A. Pazouki, Pressure drop increase of forced convective condensation inside horizontal coiled wire inserted tubes, Int. Commun. Heat Mass Transf. 35 (2008) 1220–1226.
Contents lists available at ScienceDirect
International Communications in Heat and Mass Transfer j o u r n a l h o m e p a g e : w w w. e l s ev i e r. c o m / l o c a t e / i c h m t
Heat transfer and pressure drop characteristics of forced convective evaporation in horizontal tubes with coiled wire inserts☆ M.A. Akhavan-Behabadi ⁎, S.G. Mohseni, H. Najafi, H. Ramazanzadeh School of Mechanical Engineering, University College of Engineering, University of Tehran, Tehran, Iran
a r t i c l e
i n f o
Available online 12 August 2009 Keywords: Heat transfer Evaporation Pressure drop R-134a Coiled wire
a b s t r a c t Heat transfer enhancement and pressure drop increasing during evaporation of R-134a due to the presence of coiled wire insert inside a horizontal evaporator was studied experimentally. The test evaporator was an electrically heated copper tube of 1200 mm length and 7.5 mm inside diameter. Helically wire coils with different wire diameters of 0.5, 0.7, 1.0 and 1.5 mm and different coil pitches of 5, 8, 10 and 13 mm were made and used in full length of the test evaporator. For each inserted tube and also the plain tube, several test runs were carried out with different mass velocities and heat fluxes. From analysis of acquired data, it was found that the coiled wire inserts enhance the heat transfer coefficient but with a higher penalty due to the increasing of pressure drop, in comparison to that for the plain flow. An empirical correlation has been developed to predict the heat transfer coefficient during evaporation inside a horizontal tube in the presence of a coiled wire insert. © 2009 Elsevier Ltd. All rights reserved.
1. Introduction The limited availability of the world's conventional energy resources and the ever increasing cost of energy over the last several years has accelerated research in the field of energy conservation with the aim of possible reduction in its consumption in a given process. In the heat transfer field, extensive efforts are made to develop compact and more efficient heat exchangers. Evaporator is an important and widely used heat exchanger in air conditioning and refrigeration industries. Different methods have been used by investigators to increase the heat transfer rates in evaporators [1–3]. Webb [4] summarized various kinds of heat transfer enhancement techniques and has classified them into two main categories, viz., active technique and passive technique. The use of various types of insertions, such as, wire mesh or brush, twisted tape and coil or spiral spring, is considered as the passive techniques of enhancing the heat transfer, which does not require any external energy. The internal fin tubes are also used to enhance the heat transfer coefficient but they are expensive in comparison to the in-tube inserts. Moreover, inserts are easy to install inside the tubes. It should be also noted that, the pressure drop is also usually increased by using these techniques. The main objective of all the research work in this field is to obtain the highest heat transfer coefficient with the least pressure drop. A review of the existing literature, revealed that the use of coiled wire inserts to enhance the heat transfer rates during evaporation of
☆ Communicated by W.J. Minkowycz. ⁎ Corresponding author. E-mail address: [email protected] (M.A. Akhavan-Behabadi). 0735-1933/$ – see front matter © 2009 Elsevier Ltd. All rights reserved. doi:10.1016/j.icheatmasstransfer.2009.07.009
refrigerants flowing inside a horizontal tube, has been studied for quite some time [5]. Furthermore, the refrigerant R-134a, nowadays, is seen to be the eco-friendly substitute of R-12. It has zero ozone depletion potential (ODP) and 10% global warming potential (GWP) in comparison to R-12. Therefore, an experimental investigation was carried out to study the effects of wire coil inserts on evaporation heat transfer and pressure drop characteristics of R-134a in a horizontal tube. Finally, the performance evaluation of the wire coil inserts from the point view of heat transfer enhancement and pressure drop increasing should be done. 2. Experimental apparatus and procedure The test apparatus was a well instrumented vapor compression refrigeration system. The schematic diagram of experimental apparatus is shown in Fig. 1. It consisted of a compressor, a condenser, an expansion valve, a pre-evaporator, a test evaporator or test section, an after-evaporator, and necessary instruments for measurements and controls. The test evaporator was a copper tube of 7.5 mm inside diameter, di, 9.54 mm outside diameter, do, and 1200 mm length, L. Refrigerant R-134a flowing inside the test-evaporator tube was heated by a flexible electrical heating tape of 2 kW capacity, wrapped around it. During the experiments, a short range of vapor quality could be achieved in a single test run. Therefore, to cover the entire ranges of vapor qualities a pre-evaporator was installed before the test evaporator. By regulating the voltage to the electrical coil around the pre-evaporator, the vapor quality at the inlet of the test evaporator was controlled. It was ensured that the refrigerant coming from the test evaporator was superheated before it entered the compressor. This was attained by installing an after-evaporator
1090
M.A. Akhavan-Behabadi et al. / International Communications in Heat and Mass Transfer 36 (2009) 1089–1095
Nomenclature Bo C1,C2 cp d e g G h h̅ L ṁ P p q Q Rh RΔp T k x
boiling number, q/Gλ constants of Eq. (11) specific heat, kJ/kg K diameter, m wire diameter, m gravitational acceleration, m/s2 mass velocity, kg/s m2 heat transfer coefficient, W/m2 K average heat transfer coefficient, W/m2 K tube length, m mass flow rate, kg/s pressure, kPa coil pitch, m heat flux, kW/m2 electrical work on evaporator, W ratio of heat transfer coefficient of rough tube to plain tube ratio of pressure loss of rough tube to plain tube temperature, K thermal conductivity, W/m K vapor quality
Greek symbols α helix angle of coiled wire, degree λ latent heat, kJ/kg Δp pressure drop, kPa ΔT temperature difference, K ρ density, kg/m3
Different types of full length wire coils were inserted, one by one, having different wire diameters of 0.5, 0.7, 1.0 and 1.5 mm and different coil pitches of 5, 8, 10 and 13 mm. Table 1 represents the specification of the used coiled wire inserted tubes. After collecting experimental data for plain flow (tube set H), the data were acquired for the all the roughed tubes (coiled wire inserted tubes) given in Table 1. The data were acquired at refrigerant mass velocities of 54, 85, 114 and 136 kg/s m2. All together 194 test runs were performed. The ranges of operating parameters are as follows: Refrigerant mass velocity: 54–136 kg/s m2 Heat flux: 1.8–5.3 kW/m2 Boiling temperature: − 3 to − 19 °C Inlet vapor quality: 0.2–0.9 Exit vapor quality: 0.3–1.0 Liquid Reynolds number: 1250–3500 Twist ratio (p/di): 0.66–1.73. The thermo physical properties of R-134a were taken from [6]. The vapor quality at the inlet of pre-evaporator was computed by considering the iso-enthalpic expansion of refrigerant in the needle valve. An energy balance technique was used to get the vapor quality of the inlet and the exit of the test evaporator. The mean vapor quality was taken as the average of inlet and outlet vapor qualities of the test section. The heat transfer coefficient of the test section was determined using the heat gain from the electrical heater and the temperature difference between the evaporator inside wall surface and the boiling refrigerant. The average outside tube-wall temperature of the test evaporator at a particular station, Tws, was calculated by using the following equation: Tws =
Subscripts b bottom c wire coiled inserted tube e equivalent f liquid g vapor i inside tube o outside tube p plain tube s saturated, side t top w tube wall wi inside wall wo outside wall
ð1Þ
The average outside tube-wall temperature of the test evaporator, Two, was computed by taking the average temperatures of six axial stations as: 6
∑Tws Two =
1
6
:
ð2Þ
The radial heat flux, q, for the test evaporator was calculated by: q = Q =ðπdi LÞ:
ð3Þ
The temperature drop across the tube wall, ΔTw, was evaluated by: ΔTw =
downstream of the test evaporator. All the three evaporators were thermally insulated. An accumulator was also provided in the suction line of the compressor. The refrigerant mass flow rate was measured by a pre-calibrated rotameter installed downstream of the condenser. The outside wall temperatures of the test-evaporator tube were measured at six axial locations. At each location the temperature of the tube was measured at the top, both sides in the middle of the tube and bottom positions. The refrigerant temperatures at the inlet and outlet of the test evaporator were also measured. Fig. 2 shows the schematic diagram of the test section. All the temperature measurements were done by T-type thermocouples with a calibrated accuracy of 0.1 °C. To measure the pressure drop across the test evaporator, a differential pressure transducer apparatus was utilized. This apparatus was calibrated to measure the pressure loss in the ranges of 0 to 150 kPa. All other necessity parameters were also measured by accurate instruments.
Tt + 2Ts + Tb : 4
qdi lnðdo = di Þ : 2kw
ð4Þ
The average inside tube-wall temperature, Twi, was calculated by subtracting the temperature drop across the wall, ΔTw, from measured average outside temperature, Two. Twi = Two −ΔTw :
ð5Þ
The average static pressure in the test evaporator was taken to be the mean of the inlet and outlet pressure. The vapor saturation temperature in the test evaporator, Ts, was taken as the saturation temperature corresponding to this average static pressure. The heat transfer coefficient of the test evaporator was calculated by the following equation: h=
q : ðTwi −Ts Þ
ð6Þ
M.A. Akhavan-Behabadi et al. / International Communications in Heat and Mass Transfer 36 (2009) 1089–1095
1091
Fig. 1. Schematic diagram of experimental set-up.
3. Results and discussion 3.1. Heat transfer The Gungor and Winterton correlation for estimation of heat transfer coefficient during flow boiling inside the horizontal plain tube is as follows [7]: " 0:86
hp = hfp 1 + 3000Bo
x 0:75 ρ f + 1:12 ρg 1−x
!0:41 # ð7Þ
where, Gð1−xÞdi 0:8 0:4 kf : Prf hfp = 0:023 μf di
ð8Þ
The above mentioned correlation predicts the plain flow experimental data within an error band of −14% to +9%. This establishes the integrity of experimental apparatus. The results of plain flow are also used in comparing the performance of different coiled wire inserts. The variation of heat transfer coefficient with vapor quality for plain tube and the tubes with coiled wire inserts has been shown in Figs. 3–6. The effects of coiled wire inserts having different coil pitches and a constant wire thickness of 1 mm are shown in Figs. 3 and 4 for the mass velocities of 54 kg/m2 s and 136 kg/s m2, while the effects of coiled wire inserts having different wire thicknesses and a constant coil pitch of 10 mm on the enhancement of heat transfer coefficients are shown in Figs. 5 and 6 for mass velocities of 85 kg/s m2 and 114 kg/s m2. From these figures it is observed that, the insertion of a coiled wire inside the test-evaporator tube has produced higher heat transfer coefficient compared to plain tube value. The magnitude of the augmentation is not consistent for a special coiled wire and it is a complex function of mass velocity, vapor quality, and the geometry of a coil. For all the flow rates the best performing coiled wire insert is that having a coil pitch of 10 mm and wire thickness of 1.5 mm (the thickest coiled wire insert). In the high vapor quality region the insert
Table 1 Characteristic parameters of the coiled wire inserted tubes.
Fig. 2. Details of the test section.
Tube set
di (mm)
de (mm)
e (mm)
p (mm)
α (degree)
A B C D E F G H
7.5 7.5 7.5 7.5 7.5 7.5 7.5 7.5
6.44 6.09 5.61 4.95 4.48 5.29 5.94 7.5
0.5 0.7 1.0 1.5 1.0 1.0 1.0
10 10 10 10 5 8 13 Plain tube
73 73 73 73 81.5 76.6 69
1092
M.A. Akhavan-Behabadi et al. / International Communications in Heat and Mass Transfer 36 (2009) 1089–1095
Fig. 3. Heat transfer coefficient versus vapor quality for the plain tube and the roughed tubes with constant wire diameter of 1 mm at mass velocity of 54 kg/s m2.
Fig. 5. Heat transfer coefficient versus vapor quality for the plain tube and the roughed tubes with constant coil pitch of 10 mm at mass velocity of 85 kg/s m2.
enhances the heat transfer coefficient in a range of 72 to 98% in comparison to that for the plain flow. For the same insert at vapor quality, x, of 0.2 the enhancement in heat transfer coefficient is in a range of 0 to 49% and at the vapor quality, x, of 0.8 the enhancement in heat transfer coefficient is in a range of 74 to 88%. Akhavan-Behabadi et al. [8] investigated the effect of coiled wire inserts during forced convective condensation of R-134a inside a 10.7-mm I.D. horizontal copper tube. They reported a maximum of 85% heat transfer enhancement for coiled wire having a coil pitch of 10 mm and wire thickness of 1.5 mm. Since, the present investigation has been carried out for the evaporation of R-134a, the highest enhancement in heat transfer coefficient is 98%. From Figs. 3 and 4, it is seen that, the heat transfer coefficient increases with the decrease of coil pitch, so that the highest heat transfer coefficient is achieved by the coil with the pitch of 8 mm (tube set ‘F’); further reduction in the coil pitch reduces the heat transfer coefficient. This trend had been observed previously and was explained by this fact that in coil pitches smaller than a particular value, coil resists to heat transfer like a solid shell [9]. Also, it is revealed from Figs. 5 and 6 that in the high vapor quality region, heat transfer coefficient increases with the increase of the wire thickness but in the low vapor quality region,
although the maximum heat transfer coefficient is related to the thickest coiled wire this trend is not general. Yun et al. [10] also reported the same observations in their study of flow boiling heat transfer characteristics of nitrogen in coiled wire inserted tubes. In Fig. 7, the variation of heat transfer coefficient with vapor quality for different mass velocities of R-134a is shown for the tube having a coiled wire insert of 0.5 mm diameter and 10 mm pitch. It is observed that for a given vapor quality, the heat transfer coefficient is increased with the increase of vapor mass velocity. This phenomenon is simply contributed to the fact that at higher mass velocities, the Reynolds number is also higher, which in turn, increases the Nusselt number and, hence, the heat transfer coefficient. In addition, as the mass velocity increased, the swirl flow became much stronger and the dryout vapor quality increased. The same observation has also been made by Yun et al. [10]. The coiled wire insert of 0.5-mm diameter wire and 10 mm pitch has enhanced the average heat-transfer coefficient, h ,̅ in a range of 11– 20% in comparison to that for the plain flow. Literature survey showed that there is no correlation to predict the heat transfer coefficient during forced convective boiling of R-134a in spring inserted tubes. Therefore, a new correlation was developed to predict the heat transfer coefficient in coiled wire inserted tubes. For
Fig. 4. Heat transfer coefficient versus vapor quality for the plain tube and the roughed tubes with constant wire diameter of 1 mm at mass velocity of 136 kg/s m2.
Fig. 6. Heat transfer coefficient versus vapor quality for the plain tube and the roughed tubes with constant coil pitch of 10 mm at mass velocity of 114 kg/s m2.
M.A. Akhavan-Behabadi et al. / International Communications in Heat and Mass Transfer 36 (2009) 1089–1095
Fig. 7. Heat transfer coefficient versus vapor quality for tube set “A” at different mass velocities.
this purpose, Eq. (7) was used as a base equation and three additional parameters are added to it, i.e., ratio of the wire diameter to the tube inside diameter (e/di), the ratio of the coil pitch to tube inside diameter (p/di), and equivalent diameter (de). The equivalent diameter was calculated by using Eq. (9): 2
de = ðdi −eγÞ = ðdi + γÞ
ð9Þ
where, γ = πeðdi −eÞ = ðp sin αÞ:
ð10Þ
These variables have also been used by Akhavan-Behabadi et al. [8]. The following form of the functional relationship was developed as follows: 2
hc = C1
e pdi
!C
2
0:86
hfc 1 + 3000Bo
+ 1:12
x 0:75 ρ f ρg 1−x
!0:41 ! ð11Þ
where, Gð1−xÞde 0:8 0:4 kf : Prf hfc = 0:023 μL de
ð12Þ
1093
Fig. 8. Comparison of the experimental heat transfer coefficient with that predicted by Eq. (13).
3.2. Pressure drop Since insertion of coiled wire in the plain tube also increases the pressure drop across the test evaporator, the effect of wire geometry on pressure drop rates is also required to be investigated. The variation of pressure drop with vapor quality for the plain tube and the tubes with coiled wire inserts at different mass velocities has been shown in Figs. 9–12. It is found that, in general, by insertion of coiled wire inside horizontal tubes, the pressure drop is increased in comparison to that of plain flow. It is also observed that, the pressure drop is increased with the rise of vapor quality. This is due to the fact that, as the quality of vapor inside the tube grows; the velocity of fluid increases which in its own turn leads in more shear stress and more pressure loss. Furthermore, it is also observed that for a given vapor quality, the pressure drop is reduced with the decrease in mass velocity. The effects of the coiled wire inserts having different pitches and a constant wire thickness of 1.0 mm on the pressure drop are shown in Figs. 9 and 10 for mass velocities of 54 kg/s m2 and 136 kg/s m2. It is seen that, the lower pressure drop is achieved for the tube set with larger coil pitch; this can be explained as for larger coil pitches, the frictional surface per length is reduced, results in, lower frictional
In which, hc is the coiled wire inserted tube heat transfer coefficient. Using the present data and with the help of least squares regression analysis, the following correlation was obtained: hc = 1:6
e2 pdi
!0:065 0:86
hfc 1 + 3000Bo
+ 1:12
x 0:75 ρ f ρg 1−x
!0:41 ! : ð13Þ
The mean deviation and standard deviation of heat transfer coefficients calculated by Eq. (13) from the experimental data are 2.8% and 12.5%, respectively. The comparison between the computed heat transfer coefficient from developed correlation and the present experimental data has been shown in Fig. 8. As it is observed, most of the estimated values are within an error band of ±25% of the experimental data. So, the above mentioned correlation is in good agreement with the present experimental data. Also, uncertainty analysis done by the method described by Schultz and Cole [11] showed that the uncertainty in the determination of heat transfer coefficient is less than 10% for all test runs.
Fig. 9. Pressure gradient versus vapor quality for smooth tube and the roughed tubes with constant wire diameter of 1 mm at mass velocity of 54 kg/s m2.
1094
M.A. Akhavan-Behabadi et al. / International Communications in Heat and Mass Transfer 36 (2009) 1089–1095
Fig. 10. Pressure gradient versus vapor quality for the plain tube and the roughed tubes with constant wire diameter of 1 mm at mass velocity of 136 kg/s m2.
Fig. 12. Pressure gradient versus vapor quality for the plain tube and the roughed tubes with constant coil pitch of 10 mm at mass velocity of 114 kg/s m2.
pressure loss. The tube set ‘E’ which has the coil with the lowest pitch, increases the pressure drop by 1000% in comparison to that for the plain tube, at high vapor quality region and the mass velocities of 54 and 85 kg/s m2. Varma et al. [9] reported an increase of 1000–1200% in pressure drop compared to plain tube value in the boiling of R-22 inside coiled wire inserted tubes. Figs. 11 and 12 have also been drawn to illustrate the effect of the coiled wire inserts having different wire thicknesses and a constant coil pitch of 10 mm on the pressure drop for mass velocities of 85 kg/s m2 and 114 kg/s m2. From these figures, it is observed that the increase of wire diameter causes more pressure loss. Therefore, tube set ‘D’ which has the thickest coiled wire insert has the most pressure loss. This can be contributed to the fact that the rise in wire thickness increases the frictional surface which results in more frictional pressure loss. Also, thicker coils promote turbulence of the vapor core and liquid film which causes increased frictional pressure loss. For these coils, the momentum pressure loss is also increased because of the increased velocity resultant from reduced flow cross section area. The minimum value of pressure drop enhancement is due to tube set ‘A’ which has the lowest wire thickness. This coiled wire increases the pressure drop by 270% in comparison with the plain tube values at mass velocities of 54 and 85 kg/s m2.
3.3. The performance evaluation of coiled wire inserted tubes Analysis of the present experimental data showed that the use of coiled wire inserts inside the horizontal tubes increases the heat transfer coefficient with a pressure drop increasing penalty. The heat transfer coefficient and pressure drop are two independent parameters which do not relate by an equation, so comparison of two different cases such as two different coil pitches is difficult. Therefore, a third parameter is to be considered which relates to both of them and can produce the conditions for the comparison of these two parameters. For this purpose, Agrawal and Varma [12] proposed the ratio of pumping power to enhanced heat transfer rate or enhanced heat transfer coefficient as an alternative to be used as the criterion of performance evaluation. They calculated the increased pumping power due to the presence of twisted tape in the test evaporator by multiplying the volumetric flow rate and produced pressure drop in the test evaporator: ˙ W = vΔP:
In the present work, firstly, the ratio of the power consumption of compressor to heat transfer coefficient of plain flow (W/h)p is calculated, then the same ratio for coiled wire inserted tubes (W/h)c is computed. The compressor power is calculated from Eq. (14). Finally, the ratio from plain flow is divided by that for the flow with insert:
˙ vΔP h p
Fig. 11. Pressure gradient versus vapor quality for the plain tube and the roughed tubes with constant coil pitch of 10 mm at mass velocity of 85 kg/s m2.
ð14Þ
=
˙ vΔP h c
hc hp
ðΔPÞc ðΔPÞp
=
Rh RΔp
ð15Þ
where, Rh is the ratio of rough tube heat transfer coefficient to plain tube heat transfer coefficient and RΔp is the ratio of rough tube pressure loss to plain tube pressure loss. If Rh/RΔp is greater than one, using in-tube insert is beneficial, otherwise it is not recommended to utilize the inserts unless under specific circumstances. Fig. 13 illustrates this ratio for different mass flow rates and different coiled wire inserted tubes. From this figure, it is clear that for all the conditions the performance ratio, Rh/RΔp, is lower than one viz. the rough tubes are unfavorable. Therefore, the coiled wire inserted tubes should only been used in special cases when the compact heat exchangers are required e.g. for the scarcity of space, and more pumping power is justifiable. Akhavan-Behabadi et al. [13] also reported the same observations in their study of forced convective condensation in coiled wire inserted tubes. Comparing the different
M.A. Akhavan-Behabadi et al. / International Communications in Heat and Mass Transfer 36 (2009) 1089–1095
1095
2. The coiled wire inserts also increase the pressure drop. The evaporation pressure drop increases by as much as 1000% above that for plain flow. 3. Based on the present experimental data, a correlation was developed to predict the heat transfer coefficient during evaporation of R-134a inside a coiled wire inserted tube. 4. By considering the heat transfer and pressure drop, the plain tube has the best performance. The coiled wire inserts are only recommended under special conditions by considering the appropriate consistence between the heat transfer performance and the rate of pumping power increasing. References
Fig. 13. Ratio of Rh to RΔp for different mass velocities and various coiled wire inserted tubes.
coiled wire inserts of Fig. 13, it is revealed that at low mass velocities of 54 and 85 kg/s m2 the wire with coil pitch of 13 mm and wire thickness of 1.0 mm has a better performance relative to other wires. The reason for better performance of this wire at low mass velocities is that, at low mass flow rates the flow pattern in plain tube is separatedwavy and by insertion of coiled wire inside it, the flow regime changes to annular. This flow pattern changing is one of the most important reasons to enhance the heat transfer rates and all of the coiled wires can produce this flow pattern changing. On the other hand, the wire with coil pitch of 13 mm and wire thickness of 1.0 mm produces the least pressure drop among all other coiled wires. Thus, at low mass velocities, the tube set ‘G’ shows a better performance. From Fig. 13, it is also seen that at mass velocities of 114 and 136 kg/s m2, the best performing coiled wire insert is that having a coil pitch of 8 mm and wire thickness of 1.0 mm (tube set ‘F’) and the tube set ‘E’ which has the lowest coil pitch has the worst performance at all mass velocities. 4. Conclusions The following conclusions have been drawn from the present study: 1. The use of coiled wire inside horizontal tubes increases the evaporation heat transfer coefficient. This enhancement depends upon the test conditions and the geometry of a coiled wire insert.
[1] M.A. Akhavan-Behabadi, R. Kumar, M. Jamali, Investigation on heat transfer and pressure drop during swirl flow boiling of R-134a in a horizontal tube, Int. J. Heat Mass Transfer 52 (2009) 1918–1927. [2] R.S. Reid, M.B. Pate, A.E. Bergles, A comparison of augmentation techniques during in-tube evaporation of R-113, J. Heat Transfer 113 (1991) 451–458. [3] Waldemar Targanski, Janusz T. Cieslinski, Evaporation of R407C/oil mixtures inside corrugated and micro-fin tubes, Appl. Therm. Eng. 27 (2007) 2226–2232. [4] R.L. Webb, Principles of Enhanced Heat Transfer, Wiley, New York, 1994, pp. 3–11. [5] J. Lan, P.J. Disimile, J. Weisman, Two-phase flow patterns and boiling heat transfer in tubes containing helical wire inserts – part I – flow patterns and boiling heat transfer coefficients, J. Enhanc. Heat Transf. 4 (1997) 269–282. [6] American Society of Heating Refrigerating and Air-conditioning Engineers Handbook, Fundamentals, 2005, pp. 20–17. [7] K.E. Gungor, R.H. Winterton, Simplified general correlation for saturated flow boiling and comparison of correlations to data, Chem. Eng. Res. Des. 65 (1987) 148–156. [8] M.A. Akhavan-Behabadi, M.R. Salimpoor, R. Kumar, K.N. Agrawal, Augmentation of forced convection condensation heat transfer inside a horizontal tube using spiral spring inserts, J. Enhanc. Heat Transf. 12 (2005) 373–384. [9] H.K. Varma, K.N. Agrawal, M.L. Bansal, Turbulence promoters in horizontal R22 evaporators, AIRAH J. (Aust.) 46 (1992) 21–30. [10] R. Yun, J. Hwang, J.T. Chung, Y. Kim, Flow boiling heat transfer characteristics of nitrogen in plain and wire coil inserted tubes, Int. J. Heat Mass Transfer 50 (2007) 2339–2345. [11] R.R. Schultz, R. Cole, Uncertainty analysis in boiling nucleation, AICHE Symp. Ser. 75 (189) (1979) 32–38. [12] K.N. Agrawal, H.K. Varma, Experimental study of heat transfer augmentation versus pumping power in a horizontal R-12 evaporator, Int. J. Refrig. 14 (1991) 273–281. [13] M.A. Akhavan-Behabadi, M.R. Salimpour, V.A. Pazouki, Pressure drop increase of forced convective condensation inside horizontal coiled wire inserted tubes, Int. Commun. Heat Mass Transf. 35 (2008) 1220–1226.