- Email: [email protected]
Experimental study of forced convective heat transfer in a different arranged corrugated channel
International Communications in Heat and Mass Transfer 46 (2013) 106–111
Contents lists available at SciVerse ScienceDirect
International Communications in Heat and Mass Transfer journal homepage: www.elsevier.com/locate/ichmt
Experimental study of forced convective heat transfer in a different arranged corrugated channel☆ H. Pehlivan, I. Taymaz ⁎, Y. İslamoğlu Sakarya University, Mechanical Engineering Department, Sakarya, Turkey
a r t i c l e
i n f o
Available online 2 June 2013 Keywords: Corrugated channel Convective heat transfer Wavy geometry
a b s t r a c t In this study, heat transfer rate for sinusoidal corrugated channel has been experimentally investigated. Three different type sharp corrugation peak fins and a plain surface were used in the experiment. Results were carried out for constant heat flux of 616 W/m2, varied Reynolds number Re 1500 to 8000 for the corrugation angle (27, 50 and 22/60°) and channel height of 5 and 10 mm. Nusselt number (Nu), convection heat transfer coefficient (h), Colburn factor (j) and enhancement ratio (E) against Reynolds number (Re) have been studied. The effects of the wavy geometry and channel height have been discussed. The increase of corrugated angle gave rise to a heat transfer rate. © 2013 Elsevier Ltd. All rights reserved.
1. Introduction In recent years, due to the increasing demand by industries for heat exchangers that are more efficient, compact and less expensive, heat transfer enhancement has gained great momentum. For this purpose, two techniques have been identified: passive and active [1]. Convective heat transfer can be enhanced passively by changing flow geometry [2]. The symmetrical converging–diverging channel is one of several devices employed for enhancing the heat and mass transfer [3]. Enhancement techniques based on artificial roughness are used in numerous applications [4]. Rough walls exist in all flow systems, where they may lead to either deterioration or improvement of the desired functionality. Wall roughness can be increased to promote mixing of the fluid or reduced to eliminate flow disturbances [5]. In view of their turbulence promoting ability, corrugated surfaces have been used widely to improve the rate of forced convection heat and mass transfer in equipment such as heat exchangers, solar collectors, electrochemical and catalytic reactors and membrane processes [6]. 2. Background Many experimental or numerical studies for fluid flow and heat transfer in corrugated channels have been done. Experimentally: Chang and Huang [7] studied the properties of heat transfer, pressured drop and thermal performance factor of a corrugated wavy channel with Re numbers from 350 to 5500. Zhang and Chen [8] used crosscorrugated triangular duct under uniform heat flux boundary conditions
☆ Communicated by W.J. Minkowycz. ⁎ Corresponding author. E-mail address: [email protected] (I. Taymaz). 0735-1933/$ – see front matter © 2013 Elsevier Ltd. All rights reserved. http://dx.doi.org/10.1016/j.icheatmasstransfer.2013.05.016
in the Re range of 500–5000. Faizal and Ahmed [9] investigated the channel height between the plates to determine the configuration that gives the optimum heat transfer. Three different heights, 6 mm, 9 mm, and 12 mm, were used. The optimum heat transfer between the two streams is obtained for the minimum spacing of 6 mm. Chang et al. [10] examined the detailed Nusselt number (Nu) distributions, pressure drop coefficients (f) and thermal performance factors (g) for two furrowed rectangular channels with transverse and skewed sinusoidal wavy walls. García et al. [11] analyzed the thermal-hydraulic behavior of three types of enhancement techniques based on artificial roughness: corrugated tubes, dimpled tubes and wire coils. Pirompugd et al. [12] presented the heat and mass transfer characteristics of 18 wavy fin-and-tube heat exchangers under dehumidifying conditions. Naphon [13] researched heat transfer characteristics and pressure drop in the channel with V corrugated upper and lower plates under constant heat flux. Corrugated plates with three different corrugated tile angles of 20°, 40°and 60° are tested. The experiments are performed for the Reynolds number and heat flux in the ranges of 2000–9000 and 0.5–1.2 kW/m2, respectively. Nilpueng and Wongwises [14] studied flow patterns and pressure drop of upward liquid single-phase flow and air–water two-phase flow in sinusoidal wavy channels. Different phase shifts between the side walls of the wavy channel of 0°, 90° and 180° are investigated. The flow phenomena, which are bubbly flow, slug flow, churn flow, and dispersed bubbly flow are observed and recorded by high-speed camera. Pressure drop characteristics of flow in a periodically grooved channel are investigated by Adachi et al. [15]. The Reynolds number regime was 100 b Re b 1000. A novel heat transfer enhancement roughness with V-shaped ribs and deepened scales is devised by Chang et al. [16]. Performances of heat transfer and pressure drop in a rectangular channel were examined for both forward and backward flows in the Re range of 1000–30,000. Solar air heaters were analytically and experimentally studied El-Sebaii et al. [17] and Gao et al. [18]
H. Pehlivan et al. / International Communications in Heat and Mass Transfer 46 (2013) 106–111
Nomenclature Acycle cp Dh E f H h j k L _ m Nu Pr Qcycle P Re S T V W
convection heat transfer area per cycle, m2 specific heat, kJ/kg·K hydraulic diameter, m enhancement ratio, fully developed friction factor, fin channel height, m cycle average fully developed heat transfer coefficient, W/m2·K Colburn factor, thermal conductivity, W/m·K length, m mass flow rate, kg/s cycle average fully developed Nusselt number, Prandtl number, heat transferred to fluid per cycle, W Pressure, Pa Reynolds number, pitch (axial length of cycle), m temperature, °C mean velocity, m/s width, m
Greek symbols ρ density, kg/m3 υ kinematic viscosity, m2/s θ corrugation angle, °
Subscript av b fd min max p w
quantity evaluated at average cross-section fluid bulk fully developed minimum, maximum, plain surface, wall.
using the double pass flat and V-corrugated plate and cross-corrugated solar air heaters, respectively. Various numerical studies and approaches to the flow and heat transfer characteristics in corrugated or wavy ducts are conducted. Sawyers et al. [19] studied the effect of three dimensional hydrodynamics on the enhancement of steady, laminar heat transfer in corrugated channels. Reynolds numbers are considered in the range of 0 b Re b 250 to avoid unsteady flow. Hwang et al. [20] investigated the flow and local mass transfer characteristics inside a corrugated duct. The flow visualization technique and a numerical analysis using a commercial code, FLUENT, are used to understand the overall flow structure inside the duct. The corrugation angle of the wavy duct is 145°and the duct aspect ratio is 7.3. The Reynolds numbers, based on the duct hydraulic diameter, vary from 100 to 5000. Fabbri [21] analyzed the heat transfer in a channel composed of a smooth and a corrugated wall under laminar flow conditions. The velocity and temperature distributions are determined with the help of a finite element model. Desrues et al. [22] calculated Nusselt numbers and friction factor of a ribbed channel for Reynolds numbers from 75 to 2000 and geometric parameters, by the mean of three-dimensional incompressible computational fluid dynamics. The geometry is a rectangular duct with stream wise-periodic transverse rectangular ribs, alternated on its two smaller walls. The aim of this study, which
107
includes heat transfer rates for three different types of sharp corrugation peak fins and, sinusoidal converging–diverging channel, will be experimentally investigated. 3. Experimental program 3.1. Experimental apparatus A test ring is used to perform heat transfer and pressure drop experiments as shown in Fig. 1. In the experiment, three different types corrugated channels and one plain channel has been used. Their schematic and geometrical configurations are shown in Fig. 2 and Table 1, respectively. The test section is assembled with two copper plates of 10 mm thickness. The plate geometry is 50 × 278 mm. A forced-convection air flow was used for the experiments and equal power input for heated length was established in both top and bottom walls. To measure the entering bulk temperature of the air, two K type thermocouples were positioned at the inlet of the duct. Also, to measure the temperature distribution of each corrugated walls, six K type thermocouples are used. The mass flow rate of air flowing through the systems was measured by the orifice. Two static pressure taps were located at the valleys of the top to measure axial pressure drops, used to evaluate average friction factor and the pressure drops were measured by a U-manometer. The test section is also isolated to avoid thermal losses (Table 3). 3.2. Data collection The experimental procedure involved adjusting the flow rate to the desired value. After the exhaust fan was turned on and the desired Reynolds number was obtained, the power input of plate heaters gradually increased and maintained at 616 W/m2 to provide sufficient measurement while the fluid property varies. The supplied heat into the corrugated walls was adjusted to achieve the desired level by using electric heaters, which are 2 mm thick, 50 mm wide and 278 mm long. They were located to the back of each plate. The voltage and current of electric input to the plate type heaters were controlled by a DC power supply unit. Temperatures were recorded at intervals of 15 min until a steady state was reached. Steady-state conditions were assumed to prevail when the temperature measurement on the plates and fluid inlet were within ± 0.2 °C. For a duct of periodic geometry, the heat transfer coefficient may vary axially in the thermally developed regions, but the same value of the coefficient recurs periodically at successive axial stations separated by an axial distance S, and the cycle-average heat transfer coefficient is the same from cycle to cycle [23]. The schematic diagram of the test section for which the details were given in the overview of the measurement section including the position where temperature and pressure are measured is shown in Fig. 3 and Table 2. 4. Heat transfer works The fully developed condition has been verified experimentally in this study by observing the temperature distribution. The slope of these straight lines is dT Q cycle : ¼ _ p dX Smc
ð1Þ
For constant surface heat flux, the axial temperature variation (as seen in Fig. 4) of the fluid is
Tb ðXÞ ¼ Tb þ
Q cycle : _ p Smc
ð2Þ
108
H. Pehlivan et al. / International Communications in Heat and Mass Transfer 46 (2013) 106–111
Fig. 1. Experimental setup (1. Airline and pressure measurement line, 2. Orifice, 3. Upstream plenum, 4. Isolated test section and thermocouple connection points, 5. Downstream plenum, 6. Valve, 7. Exhaust Fan, 8. Digital Scanner, 9. Multi meter, 10. Power, 11. Ice Bath).
Fig. 2. The wall with sharp corrugation peak.
The cycle-average full development heat transfer coefficients were evaluated from the measured temperatures and heat inputs
h¼
Q cycle : ðTw −Tb Þfd Acycle
ð3Þ
The enhancement ratio E, the ratio of the (hA) of an enhanced surface to that of a plain surface (hA)P is defined by Webb [25] as: E¼
hDh k
ð4Þ
Re ¼
VDh : υ
ð5Þ
The hydraulic diameter is defined in [24] as: Dh ¼ ðHmin þ Hmax Þ:
ð6Þ
The pressures at successive points lie on a straight line as well as temperatures of the same set of points. With this, the friction factor is evaluated using the defining equation.
f¼
dP − dX Dh
ð7Þ
2 1 2 ρV
Table 1 Geometric configuration of corrugated walls (unit mm). f
θ
S
a
b
1 2 3 4
50 22/60 27 0
10 15 10 17.32
5 5 5 5
7.5 7.5 7.5 5
1 2 3 4
L
278 278 278 278
X2
X1
j¼
Nu 1 RePr =3
:
ð9Þ
In this experimental study, to calculate experimental uncertainty analysis Kline and McClintock's [26] method was used. The uncertainties of calculated parameters are summarized in Table 4. Uncertainty analysis not only comments on experimental results but also plays an important role in picking a suitable measurement technique. 5. Results and discussions The results are presented and discussed in this section. In Fig. 5, fully development flows on 5 and 10 mm channel heights for three different fin types and plain surface are shown. Nusselt number for the corrugated channels is higher than the plain channel. Reynolds number increases with the Nusselt number. However, the Nusselt number is approximately constant in the plain surface because the Nusselt number depends on the heat transfer rate. Best results are obtained from fin 1 with 50° corrugated angle and 10 mm axial length of cycle. This fin configuration has a minimum pitch and maximum corrugation angle than other fins. In generally, decrease of the channel height has positive effect of heat transfer. The Nusselt number for H = 5 mm is about 80% higher than for H = 10 mm at Re = 4000, and the trend is that at a higher Reynolds number the Nusselt number is higher, reaching approximately 100% at Re = 6000. These results are contained in fins 2, 3 and 4. For fin 1, difference is much bigger. Table 3 Technical properties of experimental apparatus.
Table 2 Temperature and pressure measurement point (unit mm). fin
ð8Þ
The Colburn factor is defined as
Nusselt number (Nu) and Reynolds number (Re) are evaluated by Nu ¼
hA : ðhAÞp
X3
Bottom
Upper
Bottom
Upper
Bottom
Upper
192 215 213 209
198 218 214 215
26 22 21 19
25 15 20 21
25 16 20 18
23 14 20 22
Apparatus
Model-type
Thermocouple Microvolt multi meter Orifice Inclined manometer DC power supply Exhaust Fan
Comark AK 28 M, K type (Cr–Al) Keithley, 1997 Manufactured,TSE Airflow, 504 type M890G, Mastech Gurvent, GRV 350/4.T
H. Pehlivan et al. / International Communications in Heat and Mass Transfer 46 (2013) 106–111
109
Fig. 3. Front elevation and top view of the test section.
Fig. 6 demonstrates the variation of the convection heat transfer coefficient with 3 different sharp corrugation peak fins and plain surface at a constant heat flux of 616 W/m2. In this figure two different tube heights (5–10 mm) are compared. According to the value for the plain surface in [27], the variation of convection heat transfer coefficient up to Re 4000 was between 25 and 27 W/m2·K for H = 5 mm and also between 19 and 21 W/m2·K for H = 10 mm. Fig. 6 shows that the heat transfer coefficient increases with the increase in the Reynolds number. The reason for this is similar to the explanation concerning the corrugation pitch mentioned above. The maximum heat transfer coefficient (154 W/m2·K) is obtained in rounded corrugation peak fins for H = 5 mm at Re = 5007. For channel height at 10 mm, the peak value acquired is 76 W/m2·K at Re = 7380. The maximum heat transfer coefficient of the corrugated tube for 5 mm is up to 102.63% higher than that of the 10 mm. The effect of the corrugation
angle on the heat transfer coefficient can be seen: the heat transfer coefficient tends to increase as the corrugation pitch decreases. This is because the lower corrugation pitch promotes turbulence of the refrigerant flow and mixing of the condensate film. These agitations are mainly due to corrugation of the surface. The ratio of the convective heat transfer coefficient of the corrugated tube to that of the plain surface varies from 1.3 to 2.2 under the same average quality. This means that the corrugated tubes had a heat transfer coefficient approximately 100–280% minimum higher than that of the plain surface for H = 10 mm and also 10–50% higher for H = 5 mm. The two roughened tubes (f2–f3) and plain plate (f4) have approximately the same heat transfer coefficients at H = 5 mm, but fin 1 has two times their heat transfer coefficients. Colburn factor is a dimensionless heat-transfer equation to calculate the convection movement of heat. The characteristics of fluid flow and heat transfer in the periodic fully developed region of the corrugated duct can be shown using the Colburn factor in Fig. 7. As it can be clearly seen in Fig. 7, the Colburn factor decreases with the Reynolds number. The value of the Colburn factor for H = 5 mm is about 160% higher than for H = 10 mm. Variation of the total pressure drop with Reynolds number for different fins is shown in Fig. 8. It is clearly shown that the pressure gradient increase with Reynolds number for 5 and 10 mm channel heights. The ratio of the frictional pressure drop for 5 mm channel is slightly higher than 10 mm. The frictional pressure drop can be obtained by subtracting the acceleration pressure drop from the total measured pressure drop. The pressure drops are obtained by dividing the measured pressure drop by the length between pressure taps in Fig. 3. Also, frictional pressure drop increases with increasing corrugated angle for three different corrugated tubes.
Table 4 Total uncertainty of calculated parameters.
Fig. 4. Axial temperature variations of fluids at constant heat flux.
Re
Nu
h
f
V
Qcycle
ΔT
±3.57
±10.22
±10.77
±7.02
±2.36
±7.08
±6.98
H. Pehlivan et al. / International Communications in Heat and Mass Transfer 46 (2013) 106–111
160
80
120
60
Nu
Nu
110
80
40 20
40 0
0
3000
6000
0
9000
0
3000
Re
6000
9000
Re
Fig. 5. The relationships between the Reynolds number and the Nusselt number for different fin types.
200
80
h (W/m2.K)
h (W/m2.K)
160 120 80 40
60 40 20 0
0 0
3000
6000
9000
0
3000
Re
6000
9000
Re
Fig. 6. Distributions of local convection heat transfer coefficients for different channel heights as a function of Reynolds number.
Fig. 9 shows the variation of the enhancement ratio which is the ratio of the (hA) of an enhanced surface to a plain surface (hA)p, the ratio calculated from the corrugated surface to that calculated from the plain surface, with Reynolds number. Fig. 9 also shows the effect of channel height on heat transfer enhancement ratio. The enhancement ratio for smaller channel height is better than that for the larger for fluid flowing past the corrugated surface, fluid re-circulation or/and swirl flows are generated in the corrugation troughs. The improvement in heat transfer is up to 4 times for H = 10 mm and 11 times for H = 5 mm over flat plate channels for a value of Reynolds number amounting 5000. After 5000, as seen from Fig. 9, the relation between the Reynolds number and the enhancement ratio is a horizontal line for H = 10 mm and H = 5 mm in case of the two roughened tubes (f2–f3).
6. Conclusions New experimental data on the heat transfer and pressure drop characteristics in converging–diverging wall channels are presented. • Nusselt number depends strongly on the inclination angle (θ). With an increasing angle, the Nusselt number increases • Three different fin types and one plain surface for two different channel heights are investigated in this study; fin 1 gives the best results for different result parameters. • These channels can lengthen the flow path and cause better mixing; higher heat transfer performance is obtained compared to straight ducts.
0,12
0,04
0,1 0,03
0,06
j
j
0,08
0,02
0,04 0,01 0,02 0
0 0
3000
6000
9000
0
3000
Re
6000
9000
Re
0,5
0,5
0,4
0,4
0,3
0,3
dp/dx
dp/dx
Fig. 7. Comparison of Colburn factors at different fins.
0,2
0,1
0,1 0
0,2
0 0
3000
6000
Re
9000
0
3000
6000
Re
Fig. 8. Relationship between pressure drop and Reynolds number at various fins.
9000
H. Pehlivan et al. / International Communications in Heat and Mass Transfer 46 (2013) 106–111
111
8 12 8
E
E
6
4
4 2
0 0
3000
6000
9000
Re
0
0
3000
6000
9000
Re
Fig. 9. The enhancement ratio for different fin types as a function of Reynolds number.
• Corrugated channel is a good alternative for high heat flux applications or for more efficient heat exchange devices used in a wide variety of engineering applications like heating and air conditioning units. • The heat transfer performance for a converging–diverging channel has a relatively positive effect than in phase arrangement. • The enhancement of the heat transfer in channel 1 exceeds up to 11 times between Re = 2000 and 5000. • Local heat transfer characteristic can be evaluated through a numerical approach. Therefore further research should be focused on numerical simulation and CFD analyses for converging–diverging channels. References [1] E.A.M. Elshafei, M.M. Awad, E. El-Negiry, A.G. Ali, Heat transfer and pressure drop in corrugated channels, Energy 35 (2010) 101–110. [2] M. Rahimi-Esbo, A.A. Ranjbar, A. Ramiar, A. Arya, M. Rahgoshay, Numerical study of turbulent forced convection jet flow in a converging sinusoidal channel, International Journal of Thermal Sciences 59 (2012) 176–185. [3] M.A. Habib, Ikramul-Haq, H.M. Badr, S.A.M. Saıd, Calculation of turbulent flow and heat transfer in periodically converging-diverging channels, Computers and Fluids 27 (1998) 95–120. [4] H.A. Mohammed, A.K. Abbas, J.M. Sheriff, Influence of geometrical parameters and forced convective heat transfer in transversely corrugated circular tubes, International Communications in Heat and Mass Transfer 44 (2013) 116–126. [5] A. Cabal, J. Szumbarski, J.M. Floryan, Numerical simulation of flows over corrugated walls, Computers and Fluids 30 (2001) 753–776. [6] I. Hassan, I. Nirdosh, G.S. Free, Convective mass transfer at corrugated surfaces in relation to catalytic and electrochemical reactor design, Chemical Engineering and Processing 48 (2009) 1341–1345. [7] S.W. Chang, B.J. Huang, Thermal performances of corrugated channel with skewed wall waves at rolling and pitching conditions, International Journal of Heat and Mass Transfer 55 (2012) 4548–4565. [8] L.-Z. Zhang, Z.-Y. Chen, Convective heat transfer in cross-corrugated triangular ducts under uniform heat flux boundary conditions, International Journal of Heat and Mass Transfer 54 (2011) 597–605. [9] M. Faizal, M.R. Ahmed, Experimental studies on a corrugated plate heat exchanger for small temperature difference applications, Experimental Thermal and Fluid Science 36 (2012) 242–248. [10] S.W. Chang, A.W. Lees, T.C. Chou, Heat transfer and pressure drop in furrowed channels with transverse and skewed sinusoidal wavy walls, International Journal of Heat and Mass Transfer 52 (2009) 4592–4603.
[11] A. García, J.P. Solano, P.G. Vicente, A. Viedma, The influence of artificial roughness shape on heat transfer enhancement: corrugated tubes, dimpled tubes and wire coils, Applied Thermal Engineering 35 (2012) 196–201. [12] W. Pirompugd, S. Wongwises, C.-C. Wang, Simultaneous heat and mass transfer characteristics for wavy fin-and-tube heat exchangers under dehumidifying conditions, International Journal of Heat and Mass Transfer 49 (2006) 132–143. [13] P. Naphon, Heat transfer characteristics and pressure drop in channel with V corrugated upper and lower plates, Energy Conversion and Management 48 (2007) 1516–1524. [14] K. Nilpueng, S. Wongwises, Flow pattern and pressure drop of vertical upward gas–liquid flow in sinusoidal wavy channels, Experimental Thermal and Fluid Science 30 (2006) 523–534. [15] T. Adachi, Y. Tashiro, H. Arima, Y. Ikegami, Pressure drop characteristics of flow in a symmetric channel with periodically expanded grooves, Chemical Engineering Science 64 (2009) 593–597. [16] S.W. Chang, T.-M. Liou, K.F. Chiang, G.F. Hong, Heat transfer and pressure drop in rectangular channel with compound roughness of V-shaped ribs and deepened scales, International Journal of Heat and Mass Transfer 51 (2008) 457–468. [17] A.A. El-Sebaii, S. Aboul-Enein, M.R.I. Ramadan, S.M. Shalaby, B.M. Moharram, Investigation of thermal performance of-double pass-flat and v-corrugated plate solar air heaters, Energy 36 (2011) 1076–1086. [18] W. Gao, W. Lin, T. Liu, C. Xia, Analytical and experimental studies on the thermal performance of cross-corrugated and flat-plate solar air heaters, Applied Energy 84 (2007) 425–441. [19] D.R. Sawyers, M. Sen, H.-C. Chang, Heat transfer enhancement in three dimensional corrugated channel flow, International Journal of Heat and Mass Transfer 41 (1998) 3559–3573. [20] S.D. Hwang, I.H. Jang, H.H. Cho, Experimental study on flow and local heat/mass transfer characteristics inside corrugated duct, International Journal of Heat and Fluid Flow 27 (2006) 21–32. [21] G. Fabbri, Heat transfer optimization in corrugated wall channels, International Journal of Heat and Mass Transfer 43 (2000) 4299–4310. [22] T. Desrues, P. Marty, J.F. Fourmigué, Numerical prediction of heat transfer and pressure drop in three-dimensional channels with alternated opposed ribs, Applied Thermal Engineering 45–46 (2012) 52–63. [23] E.M. Sparrow, J.W. Comb, Effect of inter-wall spacing and fluid flow inlet conditions on a corrugated-wall heat exchanger, International Journal of Heat and Mass Transfer 26 (1983) 993–1005. [24] B. Niceno, E. Nobile, Numerical analysis of fluid flow and heat transfer in periodic wavy channels, International Journal of Heat and Fluid Flow 22 (2001) 156–167. [25] R.L. Webb, Principles of Enhanced Heat Transfer, John Wiley & Sons Inc., New York, 2000. [26] S.J. Kline, F.A. McClintock, Describing uncertainties in single-sample experiments, Mechanical Engineering 1 (1953) 3–8. [27] I. Taymaz, I. Koç, Y. Islamoğlu, Experimental study on forced convection heat transfer characteristics in a converging diverging heat exchanger channel, Heat and Mass Transfer 44 (2008) 1257–1262.
Contents lists available at SciVerse ScienceDirect
International Communications in Heat and Mass Transfer journal homepage: www.elsevier.com/locate/ichmt
Experimental study of forced convective heat transfer in a different arranged corrugated channel☆ H. Pehlivan, I. Taymaz ⁎, Y. İslamoğlu Sakarya University, Mechanical Engineering Department, Sakarya, Turkey
a r t i c l e
i n f o
Available online 2 June 2013 Keywords: Corrugated channel Convective heat transfer Wavy geometry
a b s t r a c t In this study, heat transfer rate for sinusoidal corrugated channel has been experimentally investigated. Three different type sharp corrugation peak fins and a plain surface were used in the experiment. Results were carried out for constant heat flux of 616 W/m2, varied Reynolds number Re 1500 to 8000 for the corrugation angle (27, 50 and 22/60°) and channel height of 5 and 10 mm. Nusselt number (Nu), convection heat transfer coefficient (h), Colburn factor (j) and enhancement ratio (E) against Reynolds number (Re) have been studied. The effects of the wavy geometry and channel height have been discussed. The increase of corrugated angle gave rise to a heat transfer rate. © 2013 Elsevier Ltd. All rights reserved.
1. Introduction In recent years, due to the increasing demand by industries for heat exchangers that are more efficient, compact and less expensive, heat transfer enhancement has gained great momentum. For this purpose, two techniques have been identified: passive and active [1]. Convective heat transfer can be enhanced passively by changing flow geometry [2]. The symmetrical converging–diverging channel is one of several devices employed for enhancing the heat and mass transfer [3]. Enhancement techniques based on artificial roughness are used in numerous applications [4]. Rough walls exist in all flow systems, where they may lead to either deterioration or improvement of the desired functionality. Wall roughness can be increased to promote mixing of the fluid or reduced to eliminate flow disturbances [5]. In view of their turbulence promoting ability, corrugated surfaces have been used widely to improve the rate of forced convection heat and mass transfer in equipment such as heat exchangers, solar collectors, electrochemical and catalytic reactors and membrane processes [6]. 2. Background Many experimental or numerical studies for fluid flow and heat transfer in corrugated channels have been done. Experimentally: Chang and Huang [7] studied the properties of heat transfer, pressured drop and thermal performance factor of a corrugated wavy channel with Re numbers from 350 to 5500. Zhang and Chen [8] used crosscorrugated triangular duct under uniform heat flux boundary conditions
☆ Communicated by W.J. Minkowycz. ⁎ Corresponding author. E-mail address: [email protected] (I. Taymaz). 0735-1933/$ – see front matter © 2013 Elsevier Ltd. All rights reserved. http://dx.doi.org/10.1016/j.icheatmasstransfer.2013.05.016
in the Re range of 500–5000. Faizal and Ahmed [9] investigated the channel height between the plates to determine the configuration that gives the optimum heat transfer. Three different heights, 6 mm, 9 mm, and 12 mm, were used. The optimum heat transfer between the two streams is obtained for the minimum spacing of 6 mm. Chang et al. [10] examined the detailed Nusselt number (Nu) distributions, pressure drop coefficients (f) and thermal performance factors (g) for two furrowed rectangular channels with transverse and skewed sinusoidal wavy walls. García et al. [11] analyzed the thermal-hydraulic behavior of three types of enhancement techniques based on artificial roughness: corrugated tubes, dimpled tubes and wire coils. Pirompugd et al. [12] presented the heat and mass transfer characteristics of 18 wavy fin-and-tube heat exchangers under dehumidifying conditions. Naphon [13] researched heat transfer characteristics and pressure drop in the channel with V corrugated upper and lower plates under constant heat flux. Corrugated plates with three different corrugated tile angles of 20°, 40°and 60° are tested. The experiments are performed for the Reynolds number and heat flux in the ranges of 2000–9000 and 0.5–1.2 kW/m2, respectively. Nilpueng and Wongwises [14] studied flow patterns and pressure drop of upward liquid single-phase flow and air–water two-phase flow in sinusoidal wavy channels. Different phase shifts between the side walls of the wavy channel of 0°, 90° and 180° are investigated. The flow phenomena, which are bubbly flow, slug flow, churn flow, and dispersed bubbly flow are observed and recorded by high-speed camera. Pressure drop characteristics of flow in a periodically grooved channel are investigated by Adachi et al. [15]. The Reynolds number regime was 100 b Re b 1000. A novel heat transfer enhancement roughness with V-shaped ribs and deepened scales is devised by Chang et al. [16]. Performances of heat transfer and pressure drop in a rectangular channel were examined for both forward and backward flows in the Re range of 1000–30,000. Solar air heaters were analytically and experimentally studied El-Sebaii et al. [17] and Gao et al. [18]
H. Pehlivan et al. / International Communications in Heat and Mass Transfer 46 (2013) 106–111
Nomenclature Acycle cp Dh E f H h j k L _ m Nu Pr Qcycle P Re S T V W
convection heat transfer area per cycle, m2 specific heat, kJ/kg·K hydraulic diameter, m enhancement ratio, fully developed friction factor, fin channel height, m cycle average fully developed heat transfer coefficient, W/m2·K Colburn factor, thermal conductivity, W/m·K length, m mass flow rate, kg/s cycle average fully developed Nusselt number, Prandtl number, heat transferred to fluid per cycle, W Pressure, Pa Reynolds number, pitch (axial length of cycle), m temperature, °C mean velocity, m/s width, m
Greek symbols ρ density, kg/m3 υ kinematic viscosity, m2/s θ corrugation angle, °
Subscript av b fd min max p w
quantity evaluated at average cross-section fluid bulk fully developed minimum, maximum, plain surface, wall.
using the double pass flat and V-corrugated plate and cross-corrugated solar air heaters, respectively. Various numerical studies and approaches to the flow and heat transfer characteristics in corrugated or wavy ducts are conducted. Sawyers et al. [19] studied the effect of three dimensional hydrodynamics on the enhancement of steady, laminar heat transfer in corrugated channels. Reynolds numbers are considered in the range of 0 b Re b 250 to avoid unsteady flow. Hwang et al. [20] investigated the flow and local mass transfer characteristics inside a corrugated duct. The flow visualization technique and a numerical analysis using a commercial code, FLUENT, are used to understand the overall flow structure inside the duct. The corrugation angle of the wavy duct is 145°and the duct aspect ratio is 7.3. The Reynolds numbers, based on the duct hydraulic diameter, vary from 100 to 5000. Fabbri [21] analyzed the heat transfer in a channel composed of a smooth and a corrugated wall under laminar flow conditions. The velocity and temperature distributions are determined with the help of a finite element model. Desrues et al. [22] calculated Nusselt numbers and friction factor of a ribbed channel for Reynolds numbers from 75 to 2000 and geometric parameters, by the mean of three-dimensional incompressible computational fluid dynamics. The geometry is a rectangular duct with stream wise-periodic transverse rectangular ribs, alternated on its two smaller walls. The aim of this study, which
107
includes heat transfer rates for three different types of sharp corrugation peak fins and, sinusoidal converging–diverging channel, will be experimentally investigated. 3. Experimental program 3.1. Experimental apparatus A test ring is used to perform heat transfer and pressure drop experiments as shown in Fig. 1. In the experiment, three different types corrugated channels and one plain channel has been used. Their schematic and geometrical configurations are shown in Fig. 2 and Table 1, respectively. The test section is assembled with two copper plates of 10 mm thickness. The plate geometry is 50 × 278 mm. A forced-convection air flow was used for the experiments and equal power input for heated length was established in both top and bottom walls. To measure the entering bulk temperature of the air, two K type thermocouples were positioned at the inlet of the duct. Also, to measure the temperature distribution of each corrugated walls, six K type thermocouples are used. The mass flow rate of air flowing through the systems was measured by the orifice. Two static pressure taps were located at the valleys of the top to measure axial pressure drops, used to evaluate average friction factor and the pressure drops were measured by a U-manometer. The test section is also isolated to avoid thermal losses (Table 3). 3.2. Data collection The experimental procedure involved adjusting the flow rate to the desired value. After the exhaust fan was turned on and the desired Reynolds number was obtained, the power input of plate heaters gradually increased and maintained at 616 W/m2 to provide sufficient measurement while the fluid property varies. The supplied heat into the corrugated walls was adjusted to achieve the desired level by using electric heaters, which are 2 mm thick, 50 mm wide and 278 mm long. They were located to the back of each plate. The voltage and current of electric input to the plate type heaters were controlled by a DC power supply unit. Temperatures were recorded at intervals of 15 min until a steady state was reached. Steady-state conditions were assumed to prevail when the temperature measurement on the plates and fluid inlet were within ± 0.2 °C. For a duct of periodic geometry, the heat transfer coefficient may vary axially in the thermally developed regions, but the same value of the coefficient recurs periodically at successive axial stations separated by an axial distance S, and the cycle-average heat transfer coefficient is the same from cycle to cycle [23]. The schematic diagram of the test section for which the details were given in the overview of the measurement section including the position where temperature and pressure are measured is shown in Fig. 3 and Table 2. 4. Heat transfer works The fully developed condition has been verified experimentally in this study by observing the temperature distribution. The slope of these straight lines is dT Q cycle : ¼ _ p dX Smc
ð1Þ
For constant surface heat flux, the axial temperature variation (as seen in Fig. 4) of the fluid is
Tb ðXÞ ¼ Tb þ
Q cycle : _ p Smc
ð2Þ
108
H. Pehlivan et al. / International Communications in Heat and Mass Transfer 46 (2013) 106–111
Fig. 1. Experimental setup (1. Airline and pressure measurement line, 2. Orifice, 3. Upstream plenum, 4. Isolated test section and thermocouple connection points, 5. Downstream plenum, 6. Valve, 7. Exhaust Fan, 8. Digital Scanner, 9. Multi meter, 10. Power, 11. Ice Bath).
Fig. 2. The wall with sharp corrugation peak.
The cycle-average full development heat transfer coefficients were evaluated from the measured temperatures and heat inputs
h¼
Q cycle : ðTw −Tb Þfd Acycle
ð3Þ
The enhancement ratio E, the ratio of the (hA) of an enhanced surface to that of a plain surface (hA)P is defined by Webb [25] as: E¼
hDh k
ð4Þ
Re ¼
VDh : υ
ð5Þ
The hydraulic diameter is defined in [24] as: Dh ¼ ðHmin þ Hmax Þ:
ð6Þ
The pressures at successive points lie on a straight line as well as temperatures of the same set of points. With this, the friction factor is evaluated using the defining equation.
f¼
dP − dX Dh
ð7Þ
2 1 2 ρV
Table 1 Geometric configuration of corrugated walls (unit mm). f
θ
S
a
b
1 2 3 4
50 22/60 27 0
10 15 10 17.32
5 5 5 5
7.5 7.5 7.5 5
1 2 3 4
L
278 278 278 278
X2
X1
j¼
Nu 1 RePr =3
:
ð9Þ
In this experimental study, to calculate experimental uncertainty analysis Kline and McClintock's [26] method was used. The uncertainties of calculated parameters are summarized in Table 4. Uncertainty analysis not only comments on experimental results but also plays an important role in picking a suitable measurement technique. 5. Results and discussions The results are presented and discussed in this section. In Fig. 5, fully development flows on 5 and 10 mm channel heights for three different fin types and plain surface are shown. Nusselt number for the corrugated channels is higher than the plain channel. Reynolds number increases with the Nusselt number. However, the Nusselt number is approximately constant in the plain surface because the Nusselt number depends on the heat transfer rate. Best results are obtained from fin 1 with 50° corrugated angle and 10 mm axial length of cycle. This fin configuration has a minimum pitch and maximum corrugation angle than other fins. In generally, decrease of the channel height has positive effect of heat transfer. The Nusselt number for H = 5 mm is about 80% higher than for H = 10 mm at Re = 4000, and the trend is that at a higher Reynolds number the Nusselt number is higher, reaching approximately 100% at Re = 6000. These results are contained in fins 2, 3 and 4. For fin 1, difference is much bigger. Table 3 Technical properties of experimental apparatus.
Table 2 Temperature and pressure measurement point (unit mm). fin
ð8Þ
The Colburn factor is defined as
Nusselt number (Nu) and Reynolds number (Re) are evaluated by Nu ¼
hA : ðhAÞp
X3
Bottom
Upper
Bottom
Upper
Bottom
Upper
192 215 213 209
198 218 214 215
26 22 21 19
25 15 20 21
25 16 20 18
23 14 20 22
Apparatus
Model-type
Thermocouple Microvolt multi meter Orifice Inclined manometer DC power supply Exhaust Fan
Comark AK 28 M, K type (Cr–Al) Keithley, 1997 Manufactured,TSE Airflow, 504 type M890G, Mastech Gurvent, GRV 350/4.T
H. Pehlivan et al. / International Communications in Heat and Mass Transfer 46 (2013) 106–111
109
Fig. 3. Front elevation and top view of the test section.
Fig. 6 demonstrates the variation of the convection heat transfer coefficient with 3 different sharp corrugation peak fins and plain surface at a constant heat flux of 616 W/m2. In this figure two different tube heights (5–10 mm) are compared. According to the value for the plain surface in [27], the variation of convection heat transfer coefficient up to Re 4000 was between 25 and 27 W/m2·K for H = 5 mm and also between 19 and 21 W/m2·K for H = 10 mm. Fig. 6 shows that the heat transfer coefficient increases with the increase in the Reynolds number. The reason for this is similar to the explanation concerning the corrugation pitch mentioned above. The maximum heat transfer coefficient (154 W/m2·K) is obtained in rounded corrugation peak fins for H = 5 mm at Re = 5007. For channel height at 10 mm, the peak value acquired is 76 W/m2·K at Re = 7380. The maximum heat transfer coefficient of the corrugated tube for 5 mm is up to 102.63% higher than that of the 10 mm. The effect of the corrugation
angle on the heat transfer coefficient can be seen: the heat transfer coefficient tends to increase as the corrugation pitch decreases. This is because the lower corrugation pitch promotes turbulence of the refrigerant flow and mixing of the condensate film. These agitations are mainly due to corrugation of the surface. The ratio of the convective heat transfer coefficient of the corrugated tube to that of the plain surface varies from 1.3 to 2.2 under the same average quality. This means that the corrugated tubes had a heat transfer coefficient approximately 100–280% minimum higher than that of the plain surface for H = 10 mm and also 10–50% higher for H = 5 mm. The two roughened tubes (f2–f3) and plain plate (f4) have approximately the same heat transfer coefficients at H = 5 mm, but fin 1 has two times their heat transfer coefficients. Colburn factor is a dimensionless heat-transfer equation to calculate the convection movement of heat. The characteristics of fluid flow and heat transfer in the periodic fully developed region of the corrugated duct can be shown using the Colburn factor in Fig. 7. As it can be clearly seen in Fig. 7, the Colburn factor decreases with the Reynolds number. The value of the Colburn factor for H = 5 mm is about 160% higher than for H = 10 mm. Variation of the total pressure drop with Reynolds number for different fins is shown in Fig. 8. It is clearly shown that the pressure gradient increase with Reynolds number for 5 and 10 mm channel heights. The ratio of the frictional pressure drop for 5 mm channel is slightly higher than 10 mm. The frictional pressure drop can be obtained by subtracting the acceleration pressure drop from the total measured pressure drop. The pressure drops are obtained by dividing the measured pressure drop by the length between pressure taps in Fig. 3. Also, frictional pressure drop increases with increasing corrugated angle for three different corrugated tubes.
Table 4 Total uncertainty of calculated parameters.
Fig. 4. Axial temperature variations of fluids at constant heat flux.
Re
Nu
h
f
V
Qcycle
ΔT
±3.57
±10.22
±10.77
±7.02
±2.36
±7.08
±6.98
H. Pehlivan et al. / International Communications in Heat and Mass Transfer 46 (2013) 106–111
160
80
120
60
Nu
Nu
110
80
40 20
40 0
0
3000
6000
0
9000
0
3000
Re
6000
9000
Re
Fig. 5. The relationships between the Reynolds number and the Nusselt number for different fin types.
200
80
h (W/m2.K)
h (W/m2.K)
160 120 80 40
60 40 20 0
0 0
3000
6000
9000
0
3000
Re
6000
9000
Re
Fig. 6. Distributions of local convection heat transfer coefficients for different channel heights as a function of Reynolds number.
Fig. 9 shows the variation of the enhancement ratio which is the ratio of the (hA) of an enhanced surface to a plain surface (hA)p, the ratio calculated from the corrugated surface to that calculated from the plain surface, with Reynolds number. Fig. 9 also shows the effect of channel height on heat transfer enhancement ratio. The enhancement ratio for smaller channel height is better than that for the larger for fluid flowing past the corrugated surface, fluid re-circulation or/and swirl flows are generated in the corrugation troughs. The improvement in heat transfer is up to 4 times for H = 10 mm and 11 times for H = 5 mm over flat plate channels for a value of Reynolds number amounting 5000. After 5000, as seen from Fig. 9, the relation between the Reynolds number and the enhancement ratio is a horizontal line for H = 10 mm and H = 5 mm in case of the two roughened tubes (f2–f3).
6. Conclusions New experimental data on the heat transfer and pressure drop characteristics in converging–diverging wall channels are presented. • Nusselt number depends strongly on the inclination angle (θ). With an increasing angle, the Nusselt number increases • Three different fin types and one plain surface for two different channel heights are investigated in this study; fin 1 gives the best results for different result parameters. • These channels can lengthen the flow path and cause better mixing; higher heat transfer performance is obtained compared to straight ducts.
0,12
0,04
0,1 0,03
0,06
j
j
0,08
0,02
0,04 0,01 0,02 0
0 0
3000
6000
9000
0
3000
Re
6000
9000
Re
0,5
0,5
0,4
0,4
0,3
0,3
dp/dx
dp/dx
Fig. 7. Comparison of Colburn factors at different fins.
0,2
0,1
0,1 0
0,2
0 0
3000
6000
Re
9000
0
3000
6000
Re
Fig. 8. Relationship between pressure drop and Reynolds number at various fins.
9000
H. Pehlivan et al. / International Communications in Heat and Mass Transfer 46 (2013) 106–111
111
8 12 8
E
E
6
4
4 2
0 0
3000
6000
9000
Re
0
0
3000
6000
9000
Re
Fig. 9. The enhancement ratio for different fin types as a function of Reynolds number.
• Corrugated channel is a good alternative for high heat flux applications or for more efficient heat exchange devices used in a wide variety of engineering applications like heating and air conditioning units. • The heat transfer performance for a converging–diverging channel has a relatively positive effect than in phase arrangement. • The enhancement of the heat transfer in channel 1 exceeds up to 11 times between Re = 2000 and 5000. • Local heat transfer characteristic can be evaluated through a numerical approach. Therefore further research should be focused on numerical simulation and CFD analyses for converging–diverging channels. References [1] E.A.M. Elshafei, M.M. Awad, E. El-Negiry, A.G. Ali, Heat transfer and pressure drop in corrugated channels, Energy 35 (2010) 101–110. [2] M. Rahimi-Esbo, A.A. Ranjbar, A. Ramiar, A. Arya, M. Rahgoshay, Numerical study of turbulent forced convection jet flow in a converging sinusoidal channel, International Journal of Thermal Sciences 59 (2012) 176–185. [3] M.A. Habib, Ikramul-Haq, H.M. Badr, S.A.M. Saıd, Calculation of turbulent flow and heat transfer in periodically converging-diverging channels, Computers and Fluids 27 (1998) 95–120. [4] H.A. Mohammed, A.K. Abbas, J.M. Sheriff, Influence of geometrical parameters and forced convective heat transfer in transversely corrugated circular tubes, International Communications in Heat and Mass Transfer 44 (2013) 116–126. [5] A. Cabal, J. Szumbarski, J.M. Floryan, Numerical simulation of flows over corrugated walls, Computers and Fluids 30 (2001) 753–776. [6] I. Hassan, I. Nirdosh, G.S. Free, Convective mass transfer at corrugated surfaces in relation to catalytic and electrochemical reactor design, Chemical Engineering and Processing 48 (2009) 1341–1345. [7] S.W. Chang, B.J. Huang, Thermal performances of corrugated channel with skewed wall waves at rolling and pitching conditions, International Journal of Heat and Mass Transfer 55 (2012) 4548–4565. [8] L.-Z. Zhang, Z.-Y. Chen, Convective heat transfer in cross-corrugated triangular ducts under uniform heat flux boundary conditions, International Journal of Heat and Mass Transfer 54 (2011) 597–605. [9] M. Faizal, M.R. Ahmed, Experimental studies on a corrugated plate heat exchanger for small temperature difference applications, Experimental Thermal and Fluid Science 36 (2012) 242–248. [10] S.W. Chang, A.W. Lees, T.C. Chou, Heat transfer and pressure drop in furrowed channels with transverse and skewed sinusoidal wavy walls, International Journal of Heat and Mass Transfer 52 (2009) 4592–4603.
[11] A. García, J.P. Solano, P.G. Vicente, A. Viedma, The influence of artificial roughness shape on heat transfer enhancement: corrugated tubes, dimpled tubes and wire coils, Applied Thermal Engineering 35 (2012) 196–201. [12] W. Pirompugd, S. Wongwises, C.-C. Wang, Simultaneous heat and mass transfer characteristics for wavy fin-and-tube heat exchangers under dehumidifying conditions, International Journal of Heat and Mass Transfer 49 (2006) 132–143. [13] P. Naphon, Heat transfer characteristics and pressure drop in channel with V corrugated upper and lower plates, Energy Conversion and Management 48 (2007) 1516–1524. [14] K. Nilpueng, S. Wongwises, Flow pattern and pressure drop of vertical upward gas–liquid flow in sinusoidal wavy channels, Experimental Thermal and Fluid Science 30 (2006) 523–534. [15] T. Adachi, Y. Tashiro, H. Arima, Y. Ikegami, Pressure drop characteristics of flow in a symmetric channel with periodically expanded grooves, Chemical Engineering Science 64 (2009) 593–597. [16] S.W. Chang, T.-M. Liou, K.F. Chiang, G.F. Hong, Heat transfer and pressure drop in rectangular channel with compound roughness of V-shaped ribs and deepened scales, International Journal of Heat and Mass Transfer 51 (2008) 457–468. [17] A.A. El-Sebaii, S. Aboul-Enein, M.R.I. Ramadan, S.M. Shalaby, B.M. Moharram, Investigation of thermal performance of-double pass-flat and v-corrugated plate solar air heaters, Energy 36 (2011) 1076–1086. [18] W. Gao, W. Lin, T. Liu, C. Xia, Analytical and experimental studies on the thermal performance of cross-corrugated and flat-plate solar air heaters, Applied Energy 84 (2007) 425–441. [19] D.R. Sawyers, M. Sen, H.-C. Chang, Heat transfer enhancement in three dimensional corrugated channel flow, International Journal of Heat and Mass Transfer 41 (1998) 3559–3573. [20] S.D. Hwang, I.H. Jang, H.H. Cho, Experimental study on flow and local heat/mass transfer characteristics inside corrugated duct, International Journal of Heat and Fluid Flow 27 (2006) 21–32. [21] G. Fabbri, Heat transfer optimization in corrugated wall channels, International Journal of Heat and Mass Transfer 43 (2000) 4299–4310. [22] T. Desrues, P. Marty, J.F. Fourmigué, Numerical prediction of heat transfer and pressure drop in three-dimensional channels with alternated opposed ribs, Applied Thermal Engineering 45–46 (2012) 52–63. [23] E.M. Sparrow, J.W. Comb, Effect of inter-wall spacing and fluid flow inlet conditions on a corrugated-wall heat exchanger, International Journal of Heat and Mass Transfer 26 (1983) 993–1005. [24] B. Niceno, E. Nobile, Numerical analysis of fluid flow and heat transfer in periodic wavy channels, International Journal of Heat and Fluid Flow 22 (2001) 156–167. [25] R.L. Webb, Principles of Enhanced Heat Transfer, John Wiley & Sons Inc., New York, 2000. [26] S.J. Kline, F.A. McClintock, Describing uncertainties in single-sample experiments, Mechanical Engineering 1 (1953) 3–8. [27] I. Taymaz, I. Koç, Y. Islamoğlu, Experimental study on forced convection heat transfer characteristics in a converging diverging heat exchanger channel, Heat and Mass Transfer 44 (2008) 1257–1262.