International Communications in Heat and Mass Transfer

International Communications in Heat and Mass Transfer 39 (2012) 1505 1512 Contents lists available at SciVerse ScienceDirect International Communications in Heat and Mass Transfer journal homepage: www.elsevier.com/locate/ichmt Heat transfer performance evaluation for turbulent flow through a tube with twisted wire brush inserts M.M.K. Bhuiya a,b,, M.S.U. Chowdhury b, M. Islam b,c, J.U. Ahamed b, M.J.H. Khan d, M.R.I. Sarker a, M. Saha a a School of Mechanical Engineering, The University of Adelaide, Adelaide, SA 5005, Australia b Department of Mechanical Engineering, Chittagong University of Engineering and Technology (CUET), Chittagong 4349, Bangladesh c School of Physics, Chemistry and Mechanical Engineering, Queensland University of Technology, Brisbane, Queensland 4001, Australia d Department of Chemical Engineering, University of Malaya, 50603 Kuala Lumpur, Malaysia article info abstract Available online 15 October 2012 Keywords: Twisted wire brush insert Nusselt number Friction factor Thermal performance factor In the present study, the heat transfer performance and friction factor characteristics in a circular tube fitted with twisted wire brush inserts were investigated experimentally. The twisted wire brush inserts were fabricated with four different twisted wire densities of 100, 150, 200, and per centimeter by winding a 1 mm diameter of the copper wire over a 5 mm diameter of two twisted iron core-rods. Heat transfer and friction factor data in tubes were examined for Reynolds number ranging from 7,200 to 50,200. The results indicated that the presence of twisted wire brush inserts led to a large effect on the enhancement of heat transfer with corresponding increase in friction factor over the plain tube. The Nusselt number and friction factor of using the twisted wire brush inserts were found to be increased up to 2.15 and 2.0 times, respectively, than those over the plain tube values. Furthermore, the heat transfer performance was evaluated to assess the real benefits of using those type of inserts and the performance was achieved 1.85 times higher compared to the plain tube based on the constant blower power. Finally, correlations were developed based on the data generated from this work to predict the heat transfer, friction factor, and thermal performance factor for turbulent flow through a circular tube fitted with the twisted wire brush inserts in terms of wire density (y), Reynolds number (Re), and Prandtl number (Pr). 2012 Elsevier Ltd. All rights reserved. 1. Introduction Many efforts have been made on heat transfer enhancement according to the progress of thermal systems. The recent researches in heat transfer enhancement lead to the development of currently used heat transfer techniques. The turbulent generators with different geometrical configurations have been used as one of the passive heat transfer enhancement techniques and are the most widely used in tubes in several heat transfer applications, for example, heat recovery processes, air conditioning and refrigeration systems, chemical reactors, and food and dairy processes. Extensive studies have been performed from the beginning of the 20th century to determine the heat transfer characteristics inside the tubes. Agarwal and Rao [1] studied the isothermal and nonisothermal friction factor and mean Nusselt number under uniform wall temperature conditions. They also proposed the isothermal friction factor and the Nusselt number correlations. Hsieh et al. [2] Communicated by W.J. Minkowycz. Corresponding author at: School of Mechanical Engineering, The University of Adelaide, Adelaide, SA 5005, Australia. E-mail addresses: mkamalcuet@yahoo.com, muhammadmostafakamal.bhuiya@adelaide.edu.au (M.M.K. Bhuiya). experimentally studied the turbulent heat transfer and flow characteristics in a horizontal circular tube with strip-type inserts (longitudinal and cross strips) using air as working fluid. Eiamsa-ard et al. [3] investigated the heat transfer and fluid friction characteristics in a circular tube fitted with regularly spaced twisted tape elements. Mafiz et al. [4] studied the turbulent flow heat transfer performance of circular tubes having six integral internal longitudinal fins. The study indicated that significant enhancement of heat transfer was possible by using internal fins without requiring much additional pumping power. Naphon [5] considered effect of coil wire insert on heat transfer enhancement and pressure drop of the horizontal concentric plain tubes. Garc et al. [6] investigated the laminar transition turbulent heat transfer enhancement and flow patterns in the tube with wire coil inserts. Experimental correlations of fanning friction factor and Nusselt number were proposed. Sarma et al. [7] presented a new approach for predicting the convective heat transfer coefficient in a tube with twisted tape inserts for different pitch to diameter ratios. Eiamsa-ard and Promvonge [8] studied the heat transfer characteristics in a tube fitted with helical screw-tape with/without core-rod inserts. Sarkar et al. [9] investigated the heat transfer enhancement in turbulent flow through a tube with wire-coil inserts. The heat transfer coefficient for the tube with wire-coil inserts increased between 1.2 and 2.0 times that for a plain tube with a comparable 0735-1933/$ see front matter 2012 Elsevier Ltd. All rights reserved. http://dx.doi.org/10.1016/j.icheatmasstransfer.2012.10.005

1506 M.M.K. Bhuiya et al. / International Communications in Heat and Mass Transfer 39 (2012) 1505 1512 Nomenclature A x cross sectional area of test section [m 2 ] C p specific heat at constant pressure [J/(kg K)] d core-rod diameter [m] D i tube inside diameter [m] D o tube outer diameter [m] d w wire diameter [m] f friction factor, dimensionless f p predicted friction factor for the tube with wire brush inserts, dimensionless h convective heat transfer coefficient [W/(m 2 K)] h x local convective heat transfer coefficient [W/(m 2 K)] I current [ampere] k thermal conductivity [W/(m K)] L tube length [m] _m mass flow rate [kg/s] ΔP pressure drop along axial length of tube [N/m 2 ] Q average heat transfer rate [W] q heat flux [W/m 2 ] Q loss heat loss [W] Q t generated total heat [W] Q 1 actual heat supplied [W] Q 2 heat absorbed by the fluid [W] T i inlet temperature [K] T o outlet temperature [K] T b mean bulk temperature [K] T w mean wall temperature [K] T bx local bulk fluid temperature [K] T wx local wall temperature [K] V mean velocity in the test section [m/s] _V mass flow flux [kg/s.m 2 ] V i mean velocity at inlet section [m/s] V v voltage [volt] W wetted perimeter of the tube [m] X axial distance [m] y wire density, dimensionless Greek symbols η thermal performance factor, dimensionless η p predicted thermal performance factor, dimensionless ρ density [kg/m 3 ] Subscripts b bulk i inlet o outlet p wire brush s plain w wall x local Dimensionless numbers Nu Nusselt number, dimensionless Nu p predicted Nusselt number for the tube with wire brush insert, dimensionless Nu x local nusselt number, dimensionless Pr Prandtl number, dimensionless Re Reynolds number, dimensionless Re p equivalent Reynolds number for the tube with wire brush insert, dimensionless Re s equivalent Reynolds number of plain tube, dimensionless Reynolds number. A comparison of thermal performance factor of helical screw tape inserts of different twist ratios in laminar flow region of using Al 2 O 3 /water and CuO/water nanofluids through a circular duct with constant heat flux condition was investigated by Suresh et al. [10]. Sivashanmugam and Suresh [11,12] studied the heat transfer and friction factor characteristics in a circular tube fitted with regularly spaced helical screw-tape inserts and helical screw-tape inserts. Hsieh and Huang [13] investigated plain tubes and tubes with square and rectangular as well as crossed strip inserts using water as the working fluid. Zimparov [14] predicted the friction factors and heat transfer coefficients for turbulent flow in corrugated tubes with twisted tape inserts. Promvonge [15] experimentally investigated the thermal enhancement in a round tube with snail entry and coiled-wire inserts. Ahamed et al. [16] studied the prediction of heat transfer in turbulent flow through a tube with perforated twistedtape inserts and also developed a new correlation. The study revealed that the perforated twisted-tape-inserts caused an increase of heat transfer rate at the cost of increased pumping power. Bhuiya et al. [17] investigated the heat transfer enhancement and developed new correlations for turbulent flow through a tube with triple helical tape inserts. Sreenivasulu and Prasad [18] numerically studied the flow and heat transfer characteristics in an annulus wrapped with a helical wire for constant heat flux boundary condition. Behabadi et al. [19] studied the heat transfer and pressure drop characteristics of forced convective evaporation in horizontal tubes with coiled-wire inserts. Behabadi et al. [20] carried out the enhancement in heat transfer coefficient in the double horizontal tube with coiled-wire inserts. Two empirical correlations have been developed for predicting the heat transfer enhancement of these coiled-wire inserts. Karwa et al. [21] investigated the effect of relative roughness pitch and perforation of the spring roughness on heat transfer and friction factor for turbulent flow in an asymmetrically heated annular duct (radius ratio=0.39) with a heated tube having a spirally wound helical spring. Gunes et al. [22,23] studied the heat transfer enhancement in a tube with equilateral triangle cross sectioned coiled wire inserts for uniform heat flux boundary condition. The use of coiled wire inserts has a significant effect on the heat transfer and pressure drop. Wazed et al. [24] investigated the enhancement of heat transfer in turbulent flow through tube with perforated twisted tape inserts and found a significant enhancement of heat transfer at the cost of increased pumping power. Naphon and Suchana [25] experimentally studied the heat transfer enhancement and pressure drop of the horizontal concentric tube with twisted wire brush inserts, and showed that the twisted wire brush inserts have a large effect on the enhancement of heat transfer with the corresponding increase in pressure drop. In the avobe literature review, the numerous research articles were reported on heat transfer enhancement and pressure drop characteristics in tubes with various geometrical configurations of turbulence creator. However, limited research works were found related to heat transfer performance and friction factor characteristics through a tube with twisted wire brush inserts. In this study, the effects of twisted wire brush inserts on heat transfer performance and friction factor characteristics for turbulent flow through a circular tube were evaluated. Moreover, new correlations were developed for predicting the heat transfer, friction factor, and thermal performance factor with different twisted wire densities ranging from 100 to. In this study, some assumptions were made in order to make easy experiments, comparison and analysis. These were: i. Inside diameter of the tube (D i ) was used instead of hydraulic diameter (D h )indefining Reynolds number (Re), Nusselt number (Nu), and friction factor (f). ii. All the fluid properties were calculated at local bulk temperature (T bx ) and at atmospheric pressure instead of local pressure in the test section which was slightly less than the atmospheric pressure.

M.M.K. Bhuiya et al. / International Communications in Heat and Mass Transfer 39 (2012) 1505 1512 1507 iii. The heat transfer mode was considered only by forced convection from the inside wall of the tube to the fluid. However, there were points of contact between the inserts and the inside of wall of the tube. Thus there was the potential for heat transfer to occur through the inserts by conduction. It was not possible to quantify this; also, heat was conducted through the ends of the test section to adjacent sections. 2. Mathematical formulations The experimental data were used to calculate the Nusselt number, friction factor, and thermal performance factor at different Reynolds number in turbulent flow region for both the cases with and without using helical tape inserts: Mass flow rate was calculated by, _m ¼ ρa x V i where ρ is the density of air, A x is the cross sectional area of test section and V i is the mean inlet velocity. In the test section the velocity of air was obtained from, ð1þ The local bulk fluid temperature was determined by the following energy balance equation, T bx ¼ T i þ qwx _m C p ð9þ where W is the wetted perimeter and X is the axial distance of the tube The local Nusselt number was calculated as, Nu x ¼ h xd i k ð10þ where k is the thermal conductivity of air. The average heat transfer coefficient was obtained from, q h ¼ T w T ð11þ b where T w and T b are the mean wall and bulk fluid temperatures. The average Nusselt number was calculated according to the following way, V ¼ _m ρ b A x ð2þ Nu ¼ hd i k ð12þ where ρ b is the density at bulk fluid temperature. The total heat generated by the electrical winding was calculated as, Q t ¼ V v I where V v is the voltage and I is the current. The heat loss (Q loss ) through the insulation was calculated by measuring the average wall and the ambient temperature and estimated as 2 4% of the total heat supplied. Therefore actual heat supplied by the electrical winding, Q 1 ¼ Q t Q loss The heat absorbed by the fluid was calculated as, Q 2 ¼ _mc p ðt 0 T i Þ ð5þ where C p is the specific heat of air, T i and T o are the inlet and outlet temperatures of air, respectively. Heat balance between the actual heat input (Q 1 ) and the heat carried out by the fluid (Q 2 ) was within 1 2% for all runs. The average value of heat transfer (Q) rate was obtained from the actual heat supplied by electrical winding and the heat absorbed by the fluid for convective heat transfer calculation. Therefore, Q ¼ Q 1 þ Q 2 2 And the heat flux was calculated by, q ¼ Q πd i L where D i is the inner diameter and L is the length of the tube. Local convective heat transfer coefficient was obtained from, q h x ¼ ðt wx T bx Þ where T wx and T bx are the local wall and bulk fluid temperatures. ð3þ ð4þ ð6þ ð7þ ð8þ Friction factor was obtained from, ΔP f ¼ ð13þ L D i ρ b V 2 2 where ΔP is the pressure drop along length of the tube. 3. Experimental setup The experimental setup consisted of an inlet section, a test section, an air supply system (electric blower) and a heating arrangement. The schematic diagram of the experimental facility is shown in Fig. 1. The tube shaped inlet section, 533 mm long, was made as integral part of the test section to avoid any flow disturbances upstream of the test section and to get fully developed flow in the test section as well. The inlet section shape of the experimental setup was made as per suggestions of Owner and Pankhurst [26] to avoid separation and stratification of the flow. Geometry of the test section fitted with the twisted wire brush insert over a 5 mm diameter two twisted iron core-rods and geometric parameters of the wire brush insert are shown in Fig. 2(a) and (b), respectively. The plain tube (test section) was made of brass having 70 mm inside diameter, 90 mm outside diameter, and 1500 mm in length. The twisted wire brush inserts were fabricated by winding a 1 mm diameter of the copper wire over a 5 mm diameter of two twisted iron core-rods. The four different twisted wire densities of 100, 150, 200, and were considered per centimeter by winding the copper wire over a two twisted iron core-rods. Nichrome wire (resistance 1.2 Ω/m) was used as an electric heater to heat the test section at constant heat flux condition. Nichrome wire was spirally wounded uniformly around the tube. The terminals of the Nichrome wire heating coil were connected to the variac transformer. The electrical output power was controlled by a variac transformer to obtain a constant heat flux condition throughout the entire test section. The outer surface of the test section was well insulated to minimize heat leak to the surroundings. Sixteen K-type thermocouples were tapped along the tube wall for monitoring the local wall temperatures, while the bulk air temperatures were measured with the help of RTDs. Data logger was used to record the inlet and outlet bulk fluid temperatures as well as to measure the tube wall temperatures of the test section. The calibrated

1508 M.M.K. Bhuiya et al. / International Communications in Heat and Mass Transfer 39 (2012) 1505 1512 Fig. 1. Schematic diagram of the experimental facility. thermocouples were used to measure the temperatures of the wall as well as bulk fluid temperature of the test section. The pressure drop across the test section was measured with an inclined U-tube manometer. The heat transfer and pressure drop experiments were carried out individually. The heat transfer experiment was performed under a constant heat flux condition. In contrast, the pressure drop (friction) test was conducted under an isothermal condition without turning on the heater. The air flow rate was measured by using an orifice meter was built built according to the ASME standard [27] and was calibrated with a hot-wire anemometer to measure the flow velocities across the tube section. The experiments were conducted for the Reynolds number ranging from 7,200 to 50,200. The uncertainties in the experimental measurements were determined by using the method introduced by Kline and McClintock [28]. The uncertainty calculation method used the calculation of derivatives of the desired variables with respect to the individual experimental quantities and applied with the known uncertainties. The maximum uncertainties of non-dimensional parameters were found to be ±1.6% for Reynolds number, ±4% for Nusselt number, and ±4.2% for friction factor. 4. Results and discussion 4.1. Validation test of the plain tube results The results obtained from present experiments on heat transfer and friction factor characteristics of the plain tube were verified in terms of Nusselt number and friction factor. The Nusselt number and friction factor data obtained from the present plain tube were validated with those from the proposed correlations by Gnielinski [29] and Petukhov [30] for the Nusselt number and friction factor in Fig. 3(a) and (b), respectively. The results obtained from the present plain tube were agreed well with those from the proposed correlations within ±5% and ±4% deviations for the Nusselt number and Fig. 2. (a) Geometry of the test section fitted with the twisted wire brush insert. (b) Geometric parameters of the twisted wire brush insert.

M.M.K. Bhuiya et al. / International Communications in Heat and Mass Transfer 39 (2012) 1505 1512 1509 friction factor, respectively. These results revealed the accuracy of the present experimental facility and the measurement technique. The correlations obtained from the present plain tube results for the Nusselt number and the friction factor, respectively, were given as follows: Nu ¼ 0:0137Re 0:843 Pr 0:33 f ¼ 0:431Re 0:292 4.2. Heat transfer characteristics ð14þ ð15þ Fig. 4(a) and (b) shows the relationship between the Nusselt number and the Reynolds number of the twisted wire brush inserts of different wire densities. It could be shown from Fig. 4(a) that the trend of Nusselt number was similar for both the plain tube and the tube with twisted wire brush inserts. It was shown from Fig. 4(a) that for all cases, the Nusselt number increased with increasing Reynolds number. This was attributed to the increase of turbulent intensity as the Reynolds number was increased, which led to an amplification of convective heat transfer. As expected from Fig. 4(a), the Nusselt number obtained from the tube with twisted wire brush inserts was significantly higher than those of the plain tube. The twisted wire brush inserts caused swirl a Nusselt number, Nu b 110 90 70 50 30 Experimental Gnielinski correlation 10 0.040 0.035 Experimental Petukhov correlation flow or secondary flow and pressure gradient might be created along the radial direction through the tube. Furthermore, the swirl enhanced the flow turbulence, which led to even better convection heat transfer. From the experimental results, it could be observed that the heat transfer rate increased with the increase of wire density of the twisted wire brush inserts. This could be explained by the fact that at higher wire density, stronger swirl intensity was generated, which led to more efficient interruption of boundary layer along the flow path. It could be shown from Fig. 4(a) that the the tube with the higher wire density () provided the higher heat transfer rate than those of the tube with the lower ones i.e. wire densities of 100, 150, and. The effectiveness of heat transfer enhancement of the tube equipped with the twisted wire brush inserts compared to that of the plain tube, in terms of Nusselt number ratio (Nu p /Nu s ), is presented in Fig. 4(b). It was observed from Fig. 4(b) that the Nusselt number ratio of all the investigated cases was consistently higher than the unity. This implied the beneficial gain for heat transfer enhancement of using the twisted wire inserts over the plain tube. From Fig. 4(b), it could be noted that the Nussselt number ratio tended to decrease with increasing Reynolds number. This was because of the influence of twisted wire brush insert on heat transfer enhancement was less significant for increasing Reynolds number. According to the experimental results, the Nusselt numbers of the tube with twisted a Nusselt number, Nu b 240 210 180 150 120 90 60 30 Plain tube 0 2.75 2.35 Friction factor, f 0.030 0.025 0.020 Nup/Nus 1.95 1.55 0.015 0.010 Fig. 3. Validation of the plain tube: (a) Nusselt number and (b) friction factor. 1.15 Fig. 4. Relationship between the Nusselt number and Reynolds number: (a) Nu and (b) Nu p /Nu s.

1510 M.M.K. Bhuiya et al. / International Communications in Heat and Mass Transfer 39 (2012) 1505 1512 wire brush inserts varied from 1.25 to 2.15 times than those over the plain tube values. 4.3. Fluid flow characteristics The effect of the wire density of the tube fitted with the twisted wire brush inserts on friction factor characteristics at different Reynolds number is presented in Fig. 5(a). It could be clearly depicted from Fig. 5(a) that the friction factor continue to decrease with increasing Reynolds number. As expected from Fig. 5(a), the friction factors obtained from the tube with twisted wire brush inserts were higher than those of the plain tube. This was because of the swirl flow, drag forces, and the turbulence augmentation produced by the twisted wire brush inserts [25]. As shown in Fig. 5(a), the friction factor tended was increased with increasing twisted wire density. This could be attributed to the use of twisted wire brush insert with higher wire density which caused stronger swirl or turbulence flow and long residence time in the tube. From Fig. 5(a), it could be shown that the tube with the twisted wire brush insert with higher wire density () provided the higher friction factor than those of the tube with the lower ones (100,150, and ). Fig. 5(b) represents the variation of friction factor ratio (f p /f s ) with Reynolds number for different twisted wire densities. It was shown from Fig. 5(b) that the friction a Friction factor, f b fp/fs 0.07 0.05 0.03 0.01 2.50 2.25 2.00 1.75 1.50 1.25 Plain tube 1.00 Fig. 5. Relationship between the friction factor and Reynolds number: (a) f and (b) f p /f s. factor ratio tended to decrease with raising the Reynolds number for all the investigated cases. Over the range investigated, the friction factors for the tube fitted with the twisted wire inserts varied from 1.35 to 2.0 times than those of the plain tube values. 4.4. Performance evaluation In order to appraise the heat transfer augmentation performance of the four different wire densities of twisted wire brush inserts with plain tube, a constamt blower power comparison was made. The performance was evaluated on the basis of constant blower power the correlation proposed by Usui et al. [31]. According to constant blower power performance evaluation criteria [32] could be written as: _V ΔP ¼ V _ ΔP ð16þ s p From the relationship between the friction factor and Reynolds number it could be expressed as: f Re 3 ¼ f s Re3 p Re s ¼ Re p 1 f 3 p f s ð17þ ð18þ The thermal performance factor at identical blower power for turbulent flow was calculated as the following correlation proposed by Usui et al. [31] Nu p Nu s η ¼ 0:291 f p f s ð19þ The variation of thermal performance factor with the Reynolds number is represented in Fig. 6. The thermal performance factor was decreased with the increase of Reynolds number for all the investigated cases. A performance analysis was performed to evaluate the net energy gain of all the tested inserts based on the constant blower power. It was found efficient from an energy point of view by applying all of these tested inserts as the performance factors were greater than the unity. As expected from Fig. 6, the tube with inserts at lower Reynolds number provided higher thermal performance. The heat transfer performance was obtained for the tube with twisted wire brush inserts of different wire densities, and found to be 1.1 to 1.85 times higher compared to those of the plain tube values. 4.5. Correlations for prediction of heat transfer, friction factor, and thermal performance factor The correlations were developed for the turbulent flow region in a wide range of Reynolds number 7,200 to 50,200. The correlations developed for Nusselt number, friction factor, and thermal performance factor obtained from the present experimental results of the tube fitted with the twisted wire brush inserts could be written in terms of wire density (y), Reynolds number (Re), and Prandtl number (Pr) in Eqs. (20) (22), respectively. Nu p ¼ 6 10 9 y 3 þ 4 10 6 y 2 0:0006y þ 0:0945 :Re fð 110 8 Þy 3 ð610 6 Þy 2 þ0:0011yþ0:6628g 0:33 :Pr ð20þ

M.M.K. Bhuiya et al. / International Communications in Heat and Mass Transfer 39 (2012) 1505 1512 1511 Thermal performance factor, η 1.7 1.5 1.3 Predicted Nusselt number, Nup 250 210 170 130 90 50 +4% - 3% 1.1 Fig. 6. Relationship between the thermal performance factor and Reynolds number. 10 10 50 90 130 170 210 250 Experimental Nusselt number, Nu Fig. 7. Comparison between the predicted and experimental Nusselt number. f p ¼ 2 10 7 y 3 þ 0:0001y 2 0:0076y þ 1:4492 :Re fð ð 610 9 Þy 3 210 6 Þy 2 0:0001y 0:3651 g ð21þ η p ¼ 38:626:C:C 0:629 1 :Re fð 6:22410 9 Þy 3 ð4:74110 6 Þy 2 þ0:001163y 0:12196g ð22þ n o where C ¼ 6 10 9 y 3 þ 4 10 6 y 2 0:0006y þ 0:0945 n o and C 1 ¼ 2 10 7 y 3 þ 0:0001y 2 0:0076y þ 1:4492 The Nusselt number, friction factor and thermal performance factor values predicted from the above correlations Eqs. (20) (22) were compared with the experimental values, and the comparisons are shown in Figs. 7 9, respectively. From Figs. 7 9, it could be noted that the Nusselt number, friction factor, and thermal performance factor values obtained from the predicted correlations Eqs. (20) (22) agreed well with the experimental values for all the investigated cases within the range of +4% to 3%, ±3% and +6% to 3% deviations of the proposed correlations, respectively. 5. Conclusion An experimental study was conducted to investigate the heat transfer performance and friction factor characteristics for turbulent flow through a tube by means of twisted wire brush inserts. The study revealed that the twisted wire brush inserts provided significant enhancement of heat transfer with the corresponding increase in friction factor. It was found that the Nusselt number, friction factor, and thermal performance factor increased with the increase of twisted wire densities. Based on the experimental results, key findings of this study could be summarized as follows: The Nusselt number obtained for the tube with twisted wire brush inserts varied from 1.25 to 2.15 times in comparison to those of the plain tube. The twisted wire brush insert of wire density showed the highest heat transfer performance among the twisted wire inserts. The friction factor achieved for the tube with twisted wire brush inserts varied from 1.35 to 2.0 times than those of the plain tube values at the comparable Reynolds number. The thermal performance factor (η) obtained for the tube with twisted wire brush inserts varied from 1.1 to 1.85 times than those of the plain tube values at constant blower power. The empirical correlations were developed in the present study which predicted the results of the Nusselt number, friction factor, and thermal performance factor. The maximum deviations between the predicted and experimental results for Nusselt number, friction factor, and thermal performance factor were found to be+4% to 3%,±3%,and+6% to 3%, respectively. Acknowledgements The authors would like to acknowledg the Department of Chemical Engineering, University of Malaya, Malaysia for their support in this work. The Chittagong University of Engineering and Technology (CUET) authority is highly acknowledged for necessary assistance to do this research. Predicted friction factor, fp 0.080 0.070 0.060 0.050 0.040 0.030 0.020 +3% - 3% 0.010 0.010 0.020 0.030 0.040 0.050 0.060 0.070 0.080 Experimental friction factor, f Fig. 8. Comparison between the predicted and experimental friction factor.

1512 M.M.K. Bhuiya et al. / International Communications in Heat and Mass Transfer 39 (2012) 1505 1512 Predicted thermal performance factor, hp 2.00 1.80 1.60 1.40 1.20 1.00 0.80 0.60 References +6% - 3% 0.40 0.40 0.60 0.80 1.00 1.20 1.40 1.60 1.80 2.00 Experimental thermal performance factor, h Fig. 9. Comparison between the predicted and experimental thermal performance factor, η p and experimental thermal performance factor η. [1] S.K. Agarwal, M.R. Rao, Heat transfer augmentation for the flow of a viscous liquid in circular tubes using twisted tape inserts, International Journal of Heat and Mass Transfer 39 (1996) 3547 3557. [2] S.S. Hsieh, F.Y. Wu, H.H. Tsai, Turbulent heat transfer and flow characteristics in a horizontal circular tube with strip-type inserts: Part I. Fluid mechanics, International Journal of Heat and Mass Transfer 46 (2003) 823 835. [3] S. Eiamsa-ard, C. Thianpong, P. 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