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Experimental study on turbulent flow and heat transfer in an air to water heat exchanger using perforated circular-ring
Experimental Thermal and Fluid Science 70 (2016) 185–195
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Experimental Thermal and Fluid Science journal homepage: www.elsevier.com/locate/etfs
Experimental study on turbulent flow and heat transfer in an air to water heat exchanger using perforated circular-ring M. Sheikholeslami ⇑, M. Gorji-Bandpy, D.D. Ganji Department of Mechanical Engineering, Babol University of Technology, Babol, Iran
a r t i c l e
i n f o
Article history: Received 16 April 2015 Accepted 6 September 2015 Available online 16 September 2015 Keywords: Air to water heat exchanger Pressure loss Double pipe Heat transfer Circular-ring
a b s t r a c t In this study, heat transfer and pressure loss in an air to water double pipe heat exchanger are experimentally investigated. Typical circular-ring (TCR) and perforated circular-ring (PCR) turbulators are placed in annular pipe. The working fluids are air, flowing in the annular pipe, and water through the inner circular tube. The experiments are conducted for different governing parameters namely; air flow Reynolds number (6000–12,000), pitch ratio (1.83, 2.92 and 5.83) and number of perforated hole (0, 2, 4 and 8). Correlations for friction factor, Nusselt number and thermal performance are presented according to experimental data. Results indicated that using PCRs leads to obtain lower heat transfer enhancement than the CRs because of reduction of intersection angle between the velocity and the temperature field. Thermal performance increases with increase of N but it decreases with increase of Reynolds number and pitch ratio. Ó 2015 Elsevier Inc. All rights reserved.
1. Introduction One of the significant types of heat exchanger is air to water heat exchanger. This kind of heat exchanger has various applications such as: apartment buildings and condominiums, residential heating, hybrid systems, air conditioning, dehumidification. Utilize of augmentation techniques lead to increase in heat transfer coefficient but at the cost of enhance in pressure drop. To reach high heat transfer rate while taking care of the augment pumping power, various techniques have been presented in recent decade. Currently, swirl flow devices have widely been used for increasing the convective heat transfer in various industries. This application is because of their low cost and easy setting up. Vermahmoudi et al. [1] studied the overall heat transfer coefficient of water based iron oxide nanofluid in a compact air cooled heat exchanger. They indicated that the overall heat transfer coefficient and the heat transfer rate of nanofluid have been improved with increase of air flow Reynolds number. Buchlin [2] investigated effect of perforated ribs in a channel flow. He tested five types of perforated ribs made in Plexiglas and found that the optimum design of ribs combines a rib pitch ratio of 5 with an open area ratio of 0.53. Prom-
⇑ Corresponding author. E-mail addresses: [email protected], yahoo.com (M. Sheikholeslami). http://dx.doi.org/10.1016/j.expthermflusci.2015.09.002 0894-1777/Ó 2015 Elsevier Inc. All rights reserved.
mohsen.sheikholeslami@
vonge and Eiamsa-ard [3] presented the effect of a free-spacing snail entry together with conical-nozzle turbulators on turbulent heat transfer and friction characteristics in a uniform heat-flux tube. Bayrak et al. [4] studied the performance assessment of porous baffles inserted in solar air heaters (SAHs). They showed that the highest collector efficiency and air temperature rise are achieved by SAHs with a thickness of 6 mm and an air mass flow rate of 0.025 kg/s. Sheikholeslami et al. [5] studied about swirl flow devices effect on fluid flow and heat transfer. Helical-wire-coils fitted inside a round tube have been experimentally studied by Garcia et al. [6]. They found that wire coil inserts offer their best performance within the transition region. Response surface methodology (RSM) based on central composite design (CCD) was applied by Hatami et al. [7] to obtain an optimization design of finned type heat exchangers (HEX) to recover waste heat from the exhaust of a diesel engine. Parametric analysis and optimization of entropy generation in unsteady MHD flow over a stretching rotating disk was investigated by Rashidi et al. [8]. Durmus et al. [9] used snail entrance in order to increase heat transfer in concentric double-pipe heat exchangers. They concluded that the swirl flow effect of the snail caused some increase in pressure drop while this effect was unimportant compared with the improvement in heat transfer capacity. Bandos et al. [10] studied the effects of thermal storage and vertical temperature variations on energy pile heat exchangers. Promvonge and Eiamsa-ard [11] used combined conical-nozzle inserts and swirl generator in
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Nomenclature A Ai Ao cp d D f ho hi k ‘ L Nu N Pr P PR Q Re
heat transfer area inner pipe inside surface area inner pipe outer surface area specific heat at constant pressure diameter of inner pipe diameter of outer pipe friction factor (dimensionless) average heat transfer coefficient of cold fluid average heat transfer coefficient of hot fluid thermal conductivity length of pipe length of test section Nusselt number number of perforated hole Prandtl number pressure pitch ratio (=P/Do) flow rate of water Reynolds number
order to enhance in heat transfer. They showed that use of the conical nozzle in common with the snail leads to a maximum heat transfer rate that is up by 316%. Free convection heat transfer in a concentric annulus between a cold square and heated elliptic cylinders in presence of magnetic field was investigated by Sheikholeslami et al. [12]. They found that the enhancement in heat transfer increases as Hartmann number increases but it decreases with increase of Rayleigh number. Yakut and Sahin [13] used conical-ring turbulators placed inside the tube to produce reverse flows in each module of the conical rings. In their experimental study, the level of the reverse flow (re-circulation flow) was generated from the separation and reattachment of a boundary layer from different pitch lengths between the modules. The main purpose of this paper is to investigate the effect of typical circular-ring (TCR) and perforated circular-ring (PCR) turbulators on flow and heat transfer in an air to water double pipe heat exchanger. Experimental set up and formulas for calculating heat transfer rate, friction factor and thermal performance are presented. The effects of Reynolds number, pitch ratio and number of perforated hole on heat transfer rate and pressure drop are studied. 2. Experimental set-up and procedure The experimental set-up is shown in Fig. 1(a). The dimensions of the inner and outer pipes of the heat exchanger are: Di ¼ 2:8 cm; Do ¼ 3 cm; di ¼ 5 cm; do ¼ 6 cm. The length of the pipe is ‘ = 2 m and the length of test section is L = 1.2 m. Hot water is passed through the inner pipe, while cold air is flowing through
T U
fluid temperature overall heat transfer coefficient
Greek symbols thermal diffusivity DP pressure drop (Pa) l dynamic viscosity of nanofluid h dimensionless temperature q density g thermal performance
a
Subscripts i inner o outer a air w water s smooth pipe
the annulus. Heating of the water was achieved with an electrical heater at the upper tank (three heaters are used in the upper tank with the capacity of 2 kW, 2 kW and 3 kW). In the experimental work, it is intended to search for the changes in the heat transfer coefficients of the air side turbulent flow by affecting the regions near the wall of the pipe flow. The inner tube is made from copper with thermal conductivity ðk ¼ 300 kcal=ðm h CÞÞ, while the outer tube is made from Plexiglas with an outer with thermal conductivity ðk ¼ 5 104 kcal=ðm h CÞÞ. The inlet and outlet temperatures of the fluids (air and water), the temperatures of the points on the inner pipe wall (six points), the temperatures of four points in different distant from inner wall and ambient temperature were measured with Sheathe type thermocouples (element C.A; class 0.75) (Fig. 1(b)). An ST-8920 differential pressure is used to obtain the pressure drop in air side. It can measure the pressures in ±5000 Pa with 1 Pa resolution. In order to transfer the water from the lower tank to upper tank, a pump with the head of 5.5 m, is used. The inlet bulk air at 28 °C from a 0.75 kW blower was directed through an orifice meter and passed to the heat transfer test section. The volumetric air flow rates from the blower, situated before the inlet of the test tube, are adjusted by varying the motor speed through the SV008iG5A-2 inverter. The flow rates of the water are adjusted with valves and measured with rotameter. The experimental work is repeated for counter flow modes at various Reynolds numbers. The physical properties of air and water are variable with temperature as illustrated in Tables 1 and 2. In the test section, circular ring are used in order to heat transfer enhancement. Also perforated circular ring have been used in this study (Fig. 2).
Table 1 Temperature-dependent properties of air. Coefficient
A1 A2 A3 A4 A5
A1 + A2 T + A3 T2 + A4 T3 + A5 T4 Properties of air
q ðkg=m3 Þ
C p ðJ=ðkg KÞÞ
l ðkg=ðm sÞÞ
k ðW=ðm KÞÞ
4.5399557047065677 2.3244292640615217E2 5.6404522707476041E5 6.2803748539876179E8 2.3678170919661321E11
1.0540764984602797E+3 3.5067618164922393E1 5.8416753365658986E4 3.0329858178609656E7 5.2479296621138882E10
9.4680032779877928E5 1.0222587861878098E6 4.7054455296163551E9 9.1119064881185846E12 6.5461225665736524E15
1.8028147194179223E2 1.6851766935888901E4 1.3838388187738584E6 3.2630462746304979E9 2.7514584927209003E12
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Fig. 1. (a) Schematic diagram of the experimental setup; (b) test section and thermocouples.
To quantify the uncertainties of measurements the reduced data obtained experimentally were determined. The uncertainty in the data calculation was based on Schultz and Cole method [14]:
" UR ¼
n X @R i¼1
@V i
2 #1=2 UV i
ð1Þ
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Table 2 Temperature-dependent properties of water. Coefficient
A1 A2 A3 A4 A5
A1 + A2 T + A3 T2 + A4 T3 + A5 T4 Properties of water
q ðkg=m3 Þ
C p ðJ=ðkg KÞÞ
lðkg=ðm sÞÞ
kðW=ðm KÞÞ
1.6622104933785317E+2 1.2256322983468429E+1 4.6535103004960353E2 7.7101273744096163E5 5.0319235543371908E8
1.2201774895976883E+4 9.2961742884825355E+1 4.0724280562804471E1 8.033901613887863E4 6.0554273200519027E7
4.5563422230298373E1 5.266709499675417E3 2.2937228364977076E5 4.4517867607521202E8 3.2451565252286636E11
2.1117772306964272E1 4.0615080360954991E3 4.0530952053441623E5 9.5665206133793231E8 6.772213049004531E11
Table 3 Constant coefficient for using Eq. (15). aij
i=1
i=2
i=3
i=4
i=5
i=6
j=1 j=2
0.000395 7.352294
0.007348 0.737743
0.01396 0.05672
1.1E07 0.003036
0.124965 0.032623
0.0003 0.01492
Fig. 2. (a) Test section; (b) typical circular-ring (TCR) and perforated circular-ring (PCR) turbulators.
where UR is the total error, U V i is the error of each independent variable and n is the number of total variables. The uncertainty analysis showed that the measuring errors were less than 10% for all the experiments presented in this study.
3. Measurement of heat transfer coefficient and friction factor The data reduction of the measured results is summarized in the following procedures.
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Heat transferred to the air in the test section, Qa, can be calculated from
Q a ¼ ma C p;a ðT a;out T a;in Þ
ð2Þ
where ma is the mass flow rate of air, Cp,a is the specific heat of air, Ta,in and Ta,out are the inlet and outlet air temperatures, respectively. Heat transferred from the water, Qw, can be calculated from
Qw ¼
mw
C p;w ðT w;in T w;out Þ
ð3Þ
189
where mw is the mass flow rate of water, Cp,w is the specific heat of water, Tw,in and Tw,out are the inlet and outlet water temperatures, respectively. The average heat transfer rate, Qave, used in the calculation is determined from the water side and air side as follows:
Q av e ¼ ðQ a þ Q w Þ=2
ð4Þ
The water side heat transfer coefficient, hi, can be calculated from the average heat transfer rate obtained from
Fig. 3. Verification of friction factor and Nusselt number for smooth heat exchanger.
Fig. 4. Effects of Reynolds number (Rea), number of perforated hole (N) and pitch ratio (PR) on friction factor (f).
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Q av e ¼ hi Ai ðT s;av e T w;av e Þ
ð5Þ
where Ts,ave is the average inner wall temperature, Tw,ave is the average water temperature and Ai is the inner surface area of the tube (Ai = pDiL). The overall heat transfer coefficient, Ui, can be determined from
Q av e ¼ U i Ai DT LMTD
ð6Þ
where DTLMTD is the logarithmic mean temperature difference. An outside (air side) heat transfer coefficient, ho, is usually obtained from the overall thermal resistance consisting of three resistances in series: the convective resistance on the inner surface, the conductance resistance of the pipe wall and the convective resistance on the outer surface
1 1 lnðDo =Di Þ 1 ¼ þ þ U i Ai hi Ai 2pkL ho Ao
ð7Þ
Average Nusselt number along the outer pipe was calculated as follows:
Nuo ¼
ho D H kair
ð8Þ
where kair is thermal conductivity of air in bulk temperature and DH is the hydraulic diameter of the tube (DH = di Do). The friction factor (f) can be calculated from
f ¼
DP ðqu2 =2ÞðL=DH Þ
ð9Þ
where DP is the pressure drop, q is the density of air, u is the velocity of air and L is the length of the tube. For a constant pumping power,
ðV_ DPÞs ¼ ðV_ DPÞ
ð10Þ
and the relationship between friction and Reynolds number can be expressed as: 1
ðf Re3 Þs ¼ ðf Re3 Þ ! Res ¼ Reðf =f s Þ3
ð11Þ
The thermal performance factor (g) defined as the ratio as follows:
g¼
ðNu=Nus Þ 1
ðf =f s Þ3
ð12Þ
4. Results and discussion In this paper, effect of typical circular-ring (TCR) and perforated circular-ring (PCR) turbulators on the heat transfer and fluid friction characteristics in an air to water double pipe heat exchanger is investigated. The present experimental results on heat transfer and friction characteristics in a smooth heat exchanger are first validated in terms of Nusselt number and friction factor. The Nusselt number and friction factor obtained from the present smooth heat exchanger are, respectively, compared with the correlations
Fig. 5. Effects of Reynolds number (Rea), number of perforated hole (N) and pitch ratio (PR) on Nusselt number (Nu).
M. Sheikholeslami et al. / Experimental Thermal and Fluid Science 70 (2016) 185–195
Fig. 6. Effects of Reynolds number (Rea), number of perforated hole (N) and pitch ratio (PR) on friction factor ratio (f/fs).
Fig. 7. Effects of Reynolds number (Rea), number of perforated hole (N) and pitch ratio (PR) on Nusselt number ratio (Nu/Nus).
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of Gnielinski for Nusselt number and of Petukhov for friction factor found in the open literature [15] for turbulent flow in pipe. Correlation of Gnielinski,
Nu ¼
ðf =8ÞðRe 1000ÞPr 0:5
2
1 þ 12:7ðf =8Þ ðPr 3 1Þ
;
3000 < Re < 5 106
ð13Þ
Correlation of Petukhov,
f ¼ ð0:79LnðReÞ 1:64Þ2 ;
3000 6 Re 6 5 106
ð14Þ
Fig. 3 shows a comparison of the Nusselt number and friction factor obtained from the present work with those from correlations of Eqs. (13) and (14). In the figure, the present results agree very well within ±12% with the published correlations.
Fig. 4 shows the effects of Reynolds number, number of perforated hole and pitch ratio on friction factor. Obviously, friction factor decreases with rise of Reynolds number. Friction factor increases with decreasing pitch ratio. In the other word, a decrease of the distance between each pair of the turbulators causes an increase in friction factor. This is due to this fact that the smaller distance between each pair of the turbulators, the more numbers of turbulators available in the heat exchanger, thus the more blockage against the flowing stream. Clearly, friction factor decreases with the increase of the number of perforated hole due to the reduction of turbulent fluctuation or eddy motion and the appearance of reverse flow between the each pair of the turbulators. Besides, it can be indicated that friction factors in the heat exchan-
Fig. 8. Effects of Reynolds number (Rea), number of perforated hole (N) and pitch ratio (PR) on thermal performance factor (g).
Table 4 Constant coefficient for using Eq. (16). bij
i=1
i=2
i=3
i=4
i=5
i=6
j=1 j=2
0.326821 0.035122
0.017554544 0.859826928
0.0499 0.01412
0.000396 0.958071
0.00295 0.001889
0.000917 0.03401
Table 5 Constant coefficient for using Eq. (17). cij
i=1
i=2
i=3
i=4
i=5
i=6
j=1 j=2 j=3
7.74E07 0.924403 0.24951
0.000213 0.062311 1.188146
5.75E05 0.06636 0.349578
1.3E08 0.00288 1.29671
0.001121 0.005529 0.31755
3E06 7.7E05 1.334045
M. Sheikholeslami et al. / Experimental Thermal and Fluid Science 70 (2016) 185–195
ger equipped with the PCRs of N = 2, 4 and 8 holes, 1.550133– 3.578695, 1.575096–3.43694 and 1.807108–3.703923 times lower than those in the heat exchanger with typical CR, depending on pitch ratio. This indicates that the presence of the perforated hole in the conical ring possesses high potential for reduction of friction in the heat exchanger. Fig. 5 depicts the effects of Reynolds number, number of perforated hole and pitch ratio on Nusselt number. Rate of heat transfer is improved at high Reynolds numbers because the convective heat transfer is promoted more effectively at a higher turbulence level. Nusselt numbers for heat exchanger equipped with turbulators are higher than those found from the plain heat exchanger for a given Reynolds number. This is due to the interruption of flow by the turbulators which results in the destruction of thermal boundary layer near the heat exchanger wall. Nusselt number noticeably increases with decease of pitch ratio which is in similar trend found for friction factor. The quantitative results reveal that the mean heat transfer rate the in the heat exchanger with perforated circular-rings (N = 8) at smallest pitch ratio (PR = 1.83) is 1.037183, 1.140188 and 1.367915 times higher than those in the heat exchanger with PCRs at PR = 2.92, 5.83 and plain heat exchanger, respectively. Furthermore, the TCRs provide higher heat transfer rates than those offered by the PCRs. This is due to the lower turbulence intensity in the heat exchanger. The mean heat transfer rates in the heat exchanger with TCRs for all pitch ratios studied are around 1.064442–1.131967 times of those in the heat exchanger with PCRs with N = 8. The Nusselt numbers in the heat exchanger equipped with the PCRs with PR = 1.83 of N = 2, 4 and 8 holes,
193
are respectively 1.454797, 1.450367 and 1.367915 times higher than those in the plain heat exchanger. In contrast, the Nusselt numbers in the heat exchanger equipped with the PCRs of N = 2, 4 and 8 holes are respectively, 1.004903–1.064344, 1.03521– 1.067569 and 1.080317–1.131967 times lower than those in the heat exchanger with typical CR, depending on pitch ratio. Effects of Reynolds number, number of perforated hole and pitch ratio on friction factor ratio are shown in Fig. 6. Friction factor ratio increases with the rise of Reynolds number. This can be attributed to higher flow blockage, larger surface area and the act caused by the reverse flow. It can be observed that friction factor ratio shows a steeper increment at the beginning before gradually increasing for higher values of number of perforated hole. Also it can be found that friction ratio increases with decrease of number of perforated hole and pitch ratio. Effects of Reynolds number, number of perforated hole and pitch ratio on Nusselt number ratio are shown in Fig. 7. Nusselt number ratio decreases with the rise of Reynolds number and number of perforated hole while it decreases with increase of pitch ratio. Fig. 8 shows the effects of Reynolds number, number of perforated hole and pitch ratio on thermal performance factor. It is found that the thermal performance factor decreases with augment of Reynolds number and it is also observed that the turbulators with the smallest pitch ratio (PR) provide the highest thermal performance factor for both of the TCR and PCR turbulators. According to the quantitative results, the maximum thermal performance factors of heat exchanger fitted with PCRs with N = 8 at PR = 1.83, 2.99, and 5.83, are found to be 1.119634, 1.108086 and 1.031247, respectively. The performance factors of heat exchanger with the
Fig. 9. Comparison of experimental data with those calculated from the correlation for (a) friction factor; (b) Nusselt number; (c) thermal performance.
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TCRs are around 1.144314–1.366547 times lower than those of the heat exchanger with PCRs with smallest pitch ratio. This indicates that at the same pumping power the PCRs can enhance heat transfer more efficient than the TCRs and this confirms the greater benefit of the use of PCRs as energy saving device over the TCRs. The corresponding polynomial representations of Nusselt number along air side (Nu), friction factor (f) and thermal performance factor (g) are as follow:
Nu ¼ a12 þ a22 Y 1 þ a32 N þ a42 Y 21 þ a52 N2 þ a62 Y 1 N Y 1 ¼ a11 þ a21 Rea þ a31 PR þ a41 Re2a þ a51 PR2 þ a61 Rea PR f ¼ b12 þ b22 Y 1 þ b32 PR þ b42 Y 21 þ b52 PR2 þ b62 Y 1 PR Y 1 ¼ b11 þ b21 Rea þ b31 N þ b41 Re2a þ b51 N2 þ b61 Rea N
g ¼ c16 þ c26 Y 1 þ c36 Y 2 þ Y 1 ¼ c11 þ Y 2 ¼ c12 þ
c46 Y 21
þ
c56 Y 22
ð15Þ
ð16Þ
þ c66 Y 1 Y 2
c21 Rea þ c31 N þ c41 Re2a þ c51 N2 þ c61 Rea N c22 N þ c32 PR þ c42 N2 þ c52 PR2 þ c62 NPR
ð17Þ
Also aij, bij and cij can be found in Tables 3–5, respectively. Fig. 9 depicts the comparison of the friction factor, Nusselt number and thermal performance factor between experimental data and those calculated from the present correlations. It is found that the majority of the measured data falls within ±10%, ±5% and ±7%, for f ; Nu and g. Contour plots of friction factor, Nusselt number and thermal performance according to current correlations are shown in Fig. 10. The heat transfer rate and friction factor of PCRs increase with decreasing pitch ratio (PR) and decreasing number of perforated hole (N). However, the thermal performance factor increases with increasing number of perforated hole and decreasing pitch ratio. 5. Conclusion Heat transfer and friction factor characteristics in the air to water heat exchanger equipped with the typical circular-ring (TCR) and perforated circular-ring (PCR) turbulators are investigated experimentally. The effects of the pitch ratio and number of perforated hole on flow and heat transfer characteristics are considered. The correlations of the Nusselt number, friction factor and
f
(a) PR = 1.83
(b) N = 8
Nu
(c) PR = 1.83
(d) N = 8
(e) PR = 1.83
( f) N = 8
η
Fig. 10. Contour plots of friction factor, Nusselt number and thermal performance according to current correlations.
M. Sheikholeslami et al. / Experimental Thermal and Fluid Science 70 (2016) 185–195
thermal performance are presented. Results show that using PCRs leads to obtain lower heat transfer enhancement than the CRs while they have lower friction factor. Consequently, the thermal performance factor of PCRs is higher than those of the CRs over the range studied. Nusselt number increases with increase of Reynolds number and number of perforated hole and it decreases with increase of pitch ratio. Friction factor enhances with increase of Reynolds number and it decreases with increase of number of perforated hole and pitch ratio. References [1] Y. Vermahmoudi, S.M. Peyghambarzadeh, S.H. Hashemabadi, M. Naraki, Experimental investigation on heat transfer performance of Fe2O3/water nanofluid in an air-finned heat exchanger, Eur. J. Mech. B/Fluids 44 (2014) 32–41. [2] J.M. Buchlin, Convective heat transfer in a channel with perforated ribs, Int. J. Therm. Sci. 41 (2002) 332–340. [3] P. Promvonge, S. Eiamsa-ard, Heat transfer in a circular tube fitted with freespacing snail entry and conical-nozzle turbulators, Int. Commun. Heat Mass Transfer 34 (2007) 838–848. [4] F. Bayrak, H.F. Oztop, A. Hepbasli, Energy and exergy analyses of porous baffles inserted solar air heaters for building applications, Energy Build. 57 (2013) 338–345. [5] M. Sheikholeslami, M. Gorji-Bandpy, D.D. Ganji, Review of heat transfer enhancement methods: Focus on passive methods using swirl flow devices, Renew. Sust. Energ. Rev. 49 (2015) 444–469.
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[6] A. Garcia, P.G. Vicente, A. Viedma, Experimental study of heat transfer enhancement with wire coil inserts in laminar-transition-turbulent regimes at different Prandtl numbers, Int. J. Heat Mass Transfer 48 (2005) 4640–4651. [7] M. Hatami, M. Jafaryar, D.D. Ganji, M. Gorji-Bandpy, Optimization of finnedtube heat exchangers for diesel exhaust waste heat recovery using CFD and CCD techniques, Int. Commun. Heat Mass Transfer 57 (2014) 254–263. [8] M.M. Rashidi, M. Ali, N. Freidoonimehr, F. Nazari, Parametric analysis and optimization of entropy generation in unsteady MHD flow over a stretching rotating disk using artificial neural network and particle swarm optimization algorithm, Energy 55 (15) (2013) 497–510. [9] A. Durmus, A. Durmus, M. Esen, Investigation of heat transfer and pressure drop in a concentric heat exchanger with snail entrance, Appl. Therm. Eng. 22 (2002) 321–332. [10] T.V. Bandos, A.C. Celador, L.M. Lopez-Gonzalez, J.M. Sala-Lizarraga, Finite cylinder-source model for energy pile heat exchangers: effects of thermal storage and vertical temperature variations, Energy 78 (2014) 639–648. [11] P. Promvonge, S. Eiamsa-ard, Heat transfer enhancement in a tube with combined conical-nozzle inserts and swirl generator, Energy Convers. Manage. 47 (2006) 2867–2882. [12] M. Sheikholeslami, M. Gorji-Bandpy, D.D. Ganji, Numerical investigation of MHD effects on Al2O3–water nanofluid flow and heat transfer in a semiannulus enclosure using LBM, Energy 60 (2013) 501–510. [13] K. Yakut, B. Sahin, Flow-induced vibration analysis of conical rings used of heat transfer enhancement in heat exchanger, Appl. Energy 78 (3) (2004) 273–288. [14] R. Schultz, R. Cole, Uncertainty analysis in boiling nucleation, in: AIChE Symposium Series, 1979. [15] F. Incropera, P.D. Dewitt, Introduction to Heat Transfer, fifth ed., John Wiley & Sons Inc., 2006.
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Experimental Thermal and Fluid Science journal homepage: www.elsevier.com/locate/etfs
Experimental study on turbulent flow and heat transfer in an air to water heat exchanger using perforated circular-ring M. Sheikholeslami ⇑, M. Gorji-Bandpy, D.D. Ganji Department of Mechanical Engineering, Babol University of Technology, Babol, Iran
a r t i c l e
i n f o
Article history: Received 16 April 2015 Accepted 6 September 2015 Available online 16 September 2015 Keywords: Air to water heat exchanger Pressure loss Double pipe Heat transfer Circular-ring
a b s t r a c t In this study, heat transfer and pressure loss in an air to water double pipe heat exchanger are experimentally investigated. Typical circular-ring (TCR) and perforated circular-ring (PCR) turbulators are placed in annular pipe. The working fluids are air, flowing in the annular pipe, and water through the inner circular tube. The experiments are conducted for different governing parameters namely; air flow Reynolds number (6000–12,000), pitch ratio (1.83, 2.92 and 5.83) and number of perforated hole (0, 2, 4 and 8). Correlations for friction factor, Nusselt number and thermal performance are presented according to experimental data. Results indicated that using PCRs leads to obtain lower heat transfer enhancement than the CRs because of reduction of intersection angle between the velocity and the temperature field. Thermal performance increases with increase of N but it decreases with increase of Reynolds number and pitch ratio. Ó 2015 Elsevier Inc. All rights reserved.
1. Introduction One of the significant types of heat exchanger is air to water heat exchanger. This kind of heat exchanger has various applications such as: apartment buildings and condominiums, residential heating, hybrid systems, air conditioning, dehumidification. Utilize of augmentation techniques lead to increase in heat transfer coefficient but at the cost of enhance in pressure drop. To reach high heat transfer rate while taking care of the augment pumping power, various techniques have been presented in recent decade. Currently, swirl flow devices have widely been used for increasing the convective heat transfer in various industries. This application is because of their low cost and easy setting up. Vermahmoudi et al. [1] studied the overall heat transfer coefficient of water based iron oxide nanofluid in a compact air cooled heat exchanger. They indicated that the overall heat transfer coefficient and the heat transfer rate of nanofluid have been improved with increase of air flow Reynolds number. Buchlin [2] investigated effect of perforated ribs in a channel flow. He tested five types of perforated ribs made in Plexiglas and found that the optimum design of ribs combines a rib pitch ratio of 5 with an open area ratio of 0.53. Prom-
⇑ Corresponding author. E-mail addresses: [email protected], yahoo.com (M. Sheikholeslami). http://dx.doi.org/10.1016/j.expthermflusci.2015.09.002 0894-1777/Ó 2015 Elsevier Inc. All rights reserved.
mohsen.sheikholeslami@
vonge and Eiamsa-ard [3] presented the effect of a free-spacing snail entry together with conical-nozzle turbulators on turbulent heat transfer and friction characteristics in a uniform heat-flux tube. Bayrak et al. [4] studied the performance assessment of porous baffles inserted in solar air heaters (SAHs). They showed that the highest collector efficiency and air temperature rise are achieved by SAHs with a thickness of 6 mm and an air mass flow rate of 0.025 kg/s. Sheikholeslami et al. [5] studied about swirl flow devices effect on fluid flow and heat transfer. Helical-wire-coils fitted inside a round tube have been experimentally studied by Garcia et al. [6]. They found that wire coil inserts offer their best performance within the transition region. Response surface methodology (RSM) based on central composite design (CCD) was applied by Hatami et al. [7] to obtain an optimization design of finned type heat exchangers (HEX) to recover waste heat from the exhaust of a diesel engine. Parametric analysis and optimization of entropy generation in unsteady MHD flow over a stretching rotating disk was investigated by Rashidi et al. [8]. Durmus et al. [9] used snail entrance in order to increase heat transfer in concentric double-pipe heat exchangers. They concluded that the swirl flow effect of the snail caused some increase in pressure drop while this effect was unimportant compared with the improvement in heat transfer capacity. Bandos et al. [10] studied the effects of thermal storage and vertical temperature variations on energy pile heat exchangers. Promvonge and Eiamsa-ard [11] used combined conical-nozzle inserts and swirl generator in
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Nomenclature A Ai Ao cp d D f ho hi k ‘ L Nu N Pr P PR Q Re
heat transfer area inner pipe inside surface area inner pipe outer surface area specific heat at constant pressure diameter of inner pipe diameter of outer pipe friction factor (dimensionless) average heat transfer coefficient of cold fluid average heat transfer coefficient of hot fluid thermal conductivity length of pipe length of test section Nusselt number number of perforated hole Prandtl number pressure pitch ratio (=P/Do) flow rate of water Reynolds number
order to enhance in heat transfer. They showed that use of the conical nozzle in common with the snail leads to a maximum heat transfer rate that is up by 316%. Free convection heat transfer in a concentric annulus between a cold square and heated elliptic cylinders in presence of magnetic field was investigated by Sheikholeslami et al. [12]. They found that the enhancement in heat transfer increases as Hartmann number increases but it decreases with increase of Rayleigh number. Yakut and Sahin [13] used conical-ring turbulators placed inside the tube to produce reverse flows in each module of the conical rings. In their experimental study, the level of the reverse flow (re-circulation flow) was generated from the separation and reattachment of a boundary layer from different pitch lengths between the modules. The main purpose of this paper is to investigate the effect of typical circular-ring (TCR) and perforated circular-ring (PCR) turbulators on flow and heat transfer in an air to water double pipe heat exchanger. Experimental set up and formulas for calculating heat transfer rate, friction factor and thermal performance are presented. The effects of Reynolds number, pitch ratio and number of perforated hole on heat transfer rate and pressure drop are studied. 2. Experimental set-up and procedure The experimental set-up is shown in Fig. 1(a). The dimensions of the inner and outer pipes of the heat exchanger are: Di ¼ 2:8 cm; Do ¼ 3 cm; di ¼ 5 cm; do ¼ 6 cm. The length of the pipe is ‘ = 2 m and the length of test section is L = 1.2 m. Hot water is passed through the inner pipe, while cold air is flowing through
T U
fluid temperature overall heat transfer coefficient
Greek symbols thermal diffusivity DP pressure drop (Pa) l dynamic viscosity of nanofluid h dimensionless temperature q density g thermal performance
a
Subscripts i inner o outer a air w water s smooth pipe
the annulus. Heating of the water was achieved with an electrical heater at the upper tank (three heaters are used in the upper tank with the capacity of 2 kW, 2 kW and 3 kW). In the experimental work, it is intended to search for the changes in the heat transfer coefficients of the air side turbulent flow by affecting the regions near the wall of the pipe flow. The inner tube is made from copper with thermal conductivity ðk ¼ 300 kcal=ðm h CÞÞ, while the outer tube is made from Plexiglas with an outer with thermal conductivity ðk ¼ 5 104 kcal=ðm h CÞÞ. The inlet and outlet temperatures of the fluids (air and water), the temperatures of the points on the inner pipe wall (six points), the temperatures of four points in different distant from inner wall and ambient temperature were measured with Sheathe type thermocouples (element C.A; class 0.75) (Fig. 1(b)). An ST-8920 differential pressure is used to obtain the pressure drop in air side. It can measure the pressures in ±5000 Pa with 1 Pa resolution. In order to transfer the water from the lower tank to upper tank, a pump with the head of 5.5 m, is used. The inlet bulk air at 28 °C from a 0.75 kW blower was directed through an orifice meter and passed to the heat transfer test section. The volumetric air flow rates from the blower, situated before the inlet of the test tube, are adjusted by varying the motor speed through the SV008iG5A-2 inverter. The flow rates of the water are adjusted with valves and measured with rotameter. The experimental work is repeated for counter flow modes at various Reynolds numbers. The physical properties of air and water are variable with temperature as illustrated in Tables 1 and 2. In the test section, circular ring are used in order to heat transfer enhancement. Also perforated circular ring have been used in this study (Fig. 2).
Table 1 Temperature-dependent properties of air. Coefficient
A1 A2 A3 A4 A5
A1 + A2 T + A3 T2 + A4 T3 + A5 T4 Properties of air
q ðkg=m3 Þ
C p ðJ=ðkg KÞÞ
l ðkg=ðm sÞÞ
k ðW=ðm KÞÞ
4.5399557047065677 2.3244292640615217E2 5.6404522707476041E5 6.2803748539876179E8 2.3678170919661321E11
1.0540764984602797E+3 3.5067618164922393E1 5.8416753365658986E4 3.0329858178609656E7 5.2479296621138882E10
9.4680032779877928E5 1.0222587861878098E6 4.7054455296163551E9 9.1119064881185846E12 6.5461225665736524E15
1.8028147194179223E2 1.6851766935888901E4 1.3838388187738584E6 3.2630462746304979E9 2.7514584927209003E12
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Fig. 1. (a) Schematic diagram of the experimental setup; (b) test section and thermocouples.
To quantify the uncertainties of measurements the reduced data obtained experimentally were determined. The uncertainty in the data calculation was based on Schultz and Cole method [14]:
" UR ¼
n X @R i¼1
@V i
2 #1=2 UV i
ð1Þ
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Table 2 Temperature-dependent properties of water. Coefficient
A1 A2 A3 A4 A5
A1 + A2 T + A3 T2 + A4 T3 + A5 T4 Properties of water
q ðkg=m3 Þ
C p ðJ=ðkg KÞÞ
lðkg=ðm sÞÞ
kðW=ðm KÞÞ
1.6622104933785317E+2 1.2256322983468429E+1 4.6535103004960353E2 7.7101273744096163E5 5.0319235543371908E8
1.2201774895976883E+4 9.2961742884825355E+1 4.0724280562804471E1 8.033901613887863E4 6.0554273200519027E7
4.5563422230298373E1 5.266709499675417E3 2.2937228364977076E5 4.4517867607521202E8 3.2451565252286636E11
2.1117772306964272E1 4.0615080360954991E3 4.0530952053441623E5 9.5665206133793231E8 6.772213049004531E11
Table 3 Constant coefficient for using Eq. (15). aij
i=1
i=2
i=3
i=4
i=5
i=6
j=1 j=2
0.000395 7.352294
0.007348 0.737743
0.01396 0.05672
1.1E07 0.003036
0.124965 0.032623
0.0003 0.01492
Fig. 2. (a) Test section; (b) typical circular-ring (TCR) and perforated circular-ring (PCR) turbulators.
where UR is the total error, U V i is the error of each independent variable and n is the number of total variables. The uncertainty analysis showed that the measuring errors were less than 10% for all the experiments presented in this study.
3. Measurement of heat transfer coefficient and friction factor The data reduction of the measured results is summarized in the following procedures.
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Heat transferred to the air in the test section, Qa, can be calculated from
Q a ¼ ma C p;a ðT a;out T a;in Þ
ð2Þ
where ma is the mass flow rate of air, Cp,a is the specific heat of air, Ta,in and Ta,out are the inlet and outlet air temperatures, respectively. Heat transferred from the water, Qw, can be calculated from
Qw ¼
mw
C p;w ðT w;in T w;out Þ
ð3Þ
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where mw is the mass flow rate of water, Cp,w is the specific heat of water, Tw,in and Tw,out are the inlet and outlet water temperatures, respectively. The average heat transfer rate, Qave, used in the calculation is determined from the water side and air side as follows:
Q av e ¼ ðQ a þ Q w Þ=2
ð4Þ
The water side heat transfer coefficient, hi, can be calculated from the average heat transfer rate obtained from
Fig. 3. Verification of friction factor and Nusselt number for smooth heat exchanger.
Fig. 4. Effects of Reynolds number (Rea), number of perforated hole (N) and pitch ratio (PR) on friction factor (f).
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Q av e ¼ hi Ai ðT s;av e T w;av e Þ
ð5Þ
where Ts,ave is the average inner wall temperature, Tw,ave is the average water temperature and Ai is the inner surface area of the tube (Ai = pDiL). The overall heat transfer coefficient, Ui, can be determined from
Q av e ¼ U i Ai DT LMTD
ð6Þ
where DTLMTD is the logarithmic mean temperature difference. An outside (air side) heat transfer coefficient, ho, is usually obtained from the overall thermal resistance consisting of three resistances in series: the convective resistance on the inner surface, the conductance resistance of the pipe wall and the convective resistance on the outer surface
1 1 lnðDo =Di Þ 1 ¼ þ þ U i Ai hi Ai 2pkL ho Ao
ð7Þ
Average Nusselt number along the outer pipe was calculated as follows:
Nuo ¼
ho D H kair
ð8Þ
where kair is thermal conductivity of air in bulk temperature and DH is the hydraulic diameter of the tube (DH = di Do). The friction factor (f) can be calculated from
f ¼
DP ðqu2 =2ÞðL=DH Þ
ð9Þ
where DP is the pressure drop, q is the density of air, u is the velocity of air and L is the length of the tube. For a constant pumping power,
ðV_ DPÞs ¼ ðV_ DPÞ
ð10Þ
and the relationship between friction and Reynolds number can be expressed as: 1
ðf Re3 Þs ¼ ðf Re3 Þ ! Res ¼ Reðf =f s Þ3
ð11Þ
The thermal performance factor (g) defined as the ratio as follows:
g¼
ðNu=Nus Þ 1
ðf =f s Þ3
ð12Þ
4. Results and discussion In this paper, effect of typical circular-ring (TCR) and perforated circular-ring (PCR) turbulators on the heat transfer and fluid friction characteristics in an air to water double pipe heat exchanger is investigated. The present experimental results on heat transfer and friction characteristics in a smooth heat exchanger are first validated in terms of Nusselt number and friction factor. The Nusselt number and friction factor obtained from the present smooth heat exchanger are, respectively, compared with the correlations
Fig. 5. Effects of Reynolds number (Rea), number of perforated hole (N) and pitch ratio (PR) on Nusselt number (Nu).
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Fig. 6. Effects of Reynolds number (Rea), number of perforated hole (N) and pitch ratio (PR) on friction factor ratio (f/fs).
Fig. 7. Effects of Reynolds number (Rea), number of perforated hole (N) and pitch ratio (PR) on Nusselt number ratio (Nu/Nus).
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of Gnielinski for Nusselt number and of Petukhov for friction factor found in the open literature [15] for turbulent flow in pipe. Correlation of Gnielinski,
Nu ¼
ðf =8ÞðRe 1000ÞPr 0:5
2
1 þ 12:7ðf =8Þ ðPr 3 1Þ
;
3000 < Re < 5 106
ð13Þ
Correlation of Petukhov,
f ¼ ð0:79LnðReÞ 1:64Þ2 ;
3000 6 Re 6 5 106
ð14Þ
Fig. 3 shows a comparison of the Nusselt number and friction factor obtained from the present work with those from correlations of Eqs. (13) and (14). In the figure, the present results agree very well within ±12% with the published correlations.
Fig. 4 shows the effects of Reynolds number, number of perforated hole and pitch ratio on friction factor. Obviously, friction factor decreases with rise of Reynolds number. Friction factor increases with decreasing pitch ratio. In the other word, a decrease of the distance between each pair of the turbulators causes an increase in friction factor. This is due to this fact that the smaller distance between each pair of the turbulators, the more numbers of turbulators available in the heat exchanger, thus the more blockage against the flowing stream. Clearly, friction factor decreases with the increase of the number of perforated hole due to the reduction of turbulent fluctuation or eddy motion and the appearance of reverse flow between the each pair of the turbulators. Besides, it can be indicated that friction factors in the heat exchan-
Fig. 8. Effects of Reynolds number (Rea), number of perforated hole (N) and pitch ratio (PR) on thermal performance factor (g).
Table 4 Constant coefficient for using Eq. (16). bij
i=1
i=2
i=3
i=4
i=5
i=6
j=1 j=2
0.326821 0.035122
0.017554544 0.859826928
0.0499 0.01412
0.000396 0.958071
0.00295 0.001889
0.000917 0.03401
Table 5 Constant coefficient for using Eq. (17). cij
i=1
i=2
i=3
i=4
i=5
i=6
j=1 j=2 j=3
7.74E07 0.924403 0.24951
0.000213 0.062311 1.188146
5.75E05 0.06636 0.349578
1.3E08 0.00288 1.29671
0.001121 0.005529 0.31755
3E06 7.7E05 1.334045
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ger equipped with the PCRs of N = 2, 4 and 8 holes, 1.550133– 3.578695, 1.575096–3.43694 and 1.807108–3.703923 times lower than those in the heat exchanger with typical CR, depending on pitch ratio. This indicates that the presence of the perforated hole in the conical ring possesses high potential for reduction of friction in the heat exchanger. Fig. 5 depicts the effects of Reynolds number, number of perforated hole and pitch ratio on Nusselt number. Rate of heat transfer is improved at high Reynolds numbers because the convective heat transfer is promoted more effectively at a higher turbulence level. Nusselt numbers for heat exchanger equipped with turbulators are higher than those found from the plain heat exchanger for a given Reynolds number. This is due to the interruption of flow by the turbulators which results in the destruction of thermal boundary layer near the heat exchanger wall. Nusselt number noticeably increases with decease of pitch ratio which is in similar trend found for friction factor. The quantitative results reveal that the mean heat transfer rate the in the heat exchanger with perforated circular-rings (N = 8) at smallest pitch ratio (PR = 1.83) is 1.037183, 1.140188 and 1.367915 times higher than those in the heat exchanger with PCRs at PR = 2.92, 5.83 and plain heat exchanger, respectively. Furthermore, the TCRs provide higher heat transfer rates than those offered by the PCRs. This is due to the lower turbulence intensity in the heat exchanger. The mean heat transfer rates in the heat exchanger with TCRs for all pitch ratios studied are around 1.064442–1.131967 times of those in the heat exchanger with PCRs with N = 8. The Nusselt numbers in the heat exchanger equipped with the PCRs with PR = 1.83 of N = 2, 4 and 8 holes,
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are respectively 1.454797, 1.450367 and 1.367915 times higher than those in the plain heat exchanger. In contrast, the Nusselt numbers in the heat exchanger equipped with the PCRs of N = 2, 4 and 8 holes are respectively, 1.004903–1.064344, 1.03521– 1.067569 and 1.080317–1.131967 times lower than those in the heat exchanger with typical CR, depending on pitch ratio. Effects of Reynolds number, number of perforated hole and pitch ratio on friction factor ratio are shown in Fig. 6. Friction factor ratio increases with the rise of Reynolds number. This can be attributed to higher flow blockage, larger surface area and the act caused by the reverse flow. It can be observed that friction factor ratio shows a steeper increment at the beginning before gradually increasing for higher values of number of perforated hole. Also it can be found that friction ratio increases with decrease of number of perforated hole and pitch ratio. Effects of Reynolds number, number of perforated hole and pitch ratio on Nusselt number ratio are shown in Fig. 7. Nusselt number ratio decreases with the rise of Reynolds number and number of perforated hole while it decreases with increase of pitch ratio. Fig. 8 shows the effects of Reynolds number, number of perforated hole and pitch ratio on thermal performance factor. It is found that the thermal performance factor decreases with augment of Reynolds number and it is also observed that the turbulators with the smallest pitch ratio (PR) provide the highest thermal performance factor for both of the TCR and PCR turbulators. According to the quantitative results, the maximum thermal performance factors of heat exchanger fitted with PCRs with N = 8 at PR = 1.83, 2.99, and 5.83, are found to be 1.119634, 1.108086 and 1.031247, respectively. The performance factors of heat exchanger with the
Fig. 9. Comparison of experimental data with those calculated from the correlation for (a) friction factor; (b) Nusselt number; (c) thermal performance.
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TCRs are around 1.144314–1.366547 times lower than those of the heat exchanger with PCRs with smallest pitch ratio. This indicates that at the same pumping power the PCRs can enhance heat transfer more efficient than the TCRs and this confirms the greater benefit of the use of PCRs as energy saving device over the TCRs. The corresponding polynomial representations of Nusselt number along air side (Nu), friction factor (f) and thermal performance factor (g) are as follow:
Nu ¼ a12 þ a22 Y 1 þ a32 N þ a42 Y 21 þ a52 N2 þ a62 Y 1 N Y 1 ¼ a11 þ a21 Rea þ a31 PR þ a41 Re2a þ a51 PR2 þ a61 Rea PR f ¼ b12 þ b22 Y 1 þ b32 PR þ b42 Y 21 þ b52 PR2 þ b62 Y 1 PR Y 1 ¼ b11 þ b21 Rea þ b31 N þ b41 Re2a þ b51 N2 þ b61 Rea N
g ¼ c16 þ c26 Y 1 þ c36 Y 2 þ Y 1 ¼ c11 þ Y 2 ¼ c12 þ
c46 Y 21
þ
c56 Y 22
ð15Þ
ð16Þ
þ c66 Y 1 Y 2
c21 Rea þ c31 N þ c41 Re2a þ c51 N2 þ c61 Rea N c22 N þ c32 PR þ c42 N2 þ c52 PR2 þ c62 NPR
ð17Þ
Also aij, bij and cij can be found in Tables 3–5, respectively. Fig. 9 depicts the comparison of the friction factor, Nusselt number and thermal performance factor between experimental data and those calculated from the present correlations. It is found that the majority of the measured data falls within ±10%, ±5% and ±7%, for f ; Nu and g. Contour plots of friction factor, Nusselt number and thermal performance according to current correlations are shown in Fig. 10. The heat transfer rate and friction factor of PCRs increase with decreasing pitch ratio (PR) and decreasing number of perforated hole (N). However, the thermal performance factor increases with increasing number of perforated hole and decreasing pitch ratio. 5. Conclusion Heat transfer and friction factor characteristics in the air to water heat exchanger equipped with the typical circular-ring (TCR) and perforated circular-ring (PCR) turbulators are investigated experimentally. The effects of the pitch ratio and number of perforated hole on flow and heat transfer characteristics are considered. The correlations of the Nusselt number, friction factor and
f
(a) PR = 1.83
(b) N = 8
Nu
(c) PR = 1.83
(d) N = 8
(e) PR = 1.83
( f) N = 8
η
Fig. 10. Contour plots of friction factor, Nusselt number and thermal performance according to current correlations.
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thermal performance are presented. Results show that using PCRs leads to obtain lower heat transfer enhancement than the CRs while they have lower friction factor. Consequently, the thermal performance factor of PCRs is higher than those of the CRs over the range studied. Nusselt number increases with increase of Reynolds number and number of perforated hole and it decreases with increase of pitch ratio. Friction factor enhances with increase of Reynolds number and it decreases with increase of number of perforated hole and pitch ratio. References [1] Y. Vermahmoudi, S.M. Peyghambarzadeh, S.H. Hashemabadi, M. Naraki, Experimental investigation on heat transfer performance of Fe2O3/water nanofluid in an air-finned heat exchanger, Eur. J. Mech. B/Fluids 44 (2014) 32–41. [2] J.M. Buchlin, Convective heat transfer in a channel with perforated ribs, Int. J. Therm. Sci. 41 (2002) 332–340. [3] P. Promvonge, S. Eiamsa-ard, Heat transfer in a circular tube fitted with freespacing snail entry and conical-nozzle turbulators, Int. Commun. Heat Mass Transfer 34 (2007) 838–848. [4] F. Bayrak, H.F. Oztop, A. Hepbasli, Energy and exergy analyses of porous baffles inserted solar air heaters for building applications, Energy Build. 57 (2013) 338–345. [5] M. Sheikholeslami, M. Gorji-Bandpy, D.D. Ganji, Review of heat transfer enhancement methods: Focus on passive methods using swirl flow devices, Renew. Sust. Energ. Rev. 49 (2015) 444–469.
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