Applied Thermal Engineering 65 (2014) 85e93

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Applied Thermal Engineering

journal homepage: www.elsevier.com/locate/apthermeng

Experimental study of convective heat transfer from in-line cam

shaped tube bank in crossflow
Arash Mirabdolah Lavasani a, *, Hamidreza Bayat b, Taher Maarefdoost b
a
Department of Mechanical Engineering, Central Tehran Branch, Islamic Azad University, Tehran, Iran
b
Young Researchers and Elite Club, Central Tehran Branch, Islamic Azad University, Tehran, Iran

h i g h l i g h t s

Heat transfer from cam shaped-tube bank has been studied experimentally.

Tubes were mounted in an in-line arrangement with two longitudinal pitch ratios 1.5 and 2.

Pressure distribution and drag coefficient and heat transfer of each tube in tube bank were measured.

Lower pressure drop compared with circular tube bank.

Higher thermal-hydraulic performance of cam-shaped tubes compared to circular tube.

a r t i c l e i n f o a b s t r a c t

Article history: Heat transfer from cam-shaped tube bank has been studied experimentally. Tubes were mounted in an
Received 13 August 2013 in-line arrangement with two longitudinal pitch ratios 1.5 and 2. Reynolds number based on equivalent
Accepted 30 December 2013 diameter varied in range of 27,000  ReD  42,500. Drag coefficient and heat transfer from tubes in
Available online 8 January 2014
second row were obtained. Results show that drag coefficient and heat transfer of cam shaped tubes
depend on position of tubes in tube bank. Tubes that located at first and second column have the
Keywords:
maximum and minimum value of drag coefficient, respectively. In addition, Heat transfer from subse-
Experimental
quent tubes is greater than tube in first column. Thermal hydraulic performance of cam shape tube is
Cam shaped
Tube bank
about 6 times greater than circular tube with same equivalent diameter.
Heat transfer Ó 2014 Elsevier Ltd. All rights reserved.
Crossflow

1. Introduction The difficulty of predicting flow around multiple cylinders is

increased when they are placed in proximity to each other. Flow
Study of viscous flow past single and multiple bluff bodies has characteristics past each cylinder is affected by its neighbors via
wide engineering applications of which it can be mention the fol- wake interaction which results in alternation of overall flow
lowings: heat exchangers, cooling towers, oil and gas pipelines, pattern. Many researchers have been motivated by this and have
electronic cooling and so on. There are several aspects in under- immersed themselves in studying wake interaction of the two
standing the detail of fluidestructure interaction with these cylinders with equal diameter in crossflow as a basic wake inter-
structures. Such as, avoiding structural failure which could be action model. Deng et al. [7] studied 3-D transition in wake of flow
caused by flow induced vibration under several conditions. passing around two circular cylinders in tandem arrangement.
Decreasing pressure drop and drag coefficient of cylinders and Their spacing ratio L/D varied in range of 1.5e8 and their Reynolds
enhancing convective heat transfer to or from cylinders surface. number was 220e270. Their results show that flow around two
Kays and London [1], Horner [2], Zukauskas and Ziugzda [3], tandem circular cylinders can be treated as 2-D system for L/
Zukauskas and Ulinskas [4], and Zdravkovich [5,6] published books D  3.5, whereas 2-D treatment will be invalid for L/D  4. Mahir
about flow and heat transfer characteristic around bluff bodies. and Altac [8] numerically investigated unsteady laminar convective
heat transfer from two isothermal cylinders in tandem arrange-
ment. Their Reynolds numbers were 100 and 200 and their center-
* Corresponding author. Tel.: þ98 9381299907.
to-center ratios, L/D, varied from 2 to 10. They found that mean
E-mail addresses: arashlavasani@iauctb.ac.ir, arashlavasani@yahoo.com Nusselt number of the upstream cylinder approaches to that of
(A. Mirabdolah Lavasani). single cylinder for L/D  4. Sumner [9] summarized the literature

1359-4311/$ e see front matter Ó 2014 Elsevier Ltd. All rights reserved.
http://dx.doi.org/10.1016/j.applthermaleng.2013.12.078
86 A. Mirabdolah Lavasani et al. / Applied Thermal Engineering 65 (2014) 85e93

on the flow around two circular cylinders with equal diameter finned-tube heat exchanger. Their results indicate that iso-
immersed in a steady cross-flow. sectional tube increases thermal-hydraulic performance of heat
Rocha et al. [10] studied numerically elliptical and circular sec- exchanger compared to classical finned-tube heat exchangers. Tan
tion in one and two row tubes and plate fin heat exchanger. They et al. [23] studied experimentally heat transfer and pressure drop
reported that elliptic tubes and plate fin heat exchangers have performance of twisted oval tube heat exchanger. Their results
considerably better overall performance than circular one due to show that heat transfer coefficient of the twisted tube is higher
lower pressure drop and higher fin efficiency of elliptic tubes and than smooth round tube at the cost of some incensement of pres-
plate fin. sure drop.
Matos et al. [11] studied numerically forced convection of Lee et al. [24] studied numerically the effects of uneven longi-
staggered circular and elliptic tubes. Their Reynolds number varied tudinal pitch on heat transfer from in-line tube bank. It can be
in range of 300  ReL  800. They found that elliptic configuration found from their results that increasing the longitudinal space for
performs more efficiently than circular one. Bouris et al. [12] uniformly distributed cylinders enhanced overall heat transfer.
studied numerically alternate tube configurations for particle Bilirgen et al. [25] numerically study effects of fin spacing, fin
deposition rate reduction in tube bundle. Their results show that thickness, and fin material on overall heat transfer of annular fin-
elliptic-shaped tubes had the lowest fouling rates and pressure ned tubes. Their results show that increasing fin thickness leads to
drop compared with circular tubes. Matos et al. [13] numerically minor increases in heat transfer and modest increases in pressure
and experimentally optimized heat transfer from heat exchangers drop. Increasing thermal conductivity of material is also leads to
with finned circular and elliptic tubes. Their results show that by higher heat transfer. Sun and Zhang [26] used CFD model to
using optimal elliptic arrangement, heat transfer gains of up to 19% investigate simultaneously fluid flow and heat transfer on both air
relative to the optimal circular tube arrangement. side and water side of elliptical tube in finned-tube heat ex-
Ibrahim and Gomaa [14] studied experimentally and numeri- changers. Their results show that impact of axis ratio on overall
cally thermo-fluid characteristics of elliptical tube bank in cross- thermal-hydraulic performance is dependent on air velocity and
flow. Their Reynolds number varied in range of 5600 < Re < 40,000 water volumetric flow rate. At lower volumetric flow rate
and angle of attack varied from 0 to 150 . They found that elliptic increasing axis ratio is advantageous to overall thermal-hydraulic
tube bank at zero angle of attack has maximum thermal perfor- performance, while it has adverse impact on higher volumetric
mance, and using elliptic tube arrangement in heat exchangers flow rate.
could lead to energy conservation. Streamlined shaped tubes due to low hydraulic resistance need
Lam et al. [15] studied effects of wavy cylindrical tubes in a less pumping power. So, the purpose of this study is to experi-
staggered tube bundle. They used experimental measurement and mentally investigate the flow and heat transfer characteristics
large eddy simulation technique. Their Reynolds number varied in around cam shaped tube bank subject to crossflow of air.
range of 6800 < Re < 13,400. Their results showed that by using
wavy tube drag coefficient was reduced and the fluctuating lift was 2. Experimental setup
suppressed.
Nouri-Borujerdi and Lavasani [16] experimentally studied Fig. 1 shows the cross section profile of the cam shaped tube that
convective heat transfer from a cam shaped tube in crossflow. Their comprised some parts of two circles with two arcs segments
Reynolds number and angle of attack varied in ranges of tangent to them. The cylinder have identical diameters are equal to
0 <a < 180 and 15,000 < Re < 27,000, respectively. Their results d ¼ 8 mm and D ¼ 16 mm where distance between their centers is
show that thermal performance of cam shaped tube is maximum at l ¼ 15.75 mm. Tubes are made of a commercial steel plate with
a ¼ 0 and cam shaped tube had larger thermal hydraulic perfor- 0.7 mm of wall thickness.
mance compared to circular tube. For measuring drag coefficient of cam shaped tube in tube bank
Tang et al. [17] experimentally and numerically studied air-side a test tube with length of 31 cm was made. Fourteen holes (1 mm
heat transfer and friction factor of five type of fin-and-tube heat diameter) were drilled on the surface of test tube to measure the
exchanger in Reynolds number in range of 4000e10,000. Their static pressure on the tube surface by a digital differential pressure
results show that by using vortex generator with higher angle of meter. For measuring heat transfer, four test tubes with length of
attack and smaller height, overall performance of heat exchanger 22 cm were made. The two ends of test tubes were insulated for
will be better. decreasing heat transfer from these surfaces. In order to neutralize
Næss [18] studied heat transfer and pressure drop of ten finned the effects that sudden deformation of tube could cause on the
tube bundle using serrated fins experimentally. He found that passing flow, the insulation was formed like the cam shaped tube.
increasing the fin pitch reduced the heat transfer coefficient and By doing this, the two end’s surfaces of each tube will be insulated
increasing fin height resulted in the increase of the heat transfer without any kind of effects that sudden deformation could cause on
coefficient. Ehwany et al. [19] worked on an elbow-bend heat passing flow through tube bank.
exchanger as a heater and cooler in an alpha-type Stirling engine. Fig. 2 shows sixteen cam-shaped tubes located at wind tunnel
Their results showed that elbow-bend heat exchanger reduces the test section. The distance between the upper and lower tubes to the
hydraulic losses. Moawed [20] studied experimentally forced
convention from outside surfaces of helical coiled tube. He studied
ten helical coiled-tubes with various design parameters and Rey-
nolds number in range of 6.6  102e2.3  103. He found that with
small values of pitch ratio, higher average Nusselt number can be
achieved.
Kukulka et al. [21] investigated the overall thermal performance
of enhanced heat transfer tubes. Their results show that their tube
minimizes the fouling rate and provides heat transfer performance
gains in excess of 100% compared to smooth tubes. Simo Tala et al.
[22] performed unsteady-RANS simulation to investigated effects
of tube pattern on thermal-hydraulic characteristics in a two-row Fig. 1. Schematic of a cam shape tube.
 A. Mirabdolah Lavasani et al. / Applied Thermal Engineering 65 (2014) 85e93 87

direction. As it is shown in Fig. 4 the flow is approximately uniform

in Z and Y direction of test section.
To heat up the tubes, a pump circulates hot water between a
tank and the tubes. An electric heating element supplies the hot
water and a control valve regulates the hot water at the tube inlet.
Water temperature is measured at the inlet and outlet of the tubes
using type-k thermocouple wires and saved at interval times of 1 s
by using data logger. A glass tube flow meter measures the flow rate
with 1% uncertainty in full-scale flow. A steady state condition is
reached between 5 and 15 min, depending on the ambient tem-
perature and free stream velocity, and then data collection is
started.

Fig. 2. Cam shaped tube bank with in-line arrangement.

3. Experimental technique

To estimate the pressure drag and heat transfer from the cam
upper and lower wall of test section is 9 cm. The space between two shaped tubes compared to that of a circular tube with various cross
tandem tubes is defined by longitudinal pitch SL and the space sections, it is important to select an appropriate reference length.
between side-by-side tubes is defined by transverse pitch ST. Deq is the diameter of an equivalent circular tube whose circum-
However, in this study transverse pitch ratio is ST/Deq ¼ 1.25 and ferential length is equal to that of the cam-shaped tube. Based on
longitudinal pitch ratios are SL/Deq ¼ 1.5 and 2. Fig. 1, the equivalent diameter is obtained by Deq ¼ C/p ¼ 22.44 mm
Fig. 3 shows an open circuit low speed wind tunnel where the where C is perimeter of cam shape tube.
experiments were performed. A pitot static tube is used to measure For understanding flow characteristic better, Reynolds number
the free stream velocity in front of the frame cross section. The air is defined with two equations. First, for comparing heat transfer
velocity varied from 9 to 15 m/s by controlling a variable speed from each tube in tube bank with single tube in crossflow, Reynolds
motor. In order to be assured that flow is uniform inside the test number is calculated by Reeq ¼ UN$Deq/n. Second, since the speed of
section; velocity of flow was measured in different X and Y fluid varies along its path in tube bank, a reference velocity base on

Fig. 3. Experimental apparatus.

88 A. Mirabdolah Lavasani et al. / Applied Thermal Engineering 65 (2014) 85e93

16
DP
f ¼ (2)
0:5NL rUmax
2

The rate of heat transfer from the tube to the air is obtained by
measuring mass flow rate, inlet and outlet water temperatures
through the tube as

Q_ w ¼ m
_ w cp;w ðTwi  Two Þ (3)

Where m_ w ¼ rw V_ w which cp,w, rw and V_ w are specific heat, density

and volume flow rate of water respectively.
The mean Nusselt number is determined as follows:
-0.0375 -0.025 -0.0125 0

hDeq Q_ w
Nueq ¼ ¼ (4)
k pLkðTs  TN Þ

Where temperature of tube surface is define by TS ¼ (Twi þ Two)/2.

After measuring Nueq for all tubes in tube bank, the average
Nusselt number of tube bank was calculated by the following
equation.

1
Nuave: ¼ Nueq (5)
N

Where N is number of tubes in a row of tube bank.

Symb. Heat transfer performance against the friction factor of cam
U (m/s) 9.29 10.44 11.52 12.48 13.41 14.27 14.84 defines by Nuave: =f .
Thermal hydraulic performance shaped tube bank base on cir-
-0.075 -0.05 -0.025 0 0.025 0.05 0.075 0.1 cular tube bank is defined by efficiency index h which has been
Z (m) proposed by Webb [29].

Fig. 4. Velocity distribution in front of test section of wind tunnel; (a) Y direction (b) Z Nuave:cam: =Nuave:cir:
direction. h¼ (6)
fcam: =fcir:
minimum free area available for fluid flow is being used for
Yan and Sheen [30] suggested a factor “AGF” or “area goodness
calculating of ReD ¼ Umax.Deq/n [27], which in the present work
factor” for comparing heat exchanger base on their frontal area and
varies in range of 27,000  ReD  42,500. Reynolds number base on
desired duty. Heat exchanger which has higher AGF is better
minimum free area was used for showing results of tube bank.
because it requires less frontal area.
Quarmby and Al-Fakhri [28]showed that for L/Deq > 4 this ratio
has little effect on the heat transfer as a result the test tube for
measuring heat transfer was made with L/Deq ¼ 8. AGF ¼ s2 j=f (7)
The thermophysical properties of air for both equations was
calculated from film temperature which is the average of surface In Equation (7), s is ratio of free flow area to frontal area of tube
and free stream temperature, Tf ¼ (Ts þ TN)/2. bank and j is Colburn factor.
The pressure drag coefficient CD is determined experimentally
from pressure distribution over the cam shaped-tube surface,
2
including the large and small circles as well as two tangent arcs
Circular Cylinder [2]
between them as follows: 1.75
Elliptical Cylinder [2]
( ) Cam-Shaped Tube (Present Work)
1 X
14 1.5
CD ¼ Cp;i cos qi DSi (1) Circular Tube( Present Work)
Deq i¼1
1.25
CD

1
The pressure distribution on the cam shaped is expressed in
dimensionless form by the pressure coefficient 0.75
CP;i ¼ ðpi  pN Þ=ð0:5rUN 2 Þ.
0.5
According to Fig. 1, Pi is the static pressure which was measured
by a differential pressure meter at the location of the holes drilled 0.25
perpendicular to the tube surface. PN and UN are the pressure and
velocity of the air free stream respectively and r is air density.
0
13000 15000 17000 19000 21000 23000 25000 27000
Friction factor f is determined by calculating pressure drop DP
Re
across tube bank. Where DP is the difference between the pressure
at inlet and the exit of the cam shaped tube bank and NL is number Fig. 5. Comparison of drag coefficient of a cam-shaped tube with elliptical and circular
of transverse rows [27]. cylinder.
 A. Mirabdolah Lavasani et al. / Applied Thermal Engineering 65 (2014) 85e93 89

4. Uncertainty analysis 1.2

Symb.

Wind tunnel measurements are subject to different sources of 0.8 Column 1st 1st 2nd 2nd 3rd 3rd 4th 4th
S /D 1.5 2 1.5 2 1.5 2 1.5 2
uncertainty, from instrumentation, data acquisition and data
analysis. A final result, R, is typically the combination of different 0.4
measured variables, Vi [31]. The contribution of the uncertainty in
each variable to the result can be calculated by: 0

Cp
" #12
n 
X
UR vR -0.4
¼ UVi (8)
R i¼1
vVi
-0.8
Re =21000
Uncertainty of Nusselt number was calculated by using 2nd Row

following equation: -1.2

0 30 60 90 120 150 180 210 240 270 300 330 360
( 2 " #2 θ (degree)
UNu UQw Qw
eq
¼ þ UTN
Nueq pLkðTs  TN Þ pLkðTs  TN Þ2 Fig. 7. Pressure distribution around cam shaped tube in second row.
" #2  2
Qw Qw present setup can be used for measuring pressure drag and heat
þ U T þ U
pLkðTs  TN Þ2
s
pDLk2 ðTs  TN Þ k
transfer from cam shaped tubes.
 )
2 21

Qw 5.1. Pressure distribution around cam shaped tubes in in-line tube
þ U L (9)
pL2 kðTs  TN Þ bank

The uncertainty of drag coefficient was varied from 2 to 10 Pressure coefficients of cam shaped tube in first, second, third
percent and tubes in second column had the highest uncertainty for and fourth column for second row are represented in Fig. 7 for SL/
all Reynolds number and uncertainty of pressure drop and Nusselt Deq ¼ 1.5 and 2 and Reeq ¼ 21,000. By comparing coefficients from
number varied in range of 2e5 percent and 3e7 percent, Fig. 7 for SL/Deq ¼ 1.5 with state of single cam shaped in crossflow it
respectively. can be seen that pressure coefficient at q ¼ 0 for the first tube is
approximately equal to state of single tube in crossflow. However,
this coefficient at q ¼ 0 for second, third, and fourth column is
5. Result and discussion
about 141, 140, and 147 percent lower than single tube in crossflow,
respectively. At q ¼ 180 pressure coefficient of tubes in first, sec-
For purpose of verifying data-taking process and checking
ond, and third column is about 39, 36, and 13 percent higher and for
experimental setup, drag coefficient and heat transfer from single
fourth column is about 19 percent lower than state of single tube. It
circular tube with diameter of 2.7 cm and 2.2 cm was measured,
is also clear that pressure coefficient for tubes in second, third and
respectively. All experiments were carried out with nominal tube
fourth column reaches their maximum value at about q ¼ 282 ,
surface temperature of about 79  Ce83  C and air temperature
288 and288 , respectively.
about 25  C.
For SL/Deq ¼ 2 and at q ¼ 0 , the pressure coefficient is
Fig. 5 compares drag coefficient of present study with Hoerner
approximately equal to single tube. However, this coefficient for
[2]. The difference between present work and Hoerner [2] is about 8
tubes in second, third and fourth column is about 141, 120 and 146
percent. It is also clear that drag coefficient of cam shaped tube is
percent lower compare to single tube in crossflow, respectively. At
about 64 and 25 percent lower drag coefficient of circular and
q ¼ 180 pressure coefficient of tubes in first, second, and third
elliptical tube, respectively.
column is about 32, 40, and 34 percent higher and for fourth col-
Fig. 6 compares results of heat transfer from present work with
umn is about 6 percent lower than state of single tube. It is also
Zhukauskas [3]. The difference between present work and Zhu-
clear that pressure coefficient for tubes in second, third and fourth
kauskas is about 1e7 percent. Therefore it can be concluded that
column reaches their maximum value at about q ¼ 282 , 296 and
296 , respectively.

78
Zukauskas [3]
5.2. Drag coefficient of cam shaped tubes in an in-line tube bank
74
Present Work
Drag coefficients of cam shaped tubes in an inline tube bundle
70 arrangement with two pitch ratio SL/Deq ¼ 1.5 and 2 and Reynolds
Number in ranges of 13,000  Reeq  21,000 are shown in Fig. 8.
66
Results show that for both pitch ratios, the first column tube has the
Nu

62 maximum and the second column tube has the minimum drag
coefficient.
58 All the values of drag coefficient for longitudinal pitch ratio SL/
Deq ¼ 1.5 at Reynolds numbers from 13,000 to 21,000 are approx-
54
imately 0.52, 0.23, 0.37 and 0.43, for the first, second, third and
50 fourth column tube, respectively. This coefficient for the first col-
11500 12500 13500 14500 15500 16500 17500 18500 umn tube is 15 percent greater than single cam shape tube in
Re crossflow but for tubes in second, third and fourth column is
respectively 50, 18 and 4 percent smaller within the same Reynolds
Fig. 6. Heat transfer from circular tube in crossflow. number range.
90 A. Mirabdolah Lavasani et al. / Applied Thermal Engineering 65 (2014) 85e93

62

60

58

56

Nueq
54

52

50
Column 1st 1st 2nd 2nd 3rd 3rd 4th 4th
48
1.5 2 1.5 2 1.5 2 1.5 2
2nd Row
11500 12500 13500 14500 15500 16500 17500 18500

Reeq

Fig. 10. Heat transfer from single cam shaped tube in crossflow.

Fig. 8. Drag coefficient of cam shaped tubes in an in-line tube bank.

5.5. Heat transfer from cam shaped tubes in tube bank

For longitudinal pitch ratio SL/Deq ¼ 2 all the values of drag
coefficient for Reynolds number in range of 13,000  Reeq  21,000
Nusselt number of cam shaped tubes in first, second, third and
are approximately 0.54, 0.21, 0.28 and 0.43 for the first, second,
fourth column of tube bank with in-line arrangement and Reynolds
third and fourth column tube, respectively. This coefficient for the
number in range of 11,500e18,500 for two longitudinal pitch ratios
first column tube is 19 percent greater than single cam shape tube
1.5 and 2 is presented in Fig. 11. For both pitch ratios heat transfer
in crossflow but for tubes in second, third and fourth column is
from tubes in first column is approximately equal to single cam
respectively 53, 39 and 6 percent smaller within the same Reynolds
shaped tube in crossflow. However, by moving in flow direction in
number range.
tube bank, heat transfer from inner tubes increase due to locating in
turbulent wake of upstream tube.
5.3. Pressure drop of cam shape tubes in an in-line tube bank For SL/Deq ¼ 1.5 by increasing Reynolds number from 11,500 to
18,500 Nusselt number for tube in first, second, third and fourth
A comparison of pressure drop across cam shaped tube bank
and circular tube bank with four columns and four rows is shown in
80
Fig. 9. Results show that by increasing pitch ratio from 1.5 to 2 for Symb. Column

cam shaped tubes bank friction factor increases about 7 percent. 1st
2nd
Friction factor of cam shaped tube bank is about 95 and 93 percent 3rd
70
lower than circular tube for SL/Deq ¼ 1.5 and 2, respectively, which 4th

can be attributed to lower drag coefficient of cam shaped tubes

compare to circular tube bank.
Nueq

60

5.4. Heat transfer from______

single cam shaped tube in crossflow

50
Fig. 10 shows heat transfer from single cam shaped tube placed
in crossflow of air. Result shows that as the equivalent Reynolds
S / D = 1.5
number increases from 11,500 to 18,500, heat transfer from cam
shaped tube increases about 33 percent. By comparing heat transfer 40
11500 12500 13500 14500 15500 16500 17500 18500
from cam shaped tube with circular tube, it can be concluded that
heat transfer from cam shaped tube is about 5e11 percent lower Reeq
=1.5
than circular tube with equivalent diameter.
90
Symb. Column
1 1st
2nd
80 3rd
0.8 4th

70
Nueq

0.6
Symb. S /D Tube Shape
f

60
2 Circualr [27]
0.4
1.5 Circular [27]
2 Cam Shaped
1.5 Cam Shaped 50
0.2
S /D =2

40
0 11500 12500 13500 14500 15500 16500 17500 18500
27000 29000 31000 33000 35000 37000 39000 41000 43000 Reeq
Re D (b) SL/Deq=2

Fig. 9. Friction factor of cam shape tubes in an in-line tube bank. Fig. 11. Heat transfer from cam shaped tubes (a) SL/Deq ¼ 1.5 (b) SL/Deq ¼ 2.
 A. Mirabdolah Lavasani et al. / Applied Thermal Engineering 65 (2014) 85e93 91

80

70
Nuave.

60

50
1.5

40
27000 29000 31000 33000 35000 37000 39000 41000 43000
Re D

Fig. 12. Heat transfer from cam shaped tube bank. Fig. 14. Thermal-hydraulic performance of cam shaped tube bank.

column increases about 26, 40, 39 and 29 percent, respectively. 5.7. Thermal-hydraulic performance of cam shaped tube bank
Nusselt number of tube in second, third and fourth column is about
4, 12 and 22 percent greater than single cam shaped tube in Heat transfer performance against the friction factor of cam
crossflow, respectively. shaped tube bank is presented in Fig. 13. Results show that per-
Also, for SL/Deq ¼ 2 by increasing Reynolds number from 11,500 formance of cam shaped tube with SL/Deq ¼ 1.5 is about 8 percent
to 18,500 Nusselt number for tube in first, second, third and fourth lower than SL/Deq ¼ 2. Due to aerodynamic shape of cam shaped
column increases about 45, 36, 57 and 35 percent, respectively. By tubes its pressure drop is much lower than circular tube bank.
comparing Nusselt number of single tube in crossflow with tubes Therefore, heat transfer performance against the friction factor of
that located in tube bank, it can be concluded that Nusselt number cam shaped tube is about five times greater than circular tube with
of tube in second, third and fourth column of tube bank is about 14, equivalent diameter.
26 and 41 percent greater than single tube, respectively. Fig. 14 displays thermal hydraulic performance factor of cam
As it can be seen Nusselt number of tubes in second and sub- shaped tube compare to circular tube. It is clear that this coefficient
sequent column is higher than tube in first column and even a is about 6 times greater than circular tube. This means that cam
single tube in crossflow. High turbulence in wakes of upstream shaped tube perform better compare to circular tube. In addition, in
tubes leads to thinner boundary layer of tubes in second and sub- Table 1 results of this work are compared with other works on
sequent rows and therefore heat transfer coefficient of tubes in literature. By comparing efficiency index of cam shaped tube bank
subsequent column is higher than tubes in first column [27]. with tube with other shape, it is clear that cam shaped tube per-
forms better.
5.6. Heat transfer from cam shaped tube bank A comparison of area goodness factor of cam shaped tube bank
with circular tube bank is presented in Fig. 15. Results indicate that
Fig. 12 shows mean Nusselt number of tube bank with in-line this factor for cam shaped tube bank is about 13e14 times greater
arrangement and Reynolds number in range of 27,000e42,500 for than circular tube bank for both longitudinal pitch ratios. The
two longitudinal pitch ratios 1.5 and 2. Results shows that by reason for higher area goodness factor of cam shaped tube bank is
increasing Reynolds number from 27,000 to 42,500 heat transfer the streamline shape of this tube. Higher area goodness factor of
from cam shaped tube bank increase about 37 and 43 percent for SL/ cam shaped tube means that these tube banks have lower frontal
Deq ¼ 1.5 and 2, respectively. Furthermore, by increasing longitu- area which in general could help to build more compact heat
dinal pitch ratio from 1.5 to 2, heat transfer of cam shaped tube exchangers.
bank increases about 7e14.
6. Conclusion
2000

In this study flow around cam shaped tube bank with inline
1600 arrangement has been investigated experimentally. Reynolds
number based on equivalent diameter varied in range of
1200
Nuave./f

Symb. SL/Deq Shape Table 1

Compaction of present study with other work on literature.
800 2 Cam
1.5 Cam
Ref. Tube shape 103 ReD Remarks
1.5 Circle [27]

400
2 Circle [27] [13] Elliptic 0.852 and q_ ellip =q_ cir z1:19 in
tube 1.065 optimal arrangement
bundle
0 [14] Elliptical 5.6e40 h z 4.5e6 for Ar ¼ 0.5
27000 29000 31000 33000 35000 37000 39000 41000 43000
tubes
ReD [23] Twisted Flow rate: h z 1.5e4.6
oval tubes 1.5e16 (kg/s)
Present Cam-shaped 27e42.5 h z 6.8e7.1
Fig. 13. Variation of heat transfer performance against the friction factor with Rey-
work tubes
nolds number.
92 A. Mirabdolah Lavasani et al. / Applied Thermal Engineering 65 (2014) 85e93

h thermal-hydraulic performance
s Afree flow area/Afrontal area
q hole angle (degree)

Subscripts
cam cam-shaped tube
AGF

Symb. SL/D Tube Shape

eq equivalent
1.5 Cam i inlet
2 Cam
1.5 Circular [27] o outlet
2 Circular [27] s surface
w water
ave. average
N free stream

D
References
Fig. 15. Comparison of area goodness factor with Reynolds number.
[1] W.M. Kays, A.L. London, Compact Heat Exchangers, McGraw Hill, New York,
27,000  ReD  42,500. Results show that drag coefficient is 1964.
[2] S.F. Hoerner, Fluid Dynamic Drag, Theoretical, Experimental and Statistical
maximum for the first column and minimum for the second col-
Information, AIAA, NY, USA, 1965.
umn for all ranges of Reynolds number. Furthermore, friction factor [3] A. Zukauskas, J. Ziugzda, Heat Transfer of a Cylinder in Crossflow, Hemisphere
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