Heat Transfer Analysis of Integral-Fin Tubes

Engineering and Technology
2015; 2(2): 23-34
Published online March 30, 2015 (http://www.aascit.org/journal/et)
Heat Transfer Analysis of
Integral-Fin Tubes
Laith Jaafer Habeeb1, Abdulhassan A. Karamallah1,
Ayad Mezher Rahmah2
1
2
Mech. Eng. Dept., University of Technology, Baghdad-Iraq
State Company for Oil Projects (S.C.O.P), Ministry of Oil, Baghdad-Iraq
Email address
[email protected] (L. J. Habeeb), [email protected] (L. J. Habeeb)
Citation
Keywords
Heat Transfer,
Integral - Fin Tube,
Experimental Study
Received: March 6, 2015
Revised: March 21, 2015
Accepted: March 22, 2015
Laith Jaafer Habeeb, Abdulhassan A. Karamallah, Ayad Mezher Rahmah. Heat Transfer Analysis
of Integral-Fin Tubes. Engineering and Technology. Vol. 2, No. 2, 2015, pp. 23-34.
Abstract
An experimental system has been adapted to study the heat transfer characteristics for
cross flow air cooled single aluminum tube multi passes (smooth and integral low finned
tube) and the effect of the integral low fins in enhancement the heat transfer. Also, study
all variables which have effect on heat transfer phenomena. A series of experiments was
conducted with different variables. The velocities of air across the test section are (1, 2
and 3) m/sec, the water flow rate is (5l/min) and the temperatures of the inlet water to the
test tube are (50, 60, 70, 80) oC. In this study, the integral low finned tube gave a good
enhancement in heat transfer. Hence, the experimental results showed that the air side
heat transfer coefficient of the integral low finned tube was higher than that of the
smooth tube and the enhancement ratio (( ho finned/ ho smooth ) or ( Nua finned/ Nua smooth )) was
(1.86 to 2.38) for eight passes. Also, the results showed that the increasing of air velocity
will improve the outside heat transfer coefficient. In addition to the theoretical analysis,
this work presents a suggestion to develop empirical correlations for the air side heat
transfer coefficient of an integral low finned tube, represented by the empirical
correlations for the air side Nusselt number. The results were compared with previous
works of other researchers and gave a good agreement in behavior.
1. Introduction
One of the most common methods of enhanced heat transfer is by using integral low
fin tubes and the fins usually have a two-dimensional trapezoidal or rectangular cross
section [1]. Integral finned tubes are made by extruding the fins from the tube metal. The
tube is generally made from (copper, aluminum and its alloys) that are relatively soft and
easily worked and also made of other materials (stainless steel, titanium, and its alloys,
etc.) [2, 3]. Since the fins are integral with the root tube, perfect thermal contact is
ensured under any operating conditions [2]. Integral fin tubes are commonly used in the
condensers of refrigeration, air conditioning and process industries especially where low
surface tension fluids are used [4]. It is also used in the heat exchanger, evaporator and
boiling services [5]. Fins are available in different densities ranging from 433–1675
fins/meter (11–40 fins/inch) [3]. In the last few decades, several three dimensional (3D)
enhanced surfaces were developed for condensation heat exchangers. Also, several
improvements were introduced to the standard integral finned tubes which resulted in a
performance comparable to that of the 3D enhanced surfaces [4]. Low fin heights are
ranging from about (0.66 to 1.50) mm depend on the fins density and the particular tube
metal [6].
Rich [7] performed an experimental work to determine the effect of fin spacing on
heat transfer and friction performance of multi-row fin-and-tube heat exchangers.
24
Laith Jaafer Habeeb et al.: Heat Transfer Analysis of Integral-Fin Tubes
Later, Rich [8] focused on the effect of the number of tube
rows on heat transfer performance of heat exchangers, which
was a continuation of his previous experimental work.
Brown [9] presented preliminary estimates for the thermal
design for heat exchangers. He established a procedure in a
tabulated form for the design of heat exchanger with multi
rows of circular finned tubes.
Wang et al. [10] performed a comparison study of eight
finned-tube heat exchangers. They concluded that the effect
of fin pitch on heat transfer performance is negligible for
four-row coils having ReDc> 1000, and that for ReDc< 1000,
the heat transfer performance is highly dependent on fin pitch.
Haliciand Taymaz[11] investigated experimentally the
effect of tube regulation space on the heat and mass transfer
and friction factor for heat exchangers made from aluminum
fins and copper tubes.
Chen and Hsu [12] studied theoretically and
experimentally the average heat transfer coefficient and fin
efficiency on a vertical annular circular fin of finned-tube
heat exchangers for various fin spacing in forced convection.
Choi et al. [13] investigated experimentally the heat
transfer characteristics of discrete plate finned-tube heat
exchangers with large fin pitches.
Honda et al. [14] investigated the theoretical model of film
condensation on a single horizontal low finned tube is
extended to include the effect of condensate inundation.
Cheng et al. [15] studied experimentally the condensation
heat transfer characteristics of horizontal enhanced tubes.
Kumar et al. [16] studied the heat transfer augmentation
during condensation of water and R-134a vapor on horizontal
integral-fin tubes. In This experimental investigation was
performed on two different experimental set-ups for water
and R-134a.
Tarrad [17] presented a computerized model for the
thermal-hydraulic design of a single shell – single pass low
finned tube bundle heat exchange using the step by step
technique (SST).
Fernández-Seara et al. [18] investigated experimentally the
condensation of ammonia on smooth and integral-fin (32 fins
per inch (fpi)) titanium tubes of 19.05mm outer diameter.
In this investigation, the effect of an integral low finned
tube in cross flow air cooled in a horizontal single tube multi
passes on the heat transfer behavior will be analyzed
experimentally and theoretically. Also, the effect of changing
air velocity and inlet water temperature are investigated. This
work presents a suggestion to develop empirical correlations
for the air side heat transfer coefficient of an integral low
finned tube, represented by the empirical correlations for the
air side Nusselt number.
2. Experimental Work
2.1. The Test Rig
Figures (1- a, b) show a photo and schematic diagram of
the experimental test rig. The test rig is designed and
manufactured to fulfil the requirements of the test system for
a smooth and integral low finned tube. The experimental
apparatus consist basically of:
•
The duct and test section.
•
The airflow rates supply section.
•
The water flow rates supply section.
•
The measuring devices.
(a)
Engineering and Technology 2015; 2(2): 23-34
25
(b)
Figure (1). Experimental test rig: (a) Photo, (b) Schematic diagram.
2.2. Air Circulation System
The air was supplied to the test section by centrifugal
blower of (370 W). It was supplied air at three levels of
velocity (1, 2, 3) m/sec at the test section, controlled by using
multi configurations of circular cross-section gate
manufactured for this purpose. The gate controls air mass
flow rates and air velocities at the test section. The required
velocities were obtained by replacing the configuration of the
gate between the fully opened without any gate (maximum
flow rate) and 45° partially opened (minimum flow rate). The
blower outlet is connected directly to a galvanized steel air
diffuser by bolts after inserting the rubber seal and silicon,
and the other side of diffuser is connected with the two layers
of the mesh at the face of the diffuser between the main duct
and diffuser. The mesh is designed and manufactured to
ensure damping of any disturbance in air stream before
entering the test section and to obtain a regular flow.
The air blower is fixed to the iron foundation by bolts with
thick rubber between the blower and foundation for damping
the vibration when the blower operates.
The duct is manufactured from a galvanized steel sheet at
rectangular cross section with width and height (251
mm×477mm) and length 2m with the test section part. The
duct is connected with the blower by a diffuser and the other
side ended with another diffuser opened to the atmosphere
after insert the rubber seal and silicon at the edges. The
suitable test duct length is 370 mm fixed at 2000 mm from
the beginning of inlet diffuser, the test tube passed through
the duct horizontally at 2185mm from the beginning of inlet
diffuser, as shown figure (2).
Figure (2). Schematic illustration of duct.
26
Laith Jaafer Habeeb et al.: Heat Transfer Analysis of Integral-Fin Tubes
Figure (3). (L) Photo of one of the test models, (R) Section of integral low finned tube.
2.3. Test Section
Two test sections were designed and manufactured in the
present work, each one consists of rectangular test duct (251
×477 ×350) mm width, height and length respectively, and
constructed from Perspex of (10 mm thickness) as shown in
figure(3-a). Each one has an aluminum test tube multi passes,
passing horizontally through the test duct and the distance
between center to center of passes is 55mm. the first test
section has a smooth aluminum tube of eight passes with
inner diameter 17mm and outer diameter 19mm.The second
test section has an integral low finned aluminum tube of eight
passes with inner diameter 17mm, root diameter 19mm and
outer diameter at the tip of fin 22 mm. Each pass has a length
251mm inside the duct with 125 fins, which is approximately
(500 fins per meter).The fin’s height is 1.5 mm with a
thickness of 1mm and pitch 1mm as shown in figure (3b).The finned tube was manufactured by the lathe machine.
The test duct was connected to the main duct by aluminum
flanges and bolts and manufactured in a way for easy
replacement of the test section and inserting the rubber seal
and silicon at the connections. The test pipe was connected to
the water cycle. All the pipe bends outside the test duct were
fully insulated by a thermal rubber and insulating tape.
2.4. Water Feeding System
A liquefied petroleum gas (LPG) water heater was used to
supply hot water quickly and continuously to the test section.
The water outlet temperature can be controlled by a flame
adjustment knob and a water input adjusting knob.
The other accessories used to complete the system are:
Water pump of (370 W) with a maximum volumetric flow
rate (30 l/min), insulating tank of (30 L) capacity
manufactured from galvanized steel sheet and insulated by
(glass wool ), insulating pipes of 12.7mm (1/2 inch) diameter
manufactured from galvanized steel with valves and
connections insulated by (thermal rubber ), and iron structure
foundation to support all rig parts.
2.5. The Measured Parameters
During
the
experimental
investigation,
the
main
parameters measured are:
1) The inlet and outlet temperature of water at the test tube.
2) The inlet and outlet pressure (pressure difference between
inlet and outlet of the test tube (3). The surface temperature
for the test tube. 4) The water volumetric flow rate. 5) The
temperature of air entering and leaving the test section. 6)
The atmosphere temperature. 7) The average air velocity.
Digital anemometer and flow meter were used to measure
air velocities and water flow rates respectively, and pressure
gauges were used to measure pressure drop in the water side.
Multi thermocouples and temperature probes were used to
obtain the temperatures in inlet and outlet the test section at
water and air side respectively. The thermal imager technique
(I.R. - fusion camera) was used to measure the surface
temperatures for the test tube. All of these measuring devices
were used after the calibrating.
2.6. Tests Procedure
The following procedure steps were conducted for each
experimental session after completing checking for the water
cycles and air system:
1. Switch on the circuit breaker to supply power to the
whole system when all valves of the water cycle are
opened.
2. Switch on the water heater by supply the liquefied
petroleum gas (LPG) to the heater.
3. Adjust the air velocity, regulated by using the gate at
one of the required three levels of air velocity.
4. Adjust the water flow rate in water cycle by the
control valves of the water flow through main and
bypass pipes before the test tube, or adjust by
controlling the input water flow rate adjusting knob
in the water heater at (5 l/min).
5. Adjust the required outlet temperature from the water
heater at inlet of the test section manually by
adjusting the knob of the flame or the knob of water
flow rate input to the heater.
6. Watch the reading of water inlet and outlet
temperatures till the steady state conditions reached
(40-60) minutes. Then, take the following readings:
7. Water temperatures for inlet and outlet of the test
tube. b) Air temperatures for entering and leaving the
Engineering and Technology 2015; 2(2): 23-34
8.
test duct before and after the test tube. c) The surface
temperature to the test tube, by thermal imager. d)
The atmospheric temperature. e) The inlet and outlet
pressure (pressure drop in the test tube).
Repeat the experimental procedure for every case, by
changing air velocity, inlet water temperature and by
replacing the test sections (smooth and integral low
finned tube eight passes).
3.1. Water Side
The recommended correlation presented by [22] to predict
the heat transfer coefficient in a turbulent flow in tube is:
89 = 0.023%> ?.@ AB C
where Prandtl number index (n) is equal to (0.3) for cooling
process, and this equation is valid for a turbulent flow with
(0.6 <Pr<100), then the heat transfer coefficient equal to:
3. Theoretical Analysis
The first law of thermodynamics requires that the rate of
heat transfer from the hot fluid be equal to the rate of heat
transfer to the cold one, or
=
−
=
−
27
ℎ = 0.023%>E ?.@ AB C
where the Reynolds number based on the tube inside
diameter is:
%>E = and
∆
%>E = GJK
μE
K=
·E
J
where
For counter flow
=
∆
∆
∆
∆
∆
∆
∆
=
=
∆
−
=
∆
∆
∆
and
−
!
The correction factor (Fc≤ 1) depends on the geometry of
the heat exchanger, the inlet and outlet temperatures of the
hot and cold fluid streams, number of tube rows and number
of passes. The correction factor can be expressed as function
of the dimensionless ratios (R and S), given by [20, 21].
&=
" =
'(
*)
−
−
')
')
∆
The air side heat transfer coefficient general equation is
given in the form:
For a smooth tube [19, 22]:
ℎ =
R(
−
1
U
S( [ ( ]
U)
WX
−
S(
) S)
And for an integral low finned tube [18]:
ℎ =
and
+ % + 1 ln /
% − 1 ln 3
Q"
3.2. Air Side
∆
%=
IE E
PE
then
=
then for cross flow:
=
AB = =" ∆
!
N
G
4
=
n
The actual logarithmic mean temperature difference of a
cross flow multi passes heat exchanger is obtained by [20,
21]:
∆
HE 9E G
IE
or
The rate of heat transfer in a heat exchanger can also be
expressed in the following form [2, 19]:
=
FE
G
0415
0
10
2
+ 1 5 6
0415 5+ 1 5 6
R(
7
−
1
U
S(Y [ Z ]
can be calculated using:
Q
=
WX
U)
−
Q
S(Y
) S)
28
Laith Jaafer Habeeb et al.: Heat Transfer Analysis of Integral-Fin Tubes
%>[ = 4
G =
S
H[ 9[ G
I[
3.4. Enhancement Ratio Factor
The enhancement ratio factor (EF) is given by: [24]
4 \. ]
2 \ , ]
A
then89[ for a smooth tube:
89[
D ^ G /F[
Q
a
The pressure drop caused by fluid friction in the tubes is
given by [25],
D ^ G! /F[
∆Aa
3.3. Effectiveness
The effectiveness is the ratio of the actual of heat transfer
to the maximum possible amount of heat transfer during the
operation of heat exchanger, or [23]
[ a.
`
l CCmS
3.5. Water side Pressure Drop
And for an integral low finned tube:
89[
D
D
k"
[b.
a[
∆A
∆AQa![ n
aaopmQ
,
C !l aa CnQ
where, from Darcy – Weisbach equation
∆AQa![ n
aaopmQ
qr
st . u H. 9
vw
x
G
2
For turbulent flow in a smooth pipe, the Blasius correlation
valid for Re ≤ 105, is:[26]
at
[ a. q
0.316
%> ?. z
Pressure losses due to the minor fittings is:[25]
and
[b
∆A
C.
where
[ C !l aa CnQ
sl a {H w
9
x
2
where (k) is the losses coefficient.
C
is the minimum heat capacity of hot or cold fluid.
For cross – flow heat exchanger with one of the fluids
unmixed and other mixed, the relation between effectiveness
and number of transfer unit (NTU) is given by:[22]
For Cmax mixed, Cmin unmixed
ε
1
!
d1
>e T
!
1
>
f R
4. Present Correlation
Vg
For Cmax unmixed, Cmin mixed
`
1
1
>e h
!
T1
>e 8
.
!
Vi
where
!
C
[b
is the heat capacity ratio.
The (NTU) is a function of the overall heat transfer
coefficient in the form:
8
.
C
,Q
Figure (4). Sample of curve fitting for empirical relation.
In this paper, it was suggested to develop empirical
correlations for the air side heat transfer coefficient to an
integral low finned tube based on the general correlation for
Engineering and Technology 2015; 2(2): 23-34
air side Nusselt number in cross flow over tube or cylinder.
[22]
89[
%>[ C AB
}|
where C and n are constants obtained from the experimental
results as shown in fig.( 4 ).The empirical relations are given
in table (1), valid for (20838 <Rea< 63605).
5. Results and Discussion
The experimental data and results of the measurements for
the smooth and integral low finned tube at eight passes
indicated that:
•
The temperature difference in water side (∆Tw)
increases with increase inlet water temperature.
•
The temperature difference in air side (∆Ta) increases
with increase inlet water temperature, and the outlet
air temperature increase with increase inlet water
temperature.
•
The average surface temperature (Tsave.) increases
with increase inlet water temperature, and the cooling
(a)
29
value of tube surface increases with increase inlet
water temperature.
The results of calculation for the water side pressure drop
(∆Pw) in the test tube, which indicate that the water flow rate
has the main effect on the pressure drop, i.e., the pressure
drop increases with increase the water flow rate due to
increase the friction.
Figure (5) shows the relation between the heat load and
inlet water temperature at different air velocity for smooth
and integral low finned tube eight passes. It is obvious that
the heat load increases with increase inlet water temperature
due to the increase in the temperature difference between the
air temperature and surface tube temperature. The heat load
increases with increase the air velocity due to the
improvement of the overall heat transfer coefficient of the
test tube by increasing the air side heat transfer coefficient.
The figure shows that the heat load of the integral low finned
tube is higher than that of the smooth tube. The heat load of
the finned tube increased by (1.8 to 2.13) times that of
smooth tube due to increase the heat transfer surface area.
(b)
Figure (5). The variation of the heat load with inlet water temperature at: (a) smooth tube eight passes, and (b) integral low finned tube eight passes.
30
Laith Jaafer Habeeb et al.: Heat Transfer Analysis of Integral-Fin Tubes
Figure (6). The variation of the air side heat transfer coefficient with air velocity for smooth and integral low finned tube eight passes at water flow rate (5
l/min).
Figure(6) illustrates the variation of the air side heat
transfer coefficient (ho) with air velocity for smooth and
integral low finned tube. The outside heat transfer coefficient
increased with increase the air velocity, which showed that
increasing of air velocity will improve the outside heat
transfer coefficient due to increase the turbulence. The air
side heat transfer coefficient of the integral low finned tube is
higher than that of the smooth tube. The enhancement ratio
factor (EF) in the air side heat transfer coefficient when using
the integral low finned tube (EF: the ratio between the air
side heat transfer coefficient when using the integral low
finned tube to the air side heat transfer coefficient when
using the smooth tube, ( ho finned/ ho smooth)) was ( 1.86 to 2.38)
for eight passes. This was a result of the increase in the heat
transfer surface area and the effect of the turbulence
introduced by increasing the air velocity between fins.
Figure (7) illustrates the variation of the air side
temperature difference (∆Ta) with air velocity at various inlet
water temperatures. The air side temperature difference tends
to decrease with an increase in air velocity. In addition, at the
same air velocity, the air side temperature difference at the
higher inlet water temperature is higher than at the lower one
across the range of air velocity, i.e. the air side temperature
difference increases with increase inlet water temperature due
to increase the heat load.
Figure (8) shows the variation of the air side Nusselt
number with air side Reynolds number for smooth and
integral low finned tube. The air side Nusselt number
increased with increase the air side Reynolds number. This is
because the air side Nusselt number is a function of the air
Engineering and Technology 2015; 2(2): 23-34
side heat transfer coefficient, and the air side Reynolds
number is a function of air velocity, therefore, the behavior of
this figure is similar to the behavior shown in the figure for
the relation between the air side heat transfer coefficient with
air velocity (figure 6).Hence, this figure indicates that
increasing of air side Reynolds number will improve the
outside Nusselt number due to increase the turbulence. The
air side Nusselt number of the integral low finned tube is
higher than that of the smooth tube, and the enhancement
ratio factor was approximately equal to the enhancement
ratio in the air side heat transfer coefficient. This was a result
of the increase in the heat transfer surface area and the effect
31
of the turbulence introduced by increasing the air velocity
between fins.
Figure (9) depicts the variation of the effectiveness for the
test tube with the number of transfer units (NTU) at (Cr) in
the range of (0.41 to 0.84). The figure shows that increasing
the (NTU) for a specified (Cr) caused an increase in the
effectiveness values of the test tube. This is due to the
dependence of the (NTU) and the effectiveness on the overall
heat transfer coefficient, therefore, the increasing of the
(NTU) means that the overall heat transfer coefficient
increased at the given surface area, and this led to increase
the effectiveness.
Table (1). Empirical and practical relations for integral low finned tube eight passes
Water flow rate = 5 l/min
Twin
C
n
Empirical Relations
R2
50
8.8323
0.3537
Nua = 8.8323 (Rea)0.3537 Pr1/3
0.991349
60
3.8435
0.3996
Nua = 3.8435 (Rea)0.3996 Pr1/3
0.999919
0.3685
1/3
70
4.4566
0.3685
Nua = 4.4566 (Rea)
80
7.02795
0.3148
Nua = 7.02795 (Rea)0.3148 Pr1/3
(a)
Pr
0.999998
0.999057
(b)
Figure (7). The variation of the air side temperature difference with air velocity at: (a) smooth tube eight passes, and (b) integral low finned tube eight passes.
32
Laith Jaafer Habeeb et al.: Heat Transfer Analysis of Integral-Fin Tubes
Figure (8). The variation of the air side Nusselt number with air side Reynolds number for smooth and integral low finned tube eight passes at water flow rate
(5 l/min).
(a)
(b)
Figure (9). The variation of effectiveness with (NTU) at: (a) smooth tube eight passes, and (b) integral low finned tube eight passes.
Engineering and Technology 2015; 2(2): 23-34
6. Conclusions
The following points can be concluded from the present
experimental work:
1. The heat load from the test tube is directly
proportional to both the inlet water temperature and
the air velocity.
2. The heat load of the integral low finned tube is higher
than that of the smooth tube. The heat load of finned
tube was enhanced by (1.8 to 2.13) times the heat
load of smooth tube.
3. The increasing of air velocity will improve the
outside heat transfer coefficient.
4. The air side heat transfer coefficient of the integral
low finned tube is higher than that of the smooth tube.
The enhancement ratio factor (EF) in the air side heat
transfer coefficient when using integral low finned
tube was (1.86 to 2.38) for eight passes. And this
enhancement ratio from the use of the integral low
finned tube is very useful to increase the heat load
and the effectiveness.
5. The air side temperature difference and outlet air
temperature are inversely proportional to the air
velocity, and directly proportional to the inlet water
temperature.
6. The air side Nusselt number is directly proportional
to air side Reynolds number. The air side Nusselt
number of the integral low finned tube is higher than
that of the smooth tube. And the enhancement ratio
was approximately equal to the enhancement ratio in
the air side heat transfer coefficient.
The pressure drop in the test tube is directly proportional
to water flow rate.
H
K
L
nfit
np
Nu
Pr
P∆
Q
R2
Re
T
T ci
T co
T hi
T ho
∆T
∆
u
U
W
Area [m2]
Cross section area of duct [m2]
Inner cross section area of tube [m2]
Inner surface area of tube [m2]
Outer surface area of tube [m2]
Heat capacity [ kW/oC]
Specific heat of fluid [ kJ/kg.oC ]
Heat capacity ratio
Diameter [m]
Hydraulic diameter [m]
Outer diameter of finned tube [m]
Root diameter [m]
Friction factor
Logarithmic mean temperature correction factor
Mass velocity [kg/m2.sec]
heat transfer coefficient [W/m2.oC]
Height of the duct [m]
Thermal conductivity [W/m.oC]
Length of tube [m]
Mass flow rate [kg/sec]
Number of fitting
Number of tube passes
Nusselt number
Prandtl number
Pressure drop [Pa]
Heat load [kW]
Correlation Coefficient
Reynolds number
Temperature [oC]
Inlet temperature of cold fluid [oC]
Outlet temperature of cold fluid [oC]
Inlet temperature of hot fluid [oC]
Outlet temperature of hot fluid [oC]
Temperature difference [oC]
Logarithmic mean temperature difference [oC]
Fluid velocity [m/sec]
Overall heat transfer coefficient [W/m2.oC]
Width of the duct [m]
Heat exchanger effectiveness
µ
ρ
Fluid viscosity [kg/m.sec]
Fluid density [kg/m3]
References
[1]
S. P. Sukhatme, B. S. Jagadish and P. Prabhakaran, “Film
Condensation of R-11Vapor on Single Horizontal Enhanced
Condenser Tubes “, Transactions of the ASME, Journal of
Heat Transfer, Vol.112, pp.229-234, 1990.
[2]
WessamFalih Hasan, “Theoretical and Experimental Study to
Finned Tubes Cross Flow Heat Exchange “, Master thesis,
Mech. Eng. Dept., University of Technology, 2008.
[3]
Virgil J. Lunardini and Abdul Aziz, “Effect of Condensation
on Performance and Design of Extended Surfaces “, CRREL
Report 95-20, Cold Regions Research and Engineering
Laboratory, 1995.
[4]
R. K. Al-Dadah and T. G. Karayiannis, “Passive Enhancement
of Condensation Heat Transfer“, Applied Thermal
Engineering, 18, pp.895-909, 1998.
[5]
Wolverine Tube Inc., “Wolverine Engineering Data Book II “,
2001.
[6]
Wolverine Tube Inc., “Wolverine Engineering Data Book III “,
was updated in 2007.
[7]
D.G. Rich, “The Effect of Fin Spacing on the Heat Transfer
and Friction Performance of Multi-Row, Smooth Plate Finand-Tube Heat Exchangers”, ASHRAE Transactions, Vol. 79,
No.2, pp.135-145, 1973.
Nomenclature
A
Ad c
Ai c
Ai s
Ao s
C
cp
Cr
d
dh
do f
dr
f
Fc
G
h
33
34
Laith Jaafer Habeeb et al.: Heat Transfer Analysis of Integral-Fin Tubes
[8]
D. G. Rich, “The Effect of the Number of Tube Rows on Heat
Transfer Performance of Smooth Plate Fin-and-Tube Heat
Exchangers”, ASHRAE Transactions, Vol. 81, pp. 307-317,
1975.
[17] Ali Hussain Tarrad, “A Numerical Model for ThermalHydraulic Design of a Shelland Single Pass Low Finned Tube
Bundle Heat Exchanger“, Eng. & Technology, Vol. 25, No. 4,
pp.619-645, 2007.
[9]
Brown, R., “ A Procedure for Preliminary Estimates of Air
Cooled Heat Exchangers”, in Chemical Engineering,
McGraw-Hill Publication Book Co., Newyork, pp.412-417,
1997.
[18] Jose´ Ferna´ndez-Seara, Francisco J. Uhı´a and Rube´nDiz,
“Experimental Analysis of Ammonia Condensation on Smooth
and Integral-Fin Titanium Tubes “, International Journal of
Refrigeration, 32, pp.1140-1148, 2009.
[10] C.C. Wang, K.Y. Chi, Y.J. Chang, and Y.P. Chang, “A
Comparison Study of Compact Plate Fin-and-Tube Heat
Exchangers”. ASHRAE Transactions, #TO-98-3-3, 1998.
[19] Frank B. Incropera, and David B. Doot, “Principles of Heat
Transfer “, McGraw-Hill Co., 1986.
[11] Fethi Halici and Imdat Taymaz, “Experimental Study of the
Airside Performance of Tube Row Spacing in Finned Tube
Heat Exchangers”, Heat Mass Transfer, 42, pp.817–822, 2006.
[12] Han-Taw Chen and Wei-Lun Hsu, “Estimation of HeatTransfer Characteristics on a Vertical Annular Circular Fin of
Finned-Tube Heat Exchangers in Forced Convection”,
International Journal of Heat and Mass Transfer, 51, pp.1920–
1932, 2008.
[13] Jong Min Choi, Yonghan Kim, Mooyeon Lee and Yongchan
Kim, “Air Side Heat Transfer Coefficients of Discrete Plate
Finned-Tube Heat Exchangers with Large Fin Pitch”, Applied
Thermal Engineering, 30, pp.174–180, 2010.
[14] H. Honda, S. Nozu and Y. Takeda, “A Theoretical Model of
Film Condensation in a Bundle of Horizontal Low Finned
Tubes “, Transactions of the ASME, Journal of Heat Transfer,
Vol.111, pp.525-532, 1989.
[15] W. Y. Cheng, C. C. Wang, Y. Z. Robert Hu and L. W. Huang,
“Film Condensation of HCFC-22 on Horizontal Enhanced
Tubes“, Int. Comm. Heat Mass Transfer, Vol. 23, No.1, pp.7990, 1996.
[16] Ravi Kumar, H. K. Varma, BikashMohanty and K. N. Agrawal,
“Prediction of Heat Transfer Coefficient during Condensation
of Water and R-134a on Single Horizontal Integral-Fin Tubes
“, International Journal of Refrigeration, 25, pp.111-126, 2002.
[20] R. K. Sinnott, “Chemical Engineering Design “, Volume 6,
Fourth edition, Elsevier Butterworth-Heinemann, 2005.
[21] Ali Hussain Tarrad, Fouad Alwan Saleh and Ali Ahmed
Abulrasool , “ A Simplified Numerical Model for a Flat
Continuous Triangle Fins Air Cooled Heat Exchanger Using
aStep by Step Technique “, Journal of Engineering and
Development, Vol.13, No. 3, pp.38-60, 2009.
[22] J. P. Holman, “Heat Transfer “, Ninth edition,McGraw-Hill
Co., 2002.
[23] Ali Hussain Tarrad, Ma’athe AbulWahed and Dhamia’a Saad
Khudor, “A Simplified Model for the Prediction of the Thermal
Performance for Cross Flow Air Cooled Heat Exchangers
with a New Air Side Thermal Correlation “, Journal of
Engineering and Development, Vol.12, No. 3, pp.88-119, 2008.
[24] Satesh Namasivayam and Adrian Briggs, “Effect of Vapour
Velocity on Condensation of Atmospheric Pressure Steam on
Integral-Fin Tubes “, Applied Thermal Engineering, 24,
pp.1353–1364, 2004.
[25] American Society of Heating Refrigeration, and Air
Conditioning Engineers, “ASHRAE Fundamentals Handbook
“, Chapter 22, pp.22.1-22.21, 2009.
[26] Victor L. Streeter, E. Benjamin Wylie and Keith W. Bedford,
“Fluid Mechanics”, Ninth edition, McGraw-Hill Co., 1995.