02.jnae10085

International Journal of Innovative Research in Advanced Engineering (IJIRAE) ISSN: 2349-2163
Issue 6, Volume 2 (June 2015) www.ijirae.com
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© 2014-15, IJIRAE- All Rights Reserved Page -10
Experimental Investigation on the Effect of Fluid Flow Rate on
the Performance of a Parallel Flow Heat Exchanger
Christian O. Osueke, Anthony O. Onokwai Adeyinka O. Adeoye
Department of Mechanical Department of Mechanical Department of Mechatronics
Engineering Landmark University Engineering Landmark University Engineering Afe Babalola University,
Omu-Aran, Kwara State, Nigeria Omu-Aran, Kwara State, Nigeria Ado-Ekiti,Ekiti State, Nigeria
Abstract -- The pervading industrial importance of Heat exchanger in heat transfer is one of the major motivations to carry
out this work. A plate heat exchanger is a type of heat exchanger that uses metal plates to exchange heat between two liquids
with high density fluid. This research focused on the use of an extended plate heat exchanger using water as working fluid.
This research work deals with an experimental Investigation on the effect of Fluid Flow Rate on the Performance of a Parallel
Flow Heat Exchanger. The extended plate heat exchanger consists of plates overall dimensions: 75mm by 115mm. Effective
diameter: 3.0mm, plate thickness: 0.5mm, wetted perimeter: 153.0mm and Projected heat transmission area: 0.008m2
per plate.
The study was limited to the physical characteristics and thermal performance of a parallel flow heat exchanger stationed at
the Mechanical Engineering Laboratory of landmark University. Experimental results in the form temperature distribution
and flow rates were analyzed to generate the thermal performance measures of the heat exchanger. The study was limited to
the physical characteristics and thermal performance of a parallel flow heat exchanger stationed at the Mechanical
Engineering Laboratory of landmark University. Experimental results in the form temperature distribution and flow rates were
analyzed to generate the thermal performance measures of the heat exchanger. Experimental results in the form temperature
distribution, velocity and flow rates, were analyzed to generate the Reynolds numbers, Nusselt numbers, Prandtl numbers,
thermal performance, logarithmic mean temperature difference convective and overall heat transfer coefficient of the heat
exchanger. It was deducted that rise in efficiency requires faster increase in flow rate of the hot stream than of the cold stream.
Also the heat transfer coefficient increases with Reynolds Number/Nusselt number. Increase in Reynolds and nusselt number
is an indication that flow is becoming more turbulent and results into higher heat transfer rates.With this work as foundation,
recommendations for future research included more advanced study that would involve determination of temperature
distribution by solving heat/mass transfer equation. This level of analysis will require knowledge of thermal properties and
boundary conditions. It was also recommended that counter-current flow of same facility be investigated for instructive
comparison with the studied parallel flow under the background of theoretical result that given mass flows and temperature
differences, the counter-flow heat exchanger requires less surface area (thus less length) than its parallel flow equivalent
Keywords- Extended plate heat exchanger, thermal efficiency, flow rate, Convective heat transfer coefficient, Overall heat
transfer coefficient, Reynolds number, nusselt number.
I Introduction
Heat exchanger is a device in which transfer of thermal energy takes place between two of more fluids across a solid surface.
These exchangers are classified according to construction, flow arrangement; number of fluids, compactness, etc. The use of heat
exchanger gives higher thermal efficiency to the system. In many applications like power plants, petrochemical industries, air
conditioning etc. heat exchangers are used. Plate heat exchanger is generally used in dairy industry due to its ease of cleaning and
thermal control. The plate heat exchangers are built of thin metal heat transfer plates and pipe work is used to carry streams of
fluid. Plate heat exchangers are widely used in liquid to liquid heat transfer and not suitable for gas to gas heat transfer due to high
pressure drop [1]. A plate heat exchanger is a type of heat exchanger that uses metal plates to exchange heat between two liquids.
This has a noteworthy favorable position more than a conventional heat exchanger in that the liquids are presented to a much
bigger surface range in light of the fact that the liquids spread out over the plates. This encourages the exchange of heat, and
enormously builds the pace of the temperature change. Plate heat exchanger consists of parallel metal plates that are corrugated
both to increase turbulence and to provide mechanical rigidity. These normally have four flow parts, one in each corner, and are
sealed at their outer edges and around the ports by gaskets, which are shaped to prevent external leakages and to direct the two
liquid through the relatively narrow passages between alternate pairs of heat transfer plates. The plates are clamped together in a
frame that includes connections for the fluid. All wetted parts are accessible for inspection by removing the clamping bolts and
rolling back the removable cover [2].

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R. K. Shah and S. G. Kandilkar [3], have experimentally investigated the influence of number of thermal plates on effectiveness of
heat exchanger for 1 pass 1, 2 pass 1, 3 pass 1 flow arrangements and number of plates up to 41. Results were plotted for number
of plates and F, NTU and F, for 4 different pass arrangements. They concluded, for 1pass1 flow arrangement with an even number
of thermal plates, fluid in the outermost channels is same. The heat transfer rate of multi pass arrangement may be higher or lower
than that of 1pass1 for same N and R which depends upon heat transfer characteristics of plate material. For N < 40, end effect is
considerable. When there is significant imbalance in flow rates, R < 2, 1pass1 arrangement is desirable. For (R=2, 3) 2pass1
arrangement is desirable and for R > 4, 3 pass 1 arrangement is desirable and for 1 pass 1 exchanger with an even number of
thermal plates the fluid in outermost channel is same. The exchanger effectiveness is slightly higher if outer fluid has higher heat
capacity as compared to other fluid having one less flow channel. [4]
H. Dardour, S. Mazouz, and A. Bellagi [5] had done numerical analysis of the thermal performance of a plate type heat exchanger
with parallel flow configuration. The computation is based on the effectiveness- NTU model. The numerical results illustrate the
evolution of the most important parameters of the plate heat exchanger. A parametric analysis is presented which brings out the
effect of NTU and the R parameter, the heat capacity rate ratio, on the performance of the plate heat exchanger (PHE). To check
the validity of the presented simplified model established to describe the energy balances in the PHE and the numerical scheme
adopted, simulated performance has been compared to the performance evaluated by theoretical relations. Comparison shows an
excellent agreement between them. The temperature gradients through each channel and heat fluxes through each active plate are
also evaluated. [6]
Murugesan M.P. and Balasubramani [7] Performed test for the investigation of milk adhesion and the stability of the coatings on
corrugated plates. A number of coatings and surface treatments were tested. Heat exchanger plates coated with different nano-
composites as well as electro polished plates installed in the heating section of the pasteurizer were tested. Significant differences
were observed between coated and uncoated plates. The coated plates showed that reduced deposit buildup in comparison with the
uncoated stainless steel plates. The time required for cleaning place with the coated plates was reduced by 75% compared to
standard stainless steel plates [8]. They also investigate heat transfer performance of plate type heat exchanger experimentally by
varying operating parameters and design parameters. Heat transfer coefficient was studied for various fluids like water and
ethylene glycol. The increase mass flow rate with subsequently increase in the flow velocity has led to an increased overall heat
transfer coefficient as well as individual heat transfer coefficient. [9]
T K S Sai Krishna, S G Rajasekhar, C Pravarakhya [10] modeled the plate type heat exchanger in solid works and the fluid flow
analysis is done on the modeled fluid part. The analysis stated that when the thickness of the plates decreases then the heat flow is
higher and if the number of plates increases then the outlet temperature difference of the fluids increased and the pressure contour
stated that, there is little pressure drop in the entry and outlet of the fluid, From the turbulent contour it is interfered that there is
very high turbulence in the entry and outlets due to sudden change in cross section along the plates. [11]
This paper focuses on an experimental investigation of the performance of a parallel flow heat exchanger as well as the effect of
fluid flow rate with respect to overall heat transfer coefficient.
II METHODOLOGY
A. Experimental Set Up
1) Test Procedure: The plate heat exchanger with flat plates is used for trials.The fluids used are hot and cold water. Two
flow arrangements implemented which are parallel flow and counter flow. Trials conducted with different mass flow rate of
hot and cold water and also hot water inlet flow rate was kept constant while cold water inlet flow rate varied. Procedure
repeated for getting more accurate results and results plotted
Fig.1 Hydraulic bench Fig.2 Extented plate heat exchanger

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1. Base Plate
2. Fixed endplate
3. Heat exchanger plates
4. Moving end plate
5. Frame
6. Central bolt
7. Intermediate plate
Fig. 3 Extended plate heat exchanger mounted on service unit
Fig. 4 Hydraulic bench containing the fluid and extended plate heat exchanger mounted on a services unit.
B. Equipment details
TABLE 1
EXTENDED PLATE HEAT EXCHANGER
Plates overall dimensions 75mm by 115mm
Effective diameter 3.0mm
Plate thickness 0.5mm
Wetted Perimeter 153.0mm
Projected heat transmission area 0.008m2
per plate.
TABLE 2
HYDRAULIC BENCH
Circulating Pump Type: Centrifugal
Max. Head: 21m Water
Max. Flow: 80litres/min(Using Volumetric tank)
Max. Flow: 100litres/min(Using appropriate accessory)
Pump Motor Rating 0.37Kw
Sump Tank Capacity 250litres
High-Flow Volumetric Tank Capacity 40litres
High-Flow Volumetric Tank Capacity 6litres

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TABLE 3
SERVICE UNIT
Height- 430mm
Length- 1000mm
Depth- 500mm
Hot Water Vessel Capacity 1.5litres.
C. Assumptions
1) The plate heat exchanger operates under steady state conditions,
2) No phase change occurs; both fluids are single phase and are unmixed,
3) Heat losses to surrounding are negligible,
4) The temperature in the fluid streams is uniform,
5) The fluids have constant specific heats,
6) The fouling resistance is negligible,
7) Pressure drop across heat exchanger is negligible.
TABLE 4
PROPERTIES OF WATER AT MEAN TEMPERATURE.
Property Unit(Metric) Hot Water
(Mean Temperature)
Cold Water
(Mean Temperature)
Heat Capacity(Cp) KJ/KgK 4.178 4.181
Thermal Conductivity(K) W/mK 0.6526 0.6174
Dynamic Viscosity Ns/m2
0.0006284 0.0006312
Density(ρ) Kg/m3
994.1 997.1
Specific Volume(v) M3
/Kg 1.01 1.00
Absolute Pressure KN/m2
5.6 3.2
Specific Entropy KJ/KgK 0.505 0.367
The two integrated forms of heat transfer equation of 100% efficient parallel-flow and counter-flow (with hot fluid being the
reversed flow) heat exchanger are
−
̇
±
̇
̇ = ∆ − ∆ (1)
− ̇
± ̇
= ln
∆
∆
(2)
where ̇ and ̇ are the mass flow rate of the cold and hot fluids respectively, and are the specific heat capacities of the
cold and hot fluids respectively, ̇ is the total heat exchange between the hot and cold fluid steams, ∆ and ∆ are the
temperature differences between the hot and cold fluid steams at the outlet and inlet of the heat exchanger respectively, is the
overall heat transfer coefficient and is the heat exchange area. Dividing equation (2) with equation (1) and rearranging gives
̇ =
∆ ∆
(∆ ∆⁄ )
(3)
It is seen from equation (3) that the logarithmic mean temperature difference ∆ is
∆ =
∆ ∆
(∆ ∆⁄ )
(4)
When ̇ is viewed as ∆ . At 100% efficiency all the heat emitted by the hot stream is absorbed by the cold stream.
When the heat exchange between the hot and cold fluid steams is not 100% efficient, the following nomenclature are
introduced; rate of emission of heat or heat power emitted by the hot stream ̇ , rate of absorption of heat or heat power absorbed
by the cold stream ̇ and overall efficiency . These are respectively given by
̇ = ̇ ∆ (5)
̇ = ̇ ∆ (6)
=
̇
̇
=
̇
̇
∆
∆
(7)
Where ∆ and ∆ are magnitude of the temperature differences between the outlet and inlet of the hot and cold streams
respectively. The overall heat transfer coefficient should have been given as = ̇ ( ∆ )⁄ if not for physical construction that
sometimes causes a deviation from either 100% parallel flow or 100% counter flow. This is taken care of by introduction of
correction factor such that

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=
̇
∆
=
̇
∆
(8a)
The value stipulated for in the user manual of the extended plate heat exchanger is 0.95 then
=
̇
. ∆
=
̇
. ∆
(8b)
D. Qualitative and Tabular Analysis of Experimental Results
The studied system is in parallel flow meaning that ∆ = − and ∆ = − (see figure 5) then
∆ =
( − ) − ( − )
ln[( − ) ( − )⁄ ]
(2.9)
This is better understood by a simplified diagram of the studied mode of flow as given in fig. 5 below
Fig. 5 A simplified diagram of the experimental parallel flow heat exchanger
The deduction is that ∆ is only realistic when > and > . For mathematical justification of this point suppose
> and > then the denominator of equation (9) gives
ln[− ] = ln[−1 × ] = ln[−1] + ln[ ] = (10)
Where the positive real number is given by = |( − ) ( − )⁄ |=-( − ) ( − )⁄ =( − ) ( − )⁄ .
Equation is rewritten based on the complex number theory as
ln[− ] = ln[ ( )] + ln[ ] = + ln[ ] = (11)
where is the unit magnitude complex number √−1. The deduction from equation (11) is that the condition > and >
causes the denominator to have a complex value and hence causes ∆ to have a complex value which is not supposed to be so.
The conclusion is that the realistic condition for there to be real and positive value for ∆ is > and > . These are
conditions that are consistent with the first and second laws of thermodynamics. The experimental results from the extended plate
heat exchanger are given in table 5. The experimental runs that do not meet with the necessary condition > are put in red
in table 5. This experimental runs are considered invalid and are not analyzed further in what follows. Table 5 is re-presented as
table 6 containing only the relevant experimental runs. Also in table 6 are presented the volumetric flow rates in m3
s-1
and the
computed values of the flow capacities ̇ and ̇ . This value make computations easier as will be seen in what follows.
1) Area of the flow : = 3 where 3 is the number of active plates per pack for the studied heat exchanger, = 4 is the
number of packs utilized in the experiment and = 0.008 is the projected heat transfer area of every plate then
= 3 × 4 × 0.008 = 0.096
2) Hydraulic Diameter: It is the ratio of cross sectional area of the channel to the wetted perimeter of the channel
P
A
DH
4

Where, A= Area of Flow in m2
, P= wetted Perimeter of the plate in m and HD = Hydraulic Diameter
m
x
5098.2
153.0
096.04

hot
cold
1T 2T 3T 4T 5T
6T 7T 8T 9T 10T

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3) Velocity of flow:
A
m
V


Where, A= Area of flow in m2
, ρ= Density in Kg/m3
and m = mass flow rate in Kg/sec.
 Velocity for Cold Water
sm
x
V /00022461.0
1.997096.0
0215.0

 Velocity for Hot Water
sm
x
V /000348934.0
1.994096.0
0333.0

4) Reynolds Number: It is the ratio of inertia forces to viscous forces.
Re = inertial forces/viscous forces

 HH VDD
Re
Where, V= mean velocity of the object relative to the fluid in m/s
HD =Hydraulic Diameter in m
 =dynamic viscosity of the fluid in Ns/m3
 =Kinematic viscosity )(


  in m2
/s
 =density of the fluid in Kg/m3
 Reynolds Number for Cold Water
5123.890
0006312.0
5098.200022461.01.997
Re 
xxDV
c
Hcc


 Reynolds Number for Hot Water
2686.1385
0006284.0
5098.20003489.01.994
Re 
xxDV
h
Hhh


5) Prandtl Number: It is the ratio of momentum diffusivity (kinematic viscosity) to thermal conductivity.
Pr = Viscous diffusion rate/thermal diffusion rate
K
CV
P p
r



V= Kinematic viscosity,


V (m2
/s)

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 = Thermal diffusivity,  =
pC
K

 (m2
/s)
 =Dynamic viscosity (Ns/m2
)
pC =Specific heat (J/Kgk)
K = Thermal Conductivity (W/mk)
 =Density (Kg/m3
)
 Prandtl Number for Cold Water
c
pcc
c
c
rc
K
CV
P



0004274.0
6174.0
181.40006312.0

x
rc
 Prandtl Number for Hot Water
0004027.0
6525.0
181.40006284.0

x
rh
6) Nusselt Number: It is the ratio of convective to conductive heat transfer across the boundary.
K
hD
Nu H

Where, h = Heat transfer coefficient
DH = Hydraulic viscosity in m
K = Thermal conductivity in W/mK
Nu = Nusselt Number
 Nusselt Number for Cold Water:
c
Hcc
c
K
Dh
Nu 
5271.10004274.05123.890662.0
Re662.0
33.05.0
33.05.0


xxNu
PNu
c
rccc
 Nusselt Number for Hot Water
h
h
h
K
Dhh
Nu 
8676.10004027.02686.1385662.0
Re662.0
33.05.0
33.05.0


xxNu
PNu
h
rhhh

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TABLE 5
EXPERIMENTAL RESULTS OBTAINED FROM THE EXTENDED PLATE HEAT EXCHANGER
TABLE 6
RESULTS OF HOT AND COLD FLOW RATES AND THERMAL CAPACITY
TABLE 7
RESULTS OF THERMAL EFFICIENCY AND OVERALL HEAT TRANSFER COEFFICIENT
EXP
1T
c0
/
2T
c0
/
3T
c0
/
4T
c0
/
5T
c0
/
6T
c0
/
7T
c0
/
8T
c0
/
9T
c0
/
10T
c0
/ phh
pcc
cm
cm 1T -
5T
10T
- 6T
inT


n U
)/( 2
KmW
1 40.6 39.8 38.2 36.7 34.5 26.5 29.2 32.8 33.6 28.9 0.6453 6.1 2.4 3.966 0.253 3145.687
2 40.5 39.6 36.9 36.3 31.3 26.8 30.1 31.4 33.7 30.4 0.481 9.2 3.6 6.559 0.247 5636.15
3 40.1 39.2 37.4 35.3 31.9 26.3 27.4 31.9 32.9 31.3 0.777 8.2 5 6.47 0.474 3322.069
4 39.5 38.3 35.4 32.1 32.2 26.8 27.3 30.7 31.7 33.5 0.9367 6.1 6.7 6.97 0.862 2270.65
5 40.2 38.3 36.1 33.5 32.7 26.1 28.5 31.1 31.8 36.1 0.3751 7.5 10 8.711 0.499 4688.35
6 40.5 37.6 34.3 34.3 31.5 26.3 28.6 30.8 31.9 34.2 1.729 9 7.9 6.624 0.875 828.497
7 40.2 36.3 35.3 35.3 30.8 26.7 30.3 32.6 32.7 33.1 1.9727 9.4 6.4 7.813 0.677 1234.457
8 39.8 39.3 37.4 32.8 31.9 26.2 27.3 31.3 28.6 28.3 1.2052 7.9 2.1 4.377 0.321 1527.178
9 40.2 38.3 35.3 33.9 32.7 26.3 27.9 30.4 33.6 36.8 1.2166 7.5 10.5 5.411 0.609 1161.159
10 40.7 38.6 36.4 32.8 31.8 26.5 28.4 30.8 29.4 35.1 1.8053 8.9 8.6 5.019 0.503 999.646
NO
of
Exp
1T
c0
/
2T
c0
/
3T
c0
/
4T
c0
/
5T
c0
/
6T
c0
/
7T
c0
/
8T 9T
c0
/
10T
c0
/
Liters
per sec
Liters
per sec
TT 0
42 / 1T - 5T
c0
/
10T - 6T
c0
/
1 40.6 39.8 38.2 36.7 34.5 26.5 29.2 32.8 33.6 28.9 2 1.68 6.1 6.1 2.4
2 40.5 39.6 36.9 36.3 31.3 26.8 30.1 31.4 33.7 30.4 2 1.52 9.2 9.2 3.6
3 40.1 39.2 37.4 35.3 31.9 26.3 27.4 31.9 32.9 31.3 2 1.31 8.2 8.2 5
4 39.5 38.3 35.4 32.1 32.2 26.8 27.3 30.7 31.7 33.5 2 0.99 6.1 6.1 6.7
5 40.2 38.3 36.1 33.5 32.7 26.1 28.5 31.1 31.8 36.1 2 1.92 7.5 7.5 10
6 40.5 37.6 34.3 34.3 31.5 26.3 28.6 30.8 31.9 34.2 1 1.62 9 9 7.9
7 40.2 36.3 35.3 35.3 30.8 26.7 30.3 32.6 32.7 33.1 1 1.31 9.4 9.4 6.4
8 39.8 39.3 37.4 32.8 31.9 26.2 27.3 31.3 28.6 28.3 1 0.63 7.9 7.9 2.1
9 40.2 38.3 35.3 33.9 32.7 26.3 27.9 30.4 33.6 36.8 1 0.81 7.5 7.5 10.5
10 40.7 38.6 36.4 32.8 31.8 26.5 28.4 30.8 29.4 35.1 1 0.8 8.9 8.9 8.6
NO
of
Exp
1T
c0
/
2T
c0
/
3T
c0
/
4T
c0
/
5T
c0
/
6T
c0
/
7T
c0
/
8T
c0
/
9T
c0
/
10T
c0
/
hotq
)/( 3
sm
coldq
)/( 3
sm
phhcm pcccm
phh
pcc
cm
cm
1 40.6 39.8 38.2 36.7 34.5 26.5 29.2 32.8 33.6 28.9 3.33333x10-5 2.150x10-5 139.394 89.956 0.6453
2 40.5 39.6 36.9 36.3 31.3 26.8 30.1 31.4 33.7 30.4 3.33333x10-5 1.603x10-5 139.394 67.07 0.481
3 40.1 39.2 37.4 35.3 31.9 26.3 27.4 31.9 32.9 31.3 3.33333x10-5 2.59x10-5 139.394 108.366 0.777
4 39.5 38.3 35.4 32.1 32.2 26.8 27.3 30.7 31.7 33.5 3.33333x10-5 3.121x10-5 139.394 130.583 0.9367
5 40.2 38.3 36.1 33.5 32.7 26.1 28.5 31.1 31.8 36.1 3.33333x10-5 1.25x10-5 139.394 52.3 0.3751
6 40.5 37.6 34.3 34.3 31.5 26.3 28.6 30.8 31.9 34.2 1.66667x10-5 2.79x10-5 69.78 116.734 1.729
7 40.2 36.3 35.3 35.3 30.8 26.7 30.3 32.6 32.7 33.1 1.66667x10-5 3.29x10-5 69.78 137.654 1.9727
8 39.8 39.3 37.4 32.8 31.9 26.2 27.3 31.3 28.6 28.3 1.66667x10-5 2.01x10-5 69.78 84.098 1.2052
9 40.2 38.3 35.3 33.9 32.7 26.3 27.9 30.4 33.6 36.8 1.66667x10-5 2.029x10-5 69.78 84.893 1.2166
10 40.7 38.6 36.4 32.8 31.8 26.5 28.4 30.8 29.4 35.1 1.66667x10-5 3.011x10-5 69.78 125.98 1.8053

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TABLE 8
RESULTS OF REYNOLDS NUMBER, PRANDTL NUMBER, NUSSELT NUMBER, HEAT TRANSFER COEFFICIENT AND OVERALL
HEAT TRANSFER COEFFIECIENT.
cRe hRe rcP rhP cNu hNu cH hH
866.1608 1004.45 4.222012 5.714589 31.33853 37.29299 0.007983 0.00938 3145.69
864.6954 1010.221 4.229995 5.68148 31.33154 37.32833 0.00798 0.00939 5636.15
858.1992 1001.025 4.257356 5.734454 31.28011 37.27201 0.007976 0.00937 3322.07
808.7602 1010.221 4.525615 5.68148 30.98429 37.32833 0.007889 0.00939 2270.65
859.4837 997.623 4.256147 5.754319 31.30057 37.25111 0.007972 0.00937 4688.35
858.1655 1001.025 4.262018 5.734454 31.29079 37.27201 0.00797 0.00937 828.497
858.1655 1007.899 4.262018 5.694723 31.29079 37.31406 0.00797 0.00938 1234.46
850.5794 998.7453 4.305279 5.747698 31.25618 37.2579 0.007954 0.00937 1527.18
858.1655 1001.025 4.262018 5.734454 31.29079 37.27201 0.00797 0.00937 1161.16
866.6366 1004.45 4.221509 5.714589 31.34591 37.29299 0.007981 0.00938 999.646
Fig. 6 Graph of the ratio of cold to hot thermal capacity against overall heat transfer coefficient
From fig. 6 above the overall heat transfer coefficient of the heat exchanger approximately falls with rise in the ratio of thermal
capacities
̇
̇
Fig. 7 Graph of the ratio of cold to hot thermal capacity against thermal efficiency
Fig. 7 above, efficiency of the heat exchanger approximately increases with rise in the cold stream flow rate. This is achieved
by making sure that the hot stream flow rate is stationary.

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Fig. 8 Flow rate ratio for cold/hot fluid against intermediate temperature
One important observation from fig. 8 above is that intermediate temperatures for the extended plate heat exchanger increase as
the flow rate ratio for cold/Hot fluid increases; this is as a result of increase in the cold water flow rate while the hot water flow
rate is kept constant at a low temperature. The first intermediate temperature for the heat exchanger has the maximum temperature
that is 10o
, thus possesses higher thermal efficiency. The intermediates temperatures decrease as the flow rate ratio between the
cold to hot stream increases, while that of the third intermediate temperature increase gradually. This is due to the increase in the
pressure from the hydraulic bench. The first and second intermediate temperatures are equal when the flow rate is 6.55E-06 and
2.01E-5 respectively. While that of first and second intermediate temperature are the same, when the flow rate is 4.95E-06.
Fig. 9 Graph of Logarithmic Mean Temperature Difference against Ratio of Cold to Hot flow rate
From fig. 9, the temperature driving force for heat transfer increases as the flow rate increases until it get to the maximum point
when the flow rate is 5.5m3
/s, after that, it decreases gradually as the flow rate continue to increases at 8m3
/s.
Fig. 10 Graph of cold reynolds number against overall heat transfer coefficient
2.4 2.48
5
6.7
10
4.7
6.4
2.1
3.75
2.47
6.1
9.2
8.2
7.3 7.5
9 9.4
7.9 7.5
8.9
4.4
3.6
5.5
4.4
3.3 3.3
2.4
1.3
5.7
1
8.40E-067.60E-066.55E-064.95E-069.60E-062.79E-053.29E-052.01E-052.03E-053.01E-05
0
2
4
6
8
10
12
1 2 3 4 5 6 7 8 9 10
Graph of the flowrate ratio for cold/hot fluid against
intermidiate temperature
T7-T9 T1-T5 T1-T6 Qcold/Qhot
0.00E+00
2.00E+00
4.00E+00
6.00E+00
8.00E+00
1.00E+01
1 2 3 4 5 6 7 8 9 10
Qcold/Qhot
Tin
0
2000
4000
6000
OverallHeat
Transfer
Coefficient(U)in
W/m2K)
Reynolds Number (Reh)

_________________________________________________________________________________________________________
Fig. 11 Graph of hot reynolds number against overall heat transfer coefficient
Fig. 10-11 Shows the variation of Reynolds against convective heat transfer coefficient. From the figure, it is deduce that the
convective heat transfer coefficient increases/decreases with an increase in Reynolds number. This is due to the increase/decrease
in the ratio of inertia to viscous forces in the fluid.
Fig. 12 Graph of cold nusselt number against overall heat transfer coefficient
Fig. 13 Graph of hot nusselt number against overall heat transfer coefficient
Fig. 12-13 Shows a gradual increase and decrease in overall heat transfer coefficient with an increase in Nusselt number. The
increase in overall heat transfer coefficient is as a result the corresponding increase/decrease in the ratio of convective to
conductive heat transfer across the boundary.
0
1000
2000
3000
4000
5000
6000OverallHeatTransfer
Coefficient(U)in
W/m2K)
Reynolds Number (Rec)
0
1000
2000
3000
4000
5000
6000
OverallHeatTransfer
Coefficient(W/m2K)
Nusselt Number(Nuc)
0
1000
2000
3000
4000
5000
6000
OverallHeatTransfer
Coefficient(U)
(W/m2K)
Nusselt Number(Nuh)

_________________________________________________________________________________________________________
Fig. 14 Graph of cold reynolds number against heat transfer coefficient
Fig. 15 Graph of hot reynolds number against heat transfer coefficient
Fig. 14-15 Shows the graph of convective heat transfer coefficient against Reynolds number. From the figure, it is deduce that
the convective heat transfer coefficient increases with an increase in Reynolds number due to increase in the variation of the
inertia forces applied to the heat exchanger, while the decreases is as a results of decrease in the inertia to viscous forces in the
heat exchanger. Increase in Reynolds number shows that the flow is turbulent and lead to a high rate of heat transfer.
Fig. 16 Graph of cold nusselt number against heat transfer coefficient
0.0078
0.00785
0.0079
0.00795
0.008
ConvectiveHeat
TransferCoefficient
(Hc)(W/m2K)
Cold Reynolds Number(Rec)
0.00935
0.00936
0.00937
0.00938
0.00939
ConvectiveHeat
TransferCoefficient
(Hh)W/m2)
Hot Reynolds Number (Reh)
0.0078
0.00785
0.0079
0.00795
0.008
ConvectiveHeat
Transfer
Coefficient(Hc)
Nusselt Number (Nuc)

_________________________________________________________________________________________________________
Fig. 17 Graph of hot nusselt number against heat transfer coefficient
Fig. 16-17 Shows the graph of convective heat transfer coefficient against Nusselt number. From the figure, it is deduce that the
convective heat transfer coefficient slightly increase with an increase in Nusselt Number, leading to a more active convective, with
turbulent flow. The decrease in convective heat transfer coefficient is as a result of decrease in convective heat transfer across the
boundary.
Fig. 18 Graph of mass flow rate against overall heat transfer coefficient
Fig. 18 above shows the variation of overall heat transfer coefficient against mass flow rate. From the figure, it is deduce that
the overall heat transfer coefficient increases with an increase in mass flow rate. This is due to the increase in the flow velocity
which can also lead to increase in the heat transfer rate.
III CONCLUSION
This research focuses on an experimental investigation of the effect of fluid flow rate on the performance of a parallel flow heat
exchangers in an extended plate with regard to thermal efficiency, overall heat transfer coefficient, convective heat transfer
coefficient, flow rate, and Reynolds number. Physical characteristics and thermal performance of a real heat exchanger were
studied in this work. The heat exchanger was supplied to the Mechanical Engineering laboratory of Landmark University with the
model name “HT30XC Heat exchanger Service Unit”. The detailed description of the unit is given in is as given in the previous
section. Even though the Unit can be configured for either parallel or counter-current flow by changing the direction of the pump
controlling the hot water flow, only the co-current flow was studied in this work. The experimental results that violet the laws of
thermodynamics were considered experimental outliers and discarded. Using the experimental results the thermal performance
characteristics of the heat exchanger which include; efficiency, overall heat transfer coefficient and logarithmic mean temperature
difference were calculated for all the experimental runs. The relationship between the first two and the ratio of thermal capacities
̇
̇
was presented graphically.It was seen from the graph that efficiency of the heat exchanger falls with rise in
̇
̇
. In other
words it can be stated that rise in efficiency requires faster increase in flow rate of the hot stream than of the cold stream. Also, It
was seen that the overall heat transfer coefficient approximately falls with rise in
̇
̇
. It can also be stated that rise in overall heat
transfer coefficient requires faster increase in flow rate of the hot stream than of the cold stream. There is variation of convective
heat transfer coefficient with respect to mass flow rate. Also the convective heat transfer coefficient increases with both Reynolds
and nusselt numbers, which increases the overall heat transfer coefficient.
0.00935
0.00936
0.00937
0.00938
0.00939
ConvectiveHeat
Transfer
Coefficient(Hh)
Hot Nusselt Number (NUh)
0
2000
4000
6000
0.0215 0.016 0.0259 0.0312 0.0125 0.0279 0.0329 0.0201 0.0203 0.0301
OverallHeattransfer
Coefficient(W/m2K)
Mass Flow Rate (Kg/s)

_________________________________________________________________________________________________________
ACKNOWLEDGMENT
We wish to acknowledge the efforts and contributions of the chancellors of Landmark University Omu-Aran, Kwara State,
Bishop David Oyedepo (Ph.D) and Afe-Babalola University, Ado-Ekiti, Ekiti State, Afe Babalola (SAN) for their commitment in
human capital development via procurement of laboratory equipment and training of their staffs which is evidence in this work.
We will forever remain indebted to them. To God alone be all the glory.
REFERENCES
[1]. Chi-Chuan W, Chang-Tsair C. (2012): Heat and mass transfer for plate fin-and-tube heat exchangers, with and without
hydrophilic coating. International Journal of Heat and Mass Transfer, Volume 41, Issue 20, Pages 3109-3120.Retrieved 6th
September,2014.
[2]. Shah R.K and Kandilkar S. G (1989): “The influence of the number of thermal plates on plate heat exchanger
performance”, Journal of Heat Transfer, vol.111, pp.300-313.
[3]. Ho-Ming Yeh, (2010)Effect of External Recycle on the Performance in Parallel-Flow Rectangular Heat-Exchangers,
Tamkang Journal of Science and Engineering, 13 ( 4) 405-412
[4]. Kevin M. L (1998): “Increasing Heat Exchanger Performance”, Bryan Research and Engineering, Inc. - Technical Papers
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Dairy Industry”, Research Journal of Engineering Sciences, Vol. 2(1), 5-10,
[7]. Murugesan M.P. and Balasubramanian R., “The Experimental Study on Enhanged heat Transfer Performance in Plate Type
Heat Exchanger”, Research Journal of Engineering Sciences, Vol. 2(2), 16-22,
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[9]. Sanvicente E. et al (2012): Transitional Natural Convection Flow and Heat Transfer in an Open Channel: International
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[10]. Sai K T., Rajasekhar S. G and Pravarakhya G (2013): “Design and Analysis of Plate Heat Exchanger with CO and R134a
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