Post on 09-Feb-2018
146
Chapter 5
COMPACT HEAT EXCHNAGER ANALYSIS USING NANOFLUIDS
The thermal analysis of compact heat exchanger [9] is done by
considering the fluid outlet temperatures or heat transfer rate as
dependent variables, and is related to independent parameters as
follows
parameters&iablesvartindependen
controls'designerunderparameters
iablesvarconditionsoperating
hci,ci,h
iablesvardependent
o,co,h ntarranagemeflow,A,U,C,C,T,TfqorT,T
(5.1)
Six independent and one variable which may be Th,o, Tc,o or q
dependent variable of Equation (5.1) for a given flow arrangement
can be transferred into two independent and one dependent
dimensionless groups.
Figure 5.1 Nomenclature of Heat Exchanger
(source:Shah RK [9])
147
To get the dimensionless groups a two fluid exchanger shown in
Fig. 5.1 is considered. By combining differential energy
conservation equations for the control volume we get
cchh dTCdTCdAqdq (5.2)
Where the sign depends upon whether dTc is increasing or
decreasing with increasing dA or dx.
The local overall heat transfer rate equation on a differential base
for the surface area dA is
dATUdATTUdAqdqlocalch
(5.3)
Integration of Equations (5.2) and (5.3) across the exchanger
surface area we get
i,co,cco,hi,hh TTCTTCq (5.4)
o
mm R
TTUAq
(5.5)
where Tm is the actual mean temperature difference (or MTD) that
depends upon the exchanger flow arrangement and degree of fluid
mixing within each fluid stream. The inverse of the overall thermal
conductance UA is referred to as the overall thermal resistance (Ro)
is represented as
cocso
w
hsohohA
1
Ah
1R
Ah
1
hA
1
UA
1
(5.6)
The wall thermal resistance Rw in Equation (5.6) is given by
148
walllayermultipleawithtubecircularafork
d
dln
LN2
1
walllayerglesinawithtubecircularaforLNk2
dd
ln
wallflataforkA
R
j j,w
j
1j
t
tw
f
o
ww
w
(5.7)
5.1 -NTU , P-NTU and MTD Methods
Three different methods are shown in Table 5.1 based on the
choice of three dimensionless groups. The relationship among
three dimensionless groups is derived by integrating Equations
(5.2) and (5.3) across the surface area for a specified exchanger
flow arrangement.
In the -NTU method[9], the heat transfer rate from the hot fluid to
the cold fluid in the exchanger is expressed as
i,ci,hmin TTCq (5.8)
The exchanger effectiveness (0 1) is a ratio of the actual heat
transfer rate from the hot fluid to the cold fluid to the maximum
possible heat transfer rate qmax thermodynamically permitted. The
qmax is obtained in a counterflow heat exchanger of infinite surface
area operating with the fluid flow rates and fluid inlet temperatures
149
equal to those of an actual exchanger. The exchanger effectiveness
is a function of NTU and C* in this method.
The number of transfer units NTU (0 NTU ) is a ratio of the
overall conductance UA to the smaller heat capacity rate Cmin. NTU
designates the dimensionless heat transfer size or thermal size
of the exchanger. The heat capacity rate ratio C* (0 C* 1) is
simply a ratio of the smaller to the larger heat capacity rate for the
two fluid streams.
The P-NTU method[8] represents a variant of the -NTU method.
The -NTU relationship is different depending upon whether the
shell fluid is the Cmin or Cmax fluid in the (stream asymmetric) flow
arrangements commonly used for shell-and-tube exchangers. In
order to avoid possible errors and to avoid keeping track of the Cmin
fluid side, the temperature effectiveness P is taken as a function of
NTU and R.
i,1i,222i,2i,111 TTCPTTCPq
112221 RPPRPP
112221 RNTUNTURNTUNTU
2
1 R1R (5.9)
In the MTD method[8], the heat transfer rate from the hot fluid to
the cold fluid in the exchanger is given by
mm TUAFTUAq (5.10)
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where Tm the log-mean temperature difference (LMTD), and F the
LMTD correction factor, a ratio of actual MTD to the LMTD, where
2
1
21
TT
ln
TTLMTD (5.11)
Where 1T and 2T are defined as
i,co,h2o,ci,h1 TTTTTT for all flow arrangements
except for parallel flow
o,co,h2i,ci,h1 TTTTTT for parallel flow
Table 5.1 General Functional Relationships and Dimensionless
Groups for -NTU, P-NTU, and MTD Methods (source : Shah [9])
Generally -NTU method is used by automotive, aircraft, air
conditioning, refrigeration and other industries that design or
manufacture compact heat exchangers. The MTD method is used
by process, power, and petrochemical industries that design or
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manufacture shell and tube and other noncompact heat
exchangers.
5.2 Fin Efficiency and Overall Efficiency
Extended surfaces have fins attached to the primary surface on
one or both sides of a two-fluid or a multi-fluid heat exchanger.
Fins can be of a variety of geometries plain, wavy, or interrupted
and can be attached to the inside, outside, or both sides of
circular, flat, or oval tubes, or parting sheets.
The concept of fin efficiency accounts for the reduction in
temperature potential between the fin and the ambient fluid due to
conduction along the fin and convection from or to the fin surface
depending upon the fin cooling or heating situation. The fin
efficiency is defined as the ratio of the actual heat transfer rate
through the fin base divided by the maximum possible heat
transfer rate through the fin base which would be obtained if the
entire fin were at the base temperature. Since most of the real fins
are thin, they are treated as one-dimensional (1-D) with standard
idealizations used for the analysis (Huang and Shah [177] ). This
1-D fin efficiency is a function of the fin geometry, fin material
thermal conductivity, heat transfer coefficient at the fin surface,
and the fin tip boundary condition; it is not a function of the fin
base or fin tip temperature, ambient temperature, and heat flux at
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the fin base or fin tip in general. Fin efficiency formulas for some
common fins are presented in Table 5.2.
Figure 5.2 A flat fin over (a) an in-line and (b) staggered tube
arrangement; the smallest representative segment of the fin
for (c) an in-line and (d) a staggered tube arrangement.(source:
Shah, [178])
The fin efficiency for flat fins (Figure 5.2b) is obtained by a sector
method [178]. In this method, the rectangular or hexagonal fin
around the tube or its smallest symmetrical section is divided into
n sectors (Figure 5.2). Each sector is then considered as a circular
fin with the radius re,i equal to the length of the centerline of the
sector. The fin efficiency of each sector is subsequently computed
using the circular fin formula of Table. 5.2. The fin efficiency f for
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the whole fin is then the surface area weighted average of f,i of
each sector
n
1i
i,f
n
0i
i,fi,f
f
A
A
(5.12)
In an extended-surface heat exchanger, heat transfer takes place
from both the fins (f < 100%) and the primary surface (f = 100%).
Therefore extended surface efficiency o is defined as
fff
f
p
o 1A
A1
A
A
A
A (5.13)
where Af is the fin surface area, Ap is the primary surface area, and
A = Af + Ap.
5.3 Heat Exchanger Pressure Drop
Pressure drop in heat exchangers is an important consideration
during the design stage. Since fluid circulation requires some form
of pump or fan. Pressure drop calculations are required for both
fluid streams, and in most cases flow consists of either two
internal streams or an internal and external stream. Pressure drop
is affected by a number of factors, namely the type of flow (laminar
or turbulent) and the passage geometry. Usually a fan, blower, or
pump is used to flow fluid through individual sides of a heat
exchanger. Due to potential initial and operating high cost, low
154
Table 5.2 Fin Efficiency expressions for Plate-Fin and Tube-Fin Geometrics (source : Shah [178])
155
fluid pumping power requirement is highly desired for gases and
viscous liquids. The fluid pumping power is approximately
related to the core pressure drop (Shah, [178]) in the exchanger as
flowTurbulentforADg2
mL4046.0
flowarminlaforRefDg2
L4
pm
8.1
o
2.1
h
2
c
8.22.0
h
2
c
(5.14)
From Equation (5.14) that the fluid pumping power is strongly
dependent upon the fluid density ( 1/2) particularly for low-
density fluids in laminar and turbulent flows, and upon the
viscosity in laminar flow. Pressure drop can be an important
consideration when blowers and pumps are used for the fluid flow
since they are head limited. Also for condensing and evaporating
fluids, the pressure drop affects the heat transfer rate. Hence, the
pressure drop determination in the exchanger is important.
Miller [179] suggested that core pressure drop may consist of one
or more of the following components depending upon the
exchanger construction (1) friction losses associated with fluid flow
over heat transfer surface (this usually consists of skin friction,
form (profile) drag, and internal contractions and expansions) (2)
the momentum effect (pressure drop or rise due to fluid density
changes) in the core (3) pressure drop associated with sudden
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contraction and expansion at the core inlet and outlet and (4) the
gravity effect due to the change in elevation between the inlet and
outlet of the exchanger. The gravity effect is generally negligible for
gases. The total pressure drop across the heat exchanger core is
obtained by taking the sum of all of these contributions is
eaci ppppp (5.15)
Combining all of the effects and rearranging, yields the following
general expression for predicting pressure drop in a heat
exchanger core
e
ie
2
e
e
i
m
i
h
e
2
i
i
2
K112D
L4fK1
2
Gp
(5.16)
The efficiency accounts for the irreversibilities in the pump, i.e.
friction losses. It is clear that a Reynolds number dependency
exists for the expansion and contraction loss coefficients. For
design purposes we may approximate the behavior of these losses
by merely considering the Re = curves. These curves have the
following approximate equations [8] are
2e 1K
22e 142.0K (5.17) The core frictional pressure drop in Equation (5.16) may be
approximated as
mhc
2 1
Dg2
fLG4p
(5.18)
157
Where lmavgm
Tp
R~
1
Here is the gas constant in R~
J/kg K, pavg = (pi + po)/2, and Tlm=
Tconst + Tlm where Tconst is the mean average temperature of the
fluid on the other side of the exchanger; the LMTD Tlm
5.4 Problem Formulation
The configuration of the radiator for the present analysis is chosen
from Charyulu et al. [159] for the purpose of establishing the
advantages of using nanofluids. The details of the radiator(Fig 5.3)
mounted on turbo charged diesel engine of type TBD 232 V-12
cross flow compact heat exchanger with unmixed fluids consisting
of 644 tubes made of brass and 346 continuous fins made of
copper are presented in table 5.3- 5.4.
The fluid parameters and normal operating conditions as
presented by Charyulu et al. [159] are given in Table 5.5. A
numerical method -NTU rating method can be applied for this
compact heat exchanger by replacing the standard coolant water+
50% ethylene glycol by Al2O3 + water nanofluids.
5.5 Compact Heat Exchanger Geometry
The core of compact heat exchanger with its principle components
coolant tubes and fins are shown in Fig 5.4. Flat tubes are more
popular for automotive applications due to their lower profile drag
158
compared with round tubes. The directions of the coolant and air
flows across each other are represented in Fig.5.4. The ultimate
Figure 5.3 Schematic of Radiator Assembly
design object of the heat exchanger is to maximize the heat
rejection rate while minimizing the flow resistance. We have
considered compact heat exchanger surface 11.32-0.737-SR (Kays
et al[8]) and the surface geometry parameters are given in Table
5.3.
Figure 5.4 Structure of typical Compact Heat Exchanger core
159
Table 5.3 Surface core geometry of flat tubes, continuous fins (surface 11.32-0.737-SR, Kays et al. ,[8])
S.No Description Air side Coolant side
1. Fin pitch 4.46 fin/cm
2. Fin Metal Thickness 0.01 cm
3 Hydraulic diameter, Dh 0.351cm 0.373cm
4. Min free flow area / Frontal area ,
0.780 0.129
5. Total heat transfer area / Total volume,
886 m2/m3
138 m2/m3
6. Fin area / Total area, 0.845
Table 5.4 Compact heat exchanger Geometric factors
Factor and Symbol
Description
A The free flow area on one side of the exchanger Af The frontal area on one side of the exchanger. It is
given as product of the overall exchanger width and height or depth and height
a The separation plate thickness b The separation plate spacing. de The equivalent diameter used to correlate flow friction
and heat transfer and is four times the hydraulic radius rh
L The flow length on one side of the exchanger P The perimeter of the passage p The porosity which is the ratio of the exchanger void
volume to the total exchanger volume. rh The hydraulic radius, which is the ratio of the
passage flow area to its wetted perimeter. S The heat transfer surface on one side of the
exchanger Sf The surface of the fins, only , on one side of the
exchanger. V The total exchanger volume, it is given as the product
of width,dpth and height. The ratio of total surface area on one side of the
exchanger to the total volume on both sides of the exchanger.
The ratio of the total surface area to the total volume on one side of the exchanger.
The fine thickness The fine efficiency 0 The overall passage efficiency The ratio of the free flow area to the frontal area on
one side of the exchanger.
160
Table 5.5 Fluid parameters and Normal Operating conditions
S.No Description Air
Coolant
1. Fluid mass rate 8 -20 kg/s 6000-10000 kg/hr
2. Fluid inlet Temperature 20-55 0C 70-95 0C
3 Fluid Temperature rise/ drop
28 0C 6 0C
4 5. 6. 7.
Core width Core height Core depth Tube size
0.6 m 0.5m 0.4m 1.872cm * .245cm
5.6 Thermal Analysis Procedure
Shah et al [9] gave a step-by-step procedure for thermal analysis
(rating) of a compact cross-flow heat exchanger. Inputs are the
exchanger construction, flow arrangement and overall dimensions,
complete details on the materials and surface geometries on both
sides including their non-dimensional heat transfer and pressure
drop characteristics (h and f vs. Re), fluid flow rates, inlet
temperatures, and fouling factors. The fluid outlet temperatures,
total heat transfer rate, and pressure drops on each side of the
exchanger are then determined.
Step 1.
Determine the surface geometric properties on each fluid side ie.
the minimum free flow area Ao, heat transfer surface area A, flow
lengths L, hydraulic diameter Dh, heat transfer surface area
density , the ratio of minimum free-flow area to frontal area , fin
161
length lf, and fin thickness for fin efficiency determination, and
any specialized dimensions used for heat transfer and pressure
drop correlations. Considering core surface as 11.32-0.737-SR,
geometric properties are as listed in table 5.3.
Step 2.
Compute the fluid bulk mean temperature and fluid
thermophysical properties on each fluid side. Since the outlet
temperatures are not known, they are estimated initially. Assume
an exchanger effectiveness as 60 to 75% for most single-pass
crossflow exchangers, or 80 to 85% for single-pass counter flow
exchangers, and calculate the fluid outlet temperatures.
i,ci,hhmini,ho,h TTC/CTT (5.19)
icihcicoc TTCCTT ,,min,, / (5.20)
The bulk mean temperatures on each fluid side will be the
arithmetic mean of the inlet and outlet temperatures on each fluid
side is calculated. Once the bulk mean temperature is obtained on
each fluid side, obtain the fluid properties, which are , cp, k, Pr,
and .
With this cp, one more iteration may be carried out to determine
Th,o or Tc,o from Equation (5.19) or (5.20) on the Cmax side, and
subsequently Tm on the Cmax side, and refine fluid properties
accordingly.
162
Step 3.
Calculate the Reynolds number Re and heat transfer and flow
friction characteristics of heat transfer surfaces on each side of the
exchanger. And compute j or Nu and f factors. Correct Nu (or j) for
variable fluid property effects (Shah, [180]) in the second and
subsequent iterations from the following equations.
m
m
w
cr
n
m
w
cr T
T
f
f
T
T
Nu
Nu
for gases
m
m
w
cr
n
m
w
cr f
f
Nu
Nu
for liquids (5.21)
where the subscript cr denotes constant properties, and m and n are empirical constants provided in Table 5.6 (a&b).
Step 4.
Compute the heat transfer coefficients for both fluids; determine
the fin efficiency f and the extended surface efficiency o, finally,
compute the overall thermal conductance UA from Equation (5.6).
Step 5.
From the known heat capacity rates on each fluid side, compute C*
= Cmin/Cmax. From the known UA, calculate NTU = UA/Cmin. With
NTU, C*, and the flow arrangement, determine the exchanger
effectiveness from the Table 5.7
163
Table 5.6 (a) Property Ratio method exponents for Laminar flow
(Shah Rk et al [180])
Fluid
Heating Cooling
Gases 1,0 mn
for 1 < Tw/Tm < 3
81.0,0 mn
for 0.5 < Tw/Tm < 1
Liquids 58.0,14.0 mn
for w/m < 1
54.0,14.0 mn
for w/m > 1
Table 5.6(b) Property Ratio method exponents for Turbulent
flow
( Shah RK et.al [180] )
Fluid
Heating Cooling
Gases 3.0T/Tlogn 4/1mw10
for 1 < Tw/Tm < 5, 0.6 < Pr< 0.9
104
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Step 2 and continue iterating Steps 2 to 6, until the assumed and
computed outlet temperatures converge within the desired degree
of accuracy.
Step 7.
Compute the heat transfer rate from
inainc TTCQ ,,min (5.22)
Step 8
Compute the pressure drop on both side of the exchanger by using
the equation (5.16)
Table 5.7 -NTU relationship for cross-flow unmixed configuration (shah [9])
Number of tube
rows
Side of
Cmin
Formula
air *)1( /1 * Ce NTUeC 1 tube ** /)1(1 CeC NTUe air 2/*2*2 1,/1(1 * NTUKC eKCKCe 2 tube 2/*2/2 ** 1,)/(1(1 CNTUCK eKCKe air 3/*42*2*3 ** 1,/))2/)(3()3(1(1 CNTUKC eKCKCKKCe
3 tube 3/*42*4*2/3 ** 1,/))2/)(2/3(/)3(1(1 CNTUCK eKCKCKCKKe
-
1exp
1exp1 78.0*22.0
*NTUCNTU
C
5.6.1 Equation used in Air side
(1) The air side heat transfer coefficient for the specified core side
geometry as presented by (Charyulu et al., [159]), ha
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3/2Pra
aaa
a
CpGjh (5.23)
Where 383.0Re
174.0
a
aj
afr
aA
WG
(5.24)
a
aha
a
DG
,Re (5.25)
(2) Fin efficiency of plate fin can be calculated as
mL
mLtanhf where
tk
h2m
fin
a (5.26)
The overall fin efficiency is determined by
fff
f
p
o 1A
A1
A
A
A
A (5.27)
(3) Pressure drop for fin side as given by Kays et al.[8]
e
ie
2
e
e
i
m
i
h
e
2
i
i
2
K112D
L4fK1
2
Gp
(5.28)
where
aoaim ,,
11
2
11
Friction factor f is given by
3565.0Re
3778.0
a
f (5.29)
(4) Air heat capacity rate , Ca
aaa CpmC (5.30)
166
5.6.2 Equations used Nanofluid side
(1) A correlation to determine the heat transfer coefficient of the
Al2O3 + H2O nanofluid in turbulent flow (Eqs. 5.31) has been
developed in previous studies by Vasu et al. ([170]-[172]). The
correlation is found in good agreement with the experimental data
with standard Deviation of 6.4% and Average deviation of 5% (see
section 4.2).
nfh
nfnf
nfD
KNuh
,
(5.31)
Where 4.08.0 PrRe0256.0 nfnfnfNu for Al2O3 + H2O
nf
nfh
nf
Du
,maxRe (5.32)
nf
nfnf
nfk
CpPr (5.33)
2324.0
05.0175.0Re
f
p
m
f
nf
k
k
k
k for Al2O3 + H2O (5.34)
The eq.(5.34) is used to calculate Thermal conductivity for
nanofluids (Vasu et al.,[170]) which is found to be in good
agreement with the experimental data with standard deviation of
4% and Average deviation of 2% (see section 3.2).
The other properties viscosity, density and specific heat of
Nanofluids are calculated by using the following equations (see
section 2.4,2.4 & 2.6)
167
)9.53311.391( 2fnf (5.35)
pfnf )(1 (5.36)
nf
ppf
nf
CpCp
f)-1(Cp (5.37)
(2) Pressure drop is given as (Kays et al.[8]).
nfhnf
nfnf
cD
HfGP
,
2
(5.38)
Where
25.0Re079.0 nfnff (5.39)
(3) Coolant heat capacity rate, Cnf
nfnfnf CpmC (5.40)
The Heat exchanger effectiveness for cross flow unmixed fluids, is
as given by Kays et al.[8] (see table 5.7)
1exp
1exp1 78.0*22.0
*NTUCNTU
C (5.41)
Where
nf
a
C
CC *
a
aa
C
AUNTU (5.42)
Overall heat transfer coefficient, based on air side is given as
nf
a
nfaa hhU
111 (5.43)
168
Figure 5.5 The overall heat transfer coefficient
Total heat transfer rate
inainc TTCQ ,,min (5.44)
For evaluation of various parameters used in the analysis of
compact heat exchanger a M.file in MATLAB is developed. This is
useful in estimating the fluid properties at different operating
temperatures, different surface core geometry of cross flow heat
exchanger and heat transfer coefficients. Pressure drops, overall
heat transfer coefficients and heat transfer rate are also estimated.
The flowchart of the numerical analysis is shown in Fig.5.6.
5.7 Results and Discussions
The numerical results thus obtained are presented in graphical
form through fig 5.7 to 5.14. Figure 5.7 & 5.8 indicates that
nanofluids possess higher heat transfer characteristics than
169
conventional coolant water and 50% ethylene glycol. The overall
heat transfer coefficient for nanofluids is found higher than water
and increasing with increase of the volume fraction of
nanoparticles.
5.7.1 Effect of Air inlet Temperature
One of the most important factors governing the performance of an
automotive radiator system is the air inlet temperature. At different
air inlet temperatures the cooling capacity and overall heat
transfer coefficient of the radiator is found and is shown in Fig.
5.9- 5.10 for the two limiting air flows (12kg/sec and 6 kg/sec) for
a range of temperature from 00C to 500C. As expected the heat
transfer rate clearly decreases with air inlet temperature rise, as
the cooling temperature difference is being reduced. It is
interesting to point out that the Al2O3+H2O nanofluids is posing
higher cooling capacity than that of water as coolant. The influence
of air inlet temperature on the overall heat transfer coefficient is
very small whereas the air pressure drop is found moderate.
170
Figure 5.6 Schematic of the Numerical method
171
5.7.2 Effect of Air and Coolant mass flow rate
The cooling capacity of the compact heat exchanger is strongly
dependent on both fluids mass flow rate. Figure 5.11 & 5.12 shows
the heat transfer characteristic of the selected radiator over a wide
range, by fixing the geometry and temperature levels at the normal
situation. It is observed that cooling capacity is increasing with
both air and coolant flow rates. The cooling capacity is more with
increasing air flow rate. The pressure drop also increases
quadratically with both air and coolant mass flow rates and is
found almost same for all flow rate of air(6-12 kg/sec) and
coolant(6000- 10000 kg/hr). It is observed that the cooling
capacity and overall heat transfer coefficient of the radiator is very
high when Al2O3 + H2O nanofluids is used as coolant as shown in
Fig. 5.7 & 5.8. However the particle concentration in the fluid
causes more pressure drop when compared to conventional
coolants.
5.7.3 Effect of Coolant inlet Temperature
The heat transfer characteristics of radiator also depend as the
coolant inlet temperature. It is observed from the Fig. 5.13 that
with increases of the coolant inlet temperature the cooling capacity
is increase. From the Fig. 5.13 it is evident that the cooling
capacity of 4 vol% Al2O3 + H2O nanofluids is very high when
172
compared with water as coolant. However, the pressure drop are
nearly double than that of water.
5.7.4 Effect of Nanoparticle volume fraction
Figure 5.14 indicates the effect of nanoparticle concentration in
fluid. With increase of the volume fraction of the nanoparticle
concentration the cooling capacity and pressure drop increases in
a moderate manner. Further at given concentration the pressure
drop decreases with coolant inlet temperature.
5.8 Conclusion
In this chapter a detailed study of thermal and hydraulic behavior
of compact heat exchanger using Al2O3 + water nanofluid is given.
The calculations have been carried out by well verified and
validated detailed rating and design of compact heat exchanger
model using -NTU method. A detailed comparative analysis was
carried out by considering different parameters like flow rate and
inlet temperature of both fluids using standard coolant (water +
50% Ethylene glycol) and Al2O3 + water nanofluid on turbo
charged diesel engine of type TBD 232 V-12 radiator and are
graphically presented. It is observed that the cooling capacity has
been increased by 15 20 % with different volume fractions of the
nanoparticles will keeping the pressure drop constant. Therefore
for a given cooling capacity (say 400kw) the amount of Nanofluids
flow rate decreases and which can leads to the pumping power of
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coolant has been reduced by 75% when 4vol% Al2O3 + water is
used as coolant.
Similarly due to high heat transfer coefficients of nanofluids
compared with base fluids, it is possible to reduce the surface area
of the air side while transferring the same heat capacity. It is
observed that for a given cooling capacity (say 400kW), 5%
reduction in the air side surface area can be obtained by using
4vol% Al2O3 + water is used as coolant. Reduction in radiators size
can lead to 5% reduction on the aerodynamic drag coefficient. With
reduction of the aerodynamic drag we can achieve significant
increase in fuel efficiency.
174
Figure 5.7 Comparison of Nanofluid as coolant with conventional coolant (water)
175
Figure 5.8 Comparison of Nanofluid as coolant with conventional coolant (water)
176
Figure 5.9 Air inlet Temperature influence on the cooling capacity
177
Figure 5.10 Air inlet Temperature influence on the overall heat transfer coefficient
178
Figure 5.11 Air flow influence on cooling capacity & Pressure drop
179
Figure 5.12 Coolant flow influence on Cooling capacity & Pressure drop
180
Figure 5.13 Coolant inlet Temperature influence on Cooling Capacity
181
.
Figure 5.14 Coolant inlet Temperature and volume fraction of Nanoparticle on cooling Capacity
and Pressure drop