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)

150

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

151

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

152

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

153

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

156

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

164

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

165

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

173

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