Quantification of energy losses and performance improvement in dx cooling by exergy method

International Journal of Mechanical Engineering and Technology (IJMET), ISSN

0976 – 6340(Print), ISSN 0976 – 6359(Online) Volume 3, Issue 3, Sep- Dec (2012) ©

IAEME

137

QUANTIFICATION OF ENERGY LOSSES AND PERFORMANCE

IMPROVEMENT IN DX COOLING BY EXERGY METHOD

Dinkar V. Ghewade*1, Dr S.N.Sapali2

*Department of Mechanical Engineering, Genesis Institute of Technology, Kolhapur 416234 India

Professor in Mechanical Engineering, Govt. College of Engineering, Pune 411005 India; E-mail:

[email protected]

ABSTRACT

Direct expansion bulk milk cooling and storage tanks are found commonly in milk chilling

centers as well as in large dairy farms. These systems are used to pull down the milk

temperature from 35oC to 4

oC in 3 to 3.5 hours. This duration is excess to maintain the

quality of the milk at its original. Further, the energy consumed by bulk milk cooler is

comparatively higher, demanding the performance analysis. The refrigeration system used

for this purpose consists of standard components available in the market. Even though

these components are designed for the best individual performance, the performance of a

plant as a whole is required to be studied. The first law efficiency of the plant is higher,

but the second law efficiency is found to be low. Exergy analysis is used as a tool to

evaluate the performance of the system. Exergy flows in the system are experimentally

studied to identify and quantify exergy destruction in all components of the system. Based

on the findings, certain design changes are made in the evaporator of the new system and

tested for validation. The contributing components to exergy destruction are: (i)

compressor, (ii) condenser, (iii) evaporator and (iv) expansion valve, in decreasing order.

It is found that coefficient of performance (COP) of the new system (model) is improved

by 0.6 to 0.8 and irreversibilities in compressor, condenser and evaporator are reduced

significantly. Marginal improvement in second law efficiency of the new model is

recorded along with the saving in energy consumption rate of 0.6 - 0.8 kW. The

improvement potential in each component is determined and the scheme to achieve the

improvement is discussed.

Keywords – Exergy analysis; Exergy efficiency; Bulk Milk Cooler; Improvement

potential; Thermodynamic Analysis

INTERNATIONAL JOURNAL OF MECHANICAL ENGINEERING AND TECHNOLOGY (IJMET)

ISSN 0976 – 6340 (Print)

ISSN 0976 – 6359 (Online)

Volume 3, Issue 3, Septmebr - December (2012), pp. 137-149

© IAEME: www.iaeme.com/ijmet.html

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International Journal of Mechanical Engineering and Technology (IJMET), ISSN 0976 –

6340(Print), ISSN 0976 – 6359(Online) Volume 3, Issue 3, Sep- Dec (2012) © IAEME

138

1. INTRODUCTION

Milk chilling is the primary and one of the important processes in maintaining the good

quality of milk. The temperature of the milk at its harvest is 35oC and the bacteria count is in

the range 10000 to 25000 per ml depending upon hygiene conditions at workplace. If the

milk is not chilled within half an hour from the harvest to a temperature of 4oC, the bacteria

count increases at a faster rate. The rise in bacteria count increases the pasteurization

temperature and decreases the quality of the milk. Hence Bulk Milk Cooler (BMC) is used to

chill and store the milk at large dairy farms, chilling centers, and milk collection centers. The

size of the storage tank depends upon whether the BMC is used for two, four or six milking

conditions. BMC for two milking conditions are the most widely used and are the focus of

the present study. In the second milking condition, milk is collected by milk processing plants

once in 24 hours from dairy farms. Milk harvested in the morning (first milking) is poured in

the BMC to half of its capacity and is chilled from 35oC to 4

oC. After second milking, the

fresh raw milk at 35oC is mixed with the chilled milk and the tank is completely filled to its

capacity. This raises the temperature of milk in the tank to 19oC. The milk is further cooled to

4oC and stored till it is transported to milk processing plant.

The energy performance of the refrigeration systems is usually evaluated based on first law of

thermodynamics. However compared to energy analysis, exergy analysis can better and

accurately show the location of inefficiencies in the refrigeration system. The results of

exergy analysis can be used to assess and optimize the performance of the system. Exergy is

defined as the maximum useful work that can be obtained from the system at a given state

with respect to a reference environment (i.e. dead state) (Kotas, 1985). The total amount of

exergy is not conserved in a process or a system, but destroyed due to irreversibilities (Kotas,

1985).

BMCs are designed and used for chilling the milk in standard duration of three hours as

specified by ISO5708. BMCs are classified based on cooling time as class I, class II, class

III, and class IV. In the present study, class II BMC stipulated to chill the milk in 3.0 hours

when the tank is half-filled is analyzed for its performance.

According to the laboratory test reports of BMC obtained from different manufacturers, the

coefficient of performance (COP) of these bulk milk coolers, over its operational time is

found to be in the range of 1.95 - 2.5. A field survey was conducted by the authors to study

the conditions in which the BMCs are used. From the survey, it is found that the performance

of the BMC further declines when operating in field conditions. Due to low operational

efficiency, the bulk milk coolers are not economic for use. Another finding obtained from the

survey is that the factors like size, low operation and maintenance cost, low initial cost,

efficient heat transfer, and easy cleaning are very important in optimizing the performance of

BMC.

In an effort to understand and identify the inefficiencies in the equipment and the process,

exergy analysis of the BMC is done and exergy efficiency of each component is determined.

Energy consumption and overall performance of the BMC is major concern and needs a

scientific study to improve its energy efficiency. Exergy analysis tool is the appropriate

technique to understand the system behavior and to locate inefficiencies. Important

parameters viz. exergy destruction, coefficient of performance, work input, exergy efficiency,

second law efficiency are determined to evaluate and compare the system performance. The

work consists of two parts as mentioned below:



139

1. The testing of existing BMC system (termed as the old model) is done and its

performance in terms of COP, exergy destruction, work input, second law efficiency

is evaluated.

2. New model is developed, based on findings from the analysis of the old model, with

some design changes in evaporator and the performance is measured as above.

Results of both the models are presented in this paper, wherein it is observed that design

changes based on exergy analysis lead to improvement in performance of the system. This

paper analyzes the performance of BMC with respect to important parameters as mentioned

above in point no. (1). An attempt is made to explain the nature of irreversibilities and

practical limits to their reduction.

2. LITERATURE REVIEW

The theory of exergy analysis is discussed at length by Kotas (1985) and is applied in thermal

and chemical plant analysis by many researchers. The refrigeration systems used as heat

pumps with R22 as a refrigerant are analyzed for exergy loss by Hepbasali (2005). Entropy

generation in thermodynamic process causes exergy destruction, which is a cause of low COP

and consequently high energy consumption. It is necessary to identify, locate and quantify the

irreversibilities to improve energy efficiency of refrigeration systems. The compressor

performance is analysed using exergy method by McGovern(1995) and the refrigerant flow in

the evaporator coils and air cooled condenser coils is analysed for various mass flow

densities, inlet temperatures and tube lengths by Liang (2001). The heat transfer coefficient is

found to be low in low vapour quality two phase flow region and high in high vapour quality

two phase flow region. Ratio of irreversibility rate with augmented heat transfer in a tube to

irreversibility rate in heat transfer in smooth tube for turbulent flow is studied by Bali (2008)

and Wang (2003). Irreversibility rate depends on Reynolds number (Re) and increases with it.

Rate of increase in ratio of irreversibilities deceases towards higher Re. Non-dimensional

number of irreversibility and non-dimensional irreversibility balance is defined by Pons

(2004), and the entropy generation numbers are defined by A. Bejan (1982). The above-

mentioned studies provide an adequate framework for setting the research experiment and

analysing the refrigeration system in the present study.

3. METHODOLOGY

The methodology adopted in this study consists of two parts: (i) the experimentation scheme, and (ii) the

analysis scheme. This methodology is explained in brief in the following paragraphs.

3.1 Experimentation Scheme

The experiment is designed on two types of BMC: the old model and the new model. The old

model refers to the existing system, while the new model refers to the modified BMC design.

Refrigeration system of Bulk milk cooler consists of compressor, air cooled condenser,

receiver, thermostatic expansion valve and dimple type (jacketed) evaporator. Standard four

row air cooled condensers are used which consist of grooved copper tube of an outer

diameter of 9.525 mm, and fin density of 14 fins per inch. Evaporator, divided into two equal

parts, is at the base of the tank, which cools the milk by direct expansion of the refrigerant.

R22 is used as the refrigerant. Block diagram of 1000 liter BMC plant is shown in Figure-1.

The BMC plant consists of two refrigeration units each for half part of the evaporator. These

two units operate simultaneously to chill the milk to required temperature. Thermostatic

expansion valve is used to regulate the mass flow rate with respect to evaporator exit

temperature of refrigerant.



140

Evaporator

I1

I2

E1

E2

Compressor

Compressor

Condenser

Condenser

Expansion Valve

Expansion Valve

Evaporator

1

23

45

6

7

8

condensing unit 1

condensing unit 2

Pressure measurement points-1,2,7& 8

Temperature measurement points- 1 to 8

1-compressor inlet; 2- compressor exit; 3- condenser inlet; 4- condenser exit

5-expansion valve inlet; 6-expansion valve exit; 7-evaporator inlet; 8-evaporator exit,

I1,I2- inlets to evaporator

E1,E2-exits from evaporator Figure-1: Schematic diagram of Bulk Milk Cooler 1000 liter capacity (the old model)

The refrigerant after leaving the expansion valve enters evaporator through inlet I1 and I2,

and flows through the passage formed by a seam weld in the evaporator towards the exit E1

and E2. In existing system (the old model) there is one inlet and one exit for the refrigerant.

Experiments are done on chilling of water instead of milk as the physical properties of milk

are very similar to that of water [12]. Physically milk is a rather dilute emulsion combined

with colloidal dispersion in which the continuous phase is the solution. Table no 1 gives the

properties of water and milk at 20oC.

Table No. 1: Properties of water and milk at 20oc.

Sr. No. Name of Property Water Milk

1. Specific Gravity 1.0 1.0321

2. Specific Heat 4.183 kJ/kg 3.9315 kJ/kg

3 Thermal Conductivity 0.599 W/m-K 0.550-0.580 W/m-K

4 Viscosity 1.004x10-3

N-s/m2

2 x10-3

N-s/m2

5 Refractive index 1.3329 1.3440

The findings for the water will equally hold good for the milk without much variations.

Pressures are measured by piezoelectric transducers across the compressor and evaporator

with an accuracy of ±0.01 MPa. The pressure losses in the condenser are insignificant, and

therefore neglected. Temperatures are measured across compressor, condenser, expansion

valve and evaporator by RTD with an accuracy ±0.1oC. Data acquisition system is used to

record the temperatures and pressures at specific intervals at salient points of the cycle over

its operation. Mass flow rate is measured by Coriolis effect mass flow rate meter (in kg/min)

within an accuracy of ±0.2%. Water is used as the chilling medium instead of milk, as both

have the similar properties. For the first milking condition test the tank is half filled (i.e. 500

liter) by water and it is heated to 35oC with the auxiliary heater provided. For the second

milking condition tank is filled to its capacity (1000 liter.) and heated to 19oC. In both the

cases the water is chilled to 4oC. Data of Pressures (4 no.), Temperatures (12 no.), Mass flow

rates and work inputs to compressor are recorded over the cooling period at specific intervals.

Tests are repeated minimum once for each set of parameters and the observations are

confirmed.



141

In the new model, two major design changes are made as shown in Figure-2.

1. Seam weld provided on the evaporator to guide the fluid flow causes large pressure

drop (138 kPa to 35 kPa) in the evaporator resulting in higher irreversibility rate.

Secondly the liquid refrigerant entering in evaporator jacket evaporates immediately

and flows as gas in further portion of evaporator. Since the heat transfer coefficient

for the gas-surface interface is low, evaporator performance is low. Hence in the new

design flow restrictions in evaporator are removed.

2 Instead of one inlet and one exit for the refrigerant at the evaporator, three inlets

and three exits are provided to ensure liquid refrigerant is distributed equally

throughout the jackets at the lower side of the evaporator. Three parallel channels are

employed in the new evaporator (for the each tube in original design) The mass flux

(G) and heat flux for the new evaporator inlet of small diameter (d) tubes would be

different than those of original evaporator inlet of large diameter (D) tubes. For the

same total mass flow rate the number of small diameter tubes (n) replacing large

diameter tube is given by

2

2

d

D

G

Gn

d

D=

I1I2I3

I4I5I6

E1E2E3

E6 E5 E4

Compressor

Compressor

Condenser

Condenser

Expansion Valve

Expansion Valve

Evaporator

Evaporator

1

23

45

6

7

8

1-compressor inlet; 2- compressor exit; 3- condenser inlet; 4- condenser exit

5-expansion valve inlet; 6-expansion valve exit; 7-evaporator inlet; 8-evaporator exit,

I1-I6- inlets to evaporator

E1-E6-exits from evaporator

condensing unit 1

condensing unit 2

Figure-2: Schematic diagram of Bulk Milk Cooler 1000 liter capacity (the new model)

After making the required design changes, the new model is tested for performance by

adopting the same experimental procedure as employed for the old model.

3.2 Analysis scheme

For analyzing the data obtained from the experiments, a technique of “exergy analysis” is

used. Exergy analysis combines the first and the second laws of thermodynamics, and is a

powerful tool for analyzing both the quantity and quality of energy utilization. The maximum

work obtainable from system using environmental parameters as reference state is called

exergy and is expressible in terms of four components: physical exergy, kinetic exergy,

potential exergy and chemical exergy. However the kinetic exergy and potential exergies are

usually neglected and chemical exergy is zero as there is no departure of chemical substances

to environment. Therefore in this analysis physical exergy is only considered and is

calculated. Physical exergy of the material stream can be defined as the maximum work that

can be obtained from when it is taken to physical equilibrium state with the environment.



142

)()(.

ooo ssThhEx −+−= (1)

where h and s are enthalpy and entropy respectively and To is the dead state temperature. The

enthalpy and entropy of the substance have to be evaluated at its pressure and temperature

conditions (P, T) and the pressure and temperature at dead state (P0, T0). For a process 1-2 the

change in exergy is given by:

)()( 1212

.

ssThhxE o −+−=∆ (2)

This change in exergy represents the minimum amount of work to be added or removed to

change from state 1 to state 2 when there is an increase and decrease in internal energy or

enthalpy resulting from change.

General exergy balance can be expressed in rate form as

destoutin xExExE...

=− (3)

Considering control volume at steady state (Fig. 2) the exergy balance can be expressed as

.....

IExExWExEx QoutoutinQinin ++=++ (4)

The exergy analysis is mainly concerned for the calculation of exergy efficiency and lost

work for each unit operation.

The total exergy destruction in a cycle is simply the sum of the exergy destruction in

condenser, compressor, evaporator and expansion valve. The overall exergetic efficiency is

defined as

actual

total

actual

inout

ex

W

I

W

xExE.

.

.

..

1−=−

=η (5)

The energy efficiency is simply the ratio of useful output energy to input energy and is

referred as coefficient of performance (COP) for refrigeration system. In this context the

energy efficiency of BMC unit (COPactual) can be defined as follows:

in

e

actualW

QCOP = (6)

The ideal COP obtained from the carnot cycle is given as,

ec

eideal

TT

TCOP

−= (7)

Ideal COP is calculated on basis of effective condenser temperature (Tc) and the effective

evaporator temperature (Te).

Effective condenser temperature is defined as ( )TT

TTT

cexitcin

cexitcin

c ln

−= and effective evaporator

temperature is defined as,

( )TT

TTT

eineexit

eineexit

e ln

−= .

One of the form of representing rational efficiency (Second law efficiency) is:



143

ideal

actual

thCOP

COP=η (8)

Exergy efficiency can be written as follows:

in

dest

in

out

ex

xE

xE

xE

xE.

.

.

.

1−==η (9)

Maximum improvement in the exergy efficiency for a process or system is obviously

achieved when the exergy loss or irreversibility ( outin xExE..

− ) is minimized. It is useful to

employ the concept of improvement potential when analyzing the different processes. This

improvement potential on rate basis is given by Hammond and Stapleton as:

))(1(..

outinex xExEIP −−= η (10)

The irreversibility rates corresponding to various components of the system are calculated

using the exergy balance as follows:

I. Compressor and motor (process 1-2)

The exergy balance for this component control region is, .

2

.

1

..

ExExWI inI −+= (11)

Mechanical electrical losses can be obtained from the following relation:

)1(..

motormechinme WI ηη−= (12)

Internal irreversibility due to fluid friction is given by,

..

int

.

meI III −= (13)

II. Condenser (process 3-4)

Since the thermal exergy associated with heat transfer is zero ( 0.

=QE ), the

exergy balance in this case is written as, .

4

.

3

.

ExExI II −= (14)

III. Exergy balance for the Expansion Valve (process 5-6) .

6

.

5

.

ExExI III −= (15)

IV. Exergy balance for the Evaporator (process 7-8)

QIV EExExI..

8

.

7

.

+−= (16)

The performance of the condenser and evaporator is analyzed by defining the parameter QI /.

i.e., the ratio of irreversibility rate to heat transfer rate. The ratio indicates the relative change

in irreversibility rate with respect to heat transfer rate.

4. RESULTS

The observations were collected by conducting the experiments according to the

experimentation scheme, and the analysis was carried out based on the analysis scheme. The

results obtained from the analysis of the collected data are presented in graphical form in this

section.

Data was separately collected for both the models with same instruments, ensuring equal

accuracy. The exergy rate is calculated at inlet and exit of each component of the

refrigeration system. The reference state for R22 is taken as normal atmospheric conditions of



144

temperature and pressure as 298.16 K and 101.325 kPa respectively. The two condensing

units operate simultaneously and exhibit nearly similar performance with insignificant

variations at the end of cooling period. The refrigerant properties are calculated by using

CoolPack software.

Figure-3: Refrigeration cycle for old model at

t=72 min

Figure-4: Refrigeration cycle for new model at

t=72 min

Table-2: Results of exergy calculations at t=72 min (for old model)

Sr.

No.

Salient point Pressure

(kPa)

Temperature

(K)

Sp.

Enthalpy

(kJ/kg)

Sp.

Entropy

(kJ/kg)

Exergy

Rate

(kW)

1 Compressor

inlet 482.58 284.06 413.52 1.78 2.03

2 Compressor

exit

1840.70 368.76 461.88 1.82 3.82

3 Condenser

inlet

1840.70 368.76 461.88 1.82 3.82

4 Condenser

exit

1840.70 323.46 263.68 1.21 3.04

5 Expansion

Valve inlet

1840.70 323.46 263.68 1.19 3.36

6 Expansion

Valve exit

620.46 287.76 263.68 1.20 3.17

7 Evaporator

inlet

620.46 287.76 256.46 1.20 2.83

8 Evaporator

exit

551.52 281.76 410.49 1.76 2.21

Qcond= 9.28 kW ExQcond= 1.27 kW Teff,cond= 345.62 K

Qevap= 7.21 kW ExQevap= 0.34 kW Teff,evap= 284.75 K

Win= 2.79 kW COPactual=2.37 COPideal=4.67

Exergy loss in the system over cooling period of 180 minutes is tabulated below in Table 3.



145

Table-3: Exergy Loss in components of BMC system (old model)

Time

(min)

Compressor

Ex Loss

(kW)

Condenser

Ex loss

(kW)

Expansion

Valve

Ex loss

(kW)

Evaporator

Ex loss

(kW)

Total

Ex

Loss

(kW)

10 1.24 1.20 0.36 0.40 3.20

30 1.04 1.07 0.14 0.65 2.90

60 1.00 0.87 0.12 0.88 2.88

72 1.00 0.78 0.20 0.96 2.93

90 0.96 0.71 0.21 0.97 2.85

120 0.99 0.58 0.17 1.01 2.75

150 1.09 0.43 0.19 0.99 2.70

180 1.06 0.40 0.18 1.03 2.67

Similarly the exergy loss calculations for new model are calculated and tabulated in

table 4.

Table-4: Exergy Loss in components of BMC system (new model)

Time

(min)

Compressor

Ex Loss

(kW)

Condenser

Ex loss

(kW)

Expansion

Valve Ex

Loss (kW)

Evaporator

Ex Loss

(kW)

Total

Ex

Loss

(kW)

3 1.04 0.88 0.25 0.58 2.75

18 0.92 0.79 0.25 0.72 2.68

33 0.99 0.94 0.24 0.72 2.89

48 0.96 0.81 0.22 0.81 2.81

63 0.87 0.67 0.25 0.86 2.65

78 0.88 0.61 0.24 0.90 2.62

93 0.85 0.58 0.21 0.96 2.60

108 0.85 0.49 0.23 1.00 2.56

123 0.80 0.42 0.21 1.02 2.45

138 0.80 0.36 0.19 1.02 2.37

153 0.81 0.33 0.20 1.03 2.35

Exergy loss rate in each component of BMC refrigeration system for old and new model is

determined and presented in Figure-6 and Figure-7. Exergy loss in compressor for the old

model varies over the range 1.24 to 0.96 and that for the new model from 1.16 to 0.9. Exergy

efficiency of compressor for the new model is in the range 76% - 83% as against 70% - 81%

for the old model. The amount of exergy loss in condenser for new model lies in between

0.32-0.88 kW as against 0.4-1.2 kW for old model. In addition, the exergy destruction rate in

thermostatic expansion valve is found to be the lowest amongst all (about 10%) of the total

work input to the plant, which is considered as insignificant. The exergy loss in evaporator is

found to vary from 0.58- 1.03 kW for new model as against 0.4-1.03 kW for the old model.

The improvement in the performance is indicated by the ratio of irreversibility rate to the heat

transfer rate in condenser and evaporator as shown in Figure-8 and Figure-9. The ratio QI /.

International Journal of Mechanical Engineering and Technol

6340(Print), ISSN 0976 – 6359(Online) Volume 3, Issue 3, Sep

for the condenser in new model is found to be low in the range 0.05

of old model range 0.06-1.0. Similarly the ratio is determined for evaporator in new and old

model varying from 0.04-0.08 and 0.04

The exergy destruction rate in the condenser is primarily due to heat exchang

environment. Exergy destruction in condenser takes place due to pressure and temperature

loss. The exergy destruction rate due to pressure loss in condenser is very small and is about

0.01-0.02 kW and rest is due to temperature loss. Large pre

old model are reduced to greater extent in new model. It is observed that the amount of

pressure losses in old model were in the range 138 kPa to 35 kPa which are reduced

significantly to 35 kPa to 14 kPa, Figure

model, which is higher than that of old model by 2

consumption in components of the system is shown in Figure

the total exergy input to the system, it is

compressor, 26% in condenser, 24% in evaporator, 10% in expansion valve and 2% is

unaccounted loss.

COP of new model is found to be higher than old model by about 0.6

12. Carnot COP is determined on the basis of effective condenser and effective evaporator

temperatures. COP actual varies from 2.49 to 1.95 for old model as against the new model

from 3.5 to 2.75.

Figure- 5: Exergy destruction in

components of the BMC (the old

Figure- 7: Variation in ratio of irreversibility

rate to heat transfer rate in condenser

0

2

10 30 60 72 90 120 150

Ex. L

oss i

n K

W

Time in min

Comp Ex LossCondenser Ex. Loss

0.00

0.02

0.04

0.06

0.08

0.10

0.12

0 50 100Time in min

International Journal of Mechanical Engineering and Technology (IJMET), ISSN 0976

6359(Online) Volume 3, Issue 3, Sep- Dec (2012) © IAEME

146

for the condenser in new model is found to be low in the range 0.05-0.08 as compared to

1.0. Similarly the ratio is determined for evaporator in new and old

0.08 and 0.04-0.12 respectively.

The exergy destruction rate in the condenser is primarily due to heat exchang



0.02 kW and rest is due to temperature loss. Large pressure losses in the evaporator of



significantly to 35 kPa to 14 kPa, Figure-10 indicates the Second law efficiency for new

model, which is higher than that of old model by 2-4%. The exergy input and its

consumption in components of the system is shown in Figure-11. At particular instant, out of

the total exergy input to the system, it is found that about 38% of exergy is consumed in the


COP of new model is found to be higher than old model by about 0.6-0.8 as shown in Figure

is determined on the basis of effective condenser and effective evaporator


Exergy destruction in

components of the BMC (the old model)

Figure-6: Exergy destruction in components

of the BMC (the new model)

Variation in ratio of irreversibility

rate to heat transfer rate in condenser

Figure-8: Variation in ratio of irreversibility

rate to heat transfer rate in ev

150 180

Comp Ex LossCondenser Ex. Loss

0.00

0.20

0.40

0.60

0.80

1.00

1.20

1.40

0 50 100 150

Ex

erg

y L

oss

in

kW

Time in min

Compressor Ex Loss

Condenser Ex Loss

150 200

Old

model

0.00

0.05

0.10

0.15

0 50 100 150Time in min

ogy (IJMET), ISSN 0976 –

Dec (2012) © IAEME

0.08 as compared to that

1.0. Similarly the ratio is determined for evaporator in new and old

The exergy destruction rate in the condenser is primarily due to heat exchange with the



ssure losses in the evaporator of



Second law efficiency for new

4%. The exergy input and its

At particular instant, out of

found that about 38% of exergy is consumed in the


0.8 as shown in Figure-

is determined on the basis of effective condenser and effective evaporator


Exergy destruction in components

of the BMC (the new model)

Variation in ratio of irreversibility

rate to heat transfer rate in evaporator

150 200

Compressor Ex Loss

Condenser Ex Loss

150 200

Old …

New …

International Journal of Mechanical Engineering and Technol

6340(Print), ISSN 0976 – 6359(Online) Volume 3, Issue 3, Sep

Figure-9: Variation in second law efficiency

Figure-11: COP comparison between old and

new model

It is observed that the improvement potential in the new model has been reduced which

shows that there is overall improvement in the performance of the system. The system

performance has been improved resulting in power saving of 0.2 kW.

5 DISCUSSIONS

As observed from Figure-5 and Figure

as compared to other components of the system. Initially, the exergy loss rate is higher,

1.24kW, due to high discharge temperature because of high pressure

slowly to 0.9kW towards the end of cooling period for old model. Similar trend is observed

for new model with lower values of exergy destruction

destruction rate is due to throttling, followed by internal con

conduction and external convection and radiation also contribute to the exergy destruction in

compressor. Exergy destruction could be reduced by reducing the internal convective heat

transfer coefficient, swirl and turbu

and discharge valve ports.

Improvement in the design of evaporator leads to reduction in exergy destruction in

compressor and condenser. Isentropic efficiency of the compressor is found to be in

63% to 65%. Manufacturer’s compressor performance data and actual work input to the

compressor closely match with the theoretical work input and varies within 10%.

The exergy destruction in the condenser for new model is less as compared to that

model. Due to high discharge pressure and high discharge temperature in old model, exergy

0.00

10.00

20.00

30.00

40.00

50.00

60.00

0 50 100

Se

con

d L

aw

Eff

icie

ncy

Time in min

0.00

1.00

2.00

3.00

4.00

0 50 100

CO

P

Time in min

International Journal of Mechanical Engineering and Technology (IJMET), ISSN 0976

6359(Online) Volume 3, Issue 3, Sep- Dec (2012) © IAEME

147

Variation in second law efficiency Figure-10: Pattern of exergy consumption in

system components.

11: COP comparison between old and Figure-12: Comparison of Improvement

potential in both models.



performance has been improved resulting in power saving of 0.2 kW.

and Figure-6, exergy destruction rate is the highest in compressor


1.24kW, due to high discharge temperature because of high pressure ratio and decreases


for new model with lower values of exergy destruction The most significant exergy

destruction rate is due to throttling, followed by internal convection. Friction, mixing of fluid,



transfer coefficient, swirl and turbulence in the cylinder and by increasing the areas of suction


compressor and condenser. Isentropic efficiency of the compressor is found to be in



The exergy destruction in the condenser for new model is less as compared to that


150 200

old

mod…38%

26%10%

24%

2%

Exergy Consumption

(New Model)

150 200

old

model

ogy (IJMET), ISSN 0976 –

Dec (2012) © IAEME

Pattern of exergy consumption in

system components.

Comparison of Improvement

odels.



, exergy destruction rate is the highest in compressor


ratio and decreases


The most significant exergy

vection. Friction, mixing of fluid,



lence in the cylinder and by increasing the areas of suction


compressor and condenser. Isentropic efficiency of the compressor is found to be in the range



The exergy destruction in the condenser for new model is less as compared to that of the old


Exergy Consumption

(New Model)

Compressor

Condenser



148

loss is higher i.e.1.2kW at the starting period of cooling time, which later decreases sharply to

0.4kW as the temperature of the water in the tank approaches 4oC at the end of the cooling

time. The factors contributing to the exergy destruction in condenser are fluid friction and

turbulence caused due to change in direction of flow. Higher the refrigerant mass velocity,

higher is the exergy destruction rate. An ideal condenser coil should have a high heat transfer

coefficient with a low pressure drop. Both the heat transfer coefficient and pressure drop are

closely related to refrigerant mass velocity. Varying the refrigerant mass velocity in different

regions would balance the refrigerant side heat transfer coefficient and pressure drop.

Appropriate suitable complex refrigerant circuitry can improve the coil performance.

Significant reduction in pressure loss in the evaporator of the new model leads to decrease in

exergy losses. Exergy destruction in evaporator in the old model is comparatively high due to

the turbulent flow and mixing of the refrigerant fluid. The refrigerant side heat transfer

coefficient is low due to a vapour film with low thermal conductivity between the liquid and

the evaporator plate. In the new model, path constraint for the fluid flow is removed.

Therefore, the refrigerant vapor moves to exit after heat absorption. As a result, more amount

of liquid refrigerant comes in contact with the heat transfer surface. This design change in the

evaporator leads innovation leads to higher overall performance of the evaporator and

noticeable improvement in the second law efficiency. The study reveals that further

improvement in second law efficiency is possible by changing other design parameters,

which may be explored in future research.

The mixing of the incoming fluid with the fluid already present in the system when they have

different temperatures is one of the causes of irreversibility. Further, when the fluid is

throttled it will be at a temperature different from the fluid within the equilibrium system.

Unless the entering fluid has the same temperature as the fluid in the equilibrium system,

exergy destruction will occur.

Finally, the COP of the entire system can further be increased by reducing unuseful heat gain

in the suction line and in the condensing unit.

6. CONCLUSIONS

Significant amount of energy is lost in irreversibilities caused by improper process design and

poor design of components. The energy lost can be recovered by improving design and

process parameters. The compressor is the major contributor to the exergy destruction.

Evaporator and condenser, if redesigned to reduce the irreversibilities in flow, result in

sizeable energy savings. Proper sizing of evaporator inlets and exits for refrigerant will

further reduce the pressure loss and hence the exergy loss. The net savings in work input for

the new model is recorded as 8-10% as that of work input to old model and reduction in

cooling time is around 10%.



149

Abbreviations and symbols

Subscripts

COP : Coefficient of performance 0 : Reference state

xE.

: Exergy rate kW 1 : state 1 in process 1-2

G : Mass flux kg/m2 s 2 : state 2 in process 1-2

h : Specific enthalpy kJ/kg K c : condenser .

I : Irreversibility rate kW D : large diameter

IP : Improvement potential kW d : small diameter

Q : Heat Transfer Rate kW dest : destruction

s : Specific entropy kJ/kg K e : evaporator

T : Temperature K ex : exergetic

W : Mechanical or electrical energy

kW

in : Inlet

η : efficiency int : internal

th : thermodynamic

exit : Outlet

me : mechanical electrical

mech : mechanical

motor : electric motor

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Quantification of energy losses and performance improvement in dx cooling by exergy method

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