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

INTERNATIONAL JOURNAL OF MECHANICAL ENGINEERING
International Journal of Mechanical Engineering and Technology (IJMET), ISSN
0976 – 6340(Print), ISSN 0976 – 6359(Online) Volume 3, Issue 3, Sep- Dec (2012) ©
IAEME
AND TECHNOLOGY (IJMET)

ISSN 0976 – 6340 (Print)
ISSN 0976 – 6359 (Online) IJMET
Volume 3, Issue 3, Septmebr - December (2012), pp. 137-149
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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:
dvghewade@gmail.com

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 4oC 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

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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

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 4oC. 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:

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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.

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3 Condenser
Compressor 2

Pressure measurement points-1,2,7& 8 4
5
Temperature measurement points- 1 to 8
E1 1 Expansion Valve
8
6
Evaporator
7 condensing unit 1
I1

I2 condensing unit 2
Evaporator

Expansion Valve
E2

Compressor
Condenser

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.

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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
GD D 2
n=
Gd d 2
3 Condenser
Compressor 2

4
5

E3 E2 E1 1 Expansion Valve
8
6
Evaporator
I3 I2 I1 condensing unit 1
7

I6 I5 I4 condensing unit 2
Evaporator

Expansion Valve
E6 E5 E4

Compressor
Condenser

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

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.

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.
Ex = ( h − ho ) + To ( s − s o ) (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:
.
∆ E x = ( h2 − h1 ) + To ( s 2 − s1 ) (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
. . .
E xin − E xout = E x dest (3)
Considering control volume at steady state (Fig. 2) the exergy balance can be expressed as
. . . . .
Ex in + Ex Qin + Win = Ex out + Ex Qout + I (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
. . .
E xout − E xin I total
η ex = .
= 1− .
(5)
W actual W actual

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:

Qe
COPactual = (6)
Win
The ideal COP obtained from the carnot cycle is given as,

Te
COPideal = (7)
Tc − Te
Ideal COP is calculated on basis of effective condenser temperature (Tc) and the effective
evaporator temperature (Te).

Effective condenser temperature is defined as T c =
T cin − T cexit and effective evaporator
ln(T cin T cexit )
temperature is defined as,
T eexit − T ein .
T e = ln(
T eexit T ein )
One of the form of representing rational efficiency (Second law efficiency) is:

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COPactual
η th = (8)
COPideal
Exergy efficiency can be written as follows:
. .
E xout E x dest
η ex = .
= 1− .
(9)
E xin E xin
Maximum improvement in the exergy efficiency for a process or system is obviously
. .
achieved when the exergy loss or irreversibility ( E xin − E xout ) 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:
. .
IP = (1 − η ex )( E xin − E xout ) (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,
. . . .
I I = Win + Ex1 − Ex2 (11)
Mechanical electrical losses can be obtained from the following relation:
. .
I me = Win (1 − η mechη motor ) (12)
Internal irreversibility due to fluid friction is given by,
. . .
I int = I I − I me (13)
II. Condenser (process 3-4)
.
Since the thermal exergy associated with heat transfer is zero ( EQ = 0 ), the
exergy balance in this case is written as,
. . .
I II = Ex3 − Ex4 (14)
III. Exergy balance for the Expansion Valve (process 5-6)
. . .
I III = Ex5 − Ex6 (15)
IV. Exergy balance for the Evaporator (process 7-8)
. . . .
I IV = Ex7 − Ex8 + E Q (16)
.
The performance of the condenser and evaporator is analyzed by defining the parameter I / Q
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

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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 Figure-4: Refrigeration cycle for new model at
t=72 min t=72 min

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

Sr. Salient point Pressure Temperature Sp. Sp. Exergy
No. (kPa) (K) Enthalpy Entropy Rate
(kJ/kg) (kJ/kg) (kW)

Compressor
1 482.58 284.06 413.52 1.78 2.03
inlet
Compressor 1840.70 368.76 461.88 1.82 3.82
2
exit
Condenser 1840.70 368.76 461.88 1.82 3.82
3
inlet
Condenser 1840.70 323.46 263.68 1.21 3.04
4
exit
Expansion 1840.70 323.46 263.68 1.19 3.36
5
Valve inlet
Expansion 620.46 287.76 263.68 1.20 3.17
6
Valve exit
Evaporator 620.46 287.76 256.46 1.20 2.83
7
inlet
Evaporator 551.52 281.76 410.49 1.76 2.21
8
exit
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.

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

Expansion Total
Compressor Condenser Valve Evaporator Ex
Time Ex Loss Ex loss Ex loss Ex loss Loss
(min) (kW) (kW) (kW) (kW) (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)
Total
Compressor Condenser Expansion Evaporator
Time Ex
Ex Loss Ex loss Valve Ex Ex Loss
(min) Loss
(kW) (kW) Loss (kW) (kW)
(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 I / Q

145

Technology
Sep

for the condenser in new model is found to be low in the range 0.05-0.08 as compared to that
0.05 0.08
of old model range 0.06-1.0. Similarly the ratio is determined for evaporator in new and old
1.0.
model varying from 0.04-0.08 and 0.04
0.08 0.04-0.12 respectively.
The exergy destruction rate in the condenser is primarily due to heat exchang with the
exchange
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 pressure losses in the evaporator of
0.02 pressure
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
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
2-4%.
consumption in components of the system is shown in Figure
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
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 0.8 as shown in Figure-
0.6-0.8 Figure
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.

Comp Ex Loss 1.40 Compressor Ex Loss
2 Condenser Ex. Loss
1.20 Condenser Ex Loss
1.00
Ex. Loss in KW

Exergy Loss in kW

0.80
0.60
0.40
0 0.20
10 30 60Time in90 120 150 180
72 min 0.00
0 50 Time in min 150
100 200

Figure- 5: Exergy destruction in Figure-6: Exergy destruction in components
components of the BMC (the old model) of the BMC (the new model)

0.12 Old 0.15
0.10 model
0.10
0.08
0.06 0.05
0.04 Old…
0.00 New…
0.02
0 50 Time in min 150
100 200
0.00
0 50 100
Time in min 150 200

Figure- 7: Variation in ratio of irreversibility Figure-8: Variation in ratio of irreversibility
rate to heat transfer rate in condenser rate to heat transfer rate in evaporator
ev

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Technology
Sep

60.00

Second Law Efficiency
50.00 Exergy Consumption
40.00 (New Model)
30.00
2%
20.00 Compressor
old
10.00 mod… 24% 38%
0.00 10% Condenser
26%
0 50 100
Time in min 150 200

Figure-9: Variation in second law efficiency Figure-10: Pattern of exergy consumption in
system components.

4.00

3.00
COP

2.00

1.00
old
0.00 model
0 50 100
Time in min 150 200

Figure-11: COP comparison between old and
11: Figure-12: Comparison of Improvement
new model potential in both models.
odels.

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 , exergy destruction rate is the highest in compressor
Figure-6,
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 ratio and decreases
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 The most significant exergy
destruction rate is due to throttling, followed by internal convection. Friction, mixing of fluid,
convection.
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 turbulence in the cylinder and by increasing the areas of suction
turbulence
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 the range
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 of the old
model. Due to high discharge pressure and high discharge temperature in old model, exergy

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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%.

148


Abbreviations and symbols Subscripts

COP : Coefficient of performance 0 : Reference state
. : Exergy rate kW 1 : state 1 in process 1-2
Ex
G : Mass flux kg/m2 s 2 : state 2 in process 1-2
h : Specific enthalpy kJ/kg K c : condenser
. : Irreversibility rate kW D : large diameter
I
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 in : Inlet
kW
η : efficiency int : internal
th : thermodynamic
exit : Outlet
me : mechanical electrical
mech : mechanical
motor : electric motor

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