12 QUANTIFICATION OF ENERGY LOSSES AND … OF ENE… · QUANTIFICATION OF ENERGY LOSSES AND...
Transcript of 12 QUANTIFICATION OF ENERGY LOSSES AND … OF ENE… · QUANTIFICATION OF ENERGY LOSSES AND...
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:
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
Journal Impact Factor (2012): 3.8071 (Calculated by GISI)
www.jifactor.com
IJMET
© I A E M E
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:
International Journal of Mechanical Engineering and Technology (IJMET), ISSN 0976 –
6340(Print), ISSN 0976 – 6359(Online) Volume 3, Issue 3, Sep- Dec (2012) © IAEME
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.
International Journal of Mechanical Engineering and Technology (IJMET), ISSN 0976 –
6340(Print), ISSN 0976 – 6359(Online) Volume 3, Issue 3, Sep- Dec (2012) © IAEME
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.
International Journal of Mechanical Engineering and Technology (IJMET), ISSN 0976 –
6340(Print), ISSN 0976 – 6359(Online) Volume 3, Issue 3, Sep- Dec (2012) © IAEME
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.
International Journal of Mechanical Engineering and Technology (IJMET), ISSN 0976 –
6340(Print), ISSN 0976 – 6359(Online) Volume 3, Issue 3, Sep- Dec (2012) © IAEME
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:
International Journal of Mechanical Engineering and Technology (IJMET), ISSN 0976 –
6340(Print), ISSN 0976 – 6359(Online) Volume 3, Issue 3, Sep- Dec (2012) © IAEME
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
International Journal of Mechanical Engineering and Technology (IJMET), ISSN 0976 –
6340(Print), ISSN 0976 – 6359(Online) Volume 3, Issue 3, Sep- Dec (2012) © IAEME
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.
International Journal of Mechanical Engineering and Technology (IJMET), ISSN 0976 –
6340(Print), ISSN 0976 – 6359(Online) Volume 3, Issue 3, Sep- Dec (2012) © IAEME
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
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
0.02 kW and rest is due to temperature loss. Large pressure losses in the evaporator of
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-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
compressor, 26% in condenser, 24% in evaporator, 10% in expansion valve and 2% is
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
temperatures. COP actual varies from 2.49 to 1.95 for old model as against the new model
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
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
ssure losses in the evaporator of
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
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
compressor, 26% in condenser, 24% in evaporator, 10% in expansion valve and 2% is
0.8 as shown in Figure-
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
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.
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.
and Figure-6, exergy destruction rate is the highest in compressor
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,
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
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
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.
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
, exergy destruction rate is the highest in compressor
as compared to other components of the system. Initially, the exergy loss rate is higher,
ratio and decreases
slowly to 0.9kW towards the end of cooling period for old model. Similar trend is observed
The most significant exergy
vection. Friction, mixing of fluid,
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
lence in the cylinder and by increasing the areas of suction
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
Exergy Consumption
(New Model)
Compressor
Condenser
International Journal of Mechanical Engineering and Technology (IJMET), ISSN 0976 –
6340(Print), ISSN 0976 – 6359(Online) Volume 3, Issue 3, Sep- Dec (2012) © IAEME
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%.
International Journal of Mechanical Engineering and Technology (IJMET), ISSN 0976 –
6340(Print), ISSN 0976 – 6359(Online) Volume 3, Issue 3, Sep- Dec (2012) © IAEME
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
REFERENCES
[1] Bali,Tulin, Sarac, Betul Ayhan, Exergy Analysis of heat transfer in a turbulent pipe flow
by a decaying Swirl generator,International journal of exergy, vol. 5, No.1, 2008
[2] Bansal, P.K., Rupsinghe, A.S., An empirical model for sizing of capillary tubes,
International journal of Refrigeration, vol. 19 No. 8, pp 497-508,1996
[3] Bejan A., Entropy generation through heat and fluid flow, New York, John Wiley and
sons, 1982.
[4] Bejan A., Entropy generation minimization, New York, CRC press, 1996.
[5] Hepbasali,Arif, Thermodynamic analysis of a ground source heat pump system for district
heating, International journal of energy research, 2005, 29:671-687.
[6] Kotas,T.J., The Exergy Method of Thermal Plant Analysis, Butterworths,1985
[7] Liang,S.Y., Woong,T.N., Nathan,G.K., Numerical and experimental studies of refrigerant
circuitry of evaporator coils, International journal of refrigeration,24 (2001) 823-833
[8] McGovern, J.A., Harte,S., An exergy method for compressor performance analysis,
International Journal of Refrigeration, vol 18, No 6, pp 421-433, 1995
[9] Pons, Michel, Irreversibility in energy processes: Non-dimensional quantification and
balance, Journal of non equilibrium thermodynamics, 2004,vol 29, pp 157-175.
[10] Tirandazi,B., Mehrpooya, M., Vatani, A., Effect of valve pressure drop in Exergy
Analysis of C2+ Recovery Plants Refrigeration Cycles, Int. Journal of Electrical and
Electronics Engineering, 4:4, 2009
[11] Wang, S.P., Chen, Q.L., Yin, Q.H., Hua, B., Exergy destruction due to mean flow and
fluctuating motion in compressible turbulent flows through a tube, Energy, 28 (2003),
809-823.
[12] Robert Jenness, Noble P. Wong, Marth, Keeney, Fundamentals of Dairy Chemistry,
Springer 1998, PP 409-445.