Exergy Analysis of Vapor Compression Cycle -...

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CHAPTER 4

Exergy Analysis of Vapor Compression Cycle

4.1 Introduction

Refrigeration plays a very important role in industrial, domestic and commercial

sectors for cooling, heating and food preserving applications. There are innumerable

applications of such systems and they are the major consumer of electricity around the

world. Energy consumption is directly proportional to the economic development of any

nation, however this area is in great interest now because of increase in the cost of

conventional fuels and environmental concerns globally. The scientists are looking for

new and renewable sources of energy so as to minimize the costs. Due to the

increasing energy demand, degradation of environment, global warming and depletion

of ozone layer etc, there is urgent need of efficient energy utilization and waste heat

recovery for useful applications. The researchers are concentrating on the alternate

and environment friendly refrigerants, especially after the Kyoto and the Montreal

protocols. However, in a quest to find out alternate and environment friendly

refrigerants, the energy efficiency of the equipment having conventional refrigerants is

also very important in the present age of competitive business community. The aim of

the scientific community all over the world is to switch to new and renewable energy

sources besides, efficient utilization of all conventional sources.

Air conditioning bears a huge cost because thermal comfort is very essential as

far as domestic and industrial sectors are concerned. The big challenge is to use less

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energy for air conditioning applications in order to reduce the associated power

consumption so as to make them more efficient and environmental friendly. The

quantitative information is required to be obtained that will show the irreversibility of a

process in all the components of any plant. For effective use and proper optimization,

the detailed understanding of different thermodynamic processes in any conversion

system is very important. In order to optimize their design, a thorough thermodynamic

analysis is required. The analysis based on first law of thermodynamic is most

commonly used in engineering applications, however, it is concerned only with law of

conservation of energy and therefore it cannot show how and where irreversibility in

the system or a process occurs. On the other hand, the analysis based on second law

analysis is well known method being used to analyze all the thermodynamic cycles for

better understanding and evaluation of irreversibility associated with any process.

Unlike the first law (energy), the analysis based on second law analysis (exergy)

determines the magnitude of irreversibility associated in a process qualitatively and

thereby, provides an indication to point out the directions in which the engineers

should concentrate more in order to improve the performance of this thermodynamic

system [1-3]. Thus, the aim of second law based analysis is to determine the exergy

losses and to enhance the performance by changing the design parameters and hence,

to reduce the cost of the refrigeration cycle [4].

A lot of studies on the performance evaluation and optimization have been

carried out experimentally and theoretically are available in the literature [5-10]. Most of

the studies carried out so far on the refrigeration systems shows that the performance

analysis of refrigeration systems were investigated based on first law of

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thermodynamics. However, this approach is of limited use in view of the fact that the

actual energetic losses are difficult to make out because the first law deals with the

quantity of energy and not the quality of energy. In order to calculate the actual losses

due to irreversibility in the process, exergy analysis based on second law of

thermodynamics is the proper tool. Exergy analysis utilizes exergetic efficiency criterion

taking into account all the losses appearing in a system, for measuring the actual

performance.

The exergy analysis is widely accepted as a useful tool in obtaining the

improved understanding of the overall performance of any system and its components

[11]. Exergy analysis also helps in taking account the important engineering decisions

regarding design parameters of a system [12]. Many researchers have carried out

exergy studies of different thermal energy conversion systems describing various

approach for exergy analysis and its usefulness in a more simple and effective manner

[13-23]. Padilla et al [13] carried out the exergy analysis and the impact of direct

replacement of R12 with zeotropic mixture R413A. The performance of a domestic

vapor compression refrigeration system originally designed to work with R12 was

evaluated using a simulated modeling. They concluded that the overall energy and

exergy performance of this system working with R413A is better than that of R12.

Kumar et al [14] derived a method to carry out the exergetic analysis of a vapor

compression refrigeration system using R11 and R12 as refrigerants. The procedure

to calculate various losses as well as coefficient of performance and exergetic

efficiency of the cycle has been explained by proper example. Arora and Kaushik [15]

did a detailed exergy analysis of an actual vapor compression refrigeration cycle. They

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developed a computational model to calculate the coefficient of performance, exergy

destruction, exergy efficiency and the efficiency defects for R502, R404A and R507A

for temperature in the range of -50°C to 0°C and condenser temperature range of

40°C to 55°C. They concluded that R507A is a better substitute to R502A than that of

R404A. Nikolaidis and Probert [16] studied the behaviour of a two-stage compound

compression cycle, with flash intercooling with R22 using the exergy method and gave

some useful conclusions.

Dincer [17] asserts that conventional energy analysis, based on the first law of

thermodynamics, evaluates energy mainly on its quantity but analysis that are based

on second law considers not only the quality of energy but also quantity of energy.

In this study, the main objective is to investigate the performance of a simple

VCR system based on exergy analysis. The experimental analysis has been done on

a 2TR window air conditioning system using R-22 as refrigerant. With the objective to

find out the losses at different operating conditions for vapor compression cycle,

exergy analysis has been done by varying the quantity of refrigerant charge. The

system has been modified for experimental study to find the possible design conditions

with the minimum exergy destruction. In the present study, the effects of temperature

changes in the condenser and evaporator on the plant’s irreversibility rate was

determined. It is observed that the greater is the temperature difference between the

condenser and the environment or between the evaporator and the cold room, the

higher is the irreversibility rate. The analysis is performed by doing energy and exergy

balances for the system. The properties of refrigerant at each state point are

calculated using Forane software [18] and the results are discussed.

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The present investigation will help us to improve the understanding of vapor

compression cycle and will enhance its efficiency. Further study on the design

parameters, (external and internal) operating conditions is in process to obtain the

optimum performance of the system.

4.1.1 Basic thermodynamic considerations

First law of thermodynamics is used and is based on law of conservation of

energy. This law only quantifies the energy at various components and tells nothing

about how to reduce the energy consumption in these systems. In addition to this the

second law of thermodynamics in its fundamental form is not used in energy utilization

analysis and therefore, some of its derived principles are used in energy

conservations. In the interest of energy conservation, minimization of irreversible

effects is sought in any thermodynamic process. Any process that occurs without any

change of initial state. A number of physical effects which cause a loss of energy

available for doing work, or cause an increase in energy required to produce desired

output are as follows:

Temperature difference: Large temperature difference causes greater losses during

heat transfer, therefore, temperature difference should be kept as small as practical

e.g. in evaporators and condensers.

Friction: Friction causes loss of useful energy and therefore should be minimized. For

example, regular cleaning of tube walls prevents scale build-up. Thus fluid friction will

be less and less energy will be lost in pumping power.

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Rapid expansion: The adiabatic expansion of a high pressure fluid to a low pressure,

sometimes called throttling, is a process that wastes some energy available in the high

pressure fluid which could have been used to do some work.

Mixing: Mixing of fluids will results in a loss of the useful available energy.

The term entropy is a physical property of substances related to energy

utilization and conservation. Entropy is a measure of the energy that is not available to

do work. For a fluid with isentropic nature the amount of work done in compressor is

minimum. A constant entropy (i.e. isentropic) process is an ideal reversible process

which can never take place. Hence, in a real process where work is required, the

entropy increases and efforts are made to minimize this increase.

4.1.2 Availability

Availability is the maximum useful work that can be obtained from a system at a

given state. A system is said to be in the dead state when it is in thermodynamic

equilibrium with the surroundings. At the dead state, a system is at the temperature

and pressure of its surroundings as well as in thermal, chemical and mechanical

equilibrium. It has no kinetic or potential energy relative to its surroundings and does

not reacts with the surroundings. A system has zero availability at the dead state. The

notion that a system must go, the dead state at the end of the process to maximize the

work output can be explained as follows:

If the system temperature at the final state is greater than or less than the

temperature of the surroundings, we can always produce additional work by running a

heat engine between these two temperature levels. If the final velocity of the system is

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not zero, we can catch that extra kinetic energy by a turbine and convert it to rotating

shaft work and so on. However, no work can be produced from a system that is initially

at the dead state. The atmosphere around us contains a tremendous amount of

energy. However, the atmosphere is in the dead state and the energy it contains has

no work potential therefore, we conclude that a system will deliver the maximum

possible work as it undergoes a reversible process from the specified initial state to the

state of its environment, the dead state. This represents the maximum work potential

of the system at the specified state and it is called availability.

It is important to realize that availability does not represent the amount of work

that a work producing device will actually deliver upon installation. Rather it represents

the upper limit on the amount of work a device can deliver without violating any

thermodynamic laws. There will always be a difference, large or small between

availability and actual work delivered by a device. This difference represents the scope

for improvement. It may be noted that the availability of a system at a specified state

depends on the conditions of the environment as well as the properties of the system.

Therefore, availability is a property of the system-surroundings combination and not of

the system alone. Altering the environment is another way of increasing availability but

it is definitely not an easy alternative. Thus it would be desirable to have a property to

enable us to determine the useful work potential of a given amount of energy at some

specified state.

The work potential of the energy contained in a system at a specified state is

simply the maximum useful work that can be obtained from the system. The work done

during a process depends on the initial state, final state and the process path. In an

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availability analysis, the initial state is specified and thus it is not a variable. The work

output is maximized when the process between the two specified states is executed in

a reversible manner. Therefore all the irreversibilities are disregarded in determining

the work potential. Finally the system must be in the dead state at the end of the

process to maximize the work output.

4.1.3 Reversible work

The availability analysis is useful for determining the useful work potential of the

energy of a system at a specified state. It serves as a valuable tool in determining the

quality of energy and comparing the work potential of different energy sources. The

availability analysis alone however, is not of great value for studying energy devices

operating between two fixed states. This is because in an Availability analysis, the final

state is always assumed to be the dead state, which is hardly ever the case for actual

thermodynamic systems.

The adiabatic efficiencies are of limited use because the exit state of the model

(isentropic process) is not the same as the actual exit state. There are two new

quantities that are related to the exact initial and final states of the actual process.

These two quantities are reversible work and irreversible work. They serve as valuable

tools in the optimization studies of the components in complex thermodynamic

systems. Reversible work (Wrev) is defined as the maximum amount of work that can

be obtained as a system undergoes a process between the specified initial and final

states. This is the useful work output (or input) obtained when the process between

the initial and final states is executed in a totally reversible manner. That is, any heat

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transfer between the system and the surroundings must take place reversibly, and no

irreversibilities should be present within the system during the process. When the final

state is the dead state, the reversible work equals availability.

4.1.4 Irreversibility

For processes, that require work, reversible work represents the minimum

amount of work necessary to carry out that process. Any difference between the

reversible work Wrev and the useful work Wu is due to the irreversibilities present

during the process and this difference is called irreversibility I, it is expressed as lost

work potential i.e.

I = Wrev – Wu

For a totally reversible process, the actual and reversible work terms are

identical and thus irreversibility is zero. This is expected since totally reversible

processes generate no entropy which is a measure of irreversibilities occurring during

a process. For all actual (irreversible processes, irreversibility is a positive quality since

the work term is positive and Wrev > Wu for work producing devices ad the work term is

negative since Wrev < Wu for work consuming devices. Irreversibility can be used as

the lost opportunity to do work. It represents the entropy that could have been

converted to work. The smaller the irreversibility associated with a process, the greater

the work that will be produced (or the smaller the work that will be consumed). To

improve the performance of thermodynamic systems, the primary sources of

irreversibility associated with each component in the system should be located and

effort should be made to minimize them.

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4.2 Energy conservation aspects in VCR systems

Refrigeration and air conditioning system mainly consists of four major

components, i.e. compressor, condenser, evaporator and throttling device along with

piping network, suitable refrigerant and heat transfer liquids. It is imperative that

system operating criteria should be analyzed completely to assure improvements in

components, protective devices, advanced and complex thermodynamic cycle designs

and proper working fluid selection. However, the working of these components has to

be understood deeply in order to understand the working of the overall system and to

look for the options of energy conservation and savings. A summary of cyclic effects

that result in reduced energy consumption per unit of refrigeration capacity are

described as follows:

4.2.1 Energy conservation in evaporator

The function of the evaporator is to cool the space by transferring heat into

refrigerant in terms of its latent heat of vaporization. Evaporators may be of the types

as, water-cooling, air-cooling shell and tube heat exchangers. It has been observed

that when evaporator temperature changes from +5 to -20°C, the refrigeration tons

drop to nearly one third its value and power/ton increases from 0.81 to 1.67 which is

more than double. A high value of evaporator temperature will increase cooling output

and decrease the input power and hence, will have a more pronounced effect on

improving COP of the system. Performance of an actual practical machine also shows

this effect very clearly. Thus, it is advantageous to operate machine at as high as

evaporator temperature and as low as condensing temperature as possible. The

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correct choice for designing an evaporator and their sizing is the most important factor

for an efficient and economic refrigerating system. For energy conservation, promotion

of increased heat transfer and maintaining of the highest reasonable evaporating

temperature is required. It should be noted that under no circumstances these two

factors are opposing. That is, one way of increasing evaporator heat transfer without

decreasing evaporating temperature. Effectively this means that for a given heat

transfer (refrigeration) capacity, the evaporating temperature is higher, and hence less

power is used. The energy conservation and saving options in the use of evaporators

are as follows:

1. Counter flow arrangement increases heat transfer over parallel flow, since

mean effective temperature difference is greater.

2. A flooded evaporator surface is more effective than a dry expansion surface,

which is only partially wetted.

3. Refrigerant sprays may increase surface wetting.

4. Maintenance of clean surfaces reduces the heat transfer coefficient.

5. Increased surface area, through use of fins/more tubing, increases heat

transfer.

6. Having higher fluid velocities increase the heat transfer coefficient, in both the

refrigerant and air or water side. In liquid chillers this is achieved by increasing

the number of passes.

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4.2.2 Energy conservation in compressor

The compressor pumps refrigerant round the circuit as well as produces

the required substantial increase in the pressure of the refrigerant i.e. its function is to

raise the gas refrigerant pressure and temperature and they are of various types such

as reciprocating, centrifugal, positive displacement, screw and fully or partially

hermetically sealed. The desired operating conditions appropriate for energy

conservation in the use of compressor are given as below:

1. Compressor capacity decreases with higher suction pressure and lower

discharge pressure.

2. Compressor power requirement decreases with decreasing compression ratio

and increases with greater clearance volume.

3. Speed control results in excellent reduction of power with capacity decrease.

4. Cylinder unloading provides good power reduction with capacity decrease.

5. Screw compressors have excellent full load (and part load with slide valve

control) power requirements.

6. An economizer arrangement with screw compressors reduces energy

requirements considerably.

Using multiple compressors in a system can also save energy, since at partial load

some of the compressors can be operated and at full load (generally the most efficient

operating condition) while others are shut down. Of course this also results in the extra

benefit of stand-by capacity in case of failure of a compressor. Energy conservation

methods applied to centrifugal chillers have become an especially important

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consideration because of the vast quantity of power used. These energy saving

methods can be summarized as:

1. Multistage compression with the economizer cycle decreases power

requirements over single-stage compression.

2. Variable inlet guide vane capacity control will result in good reduction of power

with capacity reduction.

3. Increasing the number of water passes in the heat exchangers (evaporator and

condenser) will reduce compressor power requirement. However, increased

pumping cost must be balanced against this.

4. Demand of limiting devices that limit the current drawn by the motor will reduce

utility company demand charges.

5. Non condensed hot gas discharge temperature is high which can be used to

recover heat energy.

4.2.3 Energy utilization in condensers

Condenser’s function is to reject the heat of condensation and change the state

of the refrigerant from gas to a liquid. Condensers may be of the types such as, water-

cooled/air-cooled shell and tube and water-cooled evaporative condenser. In many

applications of refrigeration, temperatures lower than necessary are specified. A low

value of the condensing temperature can reduce the hot gas discharge temperature

which is too high. Design and operation of condenser can contribute for the energy

conservation in many ways as given below:

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1. Maintain clean heat transfer surfaces, through proper water treatment and

maintenance.

2. Compare the required power for air cooled condenser with that of either

evaporative or water cooled condenser and then select the condenser design.

3. Select condensers with a large heat transfer surface area.

4.2.4 Sub-cooling of liquid before the expansion valve

It is possible to sub-cool the condensed liquid after it comes out of the

condenser. This sub-cooling can be provided by using the surplus cooling of the

refrigerant gas which might otherwise be lost. When evaporator coils are defrosted by

water sprays, the water of melted ice can be used for sub-cooling. The defrost water is

also heated in the process. For short term increase of capacity, sub-cooling by liquid

nitrogen can be used. In general, it can be said that sub-cooling improves the

compressor efficiency, provided the heat given up during the sub-cooling leaves the

system and does not re-enter. Following are some of the points relating to energy

utilization and conservation as affected by flow control devices:

1. The thermostatic expansion valve (TEV) provides efficient use of the evaporator

surface over a wide range of loads.

2. If there is a significant pressure drop in either the evaporative or the

refrigerant distributor, an external equalizer type TEV must be used.

3. The internal equalizer valve will result in inefficient use of evaporator surface.

If the system is subject to hunting a cross charge or other anti- hunting type

TEV must be used.

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4. The capillary tube is a less-energy-efficient flow control device than the

TEV. However, for small capacity units, capillary tube is used.

5. The constant pressure expansion valve is a very inefficient flow control

device if there is significant load variation.

4.2.5 Energy conservation in connecting pipe network

Pipes are used to circulate refrigerant and cooling liquid through various

components of the system. The major considerations in refrigerant piping that affect

energy use are summarized as follows:

1. Suction and hot gas lines should be sized for reasonably low recommended

pressure drops, since compressor power increases with pressure difference.

2. Pressure drop in the liquid line should be low enough to prevent flashing.

3. If a liquid-suction heat exchanger is used for sub-cooling, it will also improve

compressor energy efficiency.

4. Excess oil in the system will coat heat transfer surfaces of the evaporator and

condenser thereby reducing their performance. It is desirable to use

recommended oil products for improved performance

4.3 Experimental setup and procedure

The experimental setup used in the present study is a 2TR window air

conditioner which was further modified to incorporate the compound gauges for

temperature and pressure measurements at suction and discharge side. The main loop

of the system under investigation is similar to that of a common VCR system and

composed of four basic components only i.e. compressor, an evaporator, condenser

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and an expansion device. The vapors from low temperature evaporator are sucked into

the compressor and compressed to increase the pressure and temperature. The

compression is assumed to be polytropic and the vapors are condensed in the air

cooled finned condenser in which the fan is driven by an electric motor which is also

driving the evaporator fan at other end. The speed of the condenser and evaporator fan

is kept constant. The compressor is hermetic type reciprocating compressor with 2100

watts nominal input power at 220V (50Hz). The evaporator was designed for a cooling

capacity of 24K BTU. The refrigerant is charged in four steps and the performance is

evaluated at each stage. The unit also comprises of other devices such as filters and

compressor protection device. The refrigerant volume flow rate was measured using a

rotameter which is specifically calibrated for R-22. The temperature and pressure were

measured by using compound gauges. The photographic view of the experimental set

up and line diagram is shown in Fig.-4.1(a) & 4.1(b) and the specifications of the VCR

system are given in Table-4.1.

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Fig.-4.1a: Experimental setup of 2 TR Window Air Conditioner based on VCR

Fig.-4.1b: Typical diagram of a vapor compression cycle.

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Table – 4.1: Specification of Experimental Set-up

Type Window Air Conditioner Capacity 2 TR (24000 BTU)

Condenser Finned Coils, Air Cooled Evaporator Finned Coils

Expansion Device Capillary Tube Compressor Hermetically Sealed, Reciprocating

4.4 Exergy Analysis

Exergy analysis has two advantages over the conventional heat balance

method for design and performance analysis of energy related systems. It provides a

more accurate measurement of the actual inefficiencies in the system and the true

location of these in efficiencies. In refrigeration cycle, with the heat balance analysis, it

is not possible to find out the true losses. Exergy analysis is based on the assumption

that there is an infinite equilibrium environment that ultimately surrounds all systems

that are to be analyzed. The exergy or available energy of a system is the maximum

work that could be derived if the system were allowed to come to equilibrium with the

environment. It is a consequence of the second law of thermodynamics that the

combined exergy of all systems can only decrease or remain unchanged. Unlike

energy, exergy is not conserved, once it is lost, it is lost forever. In other words, exergy

(quality) is degradable, while energy (quantity) is conserved. Exergy can be

exchanged between systems, but if there are thermodynamic irreversibility’s such as,

friction or heat transfer with finite temperature differences, some of the potential for the

production of work is destroyed. In all real processes, therefore, the total exergy of the

system decreases [19-23].

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For a specified system boundary a clear distinction can be made between

exergy destruction and exergy loss. Exergy loss is exergy that is passed on to some

other system often the environment and which cannot be considered useful in the

context of the purpose of the system. The term exergy destruction is used when the

potential for the production of work is destroyed within the system boundary. The

exergy of a system is a co-property of the system and the environment. In exergy

analysis of compressors the environment consists of the local surroundings of the

compressor. These local surroundings are modeled as being in equilibrium and

infinite. Given sufficient information, the exergy of all the systems can be determined

at any time. On the basis of first law, the performance of refrigeration cycle is based on

the coefficient of performance, which is defined as the ratio of net refrigerating effect

(cooling/heating load) obtained per unit of power consumed. It is expressed as:

c

e

W

QCOP = (4.1)

Exergy balance for a control volume can be expressed as [2]

( ) ( ) ( )[ ] ∑∑ ∑∑ ±−−−+−∑= WTTQTTQmemeEXoutoinoutxinxD /1()/1(( 0 (4.2)

For the present system shown in Fig.4.1b, the component wise exergy balance

equation can be written as below:

a) Compressor

))(()( 1221 ssTmEWEEX orxcxcompD −=−+= (4.3)

b) Condenser

)()()( 332232 sThmsThmEEEX ororxxcondD −−−=−= (4.4)

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c) Expansion Device

43exp)( xxD EEEX −=

)()( 4433 sThmsThm oror −−−=

))(( 12 ssTm or −= (4.5)

d) Evaporator

( ) 14 /1)( xroexevapD ETTQEEX −−+=

( ) )(/1)( 1144 sThmTTQsThm orroeor −−−+−= (4.6)

The total exergy destruction in the system is the sum of exergy destruction in different

components of the system and is given by

e) Total exergy destruction

evapDDcondDcompDtotalD EXEXEXEXEX )()()()()( exp +++= (4.7)

f) Exergy Efficiency

ro

c

eexergy TT

W

Q/1( −=η (4.8)

4.5 Results and Discussions

In order to have a comparative study of vapour compression cycle by varying the

quantity of refrigerant charged, the real time data was measured and the calculations

were made using Forane software [18]. The temperature, pressure and mass flow rate

of the refrigerant was measured using compound gauges and rotameter respectively at

different state points. The basic properties such as entropy and enthalpy of the

refrigerant at different state points were calculated using Forane software [18]. The

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coefficient of performance (COP), refrigerating effect, compressor work, exergy

destruction and exergy efficiency are calculated using simple Excel sheet.

A 2TR window air conditioner equipped with different pressure, temperature and

flow measuring devices has been studied experimentally using energy and exergy

analysis. The unit is charged with refrigerant R-22 in four steps i.e. 25%, 50%, 75% and

100% respectively and the system performance is analyzed in each case. The

reference temperature is measured to be 25oC. The coefficient of performance (COP),

cooling load, exergy destruction, exergy efficiency has been evaluated against running

time and evaporator temperature and the discussion of results is given below:

It is observed from Tables (4.2-4.5) that the evaporator temperature is least in

case when the system is 100% charged but this increases pressure ratio of the

compressor and hence compressor work also increases as a result the coefficient of

performance goes down as can be seen from Fig.-4.2. Figure-4.2 also shows the

comparison of coefficient of performance against time in all the cases and it is observed

that the COP fluctuates with time having number of peaks after certain intervals except

when the system is 75% charged. In the case when the system is 75% charged the

COP has the highest peak in the middle of running time followed by small peaks on

either sides. However, when the system is 25% charged, the COP first increases along

the highest peak in the first few minutes and then decreases gradually with small ups

and downs before finally reaching steady state condition in an hour’s time. Figure 4.2

show that when the system is 50% charged, the COP initially drops and finally achieves

a consistent value. After an hour of running the system, the steady state coefficient of

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performance was observed in all the case and hence, the data after a certain period

was not plotted as can be seen in Fig.-4.2.

Fig.-4.2: Variation of COP with respect to running time

Table – 4.2a: Parameters of the VCR system when the unit is 25% charged

Time Pdisch./

kg/cm² Psuc./

kg/cm² m/

kg/s Cond. In/°c

Evap. Inlet / °c

5 4 -15 cm of Hg 0.0236 52 24.5

10 9.2 0.4 0.0283 52 24.5 15 9.4 0.45 0.0306 58 21 20 9.6 0.55 0.0318 58 21 25 9.7 0.6 0.033 68 20.1 30 9.8 0.61 0.0332 68 20.1 35 9.9 0.68 0.0334 75 20.8 40 9.95 0.7 0.0334 75 20.8 45 10 0.7 0.0334 76 21.1 50 10.1 0.71 0.0334 76 21.1 55 10 0.7 0.0332 75 22.1

0

1

2

3

4

5

6

7

8

9

10

5 10 15 20 25 30 35 40 45 50 55 60 65

Co

eff

icie

nt o

f P

erf

orm

an

ce

Running Time/minutes

Coefficient of Performance (25% charged)

Coefficient of Performance (50%charged)



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Table – 4.2b: Exergetic performance results of the VCR system when the unit is 25% charged

RE/ kJ

Wcomp./ kJ COP

EXD evap.

EXD comp.

EXD cond.

EXD exp.

EXD total

Exergy Efficiency

% Exergy Efficiency

4.283 0.732 5.848 2.443 3.723 1.836 2.442 10.446 0.0359 3.5 5.140 0.686 7.485 2.932 3.654 2.826 2.930 12.343 0.0431 4.31 5.547 0.878 6.313 3.109 4.068 3.092 3.169 13.440 0.1132 11.3 5.760 0.906 6.357 3.229 4.196 3.233 3.291 13.950 0.1175 11.75 5.941 1.193 4.977 3.322 4.579 3.376 3.425 14.703 0.1398 13.9 5.983 1.198 4.991 3.345 4.592 3.416 3.449 14.804 0.1408 14 5.895 1.203 4.897 3.391 4.614 3.447 3.469 14.924 0.1244 12.44 5.895 1.387 4.248 3.391 4.765 3.481 3.469 15.107 0.1244 12.44 5.915 1.411 4.191 3.392 4.785 3.484 3.474 15.137 0.1186 11.86 5.915 1.407 4.201 3.392 4.775 3.491 3.474 15.133 0.1186 11.86 5.854 1.374 4.258 3.394 4.731 3.453 3.445 15.024 0.0972 9.72


Time

Pdisch./

kg/cm²

Psuc./

kg/cm²

m/

kg/s

Cond.

In/°c

Evap. Inlet

/ °c

5 9 0.4 0.028 47 23

10 9.4 0.5 0.0306 53 20.3

15 9.6 0.6 0.0306 62 21.5

20 9.8 0.65 0.0306 62.2 21.3

25 9.9 0.7 0.0317 64 19.6

30 10 0.71 0.03175 64.4 20.3

35 10 0.71 0.03179 69 22.5

40 10.05 0.715 0.03179 67.9 21.4

45 10.05 0.7 0.03179 65 19.4

50 10.05 0.71 0.0332 62 17.3

55 10.05 0.71 0.0341 65 19.5

60 10.05 0.71 0.0353 65 19.5

125


RE/ kJ Wcomp./

kJ COP EXD

evap. EXD

comp. EXD

cond. EXD exp.

EXD

total

Exergy Efficiency % Exergy

Efficiency 5.117 0.573 8.921 2.885 3.586 2.800 2.932 12.204 0.0691 6.91 5.559 0.750 7.412 3.076 3.967 3.076 3.176 13.297 0.1270 12.7 5.556 0.961 5.780 3.104 4.142 3.110 3.175 13.532 0.1037 10.3 5.513 0.952 5.790 3.079 4.133 3.131 3.175 13.519 0.1067 10.6 5.716 1.026 5.566 3.163 4.320 3.261 3.298 14.044 0.1445 14.4 5.722 1.030 5.555 3.165 4.320 3.280 3.294 14.061 0.1307 13 5.700 1.154 4.939 3.222 4.415 3.287 3.303 14.229 0.0868 8.68 5.729 1.119 5.119 3.199 4.387 3.290 3.294 14.172 0.1089 10.89 5.697 1.052 5.413 3.168 4.311 3.277 3.308 14.066 0.1480 14.8 5.996 1.006 5.960 3.229 4.452 3.416 3.453 14.552 0.2003 20.03 6.156 1.130 5.447 3.409 4.630 3.527 3.542 15.109 0.1578 15.78 6.390 1.169 5.465 3.526 4.790 3.649 3.664 15.629 0.1638 16.38


Time Pdisch./ kg/cm²

Psuc./ kg/cm²

m/ kg/s

Cond. In/°c

Evap. Inlet / °c

5 15.16 3.2 0.03532 58 22.9

10 15.18 3.4 0.03532 60 24.3

15 15.27 3.5 0.03532 60 24

20 15.28 3.6 0.03532 64.5 23

25 15.28 3.7 0.03532 70.5 22.5

30 15.3 3.7 0.03532 56.1 22.3

35 15.3 3.7 0.03532 79.5 24.5

40 15.3 3.7 0.03532 74 22.6 45 15.3 3.7 0.03532 69.3 22.2 50 15.3 3.7 0.03532 70.1 24.8

55 15.32 3.7 0.03532 74 23

60 15.32 3.7 0.03532 74 24.7

65 15.32 3.7 0.03532 72.4 25.7

126


RE/ kJ

Wcomp./ kJ COP

EXD evap.

EXD comp.

EXD cond.

EXD exp.

EXD

total

Exergy Efficiency % Exergy

Efficiency 6.012 0.7312 8.222 3.603 3.995 3.945 3.686 15.230 0.1247 12.4 5.772 0.7806 7.393 3.592 4.058 3.945 3.702 15.299 0.1058 10.5 5.962 0.7771 7.672 3.589 4.048 3.963 3.682 15.284 0.1175 11.7 5.916 0.9290 6.368 3.645 4.164 3.974 3.685 15.469 0.0840 8.4 5.941 1.1162 5.322 3.635 4.313 3.995 3.691 15.636 0.1048 10.4 5.832 0.6676 8.735 3.635 3.921 3.935 3.697 15.189 0.1028 10.2 5.853 1.3670 4.281 3.594 4.578 4.030 3.694 15.898 0.0931 9.3 5.839 1.2222 4.777 3.647 4.408 4.009 3.682 15.747 0.0989 9.8 5.863 1.0809 5.424 3.574 4.281 3.988 3.697 15.540 0.1647 16.47 5.818 1.0809 5.382 3.554 4.334 3.984 3.704 15.578 0.1268 12.68 5.888 1.2151 4.845 3.527 4.419 3.998 3.707 15.653 0.1862 18.62 5.878 1.1975 4.908 3.609 4.440 4.002 3.684 15.736 0.0854 8.5 5.807 1.1409 5.089 3.641 4.408 3.991 3.682 15.724 0.0409 4.09


Time Pdisch./ kg/cm²

Psuc./ kg/cm²

m/ kg/s

Cond. In/°c

Evap. Inlet/ °c

5 18.6 4.1 0.03768 74 9 10 18.8 4.2 0.03768 80 7 15 18.8 4.2 0.03768 81.5 6.5 20 18.4 4.15 0.03768 82 8 25 18.6 4.15 0.03768 82 8.1 30 18.7 4.2 0.03768 81 9 35 18.7 4.2 0.03768 83 7 40 18.5 4.2 0.03768 88 7.7 45 18.5 4.2 0.03768 88 8.2 50 18.8 4.2 0.03768 88 9 55 19 4.2 0.03768 88 6.9 60 18.6 4.2 0.03768 89 7.5 65 18.3 4.15 0.03768 81 7.5

127


RE /kJ

Wcomp./kJ COP

EXD evap.

EXD comp.

EXD cond.

EXD exp.

EXD total

Exergy Efficiency

% Exergy Efficiency

6.285 1.299 4.834 3.915 4.171 3.357 3.968 15.413 0.4011 40.1 6.232 1.507 4.135 3.876 4.318 3.327 3.994 15.516 0.4451 44.5 6.262 1.552 4.033 3.855 4.363 3.319 3.993 15.532 0.4593 45.9 6.258 1.582 3.954 3.883 4.431 3.338 3.982 15.636 0.4231 42.3 6.262 1.571 3.985 3.889 4.408 3.323 3.9752 15.596 0.4210 42.1 6.273 1.522 4.121 3.892 4.385 3.327 3.970 15.575 0.4004 40.04 6.217 1.597 3.891 3.849 4.431 3.319 3.9921 15.592 0.4440 44.4 6.262 1.786 3.506 3.890 4.578 3.300 3.986 15.756 0.4305 43.05 6.228 1.767 3.524 3.888 4.589 3.308 3.981 15.767 0.4164 41.64 6.183 1.744 3.544 3.892 4.566 3.297 3.987 15.743 0.3946 39.46 6.202 1.759 3.524 3.847 4.544 3.274 3.996 15.662 0.4453 44.53 6.266 1.804 3.471 3.864 4.600 3.278 3.981 15.725 0.4356 43.56 6.266 1.556 4.026 3.884 4.397 3.345 3.9843 15.612 0.4356 43.56

Figures (4.3-4.6), shows the exergy efficiency variation with respect to

evaporator temperature. It is observed that the highest exergy efficiency is found when

the system is 100% charged. This is due to the fact that, as observed the refrigerant

temperature (evaporator) is minimum when the system is 100% charged and therefore

the term Q�| �1 − ��

� | increases. The highest exergy efficiency is found to be 45.9% at

an evaporator temperature of 7oC when the system is 100% charged while the least

exergy efficiency was found to be 3.5% at an evaporator temperature of 24.5oC and

25% system charging. The reason for reduced exergy efficiency when the refrigerant

is only 25% charged is because the refrigerant temperature is almost same as the

reference temperature and hence exergy content is very small in the later case as

compared to the former case. Thus the results are consistent with the hypothetical

128

observations. However, as far as energy analysis is concerned. This particular result is

reverse as can be seen by Carnot cycle efficiency.

)( LH

LC

TT

TCOP

−= (5.9)

Fig.-4.3: Variation of exergy efficiency with respect to evaporator temperature when 25% of the total refrigerant is charged

0

2

4

6

8

10

12

14

16

24.5 24.5 21 21 20.1 20.1 20.8 20.8 21.1 21.1 22.1

Exe

rgy

Eff

icie

ncy /%

Evaporator Temperature /oC

129



0

5

10

15

20

25

23 20.3 21.5 21.3 19.6 20.3 22.5 21.4 19.4 17.3 19.5 19.5

Exe

rgy E

ffic

iency/%


0

2

4

6

8

10

12

14

16

18

20

22.9 24.3 24 23 22.5 22.3 24.5 22.6 22.2 24.8 23 24.7 25.7

Exerg

y E

ffic

ien

cy /%

Evaporator temperature /oC

130

Fig.-4.6: Variation of exergy efficiency with respect to evaporator temperature when 100% refrigerant is charged

From Tables (4.2-4.5) above, it is observed that the exergy destruction in the

compressor is found to be the highest as compared to other components i.e.

evaporator, condenser and expansion devices. This is attributed to the reciprocating

compressor being used in the system and the losses are huge due to friction between

the cylinder and piston rings. Although, lubricant is used to minimize these losses but

still frictional losses are more prominent in reciprocating compressors. Another

component of loss is due to the wire drawing effect at entry and exit through the valve

plates which changes the entropy of the system and increases the irreversibility of the

system. The exergy destruction in each case with respect to the percentage of

refrigerant charged is given in the order as below:

36

38

40

42

44

46

48

9 7 6.5 8 8.1 9 7 7.7 8.2 9 6.9 7.5 7.5

Exe

rgy E

ffic

ien

cy

/%


131

For 25% refrigerant charged

condDDevapDcompD EXEXEXEX )()()()( exp >>>


condDevapDDcompD EXEXEXEX )()()()( exp >>>


evapDDcondDcompD EXEXEXEX )()()()( exp >>>


condDevapDDcompD EXEXEXEX )()()()( exp >>>

The exergy destruction in the evaporator is found to be the least when the

system is 75% charged. This is because of the higher evaporator temperature

observed during the test and this reduces the term Q�| �1 − ��

� | in Eq. (4.6)

The variation of total exergy destruction with respect to the evaporator

temperature for four different cases viz. 25%, 50%, 75%, and 100% charging is shown

in Figs. (4.7-4.10) respectively. It is observed that the total exergy destruction is

comparable when the system is 75% and 100% charged and it is least when the

system is 25% charged. This is because the evaporator temperature is higher due to

insufficient amount of charging and is comparable to the reference temperature.

132

Fig.-4.7: Variation of total exergy destruction with respect to evaporator temperature when 25% of the total refrigerant is charged


0

2

4

6

8

10

12

14

16

24.5 24.5 21 21 20.1 20.1 20.8 20.8 21.1 21.1 22.1

Tota

l Exe

rgy

De

str

uctio

n /kW

Evaporator Temperature/oC

0

2

4

6

8

10

12

14

16

18

23 20.3 21.5 21.3 19.6 20.3 22.5 21.4 19.4 17.3 19.5 19.5

To

tal E

xe

rgy D

estr

uction

/kW


133



Tables (4.2-4.5), also shows that the average coefficient of performance is

highest when the system is 50% charged. This is because the refrigerating effect is

14.8

15

15.2

15.4

15.6

15.8

16

22.9 24.3 24 23 22.5 22.3 24.5 22.6 22.2 24.8 23 24.7 25.7

To

tal E

xerg

y D

estr

uctio

n /kW

Evaporator Temperature/oC

15.2

15.3

15.4

15.5

15.6

15.7

15.8

9 7 6.5 8 8.1 9 7 7.7 8.2 9 6.9 7.5 7.5Tota

l Exe

rgy D

estr

uctio

n /kW


134

higher in this case and the compressor consumers less work. It is also observed from

the tables that the exergy destruction is almost comparable when the system is 75%

and 100% charged while it is least when the system is 25% charged, this is due to the

fact that when the system is 25% charged, the evaporator temperature is higher and

therefore the term Q�| �1 − ��

� | is significantly low.

5.6 Conclusions

Exergy analysis is a technique to present the process and this further aid in

reducing the thermodynamic losses occurring in the process. This is an important tool

in explaining the various energy flows in a process and in the final run helps to reduce

losses occurring in the system. In this experimental study, a window air conditioning

system based on vapor compression cycle is modified for experimental analysis. The

system comprises of four components i.e. compressor, a capillary tube (expansion

device), a condenser and an evaporator and is having a cooling capacity of 24K BTU.

Based on the experiment testing following conclusions are drawn:

1. Although the quantity of refrigerant charged do affect the exergy losses but the

maximum losses in all the cases are in the compressor. This is attributed to the

frictional losses and losses due to wire drawing effect during suction and delivery

of the refrigerant. This will augment the study of tribology to exactly study the

friction characteristics and also the design aspects needs to be improved to

reduce the wire drawing effect to have efficient compressor.

135

2. It is observed that the total exergy destruction is comparable when the system is

75% and 100% charged and it is least when the system is 25% charged

because the evaporator temperature is very close to the reference temperature.

3. The average coefficient of performance is highest when the system is 50%

charged and this is because of higher refrigerating effect and reduced

compressor work.

4. The exergy efficiency of the system varies from 3.5% to 45.9% which is mainly

due to the variation of evaporator temperature.

5. The average values of the system exergy efficiency are more when the system

is 100% charged. These values show that the overall exergy performance is

better when the system is fully charged but the compressor work is the highest

in this case and the COP is also less as compared to other situations. When the

actual requirements are less the system should be operated with variable

refrigerant flow so as to achieve optimum balance between the exergy efficiency

and energy saving.

136

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