Report 1

CHAPTER 1

INTRODUCTION

Air-Conditioning refers to the process of altering the state of the ambient air

in a system to bring it to a desired value. This may be achieved by altering one or a

combination of physical properties of the air such as dry bulb temperature, relative

humidity, molecular composition, air velocity or by disinfecting the air of undesired

impurities.

1.1 IMPORTANCE OF AIR-CONDITIONING

The main aim of air-conditioning is to provide a sufficient volume of clean

air containing a specific amount of water vapour at a temperature capable of

maintaining a pre-determined atmospheric condition within a selected enclosure.

The main areas of application of air-conditioning are:

Comfort applications

Process applications

In comfort applications, the aim is to provide a thermally comfortable

atmosphere to the occupants of a room. Though the idea of comfort is a subjective

concept, it is widely accepted that a DBT of 240C is the ideal temperature for

humans to work in and at this temperature the individual has the highest

productivity.

Process applications encompass in it all the various industrial applications

where it is necessary to maintain a specified atmosphere for carrying out a

manufacturing process or for the storage of items. These applications include ice

manufacturing, food preservation, textile industry, photography industry, clean

rooms.

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1.2 HISTORY OF AIR-CONDITIONING

The methods of production of cold by mechanical processes are quite recent.

Long back in 1748, William Coolen of Glasgow University produced refrigeration

by creating partial vacuum over ethyl ether. But, he could not implement his

experience in practice. The first development took place in 1834 when Perkins

proposed a hand-operated compressor machine working on ether. Then in 1851

came Gorrie’s air refrigeration machine, and in 1856 Linde developed a machine

working on ammonia.

The pace of development was slow in the beginning when steam engines

were the only prime movers known to run the compressors. With the advent of

electric motors and consequent higher speeds of the compressors, the scope of

application of refrigeration widened. The pace of development was considerably

quickened in 1920 decade when duPont put in the market a family of new working

substances, the fluoro-cholro derivatives of methane, ethane, etc.- popularly known

as chloro flurocarbons or CFCs- under the name of Freons. Recent developments

involve finding alternatives or substitutes for Freons, since it has been found that

chlorine atoms in Freons are responsible for the depletion of ozone layer in the

upper atmosphere. Another noteworthy development was that of the ammonia-water

vapour absorption machine by Carre. These developments account for the major

commercial and industrial applications in the field of refrigeration.

A phenomenon called Peltier effect was discovered in 1834 which is still not

commercialized. Advances in cryogenics, a field of very low temperature

refrigeration, were registered with the liquefaction of oxygen by Pictet in 1877.

Dewar made the famous Dewar flask in 1898 to store liquids at cryogenic

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temperatures. Then followed the liquefaction of other permanent gases including

helium in 1908 by Onnes which led to the discovery of the phenomenon of

superconductivity. Finally, in 1926, Giaque and Debye independently proposed

adiabatic demagnetization of a paramagnetic salt to reach temperatures near absolute

zero. Two of the most common refrigeration applications, viz., a window-type air

conditioner and a domestic refrigerator.

The application of air-conditioning for the industrial purposes has opened a

new era in the air-conditioning industry. Air-conditioning is commonly used now-a-

days for the preservation of food, in automobiles and railways, jute and cloth

industries and many others. Its varied applications have opened a new field for the

air-conditioning engineers to solve the difficult problems with full success.

Air-conditioning is a field of work in which work never stagnates. Air-

conditioning is commonly used to ease man’s environmental problems on Earth as

well as in space. The adverse problems of space environment are also successfully

solved with the advanced knowledge of air-conditioning which has made the space

travel successful.

1.3 AIR-CONDITIONING: INDIAN SCENARIO

The refrigeration and air-conditioning industry in India got the impetus to

progress with the dawn of independence in India. This industry has achieved

phenomenal growth in less than three decades in our country. The annual output has

increased from 800 tons in early fifties to about 80,000 tons in 1970. This industry

now produces a wide range of light and heavy equipments which has reduced import

from 50 percent to 5 percent. Air-conditioning industry now produces packaged air-

conditioners, water chilling units upto 200 tons capacity, hermetic compressors, air

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handling units, cooling coils and variety of other equipment apart from well known

items like room air-conditioners, refrigerators, deep freezers, food display units and

water coolers. This forms a solid base to satisfy practically all needs of refrigeration

and air-conditioning equipments in the country.

Indian engineers can confidently tackle any practical problem in the design

of refrigeration and air-conditioning equipments used for different purposes. They

have proved their competence by successfully designing complicated jobs like

cooling system for concrete dams, defense installations, dairy milk chilling units,

railway air-conditioning, quick freezing plants for fish and deep freezers for storage

of fish, air-conditioning and refrigeration for ships and many others. There has been

an increase in exports from Rs. 4 lakhs in 1955 to Rs. 60 lakhs in 1970. Presently,

the industry has been rightly placed on a priority list by the Government of India and

national panel has been formed to plan its development.

Indian atmospheric conditions are varied in different parts of the country.

Particularly the summer conditions in India are quite uncomfortable in most parts of

the country. And winter conditions are uncomfortable in a few parts of the country.

There is no doubt that air-conditioning will become a necessity for Indians in a

coming few decades with rapid industrial development and with the economic

growth of the country.

Being most developing industry in the country, air-conditioning engineers

have better prospects in the future. This industry will be able to play a more

significant role in the nation’s industrial and economic development.

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1.4 A REFRIGERATING MACHINE: THE SECOND LAW

INTERPRETATION

A refrigerating machine is a device which will either cool or maintain a body

at a temperature below that of the surroundings. Hence, heat must be made to flow

from a body at low temperature to the surroundings at high temperature. However,

this is not possible on its own. We see in nature that heat spontaneously flows from

a high temperature body to a low temperature body.

The second law of thermodynamics, like the first law, is based on the

observations of several actually existing processes and devices in nature. Figure 1.1

shows the schematic diagram of an actual refrigeration system which works on the

well-known vapour compression cycle. Most refrigeration devices/plants, including

air conditioners and refrigerators work on this cycle only.

The heat and work interactions of the processes of the cycle are as follows:

1. Heat Qo is absorbed in the evaporator by the evaporation of a liquid

refrigerant at a low pressure po and a corresponding temperature To.

2. The evaporated refrigerant vapour is compressed to a high pressure pk in

the compressor consuming work W. the pressure after compression is such

that the corresponding saturation temperature Tk is higher than the

temperature of the surroundings.

3. Heat Qk is then rejected in the condenser to the surroundings at high

temperatures Tk.

The application of the first law § δQ = § δW , to the cycle gives:

5

-Qk + Qo = -W

Qk – Qo = W

This also represents the energy balance of the system in Figure 1.1 obtained

by drawing a boundary around it.

1.5 ROOM AIR-CONDITIONER

Figure 1.2 shows the schematic diagram of a typical window-type room air-

conditioner, which works according to the principle described below:

Consider that a room air is maintained at constant temperature of 25oC. In the

air conditioner, the air from the room is drawn by a fan and is made to pass over a

cooling coil, the surface of which is maintained, say at a temperature of 10 oC. after

passing over the coil, the air is cooled (for example, to 15oC) before being supplied

to the room. After picking up the room heat, the air is again returned to the cooling

coil at 25oC.

Now in the cooling coil, a liquid working substance called a refrigerant such

as CHClF2 (monocholoro-difluoro methane) , also called Freon 22 by trade name, or

simply Refrigerant 22 (R 22), enters at a temperature of, say 5oC and evaporates,

thus absorbing its latent heat of vapourization from the room air. This equipment in

which the refrigerant evaporates is called an evaporator.

After evaporation, the refrigerant becomes vapour. To enable it to condense

back and to release the heat – which it has absorbed from the room while passing

through the evaporator – its pressure is raised by a compressor. Following this, the

high pressure vapour enters the condenser. In the condenser, the outside atmospheric

air, say at a temperature of 45oC in summer, is circulated by a fan. After picking up

the latent heat of condensation from the condensing refrigerant, the air is led out into

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Figure 1.1 Schematic of an actual vapour compression refrigeration system.

Figure 1.2 Schematic diagram of a room air-conditioner.

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the environment, say, at a temperature of 55oC. The condensation of refrigerant may

occur, for example at a temperature of 60oC.

After condensation, the high pressure liquid refrigerant is reduced to low

pressure of the evaporator by passing it through a pressure reducing device called

the expansion device, and thus the cycle of operation is completed. A partition wall

separates the high temperature side of the condenser from the low temperature side

of the evaporator.

1.6 VAPOUR COMPRESSION REFRIGERATION CYCLE

The vapour compression refrigeration systems are nowadays used to meet all

kinds of refrigeration needs ranging from a small domestic unit of 0.5 ton capacity

to an air-conditioning system of 1000 tons. The most difference in theory and

treatment of vapour refrigeration system as compared to the air refrigeration system

is that, the vapour alternately undergoes a change of phase from vapour to liquid and

liquid to vapour during the completion of a cycle. The latent heat of vapourization is

utilized for carrying heat from the refrigerator which is quite high compared with the

air-cycle, which depends only upon the sensible heat of the air.

The substances used do not leave the plant but are circulated through the

system alternately after condensing and re-evaporating. During evaporating, it

absorbs its latent heat from the brine which it is used for circulating around the cold

chamber. In condensing, it gives out its latent heat to the circulating water of the

cooler, the machine is, therefore, known as Latent Heat Pump. It absorbs its latent

heat from the brine and gives out in the condenser.

The pressure is maintained at different levels in two parts of the system by

the expansion valve (high side float valve). The function of the compressor is to

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increase the pressure and temperature of the refrigerant above atmospheric which

will be ready to dissipate its latent heat in the condenser. In passing through the

condenser, the refrigerant gives up the heat which is absorbed in the evaporator plus

the heat equivalent of the work done upon it by the compressor. This heat is

transferred to the air or water which is used as cooling medium in the condenser.

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CHAPTER 2

LITERATURE REVIEW

Air-conditioning is a relatively young and growing industry compared to

most other industries. When the idea of air-conditioning systems was first

introduced, it was considered a luxury due to the unavailability of economical

technologies that could make the machines cost competitive. The earliest air-

conditioners used simple thermodynamic cycles to achieve the required variation in

temperature and humidity. With the advancement in technology, air-conditioning

engineers were able to manufacture affordable and efficient systems and this has

made air-conditioning a necessity in today’s world.

The process of system designing of air-conditioning involves the basic steps

of analysis of requirements, selection of refrigerant and refrigeration cycle, the

designing of refrigerating equipments and the assembly and their installation [4].

Presently the most commonly used refrigeration process is the vapour compression

cycle. Refrigerants such as R-12 and R-22 which are being used in air-conditioning

systems are being slowly replaced by alternative refrigerants such as R-410 and R-

134a which pose much lesser threat to the environment [3]. Some of these

refrigerants are quite similar in their behaviour to the existing commonly used

refrigerants and hence require minimal changes in equipment design [5]. In other

cases the process of equipment design followed is similar to that for the existing

refrigerants.

In most instances air-cooled condensers for home and industrial appliances

are suitably drained cross-flow exchangers in which a cooling-air stream is utilized

to reduce the moisture content of the process-air stream. A procedure to calculate the

wet effectiveness of these exchangers was developed based on an original adaptation

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of the effectiveness vs. number of transfer units method developed for dry

exchangers [6]. Calculated results were validated successfully by employing the

experimental results obtained in a dedicated test facility. The model proposed was

thus proved accurate and reliable, both for comparing the performances of different

condensers and for analyzing the influence of design parameters. In fact, the

condensation rate is directly related to the task performed by the condenser, the total

heat transfer rate is connected to the rate of energy consumption and the ratio

between sensible and total heat transfer rates is a measure of the energy losses (the

ideal condenser does not exchange sensible heat on the process side)[6].

Airside heat and mass transfer and fluid flow characteristics of a wavy-

finned-tube direct expansion air coil under cooling and dehumidifying condition

were experimentally investigated [7]. Experiments were carried out to study the

effects of operating conditions such as: air temperature, air relative humidity, air

face velocity, and evaporator pressure on the airside performance (cooling capacity,

dehumidification capacity, pressure drop, and heat transfer coefficient) of the coil.

Charts for coil wet conditions, partially wet or totally wet, were conducted to

identify the coil wet conditions in terms of the operating conditions. Two

techniques, enthalpy potential method and equivalent dry-bulb temperature method,

were used to analyze the data and to deduce correlations for Colburn factors for the

different coil wet conditions.

For the development of empirical relations for the convective heat transfer

coefficient of refrigerants during condensation inside a horizontal cylindrical tube

test were carried out [1]. The results obtained have been widely accepted as reliable

and the errors obtained are within acceptable limits.

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Similarly for evaporation experiments were carried out for determining the

expressions for heat transfer coefficient of refrigerant R-22. Due to the similarity

between R-22 and R-134a, the expressions can be successfully utilized for

determining the heat transfer coefficient of R-134a as well [2].

The designing of the capillary tube is essentially that of cut and trial. Various

possible diameter and length combinations are available for the same required

pressure drop. Choice is made depending on the desired design conditions.

Selection of compressor is the first and the most vital task in the system

designing of air-conditioning systems. Compressors are available from

manufacturers in standard sizes and operating modes. Depending on the required

conditions, the most suited compressor is selected. The dimensional features of the

compressor can be calculated by analyzing the refrigerant flow process. But certain

assumptions have to be made since the number of relations available is limited and

the unknown variables are many.

The ASHRAE Fundamentals Handbook is the most essential literature material

that any air-conditioner or refrigeration system designer must possess. It gives complete

technical knowledge on all the fundamentals involved in refrigeration and air-

conditioning and guides in the design procedure for system designing. The handbook

also gives refrigerant tables and enthalpy diagrams for most refrigerants. A compilation

of the most widely accepted co-relations for solving heat transfer problems are available

in the book.

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2.1 OBJECTIVES

The main objectives of this project are as follows:

1. To study the working and analysis of vapour compression refrigeration cycle.

2. To design a window air-conditioner using R-134a as refrigerant and study its

feasibility as a substitute for R-22.

2.2 PROBLEM FORMULATION

To design a window type air-conditioning system for summer air-conditioner

working on an eco-friendly refrigerant R-134a which is being considered a substitute

for R-22 (presently being used in air-conditioning systems).

2.3 METHODOLOGY

The design procedure for an air-conditioning system mainly involves of the

selection of refrigerant and the designing of the various system components. The design

procedure can be briefly described in the following steps.

1. Estimation of cooling load. The cooling load is first determined by determining

the total heat load in the space to be air-conditioned. Based on the cooling load,

the tonnage requirement is determined.

2. Selection of refrigerant. The refrigerant to be used in the refrigeration cycle is

to be selected depending upon the design inside and outdoor conditions of Dry

Bulb Temperature and Relative Humidity. This depends on the application.

Presently, for window air-conditioning R-22 is the refrigerant of choice. But due

to its global warming and ozone depletion potential it is being replaced by

alternative refrigerants such as R-410a, R-134a and R-407C. The present project

carries out the design using R-134a as refrigerant.

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3. Selection of compressor. Based on the operating pressures, the compressor is

designed and selected. The important dimensions such as the bore and the length

of the compressor are calculated upon assuming parameters like speed of motor

and clearance.

4. Design of condenser. The condenser is designed based on the heat transfer to be

achieved between the flowing air and the refrigerant. The length, fin

arrangement and number of bends are determined.

5. Design of expansion valve. The expansion valve chosen is the capillary tube.

Based on the pressure drop that is to be achieved, the total length and the

diameter have been calculated.

6. Design of evaporator. The design of evaporator used in window type air-

conditioning is similar to condenser design. The length of the evaporator tube,

arrangement of fins, number of bends and number of rows have been calculated.

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CHAPTER 3

DESIGNING OF COMPRESSOR AND EXPANSION VALVE

The compressor is the main power consuming device in a vapour

compression refrigerating cycle. The main task of the compressor is to increase the

pressure of the refrigerant vapour that comes out of the evaporator so that the

condensation in the condenser can be achieved at lower temperatures.

The main function of expansion valves in a refrigeration system is to

decrease the pressure of the liquid refrigerant, thereby decrease the temperature, and

bring it down to the evaporator pressure. Additionally, expansion valves perform the

important function of controlling the flow of refrigerant to the evaporator to adjust

to the varying load demands. The main working principle behind expansion valves

is to offer resistance to the flow of refrigerant which causes pressure drop in the

refrigerant.

Before designing these components, the analysis of the refrigeration cycle is

to be performed to determine the operating conditions.

3.1 VAPOUR COMPRESSION CYCLE CALCULATIONS

The vapour compression refrigeration cycle has been selected for the

designing of the window air-conditioner. In the vapour compression refrigeration

cycle, high pressure, high temperature gas coming out of the compressor is

condensed in the condenser using outside air as coolant. A blower is used to force

the air over the condenser to improve the convection coefficient. The vapour

entering the condenser is assumed to be saturated and by the time the fluid reaches

the end of the condenser, the refrigerant is assumed to be completely condensed.

The liquid is then passed through a capillary tube to reduce the pressure which

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causes the temperature also to fall rapidly till the pressure becomes equal to the

evaporator pressure. The low pressure, low temperature fluid is then passed through

the evaporator where clean air is passed over it. The air loses heat to the low

temperature refrigerant and is directed to the room where low temperature is

required. The refrigerant vapour coming out of the evaporator is sent to a

reciprocating compressor where it is compressed to a high pressure vapour and the

cycle is completed.

The refrigerant properties and analysis of the refrigeration process is

explained below. The air-conditioning system is being designed for a load of 1.5

tons. The outside air temperature is 400C and the room temperature is 320C which is

wished to be brought down to 240C. The evaporator and condenser temperatures

have been chosen to be 160C and 520C.

3.1.1 Selection of refrigerant

The refrigerant chosen for this cycle is R-134a. Presently R-22 is widely

used in air-conditioning systems. But it poses a a threat to the environment. R-134a

is a suitable alternative. R-134a also gives a better performance as compared to R-22

as has been established.

The various properties of R-134a are given below. They were obtained from

ASHRAE Handbook of Fundamentals

R134a (CH2F-CF3)

Chemical name: 1,1,1,2-Tetrafluoroethane (R134A)

CAS Number: 811-97-2

UN3159

Common names: Freon 134a; Ethane, 1,1,1,2-tetrafluoro-; Halocarbon 134a;

1,2,2,2 Tetrafluoroethane; HFC-134a;

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Molecular weight: 102.03 g/mol

Solid phase

Melting point (1.013 bar): -101 °C

Liquid phase

Liquid density (1.013 bar and 25 °C (77 °F)): 1206 kg/m3

Boiling point (1.013 bar): -26.6 °C

Latent heat of vaporization (1.013 bar at boiling point) 215.9 kJ/kg

Vapor pressure (at 50 °C or 122 °F): 13.2 bar

Vapor pressure (at 20 °C or 68 °F): 5.7 bar

Vapor pressure (at 5 °C or 41 °F): 3.5 bar

Vapor pressure (at 15 °C or 59 °F): 4.9 bar

Critical point

Critical temperature: 100.9 °C

Critical pressure: 40.6 bar

Critical density: 512 kg/m3

Triple point

Triple point temperature: -103.3 °C

Gaseous phase

Gas density (1.013 bar at boiling point): 5.28 kg/m3

Gas density (1.013 bar and 15 °C (59 °F)): 4.25 kg/m3

Compressibility Factor (Z) (1.013 bar and 15 °C (59 °F)): 1

Specific gravity (air = 1) (1.013 bar and 15 °C (59 °F)): 3.25

Specific volume (1.013 bar and 15 °C (59 °F)): 0.235 m3/kg

Heat capacity at constant pressure (Cp) (1.013 bar and 25 °C (77 °F)): 0.087

kJ/(mol.K)

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From the refrigerant tables, the properties for R-134a at the operating

temperatures are noted down and is shown in Table 3.1

Table 3.1 Properties of R-134a at 160C and 520C.

Temperature 0C

Saturation Pressure MPa

hf kJ/kg

hg kJ/kg

sf kJ/kgK

sg

kJ/kgKvf m3/kg

vg

m3/kg

52 1.3851 124.5 273.24 0.4432 0.9004 0.000914 0.0142

16 0.50416 71.69 256.22 0.2735 0.9116 0.000806 0.0405

3.1.2Analysis of the cycle

From Table 3.1, we know the following values of enthalpies and entropies.

h1 = 256.22 kJ/kg

h2 = 273.24 kJ/kg

h3 = 124.58 kJ/kg

h4 = 71.69 kJ/kg

s1 = 0.9116 kJ/kgK

s2 = 0.9004 kJ/kgK

s3 = 0.4432 kJ/kgK

s4 = 0.2735 kJ/kgK

The refrigeration cycle being employed is represented on a T-s diagram in

Figure 3.1. It is required to calculate the C.O.P of the system and the mass flow rate.

For, this the refrigerant properties at points ‘b’ and ‘a’ also need to be known.

The expansion process 3-b is assumed to be an isenthalpic process.

Therefore, hb = h3.

hb = 124.58 kJ/kg

It can be seen from the graph that during the compression process, s2 = sa.

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Figure 3.1 Simple vapour compression cycle operating between 160C and 520C.

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sa = 0.9004 kJ/kgK

Now, assume the dryness fraction of the refrigerant as it leaves the

evaporator to be x. Thus we have,

(3.1)

x = 0.9824

Similarly, the value of enthalpy at point a can now be calculated knowing the

dryness fraction

(3.2)

ha = 252.9722 kJ/kg

The C.O.P of the system is defined as the refrigerating effect provided by the

system per unit work done by the compressor.

(3.3)

The refrigerating effect = ha – hb = 128.392 kJ/kg

Compressor work = h2 – ha = 20.2678 kJ/kg

C.O.P = 6.3347

Now, for calculating the mass flow rate we have,

3.5 T = m.(ha – hb) (3.4)

m = 0.04089 kg/sec

3.2 COMPRESSOR DESIGN

For the present design problem, a hermetically sealed type reciprocating

compressor is chosen. A hermetically sealed compressor has the motor, compressor

and all the moving parts in a sealed, gas-tight casing. This is generally employed in

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low capacity refrigeration systems where the total system size is also an important

consideration.

For the designing of the compressor, the following assumption was made

Speed of motor = 1500 r.p.m

Initially high speed compressors were assumed for the design. This was

found to be unnecessary for the required application as the dimensions were

calculated to be well within the range of acceptable values. The assumption

for the speed of motor was reduced till an optimized size of the compressor

was obtained.

3.2.1 Work done in a reciprocating compressor

The p-v diagram for the machine cycle of a reciprocating compressor is

shown Figure 3.2 along with the skeleton diagram of the cylinder and piston

mechanism. When the piston is in the extreme left position of the inner dead centre

(IDC), the volume occupied by the gas is VC = V3 called the clearance volume, i.e

the volume between the IDC position of the piston and the cylinder head. As the

piston moves outward, the clearance gas expands to 4, where the pressure inside the

cylinder is equal to the pressure at the suction flange of the compressor. As the

piston moves further, the suction valve S opens and the vapour from the evaporator

is sucked in till the extreme right position of the outer dead centre is reached. At this

point the volume occupied by the gas is V1. The stroke or swept volume or piston

displacement is

, (3.5)

where, D is the bore or diameter and L is the stroke, i.e, the distance traveled

by the piston between IDC and ODC of the cylinder. At 1, the suction valve closes

as the piston moves inwards and the compression begins. At 2, the pressure in the

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cylinder is equal to the pressure at the discharge flange of the compressor. A further

inward movement of the piston results in the pressure in the cylinder exceeding the

condenser pressure. This opens the discharge valve D and the vapour from the

cylinder flows into the condenser till the piston again reaches the IDC position. Gas

equal to the clearance volume VC remains in the cylinder and the cycle is repeated.

The work done for compression for the machine cycle is given by the cycle

integral of pdV. Hence,

(3.6)

(3.7)

W = Area 1-2-3-4

It will be seen that this area is also expressed by the term - § Vdp. Hence,

(3.8)

where, m is the mass of the suction vapour. Thus, the specific work in a

reciprocating compressor is given by

(3.9)

where, 1 and 2 are the limits of integration from the suction state 1 to the discharge

state 2 as indicated in Figure 3.2

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Figure 3.2 Cylinder and piston mechanism and p-V diagram for a reciprocating

compressor.

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3.3 CALCULATION OF VOLUMETRIC EFFICIENCY

The volume of refrigerant sucked in by the compressor is always less than

the maximum possible volume which is represented by the swept volume.

Volumetric efficiency ηv, is the term that identifies the difference in the displacement

or swept volume VP in built in the compressor and volume VS of the suction vapour

sucked and pumped. It is expressed by the ratio, ηv = VS/VP

3.3.1 Clearance Volumetric Efficiency

A gap of (0.005L + 0.5) mm is provided between the IDC position of the

piston and cylinder head in reciprocating compressors to provide for thermal

expansion and machining tolerances. This space, together with the volume of the

dead space between the cylinder head and valves, form the clearance volume. The

ratio of the clearance volume to the swept volume VP is called the clearance factor

C, i.e.,

(3.10)

This factor is normally less than 5 percent.

The purpose of clearance in reciprocating compressors is to reduce the

volume of the sucked vapour, as can be seen from Figure 4.1. The gas trapped in the

clearance space expands from the discharge pressure to the suction pressure and thus

fills a part of the cylinder space before suction begins. Considering only the effect of

clearance on volumetric efficiency, we have from Figure 4.1, for clearance

volumetric efficiency,

(3.11)

The volume occupied by the expanded clearance gases before suction begins, is

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(3.12)

(3.13)

(3.14)

It is seen that lower the value of γ, lower the ηCV, and higher the value of γ,

higher the ηCV. The expression for volumetric efficiency can also be written in the

form,

(3.15)

(3.16)

(3.17)

For the designing of the compressor, the following assumption was made

Clearance = 4%

A clearance of 4% of the total length of the compressor is generally

employed for compressors operating in this range.

3.3.2 Effect of Valve Pressure Drop

For the flow of any fluid, the pressure must drop in the direction of flow.

Both suction and discharge valves will open only when there is a pressure drop

across them. The effect of these pressure drops on the indicator diagram of the

compressor is shown in Figure 3.3. It is seen that as a result of throttling or pressure

drop on the suction side the pressure inside the cylinder at the end of the suction

stroke is pS while the pressure at the suction flange is p1. The pressure in the cylinder

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rises to the suction flange pressure p1 only after the piston has traveled a certain

distance inward during which the volume of the fluid has decreased from (VP + VC)

to V1. Assuming the compression index to be n instead of γ, as the compression

process is also polytropic due to heat exchange with cylinder walls and friction, we

have,

(3.18)

The expression for the volumetric efficiency becomes

(3.19)

(3.20)

(3.21)

Considering the effect of pressure drop at the discharge valve as well, it can

be shown that the expression for volumetric efficiency is

(3.22)

For the designing of the compressor, the following assumptions were made

Suction pressure drop = 0.02 MPa

Delivery pressure drop = 0.04 MPa

The suction and delivery pressure drops include the pressure drops in the

suction and delivery lines as well as the suction and delivery valves. The

values chosen lie within the range of pressure drops generally observed for

refrigeration units of comparable capacities.

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Figure 3.3 Effect of valve pressure drops.

3.3.3 Leakage Loss

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The effect of leakage past piston rings and under the suction valve elements

is normally accounted for by allowing 1.5 percent leakage per unit of the compressor

ratio r (p2/p1). Old worn out compressors tend to have more leakage and hence they

lose their cooling capacity.

3.3.4 Overall Volumetric Efficiency

Considering the effect of wire-drawing at the valves, polytropic

compression, re-expansion and leakage, we may write the expression for the overall

or total volumetric efficiency as follows,

(3.23)

Based on the discussions made above, the volumetric efficiency of the

assumed compressor, considering the leakage losses and pressure drops, was

calculated.

C = 0.04

ps = p1 – suction pressure drop = 0.48416 MPa

pd = p2 + delivery pressure drop = 1.4251 MPa

n = 1

ηv = 0.841524

3.4 MINIMUM REQUIRED SUCTION PRESSURE FOR THE

COMPRESSOR

Figure 3.4 shows the nature of variation of the p-v diagram of a reciprocating

compressor with the decrease in suction pressure for constant discharge pressure. It

is seen that with decreasing suction pressure, or increasing pressure ratio, the suction

volume VS and hence the volumetric efficiency ηV decrease until both become zero

28

Figure 3.4 Decrease in suction volume of a reciprocating compressor with

decreasing evaporator pressure.

29

at a certain low pressure p1”. Thus the refrigerating capacity of a reciprocating

compressor tends to zero with decreasing evaporator pressure. It can be observed

that the clearance volumetric efficiency will be zero for a pressure ratio given by

(3.24)

For a given discharge pressure p2, the above expression gives the value of

p1min, the lowest pressure possible for obtaining any capacity from a given

compressor. The minimum required suction pressure for the compressor being

designed to have any capacity has been calculated.

pmin = 0.054811 MPa

The minimum pressure required for the compressor to function is well below

the designed suction

3.5 PRINCIPAL DIMENSIONS OF RECIPROCATING COMPRESSORS

The principal dimensions of a reciprocating compressor are the bore D and

the stroke L. These are to be decided in conjunction with the rpm N or mean piston

speed Cm. Thus there are three parameters, D, L and N or Cm to be selected. Stroke

to bore ratio is a very important consideration in compressor design. However, only

one equation is available for design, i.e., the equation for the suction volume

(3.25)

where,

(3.26)

(3.27)

(3.28)

30

Assumptions are therefore necessary for two parameters. To minimize the

piston force and hence the inertia force, a large value of θ is chosen in high-pressure

compressors so that the diameter is smaller. The following values are adopted in

practice,

Vacuum pumps and high-speed air compressors ≤ 0.5

Fluorocarbon compressors ≈ 0.8

Ammonia compressors ≈ 1.0

High-pressure compressors = 0.46

For the designing of the compressor, the following assumptions were made

Stroke/Bore (L/D) = 0.8

The L/D ratio is assumed on the basis of the refrigerant employed and the

area of application. For small capacity units using fluorocarbons with

moderate compressor speeds, which is the present case, an L/D ration of 0.8

gives satisfactory performance.

The volumetric flow rate can be evaluated since the mass flow rate and the

specific density of the refrigerant are known.

(3.29)

Vp = 0.118055 m3/min

The volumetric flow rate can also be equated to the dimensional parameters

of the compressor if the speed of the motor is known.

(3.30)

Therefore, D = 50 mm

L = 40 mm

31

3.6 CAPILLARY TUBE

Capillary tubes are widely used in refrigeration systems having low loads

and systems using hermetically sealed compressors. A capillary tube is simple in

construction. It is a long tube with an extremely small diameter. It gives good

resistance to refrigerant flow and this can be used as a pressure reducing device to

meter the flow of refrigerant given to the evaporator. The selection of capillary

depends upon the application and the anticipated range of operating conditions; i.e.,

reversal of flow must not take place at the capillary inlet. When the operating

condition can be achieved with minimum resistance to flow, the capillary is said to

be working at its highest efficiency.

The small diameter of the tube offers heavy frictional resistance to the flow

of the refrigerant and this gives the required pressure drop. The resistance is directly

proportional to the length and inversely proportional to the diameter.

The rate of flow for a selected capillary tube is a function of the pressure

differential between the condenser and the evaporator. As the load increases in

summer, the tube supplies more quantity of flow as an effect of increased condenser

pressure with air-cooled condensers used on domestic units. Similarly when the load

on the unit is reduced in winter, the flow through the tube decreases as an effect of

decreased condenser pressure. The capillary is a self-compensating device over a

limited range of pressure difference. The relationship between the load, diameter

and the length is given Table 3.2

3.6.1 Advantages and disadvantages of capillary tube

The main advantages of using a capillary tube are enlisted below:

It is simple in construction and no maintenance is required

32

Table 3.2 Relationship between load, diameter and length.

H.P of

motor

Load in

kCal/hr

Suction

Temperature 0C

Tube diameter

mm

Tube length

cm

1/3 700 -3.33 1 90

1/3 700 -3.33 1.25 150

¼ 540 -3.33 1 120

¼ 540 -3.33 1.25 195

¼ 400 -3.33 1 180

1/5 400 -3.33 1.25 285

When the compressor stops, the refrigerant continues to flow from high

pressure side to low pressure side until the pressure is equalized. This

requires less starting torque to start the compressor. So a low starting torque

motor can be used with these units

System using this device does not require receiver

Its cost is also considerably compared with other devices

The main disadvantages associated with this device are:

The refrigerant must be free from moisture and dirt, otherwise it will choke

the tube and stop the flow of refrigerant

This tube cannot be used with highly fluctuating loads

For the designing of the expansion valve, capillary tube was chosen because of

its applicability in the present design problem which is that of window air-

conditioning.

33

3.6.2 Capillary tube sizing

For determining the sizing of the capillary tube, it is important to first

understand the performance factors of the tube. This is the pressure drop occurring

in the tube. The pressure drop in the capillary tube is due to two main reasons

(a) Friction, due to fluid viscosity, resulting in frictional pressure drop.

(b) Acceleration, due to the flashing of the liquid refrigerant into vapour, resulting

in momentum pressure drop.

The cumulative pressure drop must be equal to the difference in pressure at

the two ends of the tube. The mass flow through the capillary tube will, therefore

adjust so that the pressure drop through the tube just equals difference in pressure

between the condenser and the evaporator. For a given state of the refrigerant , the

pressure drop is directly proportional to the length and inversely proportional to the

bore diameter of the tube.

For achieving the desired flow rate and pressure drop, more than one design

of capillary tube is possible since there are two variable parameters upon which the

design is dependant. But, the same design cannot be used for varying condenser and

evaporator pressures.

The sizing of a capillary tube implies the selection of bore and length to

provide the desired flow for the design condenser and evaporator pressures. The

method employed by manufacturers is usually that of cut and try. The principle of

design based on methods proposed by Stocker and Hopkins and Copper [3] is

presented here.

A capillary tube of a particular bore diameter D is first selected. Step

decrements in pressure are then assumed and the corresponding required increments

34

of length calculated. These increments can be totaled to give the complete length of

the tubing for a given pressure drop.

For the designing of the capillary tube, a bore size of 2 mm has been

considered. Also, as per the refrigeration cycle chosen, the state of the refrigerant

entering the capillary tube is saturated liquid.

The mass flow rate m is known. The condenser and evaporator temperatures

are Tk and Po and the corresponding pressures are Pk and P0 respectively. Divide this

temperature drop into a smaller number of parts. Let the corresponding pressure

drops be Δp1, Δp2 and Δp3…. etc as shown in Figure 3.5 and Figure 3.6. Now there

are two approaches to design

(a) Isenthalpic expansion as shown by line k-a

(b) Adiabatic or Fanno-line expansion as shown by line k-b

In actual practice expansion takes place adiabatically, viz., according to

Fanno-line flow. Nevertheless, it may be noted that in the first steps of pressure

drop, there is not much difference between the isenthalpic and Fanno-line flow.

Thus for the present design problem, the isenthalpic process is being assumed. The

steps of calculations to be followed in both the cases are the same and are as follows

for the first element.

1. Determine the quality at the end of the decrement assuming

isenthalpic flow. Then at point 1 at pressure p1. For the present calculation, only a

single step pressure drop is considered between the condenser and evaporator

pressure

(3.31)

h3 = h2

35

(3.32)

x = 0.286620

2. Determine the specific volume

(3.33)

(3.34)

v3 = 0.0121832

3. Calculate the velocities from the continuity equation at both the

ends of the element

(3.35)

u/v = m/A = G = Constant

where G is the mass velocity.

4. Determine the pressure drop due to the acceleration, ΔpA, from the

momentum equation

(3.36)

whence,

(3.37)

ΔPA = - 1.9109 MPa

5. Determine the pressure drop due to the friction, Δpf from

(3.38)

ΔP = 0.94094 MPa

ΔPF = 2.85284 MPa

36

Figure 3.5 Incremental pressure drops in a capillary tube.

Figure 3.6 T-s graph showing isenthalpic expansion of R-134a.

37

6. Equate the required frictional pressure drop to

(3.39)

where, ρ = 1/v

ΔL = Length of the element

Substituting, m = ρ.u.A,

ρ.u = m/A = G, we have

(3.40)

where, Y = G/2D

from which the length ΔL can be calculated. For this purpose, the mean values of u

and f for the liquid and vapour phases present may be taken for the section. The

friction factor is a function of Reynolds number which in turn is expressed as

(3.41)

where, Z = DG

and, µ = Dynamic viscosity

Niaz and Davis have proposed the following correlation for evaluating the friction

factor,

(3.42)

Re = D.G/μ = 181112.6565

f = 0.324/Re0.25 = 0.0157

ΔPF = 8.109161 x 106 x ΔL

ΔL = 351.681 mm

38

CHAPTER 4

DESIGNING OF CONDENSER

The main task of the condenser is to remove heat from the refrigerant carried

from evaporator and added by compressor and convert the refrigerant vapour to

fluid. Generally, the vapour at the discharge from the compressor is superheated.

Desuperheating of the vapour takes place in the discharge line and the first few coils

of the condenser. It is followed by condensation of the vapour at the saturated

discharge temperature or condensing temperature. In some condensers sub-cooling

may also take place near the bottom where there is only liquid.

The loading on the condenser part per unit of refrigeration is called heat

rejection ratio.

(4.1)

(4.2)

Thus the heat rejection ratio depends on the C.O.P which in turn depends on

the condenser and evaporator temperatures.

4.1 TYPES OF CONDENSERS

The type of a condenser is generally characterized by the cooling medium used.

There are three types of condensers:

Air cooled condensers

Water cooled condensers

Evaporative condensers

39

4.2 AIR COOLED CONDENSERS

In air-cooled condensers, heat is removed by air using either natural or

forced circulation. The condensers are made of steel, copper or aluminium tubing

provided with fins to improve air side heat transfer. The refrigerant flows inside the

tube and the air outside.

Air-cooled condensers are designed for condensing temperatures of 15oC to

20oC above the temperature of entering air. Natural convection condensers are used

only in small capacity machines, such as refrigerators and small water coolers which

use vertical wire and tube or plate and tube construction with natural circulation, and

forced convection condensers are used in window-type and package air conditioners

which have tubes with 5 to 7 fins per cm and use forced circulation of air. The

current practice in the forced convection type is to use 10 to 15 m2 of the total

surface area per ton of refrigeration based on 2 to 5 m/s face velocity of air over the

coil.

The forced convection condensers are further classified into:

Chassis mounted condenser

Remote air-cooled condenser

The chassis mounted condensers are mounted on the same base of the

compressor and motor. In small units the compressor is belt driven from the motor

and the blower required to force the air through the condenser is mounted on the

shaft of the motor.

4.3 WATER COOLED CONDENSERS

40

Water cooled condensers are always preferred where adequate supply of

clean and inexpensive means of water disposal is available. In air-conditioning

plants it is possible to harness the exhaust air and with integrated heating and

refrigeration re-circulated outdoor air as heat source. Switchable systems for heating

in winter and cooling in summer are particularly economical. Because by fitting the

condenser in the exhaust air duct, it is possible to lower the condensation pressure

during cooling operation and hence the input to the compressor. When the heating

cycle is used, the evaporation temperature can be raised by mixing exhaust air from

ventilation and air-conditioning system to the outdoor air, especially where the

exhaust air comes from rooms with high interior cooling load. This narrows the span

between evaporation and condensing temperature, increasing the C.O.P of the

system.

The heat from the evaporator absorbing refrigerant and heat of compression

is dissipated to atmosphere by air-cooled condenser. The considerable amount of

energy lost in this way can now be recovered by directing the hot refrigerant through

a water cooled heat exchanger which can produce large quantities of warm water at

50 to 620C for domestic, commercial or industrial uses at no extra running cost.

The heat from an air-cooled condenser is lost to atmosphere and cannot be

used for any other purpose. The considerable amount of energy lost in this way can

be recovered by passing the hot refrigerant through a water cooled heat exchanger

which can produce large quantities of hot water at 50 to 600Cfor domestic,

commercial or industrial uses at no extra running cost.

Water cooled condensers can be of three types

Shell and tube

Shell and coil

41

Double tube

4.4 CONDENSATION PROCESS

In most applications that use the condensation process, condensation is

initiated by removing heat at a solid-vapour interface, either through the walls of the

vessel containing the saturated vapour or through the solid surface of a cooling

mechanism placed within the saturated vapour. If a sufficient amount of energy is

removed, the local temperature near the vapour interface will drop below its

equilibrium saturation temperature. Because the heat removal process creates a

temperature gradient with the lowest temperature near the interface, vapour droplets

most likely form at the location. This defines one type of heterogeneous nucleation

that can result in either dropwise condensation or film condensation, depending on

the physical characteristics of the solid surface and the working fluid.

Dropwise Condensation occurs on the cooling solid surface when its

surface free energy is relatively low compared to that of the liquid. Examples of this

type of interface include highly polished or fatty-acid impregnated surfaces with

contact with steam.

Film Condensation occurs when a cooling surface having relatively high

surface free energy contacts a fluid having lower surface free energy. This type of

condensation occurs in most systems.

The rate of heat transport depends on the condensate film thickness, which

depends on the rate of vapour condensation and the rate of condensate removal. At

high reduced pressures, the heat transfer coefficients for dropwise condensation are

higher than those available in the presence of film condensation at the same surface

loading. At low reduced pressure the reverse is true. For example, there is a

42

reduction of 6 to 1 in the dropwise condensation coefficient of steam when

saturation pressure is decreased from 90 to 16 kPa. One method for correlating the

dropwise condensation heat transfer coefficient employs non-dimensional

parameters, including the effect of surface tension gradient, temperature difference

and fluid properties.

When condensation occurs on horizontal tubes and short vertical plates, the

condensate film motion is laminar. On vertical tubes and long vertical plates, the

film motion can become turbulent. Grober et al. (1961) suggest using a Reynolds

number (Re) of 1600 as the critical point at which the flow pattern changes from

laminar to turbulent. This Reynolds number is based on condensate flow rate

divided by the breadth of the condensing surface. For a vertical tube, the breadth is

the circumference of the tube; for a horizontal tube, the breadth is twice the length of

the tube. In practice, condensation is usually laminar in shell-and-tube condensers

with the vapour outside horizontal tubes.

Vapour velocity also affects the condensing coefficient. When this is small,

condensate flows primarily by gravity and is resisted by the viscosity of the liquid.

When vapour velocity is high relative to the condensate film, there is appreciable

drag at the vapour-liquid interface. The thickness of the condensate film, and hence

the heat transfer coefficient, is affected. When vapour flow is upward, a retarding

force is added to the viscous shear, increasing the film thickness. When vapour flow

is downward, the film thickness decreases and the heat transfer coefficient increases.

For condensation inside horizontal tubes, the force of the vapour velocity causes the

condensate to flow. When the vapour velocity is high, the transition from laminar to

turbulent flow occurs at Reynolds numbers lower than previously described.

43

When superheated vapour is condensed, the heat transfer coefficient depends

on the surface temperature. When the surface temperature is below saturation

temperature, using the value of h for condensation of saturated vapour that

incorporates the difference between the saturation temperature and the surface

temperature leads to insignificant error. If the surface temperature is above the

saturation temperature, there is not condensation and the equations for gas

convection apply.

Correlation equations for condensing heat transfer are given below. Factors

F1 and F2, which depend only on the physical properties, have been computed for

some commonly used refrigerants in Table 4.1.

For designing the condenser, the following assumptions were made. An air

cooled condenser is being employed for the air-conditioner design as it the most

suited for the required heat extraction capacity. A blower is attached in front of the

condenser to induce forced convection of air over the tubes which gives better

results for the condensation process. The condenser coil is a continuous coil to

which fins are bonded to increase the heat transfer area.

Tube Dimensions

o Tube diameter: 3/8”

o Thickness of tube: 1.5 mm

Coolant air-condition

o Temperature of air: 400C

o Velocity of air: 12 m/s

Fins

o Number of fins: 200 fins/m

o Height of fin: 5 mm

44

o Thickness of fin: 1 mm

These assumptions were arrived at after much iteration. Mostly, the

arrangement of the fins have been redesigned. Starting of with 40 fins/m of 10 mm

height, the pipe length obtained for the condenser was close to around 8 meters

which was too large for the required application. Thus the number of fins were

increased and its dimensions modified suitably to increase the heat transfer area

without affecting the fin efficiency till a suitable condenser pipe length was arrived

at.

Table 4.1 Values of condensing coefficient factors for different refrigerants.

RefrigerantFilm temperature 0C

tf = tsat - 0.75ΔtF1 F2

R – 11 24 80.7 347.7

38 80.3 344.7

52 79.2 339.7

R -12 24 69.8 284.3

38 64.0 257.2

52 58.7 227.6

R – 22 24 80.3 347.7

38 75.5 319.4

52 69.2 285.5

Sulphur dioxide 24 152.1 812.2

38 156.8 846.0

52 166.8 917.9

Ammonia 24 214.5 1285.9

45

38 214.0 1283.8

52 214.0 1281.7

4.4.1 Condensation on Outside Surface of Horizontal Tubes

1. Vertical surfaces, height L

Laminar condensate flow, Re < 1800

(4.3)

(4.4)

(4.5)

Turbulent flow, Re >1800

(4.6)

2. Outside horizontal tubes, N rows in a vertical plane, length L, laminar

flow

(4.7)

(4.8)

Finned tubes

(4.9)

where De is determined from

(4.10)

46

For a bank of N tubes, Nusselt’s equations increased by 10% are given in

Equations (4.7) and (4.8). Experiments by Short (1951) with R-11 suggest that drops

of condensation falling from row to row cause local turbulence and increase heat

transfer.

For condensation on the outside surface of horizontal finned tubes, Equation

(4.9) is used for liquids that drain readily from the surface. For condensing steam

outside finned tubes, where liquid is retained in the spaces between the tubes,

coefficients substantially lower than those given by Equation (5.9) were reported.

4.4.2 Condensation on Inside Surface of Horizontal Tubes

1. Inside vertical tubes

(4.11)

2. Inside horizontal tubes

1000 < Re < 20000

(4.12)

20000 < Re < 100000

(4.13)

Re > 100000

(4.14)

For condensation on the inside surface of horizontal tubes, the vapour

velocity and the resulting shear at the vapour-liquid interface are major factors in

analyzing heat transfer. Hoogendoorn (1959) identified seven types of two-phase

flow patterns. For semistratified and laminar annular flow, use Equations (4.12) and

47

(4.13). Ackers et al. (1959) recommended Equation (4.14) for turbulent annular

flow.

For the calculation of the length of the condenser first the thermal resistances

offered to the flow of heat from the refrigerant to the air are individually identified.

This includes the refrigerant side film, the air side resistance and the condenser wall

which are represented in Figure 4.1. These resistances are individually calculated

and the overall coefficient of heat transfer is calculated.

It is then substituted in the equation for the overall heat transfer in which the

unknown variable, the length of the condenser pipe, is included.

4.5 CALCULATION OF LENGTH

The process starts with the calculation of the individual thermal resistances.

For this, the empirical co-relations are first used to evaluate the convective heat

transfer coefficients inside and outside the condenser tube.

4.5.1 Inside the condenser tube

To begin with, based on the state of the refrigerant, the nature of flow is

identified which is numerically expressed in terms of Reynolds number. Based on

the calculated Reynolds number, a suitable correlation for evaluating the heat

transfer coefficient of the refrigerant is chosen from pre-determined empirical

relations. For the present design problem, the following correlation suggested by

Ackers and Rosson[1] was used.

(4.15)

The value of the heat transfer coefficient of R-134a is,

h = 4539.0326 W/m2K

48

Figure 4.1 Thermal resistances in a condenser.

49

4.5.2 Outside the condenser tube

The air side coefficient is calculated on the basis of relations developed for

expressing Nusselt’s number.

(4.16)

(4.17)

(4.18)

Based on these expressions the heat transfer coefficient of air side is,

h = 18.64 W/m2K

Once the individual heat transfer coefficients have been calculated, the

overall heat transfer coefficient is calculated by adding them.

(4.19)

The value of U which will contain the unknown value of length in its expression is

then substituted in the expression for overall heat transfer and the length is

calculated.

q = U.At.Δt

L = 3.52 m

The condenser pipe shall be 3.52 m long with 9 bends.

50

CHAPTER 5

DESIGNING OF EVAPORATOR

It is the most important part of the refrigeration system. The refrigerant from

the expansion valve comes into the evaporator below the temperature required to be

maintained in the evaporator and carries the heat from the evaporator. The

evaporator is known as cooler or freezer. The evaporators are maintained in different

sizes, shapes and types as per the requirements.

The following factors are considered in the design of the evaporators

(1) Heat Transfer: The heat is carried by the refrigerant from the air or water as

per the medium used for circulation. The refrigerant boils therefore the heat

transfer coefficient of the refrigerant side is considerably high compared with

the heat transfer of the other side which is the effect of convection. The heat

transfer capacity of the evaporator is given by

Q = U.A.(Tf - Ts) kW (5.1)

where, U = Overall heat transfer coefficient

A = Area of evaporator surface.

Tf = Temperature of the fluid passing through evaporator to be

cooled.

Ts = Saturation temperature of refrigerant at evaporator pressure.

(Tf - Ts) = Temperature difference causing the heat flow.

Too high temperature differences causes excessive dehydration of the

products to be cooled and too low temperature difference (below 8oC) causes

51

slime on some products like meat and fish if primary refrigeration is used for

direct cooling.

(2) Materials: Good heat conducting materials must be used for the construction

of the evaporator. The choice of materials used depends upon the refrigerant

used in the system. Brass and copper which are good conductors of heat are

used with all the refrigerants except ammonia. Freons should not be used with

aluminium.

(3) Velocity: With an increase in the velocity of the refrigerant, the heat transfer

coefficient also increases but increased velocity causes greater pressure loss.

There are recommended velocities for different refrigerants which give high

heat transfer rates with allowable pressure loss. The velocity above the

recommended one gives uneconomical working of the unit.

5.1 REGIMES OF BOILING

The different regimes of pool boiling described by Farber and Scorah (1948)

verified those suggested by Nukiyama (1934). The regimes are illustrated in Figure

5.1. When the temperature of the heating surface is near the fluid saturation

temperature, heat is transferred by convection currents to the free surface where

evaporation occur (Region I). Transition to nucleate boiling occurs when the surface

temperature exceeds saturation by a few degrees (Region II).

In nucleate boiling (Region II), a thin layer of superheated liquid is formed

adjacent to the heating surface. In this layer, bubbles nucleate and grow from spots

on the surface. The thermal resistance of the superheated liquid film is greatly

reduced by bubble induced agitation and vapourization. Increased wall temperature

increases bubble population causing a large increase in heat flux.

52

.

Figure 5.1 Characteristic pool boiling curve.

53

A heat flux or temperature difference increases further as more vapour

forms, the flow of liquid towards the surface is interrupted and a vapour blanket

forms. This gives the maximum critical heat flux (CHF) in nucleate boiling. This

flux is often termed the burnout heat flux or boiling crisis, because for constant

power generating systems, an increase of heat flux beyond this point results in a

jump of the heater temperature (Point c, in Fig 5.1), often beyond the melting point

of a metal heating surface.

In systems with controllable surface temperatures, an increase beyond the

temperature for CHF causes a decrease of heat flux density. This is the transitional

boiling regime (Region IV in Fig 5.1); liquid alternatively falls onto the surface and

is repulsed by an explosive burst of vapour.

At sufficiently high surface temperature a stable vapour film forms at the

heater surface; this is the film boiling regime (Region V and VI). Because heat

transfer is by conduction and some radiation across the vapour film, the heater

temperature is much higher than for comparable heat flux densities in the nucleate

boiling regime.

The design procedure for the evaporator is similar to that of the condenser.

The only difference would be that in the choice of the empirical relations to be used

for calculation of the heat transfer coefficients.

The basic assumptions being made for the design of the evaporator are as

follows

Tube Dimensions

o Tube diameter: 3/8”

o Thickness of tube: 1.5 mm

54

Ambient air-condition

o Temperature of air: 320C

Fins

o Number of fins: 200 fins/m

o Height of fin: 5 mm

o Thickness of fin: 1 mm

Similar to condenser design, these assumptions were arrived after much

iteration. Even though both the coils belong to the same cycle there is a variation in

the dimensions because of two reasons.

The values of the heat transfer coefficients for the two phase change

processes are different.

The total heat transfer occurring through both the pipes are different.

5.2 CALCULATION OF LENGTH

The approach for the design of the evaporator, like in the condenser, is based

on the heat transfer between the refrigerant and the ambient air that is to be supplied

to the room at the required temperature. First, the basic thermal circuit of the

arrangement is drawn. From this, the various thermal resistances offered to the flow

of heat are identified separately. These values of thermal resistances were then

calculated.

5.2.1 Inside the evaporator tube

To begin with, based on the state of the refrigerant, the nature of flow is

identified which is numerically expressed in terms of Reynolds number. Based on

which, a suitable correlation for evaluating the heat transfer coefficient of the

55

refrigerant is chosen from pre-determined empirical relations. For the present design

problem, the following correlation suggested by Anderson [2] was used.

(5.2)

The value of the heat transfer coefficient of R-134a is,

h = 943.546 W/m2K

5.2.2 Outside the condenser tube

The air side coefficient is calculated on the basis of relations developed for

expressing Nusselt’s number.

(5.3)

(5.4)

(5.5)

h = 12.56 W/m2K

Once the individual heat transfer coefficients have been calculated, the

overall heat transfer coefficient is calculated by adding them.

(5.6)

The value of U which will contain the unknown value of length in its

expression is then substituted in the expression for overall heat transfer and the

length is calculated.

q = U.At.Δt

L = 3.27 m

The evaporator pipe shall be 3.27 m long with 8 bends.

56

CHAPTER 6

DISCUSSION OF RESULTS

This project was a design problem that aimed to study the basics of

refrigeration and air-conditioning and the methodology involved in the designing of

an air-conditioning system. A comprehensive study was conducted on the operation

of the vapour compression refrigeration system and on the analysis of the system.

The design procedure for air-conditioning systems based on the vapour compression

refrigeration system were studied. This included mainly, the calculation of cooling

loads, selection of refrigerant, analysis of the refrigeration cycle and equipment

selection and design.

Based on a sample requirement of 1.5 tons for a room, a window air-

conditioner was designed. The details of the results are being discussed below

6.1 DISCUSSION OF RESULTS FOR REFRIGERATION CYCLE

For the design, a simple vapour compression refrigeration cycle was chosen

for the design process. The sample space had a cooling load of 1.5 tons and a DBT

of 240C was required to be maintained. The refrigerant chosen for this design was R-

134a and the evaporator and condenser pressures were 160C and 520C respectively.

The results are tabulated in Table 6.1.

During the analysis of the simple vapour compression cycle, it was

calculated that the dryness fraction of the liquid-vapour refrigerant mix leaving the

evaporator is close to 0.98 which almost corresponds to a fully saturated vapour.

The compressor work required for the process has been calculated to be quite

low as compared to general refrigeration systems. Thus the system has a high C.O.P.

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Table 6.1 Results for refrigeration cycle analysis.

REFRIGERATION CYCLE

Load capacity 1.5 tons

Refrigerant used R-134a

Refrigeration cycle employed Vapour compression cycle

Operating temperatures 160C (evaporator), 520C (condenser)

Required comfort conditions 240C

Refrigerating effect 128.392 KJ/kg

Compressor work 20.26 KJ/kg

C.O.P 6.33

Refrigerant mass flow rate 2.453 kg/min

6.2 DISCUSSION OF RESULTS FOR COMPRESSOR

The compressor designed for the window air-conditioning system has the

specifications as detailed in Table 6.2.

From the discussions on compressor design and operation the following

conclusions were made. The volumetric efficiency and thus the performance of the

reciprocating compressor can be improved in the following ways:

o Providing minimum clearance. Though it has to be kept in mind that

the main aim of providing clearance is for expansion allowance,

which is more important than obtaining higher values of volumetric

efficiencies.

o Maintaining low pressure ratio. Thus for a system where compression

is to be achieved over a large pressure range, it is most suitable to do

this by employing multi-compression stages instead of a single large

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compressor. In the present design, however, this problem was not

faced since the pressure ratio was considerably small.

o Cooling during compression.

o Reducing pressure drops at the valves by designing a light-weight

valve mechanism, minimizing valve overlaps and choosing suitable

lubricating oils. The pressure drop in the lines can only be reduced by

reducing the length of the pipe line between the compressor and

evaporator or compressor and condenser. In the case of a window air-

conditioner, this is of not much importance since the unit in itself is

highly compact. However, it is extremely desirable to prevent bends

and turns in the piping to avoid these losses.

Table 6.2 Compressor specifications.

COMPRESSOR SPECIFICATIONS

Model used Hermetically-sealed

reciprocating compressor

Speed of motor 1500 r.p.m

Bore 50 mm

Stroke 40 mm

Clearance 0.04

Volumetric efficiency 84.15%

Suction pressure 0.48416 MPa

Delivery pressure 1.4251 MPa

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6.3 DISCUSSION OF RESULTS FOR CONDENSER

The condenser to be used for the window air-conditioning system has the

design specifications as elaborated in Table 6.3.

Table 6.3 Condensor design.

CONDENSER DESIGN

Tube diameter 3/8”

Thickness of tube 1.5 mm

Velocity of air 12 m/s

Number of fins 400 fins/m

Height of fin 5 mm

Thickness of fin 1 mm

Length of tube 3.52 m

Number of bends 9

The design of the condenser is the most tedious process in the designing of

the whole refrigeration system. This is mainly because there are no set of definite

equations or relations available for the designing of the equipment. Like most heat

transfer design problems, the process starts with modelling the system in heat

transfer circuits and identifying the thermal resistances.

Many empirical relations have been devised by many researchers for the

calculation of the convective heat transfer coefficients of fluids. The present design

case is that of two-phase heat transfer. It was noticed that the relations obtained by

different researchers may vary. This can be attributed to the randomness of fluid

dynamics. Thus choosing the correct relation for the design purpose is as important

as the solution of the expressions itself.

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For refrigeration and air-conditioning designers, such co-relations have been

catalogued and are available in handbooks such as ASHRAE. The co-relations used

for the designing of the condenser have been obtained from the same.

The design process started with assuming come of the parameters of the

condenser based on previous designs. When the results obtained using these

assumptions were found to be unsatisfactory, suitable modifications were made in

the assumptions. This process of iteration was continued till the final results

obtained confirmed with existing design standards and met the needs required from

the system.

6.4 DISCUSSION OF RESULTS FOR EXPANSION VALVE

The expansion valve to be used for the window air-conditioning system has

the specifications as given in Table 6.4.

Table 6.4 Expansion valve specifications.

EXPANSION VALVE DESIGN

Model used Capillary tube

Diameter of capillary 2 mm

Length of capillary 351.681 mm

The capillary tube was chosen because it was the simplest kind of expansion

valve and it easily met with all the requirements of the system. A lower diameter

would have given better performance for a lesser length. But other factors such as

choking of the tube are also to be considered during the design process.

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6.5 DISCUSSION OF RESULTS FOR EVAPORATOR

The evaporator to be used for the window air-conditioning system has the

specifications as given in Table 6.5

The design procedure for the evaporator was similar to that of the condenser.

The only difference would be the co-relations to be used to obtain the heat transfer

coefficients for the refrigerant and air.

Table 6.5 Evaporator design.

EVAPORATOR DESIGN

Tube diameter 3/8”

Thickness of tube 1.5 mm

Number of fins 200 fins/m

Height of fin 5 mm

Thickness of fin 1 mm

Length of tube 3.27 m

Number of bends 8

It was also noticed that the total heat transfer area requirement for the

evaporator is much less as compared to that of the condenser. This has been

attributed to the lesser heat transfer taking place through the evaporator tubes as

compared to the condenser.

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CHAPTER 7

CONCLUSIONS

After the completion of this project, the following conclusions were made,

The design process for an air-conditioning system was studied.

The refrigerant used for the design was R-134a. Present air-conditioners use

R-22 for air-conditioning which pose a great threat to the environment.

Through this project, the equipments that would be required for an air-

conditioner running with R-134a were designed.

It was seen that the design of the equipments are similar to that of the

systems that use R-22 as the refrigerant. Thus R-134a, apart from being a

more environmentally friendly refrigerant, is a feasible replacement for the

existing refrigerants also.

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REFERENCES

1. Ackers, W.W. and Rosson, H. F., 1959. Condensation inside a horizontal tube.

Chemical Engineering Progress Symposium Series 55(29):171-176.

2. Afgan, N.H. and Schlunder, E.U. 1974. Heat Exchangers: Design and Theory

Sourcebook, Mc-Graw-Hill, New York.

3. Anderson, W., Rich, D.G. and Geary, D.F.., 1966. Evaporation of R-22 in a

horizontal 3/4” tube. ASHRAE Transactions 72(1):28.

4. Arora, C.P.. 2nd Edition. Refrigeration and air-conditioning. Tata McGraw-Hill

5. Arora, S. C., Domkundwar, S. 1985. A course in refrigeration and air-

conditioning. Dhanpat Rai and Sons, India

6. ASHRAE. 1999. ASHRAE Fundamentals Handbook. American Society of

Heating, Refrigeration and Air-Conditioning Engineers.

7. Bari, E., Noël, Y. J., Comini, G. and Cortella, G. 1986. Air-cooled condensing

systems for home and industrial appliances. ASHRAE Transactions

92(1B):739-55.

8. Carslaw, H.S. and Jaeger, J.C. 1959. Conduction of heat in solids. Oxford

University Press, England.

9. Gardner, K.A. 1945. Efficiency of extended surfaces. ASME Transactions.

67:621.

10. Incropera, F.P and Dewitt, D.P. 1996. Fundamentals of Heat Transfer. John

Wiley and Sons, New York.

11. Metais, B. and Eckert, E.R.G. 1964. Forced, mixed and free convection regimes.

ASME Journal of Heat Transfer 86(C2)(5):295.

12. Schmidt, T.E. 1949. Heat transfer calculations for extended surfaces.

Refrigeration Engineering 4:351-57.

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