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.
1
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
2
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
3
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.
4
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
6
Figure 1.1 Schematic of an actual vapour compression refrigeration system.
Figure 1.2 Schematic diagram of a room air-conditioner.
7
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
8
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.
9
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
10
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.
11
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.
12
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.
13
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.
14
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
15
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;
16
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)
17
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.
18
Figure 3.1 Simple vapour compression cycle operating between 160C and 520C.
19
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
20
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
22
Figure 3.2 Cylinder and piston mechanism and p-V diagram for a reciprocating
compressor.
23
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
25
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.
26
Figure 3.3 Effect of valve pressure drops.
3.3.3 Leakage Loss
27
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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