Design of AC system for Multiplex

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Project Report
On
Design of Air Conditioning System for Government Medical
College (Bokaro, Jharkhand)
Submitted by:
Mahfooz Alam 13 BME 0027
Abdul Khaliq 13 BME 0036
Nasir Aziz 13 BME 0046
Musawwir Alam 13 BME 0044
Iftikhar Ahmad 11 MES 0030
In partial fulfilment for the award of the degree of
Bachelor of Technology [Mechanical Engineering]
Under the supervision of
Prof. Jamshed Ahmad Usmani
Submitted to the
Department of Mechanical Engineering
Faculty of Engineering and Technology
Jamia Millia Islamia
New Delhi - 110025, India.

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Department of Mechanical Engineering
Faculty of Engineering and Technology
Jamia Millia Islamia
New Delhi - 110025, INDIA
Certificate
June 2017
This is to certify that this project report entitled “Design of AC System for
Government Medical College, Bokaro, Jharkhand” submitted by Mahfooz
Alam (Roll No: 13BME0027), Mohammad Abdul Khaliq (Roll NO: 13 BME
0027), Musawwir Alam (Roll no: 13 BME 0044), Nasir Aziz (Roll No: 13
BME 0046) & Iftikhar Ahmad (Roll No. 11-MES-0030) in partial fulfillment
of the requirement for the award of degree B.Tech Mechanical
Engineering is a bonafide work done under the supervision of Dr. Jamshed
Ahmad Usmani.
Signature Signature
Date Date
Dr.J.A.USMANI Dr. J.A.USMANI
Professor and Head Professor and Head
Mechanical Engineering Mechanical Engineering
Faculty of Engg. & Tech. Faculty of Engg. & Tech.
Jamia Millia Islamia Jamia Millia Islamia

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SELF DECLARATION
We hereby declare that
1) The project entitled “Design of AC System for Government Medical College
(Bokaro, Jharkhand) submitted by us is an authentic work of our efforts
carried out for the partial fulfilment of the requirements for the award of
B.Tech (Mechanical Engineering) degree.
2) The matter embodied in this project work has not been submitted earlier in
this university.
3) We will be responsible for any form of plagiarism committed out of the project
4) The sources of information used in our project report have been duly
acknowledge and referenced in our project.
S No. Name of the Student Roll No. Signature
1. Mahfooz Alam 13 BME 0027 ………….....
2. Mohammad Abdul Khaliq 13 BME 0036 ………….....
3. Nasir Aziz 13 BME 0046 ………….....
4. Musawwir Alam 13 BME 0044 ………........
5. Iftikhar Ahmad 11 MES 00 30 ……………..

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ACKNOWLEDGEMENT
We are pleased to submit this project report entitled “Design of AC System
for Government Medical College, Bokaro, Jharkhand”. We offer our
heartfelt thanks to all those people without whose help accomplishment of
the project of this kind would not have been possible.
In the first place we extend our gratitude to Prof .J.A.Usmani (Professor,
Department of Mechanical Engineering) for his supervision, advice and
guidance from the very early stage of the project and giving us extraordinary
experience throughout the work. Above all the most needed, he provided us
unflinching encouragement and support in various ways. His truly scientist
intuition has made him as a constant oasis and passion in science, which
exceptionally inspire and enrich our growth as student. We are indebted to
him more than he knows.
We would also like to extend our thanks to all the faculty and staff of the
department of Mechanical Engineering for their co-operation to complete
this project

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ABSTRACT
We have selected the Building (Government Medical College, Bokaro,
Jharkhand) for Central Air Conditioning. In this project total heat load has
been calculated for the above building which comes out to be 128 Tons as
per outside and inside room temperature conditions. Further we have
designed the air conditioning equipment. i.e. Rotary Screw type water-cooled
chilling machine with microprocessor based control panel (semi-hermetic
compressor, water-cooled condenser, Electronic expansion valve,
evaporator, oil separator, controls, and accessories to make it compact and
efficient unit), chilled water pump, Cooling Tower, Air-handling units,
controls, valves & ducting etc. For the above load we have selected central
air conditioning system (vapour compression refrigeration system). For
proper distribution of conditioned air for the complete building with proper
ducting system and air handling unit is also designed and selected.

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CONTENTS
Description Page no.
Certificate i
Declaration ii
Acknowledgement iii
Abstract iv
List of figure
List of table
Chapter 1
Introduction
1.1 How air conditioning works 4-5
1.2 Types of air conditioning 6-9
1.3 Factors influencing human comfort 10
Chapter 2
2.1 Building selection for air conditioning 11
2.2 Building design layout 12

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Chapter 3
3.1 Data analysis and load calculation 15
3.2 System design 17-27
3.3 Chiller package 28-29
3.4 Outside design conditions 32-65
3.5 Effective sensible heat gain
3.6 Heat gain
3.7 Effective room latent heat gain
3.8 Determining air quantity
CHAPTER 4
4.1 Compressor 51-52
4.2 Cooling Towers
4.3 AHUs
4.4 Air filter
CHAPTER 5
5.1 Duct system
5.2 E-20 Sheet 62-70
RESULTS & DISCUSSION 70-71
CONCLUSION
BIBILIOGRAPHY

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Introduction
Leonardo de Vinci build water driven fan to ventilate a suite, this could possibly be
the first attempt to automatically change the condition of air in an enclosed space.
Another device, which originated in India many years ago, was the “panka”. The
panka was a large fan, which extended from the ceiling and was operated manually
by pulling rope. Some of the later models were machine operated.
The application of the Air-condition for the industrial purpose has opened a new era
in the Air-conditioning Industry. The Air-Conditioning is very commonly used now a
days for preservation of food, in Automobile, railways, hospitals, dry manufacturing,
cloth Industries and many others its varied application have opened a new field for
air conditioning engineers to solve the difficult problems with full success. Air
conditioning is a field of work which never stagnates. It is commonly used to ease
men environmental on earth and in space. The very adverse problems of spaced
environment are also successfully solved with the advanced knowledge of Air-
conditioning, which has made the space travel successful.
Atmospheric Conditions in India are varied in different parts of the country.
Particularly the summer condition, in India is quite comfortable in few parts of the
country but quite uncomfortable in others. No doubt, A/c will become a necessity for
Indians in coming few decades with the rapid industrial development and with the
economics growth of the country. Being most developing industry in the country A/c
engineers have better prospects in future and will be able to play a more significant
part in nations industrial and economic developments.

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REFRIGRATION CYCLE
Fig1.
The vapour compression refrigeration cycle is a common method for transferring
heat from a low temperature to a high temperature. The above figure shows the
objectives of refrigerators and heat pumps. The purpose of a refrigerator is the
removal of heat, called the cooling load, from a low-temperature medium.

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1.1 AIR CONDITIONING
In wider sense this covers the complete process of controlling the physical/chemical
properties of an enclosed atmosphere within the limits required for human comfort
and the efficiency and performance.
1.2 CLASSIFICATION OF AIR CONDITIONING SYSTEM
1.2.1 Classification as to major function
(a) Comfort Air Conditioning System
(b) Industrial Air Conditioning System
1.2.2 Classification as to season of the year
(a) Winter Air Conditioning System
(b) Summer Air Conditioning System
(c) Year round Air Conditioning System
1.2.3 Classification as to equipment arrangement
(a) Unitary System
(b) Central Station System
(c) Combined System

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1.2.1 Classification as to major function
(a) Comfort Air Conditioning System
Air conditioning in the office building public auditorium homes classrooms, prayers
rooms etc. is meant for maintaining comfort condition for the occupants. In addition
to the control of temperature and relative humidity for comfort conditioning it is
necessary to clean the air and maintain proper circulation.
Human beings give of heat (around an average 400 BTU per hour per person) but to
what is called metabolism. In a healthy individual the temperature regulating
mechanisms within the body maintain the body temperature around 98.6 degrees
Fahrenheit. But the skin temperature varies according to the surrounding
temperature and relative humidity. Naturally if surrounding temperature is less than
body temperature the flow of heat from skin will be steady. But is the surrounding
temperature is very low as in cold winter Day the rate of flow from body is quite
rapid and the person feels cold. On a summer day, vice-versa.
(b) Industrial Air Conditioning System
Air conditioning is also needed for various industrial processes and installation, such
as water assembly shops. Telephone manufacturing factories and telephone
exchange tool rooms, computer room purpose of this air conditioning system are to
control the atmospheric condition primarily for the proper conduct of research and
manufacturing processes.

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1.2.2 Classification as to season of the year
(a) Winter air conditioning system
It is the system which is installed for maintain indoor condition in winter season. The
main step in winter air conditioning is to heat the air and to control the moisture
content heating is accomplished by electric heater and boilers for the humidification
we use simple pass type of spray type humidifier.
(b) Summer Air Conditioning System
Major function is to cool the air and remove excess moisture from its cooling is
accomplished by mechanical refrigeration dehumidification is accomplished as
condensation of water vapour in the air occurs on cold coil surface.
(c) Year Round Air Conditioning System
These systems consist of automatic control composed of heating and cooling
equipment to create the four automatic conditions for human comfort at all times of
years.
1.2.3 Classification as to equipment arrangement
(a) Unitary System
The systems make use of factory assembled air conditioners, when the area to be air-
conditioned is less.
(b) Central Station system
These systems are used when several rooms in the same or different buildings are
intended to be air-conditioned with same temperature and relative humidity.

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(c) Combined System
This type of systems combines both feature of central station and unitary systems.
Heat energy is supplied in pipes to several unit air-conditioners in the form of steam
or hot water. Chilled water from the central refrigerating equipment is also piped to
the unit air conditioners.
1.3 Factors Influencing Human comfort
The factors that influence include the following:
(a) Temperature
(b) Humidity
(c) Air motion
(d) Air Purity
Optimum conditions:
 TEMPERATURE : 22 ˚C – 26.5˚C •
 RELATIVE HUMIDITY : 30% - 70%
 SOUND : 40dB (FROM 1m)
 AIR VELOCITY : 0.8 m/s – 1.5 m/s
 SIZE OF PARTICLE : < 0.1 micron (operation theatre)
1.3 Concept of effective temperature
There is no precise physiological observation by which comfort can be measured.
Consideration of mean skin temperature offers a solution to some degree.
Combination of temperature humidity and air movement which induced the some
feeling of worth are called thermo equivalent condition. It really denotes the sensory
heat level. This is called the effective temperature.
1.4 Factor governing optimum effective temperature
It is desirable to analyses the factors that may change optimum effective
temperature.

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(a) Climate and seasonal difference
It is borne out by experiment that people living in cold or climate are comfortable in
effective temperature lower than those living in warmer regions
(b) Clothing
In winter much of the effect due to climate and variable occupancy is compensated,
because people can wear clothes comfortably to suit both indoor and outdoor
condition.
(c) Age and sex
The comfort chart is prepared for the performance of men only. Women required
about 0.5 degree centigrade higher effective temperature.
(d) Shock effect
This effect is due to sudden entering and leaving of people from outside to
conditioned space and vice versa.
(e) Activity
Heavy activity people needs lower temperature than those seated at rest

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2.1 Building Selection for Air-Conditioning
It is proposed to air-condition the Government Medical College in Bokaro, Jharkhand,
for maintaining comfort conditions throughout summer, as an air-conditioning
engineer.
We are required to prepare detailed project report for Air-conditioning the Building
for given details of the proposed building.
1. Heat load calculation.
2. Building Layout.
3. Machine Layout.
4. Ducting Layout.
5. Piping Layout.
6. Selection of Equipment.
7. Bill of Quantity.

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3.1 Required Data
Hospital Building
1. Ground floor Area - 31190 sq ft
2. First floor area - 14670 sq ft
3. Third floor area - 16680 sq ft
4. Fourth floor area - 16770 sq ft
5. Total area of building - 95200 sq ft
Medical College Building
1. Ground Floor Area 4075 sq ft
2. First Floor Area 7530 sq ft
3. Second Floor Area 7450 sq ft
4. Third Floor Area 3245 sq ft
5. Total area 22300 sq ft
Based on the above specifications, A.C. loads for Hospital Block are as follows:
S.No Description Area (Sq.ft.) AC Load (TR)
1.
Ground Floor
22860 129.23
2.
First Floor
12463 66.27
3.
Second Floor
12635 61.52
4.
Third Floor
131.20 67.28
5.
Fourth Floor
20565 122.09

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Total
81643 446.39
Based on the above specifications, A.C. loads for Medical College are as follows:
1.
Ground Floor
5075 32.77
1.
First Floor
18840 91.89
2.
Second Floor
14670 81.03
Total
38585 205.69
Total Tonnage (Hospital & Medical College) = 651.78 TR
Total Installed capacity of chillers = 586.6 TR (Say 600 TR)
(Considering diversity @ 90%)
Based on the above requirements, A.C. loads for Auditorium are as follows (Packaged
and Split Unit system):

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1.
All Floors
12910 100.39
Total
12910 100.39
A.C. loads for Hospital Block (Split AC’s) are as follows:
1.
Ground Floor
2063 21.75
2.
First Floor
2438 22.44
3.
Second Floor
2030 18.3
4.
Third Floor
5505 39.75
5.
Fourth Floor
880 7.72
6.
Fifth Floor
1290 14.34
Total
14206 124.34

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A.C. loads for Medical College (Split AC’s which include H.O.D Room & Dean
Room)
1.
Ground Floor
532 6.12
2.
First Floor
320 3.04
3.
Second Floor
640 6.41
4.
Third Floor
718 5.49
5.
Fourth Floor
360 2.76
6.
Fifth Floor
360 2.98
Total
2930 26.08

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3.2 System design
4.1 It is proposed to provide a central Air-conditioning system to maintain the specified inside design
conditions during summer, monsoon & winter for both the buildings.
4.2 The total peak air conditioning load works out to 600 TR for all the floors. To cater to this load, it is
proposed to install 4 Nos. Screw type water-cooled chilling machines each having 200 TR actual capacity
(3W+1S).
4.3 3 Nos.(3W+0S) of Heat Pumps of capacity 80 KW each have been proposed for the winter heating
purpose and 2 Nos. (1W+1S) of Hot Water Generator of capacity 30 KW each have been proposed for
the OTs with their own pumps.
4.4 Water chilling machines shall work in conjunction with 4 Nos. Primary chilled water pumps (3W plus
1S).
4.5 2 Nos. of Secondary water pumps (1W + 1S) shall be used for pumping chilled water to the
Hospital building.
4.6 2 Nos. of Secondary water pumps (1W + 1S) shall be used for pumping chilled water to the Medical
College building.
4.7 4 Nos. of Condenser water pumps (3W + 1S) shall be used for the cooling towers.
4.8 3 Nos. of Cooling Tower of capacity 225 TR each shall be installed at Terrace.
4.9 It is proposed to provide Stand-alone package and split Units for Auditorium Block.
4.9.1 It is proposed to provide split Units for HOD, Dean Rooms etc. for Hospital & Medical College Blocks.
4.10 Chilled water produced shall be pumped to various Air-handling units and Fan coil units. Chilled water
shall be pumped through insulated chilled water pipes installed in ceiling spaces and in vertical risers
installed in pipe shafts. At each Air-handling unit balancing valves are provided for balancing. All pipes
within plant room shall be supported from floor.
4.11 Double skin Air handling units consisting of Centrifugal plug fans with VFD control , cooling coil and filter
section shall be provided for each area. Chilled water supply and return headers shall be tapped and

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connected to cooling coils. There would be automatic controls provided for AHUs to control inside
conditions in summer and monsoon.
4.12 AHU’s with Heat Recovery Wheel has been proposed for all Operation Theatres.
4.13 AHUs would be fitted with Variable frequency drives to control the air quantity being supplied by
AHU. When the building is not fully operational, these AHU would operate on part load resulting in
substantial energy savings.
4.14 The conditioned air from the AHUs would be supplied through pre insulated ducts. The air would be
diffused through extruded aluminium Grilles and diffusers. The return air would be taken back from the
conditioned space to the AHUs through return air ducts or through ceiling spaces.
4.15 It is also proposed to provide Air washer/Scrubber for Kitchen supply & Exhaust.
4.16 The stale air from the common toilets would be exhausted by means of mechanical exhaust system.
4.17 Building shall have smoke extraction air mechanical fans. Exhaust ducts shall be provided at ceiling
level.
4.18 The capacity of mechanical exhaust and make up air fans capacity shall be of 6-12 air changes per hour.
The fans would start automatically through the fire protection system of the building and it shall be
connected to main fire alarm system of the building. These fans shall also have an arrangement to start
manually as and when required.
4.19 It is proposed to provide all lift wells, staircases and lift lobbies or corridors with a pressurization system
to keep them free of smoke and toxic gases during fire for safe escape route. However, if the staircases are
provided with open-able windows then it is not necessary to provide pressurization system.
4.20 Motorized smoke and fire dampers shall be provided in accordance with ASHRAE/NFPA within supply
air ducts and return air ducts/spaces to prevent spread of smoke / fire to adjacent areas.

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1. SCOPE
1.1 The scope of this section comprises the supply erection, testing and commissioning of the water chilling
units conforming to this specification and in accordance with the requirements of the "Schedule of
Quantities".
2. CODES & STANDARDS
The water-cooled liquid chilling packages shall conform to the latest edition of
following standards:-
ASHRAE 15 Safety code for Mechanical refrigeration
ASHRAE 23
Methods of testing and rating positive displacement
refrigerant compressors and condensing units
ASHRAE 30 Methods of testing liquid chilling packages
ASME SEC VIII DIV I Boiler and pressure vessel code
ANSI B 31.5 Code for refrigeration piping
AHRI 550/590 (2003) Standard for water chilling packages
ARI 575
Standard for method of measuring machinery sound within an
equipment space
ISO 1940
Mechanical vibration – Balance quality requirements of rigid
rotors
ISO 10816-1
Mechanical vibration – Evaluation of machine vibration of
measurements on non-rotating parts. General guidelines
3. TYPE
3.1 The water chilling machine shall consist of Multiple imported Helical Rotary Screw compressors with
motor, squirrel cage induction motor, starter, shell and Tube flooded Cooler, Shell and Tube Condenser,
refrigerant piping, wiring and automatic controls and accessories all mounted on a steel frame. Machine
shall be factory charged with refrigerant and oil.
4. COMPRESSOR
4.1 The Compressor shall be semi hermetic / hermetic type gear / direct driven rotary type using R-134a
refrigerant. The rotor shall be statically and dynamically balanced to ensure vibration free operation.
The compressor mounting shall be horizontal type. The COP of the Chiller shall be more then 5 at AHRI
conditions.
4.2 Compressor shall be equipped with slide valve to provide full modulating control compressor capacity
from 100-20% of full load. The slide valve should be actuated by oil pressure controlled by external
solenoid valves through the microcomputer-controlled center. The unit should be capable of operating

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with lower temperature cooling water during part load operation. Alternatively, the unit shall have
stepped control.
4.3 The compressor housing shall be of high-grade cast iron, machined with precision, to provide a very
close tolerance between the rotors and the housing. The rotors shall be mounted on anti-friction
bearing designed to reduce friction and power input. There shall be multiple cylindrical/-tapered roller
bearings to handle the radial and axial loads.
4.4 There shall be built in oil reservoir to ensure supply of lubricants to all bearings and a check valve to
prevent backspin during shut down. There shall be oil pump or other means of forced lubrication of all
parts during startup, running and coasting for shut down. An oil heater shall be provided in the casing.
4.5 The units shall be complete with capacity control mechanism, to permit modulation between 20% to
100% of capacity range. An oil separator shall be included to remove oil from the refrigerant and there
shall be suitable heat exchanger for oil separation.
5. MOTOR / STARTER
5.1 High efficiency continuous duty compressor motor shall be of the single speed, non-reversing, and
squirrel cage induction type suitable for 415 volts +/- 10%, 50 Hz. Motor designed speed shall be 2960
rpm at 50 Hz.
5.2 Motor shall be factory mounted and full load operation of motor shall not exceed nameplate FLA rating.
The starter should be star-delta closed transition type motor starter. The starter should be housed in a
separate free standing, housing and include all necessary safety devices i.e., overload relays, under
voltage release and single phasing preventer device.
6. CHILLER
6.1 Chiller shall be horizontal shell and tube, multi pass, direct expansion and designed for the duty
specified in the schedule of equipment. The shell shall be of welded steel construction fitted with steel
sheets on either side. The cooler shall have plain seamless copper tubes of not less than 12 mm O.D.
and 0.63 mm material thickness. The tubes shall be supported in the shell by adequate, stiff supports to
eliminate vibration and noise. The tube ends shall be properly expanded in the tube sheet to prevent
leakage of refrigerant. Tubes shall be individually replaceable from either end of the heat exchangers
without affecting the strength and durability of the tube sheet. The baffles on the waterside in the shell
are to be arranged to ensure adequate water velocity over the tubes and proper direction of flow. The
refrigerant heads shall be made of cast iron and the faces ground to a close tolerance to prevent leakage
of refrigerant between passes and between the circuits in case of a multi circuit cooler.
6.2 Chiller (Evaporator): Chiller shall be designed so as to prevent liquid refrigerant entering the
compressor. The chiller shall be provided with liquid level sight glass and a relief device to prevent
excess pressure in the heat exchanger.

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6.3 The chiller shall be provided with following connections and accessories, as separately identified in the
schedule of quantities.
I Refrigerant inlet and outlet pressure gauges.
II Water inlet and outlet connections with suitable size butterfly valves.
III Drain and vent connections with stop valves.
IV Dial type pressure gauges and stem type thermometers on inlet and outlet connections
V De-scaling valves
VI Water flow switches at the outlet
VII Ribbed rubber isolator or pads to eliminate transmission of vibration upto 90%.

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6.4 Chiller shall be insulated with 19 mm closed-cell, polyvinyl - chloride foam with a maximum K factor of
0.28. Insulation shall be applied to cooler shell, flow chamber, tube sheets, suction connection and all
the necessary parts (wherever required). The insulation shall be set with compound recommended by
the insulation manufacturer and shall be applied sealing the joints. The insulation shall be applied in
such a manner that water boxes and covers shall be removable without damaging it.
6.5 The cooler shall be designed for a pressure of 21 kg/sq.cm. and tested with nitrogen or Co2 gas against
leaks at a pressure of 14 kg/sq.cm. both on the shell and on refrigerant side and pressure shall be
maintained for a period of 24 hours when no drop in pressure should be observed indicating any leaks.
6.6 Hydraulic pressure to 10 Kg/sq.cm. Shall be applied on the waterside and shall be maintained for a
period of 24 hours when no drop in pressure should be observed indicating any leaks.
7. CONDENSER
7.1 Each unit shall have horizontal one shell and tube, water-cooled, multi pass condenser fitted with safety
valve, purge valve, and liquid line valve. The shell shall be of welded steel construction, fitted with steel
tube sheets on either side.
7.2 The tubes shall be at least 19 mm (3/4) outer dia. and 1.0 mm thick seamless copper with integral 19
fins/inch. The tubes shall be supported in the shell by adequate number of supports (at a distance not
more than one meter) to avoid noise and vibrations and the ends properly expanded in the tube sheets
to prevent leakage of refrigerant. These shall be designed for the duty specified in the schedule of
equipment.
7.3 The water head shall be of cast iron, easy to remove with suitable baffles for multipass water flow, in
and out connection and gasket to prevent water leakage. Condenser shell shall be able to hold 1.25
times the refrigerant charge in the system to which the condenser is to be connected.
7.4 The condenser shall be tested with nitrogen or CO2 gas against leaks at a pressure of 22kg/sq.cm. on the
shell side for a period of 24 hours and similarly hydraulic pressure of 10 Kg/sq.cm. on the water side
shall be applied for a period of 24 hours when there shall be no drop in pressure indicating any leak.
7.5 The condenser shall be complete in all respects and include supports for mounting, refrigerant and
water valves (isolating valve at inlet and balancing valve at outlet) at in and out connection, de-Scaling
connection, stem thermometer and pressure gauges at water inlet and outlet etc., as required).
8. MICRO -COMPUTER CONTROL CENTRE
8.1 The chiller shall be provided with a factory installed and wired microprocessor control center. The
control center shall have alphanumeric display with minimum 80 characters. The microprocessor shall
be configurable to display either English or SI metric units.

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8.2 Functions
The microprocessor based control shall provide the following functions:
a) Capacity control based on leaving chilled water fluid temperature with return fluid temperature sensing
b) 7-day time sequence of both pump and chilling machine.
c) Automatic change over between compressors.(in case of multiple compressor)
8.3 Display
System information shall include (but not limited to)
a) Return / leaving chilled water temperature
b) Return / leaving condenser water temperature
c) Evaporator refrigerant pressure
d) Condenser refrigerant pressure
e) Oil pressure at compressor
f) Oil filter differential pressure
g) Percent motor current
h) Evaporator / condenser saturation temperature
i) Compressor discharge Temp
j) Oil temperature and oil level
k) Percent slide valve position
l) Operating hours
m) Hours since last run
n) No of compressor starts
o) Time of last start and time of last stop
p) Water temp. Reset valve
Security access shall be provided to prevent unauthorized changing of set points and to select or remote
control of the chiller.
8.4 Bas Interface
Control panel shall be able to interface with building automation system such that all the dates available
on the chillers / microprocessor panel is also available on the BAS.
8.5 Safeties
Unit shall automatically shutdown when any of the following condition occurs. (Each of these protective
limits shall require manual rest and cause an alarm massage to be displayed on the LCD screen informing
the operator of the shutdown cause.)
a) Loss of refrigerant charge
b) Reverse rotation

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c) Loss of condenser / chilled water flow
d) Low chilled water temperature
e) Low oil pressure
f) High motor / oil temperature
g) Current imbalance.
h) Thermal Overload
i) High condenser refrigerant pressure
j) Electrical overload
k) Loss of phase
Ground current fault
9. AHRI CERTIFICATION
The chilling unit shall be AHRI certified as per AHRI 550 / 590 – 2003 STANDARD. All suppliers shall
furnish computer printouts along with their technical bids, giving details of capacity output at design
conditions as given in tender.
10. INSTALLATION
10.1 The chilling machine shall be installed over a cement concrete platform and shall be adequately isolated
as per manufacturer recommendations against transmission of vibrations to be building structure.
10.2 Necessary foundation bolts, nuts, leveling shims etc required for mounting of the unit shall be provided
by the contractor.
10.3 On installation, the refrigerant circuits and the connected equipment and parts shall be thoroughly
tested against leaks. The system pressure shall be reduced by a vacuum pump to within an absolute
pressure of 70 cm of water column and maintained for twenty-four (24) hours with a pressure rise not
exceeding 3.5 cm of water absolute.
10.4 Final alignment shall be checked in presence of Engineer-in-charge, with motor and compressor bolted
in position and with all the piping connections made, using a dial indicator. After alignment has been
done to the satisfaction of the Engineer-in-charge, the position of the compressor and motor shall be
fixed.
11. TESTING
11.1 During the performance testing all readings shall be jointly recorded by owner/ consultant and
contractor. A detail report with all plots to be submitted to the owner/ consultant for review and
acceptance before taking over the machine by the owner. In case the stipulated tender requirements
are not met the contractor shall make good the deficiency in every case by alerting and replacing parts

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or if required the whole equipment free of cost. Contractor shall then conduct the second performance
test to establish the tender requirements at his own cost within maximum of one month of its first test.
Equipments capacity in tons of refrigeration shall be computed from the temperature readings and
water flow measurements. Computed results shall tally with the specified capacities. The power
consumption should tally with the specified capacities according to the figures furnished in the tender.
11.2 The contractor shall provide all instruments and personnel for tests.
11.3 Performance Testing
The unit shall be selected for the lowest operating noise level, capacity ratings, and power consumption
with operating parameters given in technical sheets, shall be submitted and verified at the time of
testing and commissioning of the installation.
Power consumption shall be computed from measurements of incoming voltage and input current &
energy measure meters.
12. START UP
12.1 The Chiller manufacturer shall provide a factory trained representative, employed by the chiller
manufacturer, to perform the start-up procedures as outlined in the start-up, operation and
maintenance manual provided by the chiller manufacturer.
12.2 After the above services have been performed, the same factory trained representative shall be
available for classroom instruction not to exceed a period of 4 hours to instruct the owner's personnel
the 7proper operation and maintenance of the chiller.
12.3 Manufacturer shall supply the following literature:
a) Start-up, operation and maintenance instructions & manual
b) Installation instruction
c) Field wiring diagrams
13. PAINTING AND PROTACTIVE COATING

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Complete chiller package, pumps, all uninsulated like vessels, pipes, fittings, valves and structural items
to be painted. In case of paint of equipment got spoiled during transit/ erection, then the same shall be
repainted at site with one coat after erection.
14. LIST OF SPARES
Contractor shall provide spares needed for start up, commissioning and testing till the plant is handed
over to the owner. Contractor shall furnish separately list of recommended spares and a set of any
special tools/ tackles for 2 years normal operation and maintenance along with offer.
14.1 Guarantee
All equipment and components shall be guaranteed by the contractor against any defective material,
design, fabrication workmanship, installation and proper functioning. Free replacement/ repair/
alternation shall be included in contractor’s offer, if any defects occur during the guarantee period.
Contractor shall also guarantee to maintain the chilled water temperature irrespective of any variation
in capacity load. Pressure drops in the equipment, capacity of the units and utility consumption figures,
shall be guaranteed by vendor.
15. WARRANTY
The chiller package shall have factory warranty for five years from project completion date.

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3.3 CHILLER PACKAGE – DATA SHEET A
S.N. Description
1 Number Required 4(3W+1S)
2 Location Plant Room
3 Duty Continuous (18 hrs/day)
(Approximate)
4 Capacity required at specified design conditions
per chilling package
3 x 200 TR actual capacity
5 Refrigerant R134a
6 Full load IKW (maximum) 0.69 KW/TR
7 Maximum noise level at a distance of 1.5 meters 85 dBA
8 Compressor – type Semi-hermetic/hermetic
9 Lubrication Forced feed with an oil pump /
differential pressure
9.1 No. of Compressor (Min.) 1
10 No. of Refrigerant Circuit 2
11 Capacity control Automatic in stages
12 Static and dynamic balancing of screws As per ISO 1940
13 EVAPORATOR
13.1 Type Shell and tube, flooded/DX
13.2 Liquid to be cooled Water
13.3 Chilled water quality Potable water
13.4 Chilled water inlet temperature 12.2 º C/ 54 º F
13.5 Chilled water outlet temperature 6.7 ºC/ 44 º F
13.6 Minimum chilled water flow per chilling package 480 USGPM
13.7 Fouling factor-water side (FPS unit) 0.0005
13.8 Chiller and suction line insulation Closed cell polyvinyl chloride
foam
13.9 Maximum water side pressure drop 20 Ft of water
14 CONDENSER
14.1 Type Water cooled, shell and tube
14.2 Condenser cooling water quality Potable water
14.3 Condenser water inlet temperature 32.2º C/ 90º F
14.4 Condenser water outlet temperature 36.4º C/ 97.5º F
14.5 Minimum condenser cooling water flow per
chilling package
800 USGPM
14.6 Maximum water side pressure drop 20 Ft of water
14.7 Fouling factor-water side (FPS unit) 0.001
15 Motor 415 V +/- 10%, 3 phase, 50 Hz
16 Control Panel Microprocessor based control

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panel
17 Control panel to be interfaced with building
automation system
Provision to be available
18 Type of starter
Wye- Delta- Closed transition
type
For summer Air-conditioning, the Air conditioning space obviously has to be
maintained at a temperature lower than the surrounding temperature. The moisture
content may also have to be maintained at a level lower than the atmospherically
level. So there has to be transfer of heat as well as ingression of moisture from the
surroundings to the air conditioned space.
Further we have to take into account the heat generated by the occupants,
electric lights, Fans and other appliances. Naturally they will raise the temperature of
room, unless heat is removed as fast as it is transmitted from outside and generated
within. This heat, which is transmitted from outside and generated within and tends
to raise the dry bulb temperature of the room is called the Room sensible load.
The occupants of the room also release moisture and there may be other
sources of moisture with in the room as well. For comfort air conditioning and for
special applications not only the dry bulb temperature, but the moisture content
(Wet bulb temperature) in the room also has to be maintained at a certain level. So
it becomes necessary to remove the excess moisture, as fast as it released by
condensing it and draining it out as condensate (Water).For this latent heat of
vaporization of the excess water vapour has to be removed. So the reduction of
moisture content is also heat removal and hence is a heat load (Latent). Surrounding
air infiltrates into the conditioned space whenever the doors are opened and also
through small cracks around the doors/Windows and in the walls. The temperature
and moisture content of the infiltration (Outside) air has to be brought down and so
this also adds to sensible and latent heat load of the room.

33
Further, a small portion of the fresh air taken in for ventilation purposes, by passes the
cooling apparatus without getting cooled and dehumidified .So the load of cooling and
dehumidifying this fresh air which has by passed the cooling apparatus adds up to the
load in the room. When the supply air duct passes through the non air-conditioned
space, the supply air gains in heat due to heat transmission from outside through the
duct walls and insulation. This also adds up to the room sensible heat load. If the fan of
the air handling unit is located in the leaving side of the cooling coils (as in a draw
through arrangement), the horsepower used by the fan for moving the air reused in the
generation of heat and this heats adds up to the room sensible heat load. If the fan is on
the air entry side of cooling coil, the heat generated due to fan hp is a part of the total
heat load, but not of the room sensible heat load.
The sum of all the above sensible and latent heat loads is known as the Room
Sensible Heat Load (RSH) and the Room Latent Heat load (RLH) respectively .The
sum of the RSH and RLH is known as the Room Total Heat Load(RTH).The ratio of the
RSH to the RTH(RSH/RTH) is known as the Room Sensible Heat Factor (RSHF).
The ratio of the quantity of the by-pass air to that of the total air passing
through the coil is known as the By-pass Factor (BF).The BF depends upon the fin
pitch, Velocity of air over the cooling coil face and number of rows in the direction of
air flow. The BF increase as the coil face velocity increases but decreases as the fin
pitch decreases and also as the number of rows increase. By pass occurs in the air
washers as well and here it affects the saturation efficiency.

34
Thus we have:
 Effective room total heat (Effective Room Sensible + Latent Heat)
 Grand total heat (ERTH plus the load due to (1-BF) fresh air intake, heat gains in
return air duct/passage, chilled water system, pump hp, etc.
It is the effective room total heat load that determines the quantity and
temperature-humidity condition of the supply air and grand total heat load
determines the capacity of the refrigeration plant. The heat load estimation form is
designed to arrive at these loads and the condition and quantity of dehumidified
supply-air needed.
1) SHF =SH / (SH + LH) = SH/TH
Where,
SHF = Sensible Heat Factor
SH=Sensible Heat
LH=Latent Heat
TH=Total Heat
2) RSHF = RSH / (RSH + RLH)
Where,
RSHF = Room Sensible Heat Factor
RSH- Room Sensible Heat
RLH- Room Latent Heat
RTH- Room Total Heat
3) EFSHF = ERSH / (ERSH + ERLH)

35
Where,
ERSHF = Effective Room Sensible Heat Factor
ERSH = Effective Room Sensible Heat
ERLH = Effective Room Latent Heat
ERTH = Effective Room Total Heat
4.2 Apparatus Dew Point (ADP)
It is the effective surface temperature of the cooling coil which determines the
condition of supply air coming out of the coil. Therefore it is the surface temperature
of the cooling coil that is to be determined and controlled to obtain the desired
conditions and this effective surface temperature is termed as Apparatus Dew Point
(ADP)
4.3 Heat Load Estimation
Heat load Estimation for design and selection of the air conditioning-refrigeration
equipment should be done carefully and as accurately as possible. The Successful
performance of the air conditioning and refrigeration plants depends on the accuracy
in arriving at the loads they will have to handle.

36
3.4 Outside Design Conditions
The first important step in the exercise is fixing the peak outside condition i.e. the
date and time of the year when sustained peak conditions can prevail, based on
which the heat load can be worked out. One may be inclined to fix June 21, being the
longest summer day.
But severe hot and humid outside conditions may occur in July, when compared to
June, temperature may be lower but with much higher moisture content, which
probably will impose a higher load. Again the maximum outdoor temperature may
be around noon, but the maximum heat gain of the room from the surrounding may
occurs a few hours later due to the time lag for heat transmission through the
building structure. So selection of the design data and time has to be done
judiciously. It may be necessary to calculate the heat gain for a few selected dates (in
summer and monsoon) and also for few hours on each of these days.
3.4 Inside Design condition
Air conditioning is for maintaining the comfort conditions for people (Comfort air
conditioning) or for the industrial applications. The inside condition for the industrial
application are dictated by the products or process and the requirements will be
specified by the user.
3.5 Comfort Air conditioning and Effective Temperature
Relative Humidity and air movement within the conditioned space have to be such
that the dissipation of the heat due to body metabolism is steady, to maintain the
normal body temperature and for comfort. It has been establish by experiments that
there can be various combinations of temperature, humidity and air movement that
can include the same feeling of comfort.
With fix air movement, comfort conditions can be obtained with different
combination of temperature and relative humidity, if the temperature is raised, a

37
reduction in relative humidity can give the same comfort feeling. Similarly with a high
relative humidity, a lower temperature is required for comfort. The various
combinations of these three parameters are known as the effective temperature or
Comfort zone.
Any combination of temperature and relative humidity falling within the comfort
zone is expected to maintain comfortable conditions for majority of the occupants.
Unless otherwise specified, inside condition of 24 degree + 1 degree, 50-60 % RH and
air movement of 4.5 to 7.5 m/min (15 to 25 fpm) are satisfactory for comfort cooling
applications.
3.4 Heat load Estimation Form
A Typical format of the form is given as follows. Heat gain is separated into the room
sensible heat, room latent heat, room total heat and grand total heat, to facilitate
the calculations for the air quantity required and for the equipment selection. Heat
gain occurs because of the following reason:

38
3.5 Effective sensible heat gain
3.5.1 Solar Gain –glass
Radiations from the sun can enter the conditioned space through glass and heat is
absorbed by the material and air in the conditioned space. This gain is not confined
only to the sides facing the sun, other sides too can absorb heat, but to a much
smaller extent.
As the heat gain depends upon the orientation of each side of the building, the gain
has to be worked out individually for each side (i.e. north, north-east, east, etc.).The
values given in the below tables are generally for the exposed ordinary glass.
The heat gain is less for colored/tinted glass, double panel glass, shaded glass (with
ventilation blinds, shading, etc) and for the glass which is not directly exposed.
Multiplying factors for different types of glass, inside/outside venetion, blinds,
outside shading/awning etc are to be used.
The heat gain is
Heat gain = A x R x MF
Where,
A=Area of glass in m2
or Ft2
R=Solar Gain in Kcal/h/m2
or BTU/h/ft2
MF=Multiplying factors for the type of glass, shading etc.

39
3.5.2 Solar Transmission Gain through the walls and roofs
As a result of the absorption of direct radiation of the sun, the temperature of
walls/roof rises above the ambient temperature. This result in large temperature
difference with respect to the air conditioned space.The design tables provides the
following information.
a) Equivalent temperature difference for a standard outside temperatures (say 35 0
C
or 95 0
F) and inside temperature (say 26.7 0
C or 80 0
F).
b) Correction factors for different outside and inside temperature to be added
/Subtracted from standard values of (a)
c) Correction factors for different types of shading, to be added or subtracted to (a)
d) Transmission coefficients for the most commonly used construction materials in
heat units/h/unit area /degree temperature difference.
The heat Gain = A x U x Eq TD
Where,
A = Area of the wall or roof in m2
or ft2
U = Transmission coefficient in kcal/h/m2
/ 0
C or BTU/h/ft2
/ 0
F
Eq TD = Corrected Equivalent temperature difference in 0
C or 0
F
3.5.2.1 Transmission Gain through glass and partition
In addition to the solar gain though the glass, as described above,
There is a heat gain by transmission through the glass because of the temperature
difference between the surrounding and conditioned space. Also heat gain occurs
through the partition walls/partitions between the non-conditioned and conditioned
space

40
Heat gain = A x U x TD
Where,
A = Area of glass/Partition in m2
or Ft2
U = Transmission coefficient for glass/partitions
TD = Temperature difference between the surroundings and
Conditioned Space
3.5.2.2 Heat gain through infiltration and (by passed) Fresh Air
(a) Design Below tables give the infiltration air quantity for various type of
doors/window, for observable cracks and for infiltration due to the opening of the
doors.
This load is generally is generally ignored as this can be merged with the load due to
fresh air intake for ventilation.
(b) The room load due to the bypassed fresh air (through the cooling coil) is:
Qv X ᵨ X C X BF X TD
Heat gain in Kcal/h = m3
/min x 60 x ᵨ x C x BF xTD (0
F)
where,
ᵨ = 1.2 kg/m3
, C=0.24)
or

41
Heat gain in BTU/h = cfm X 60 X ᵨ X C X BF X TD (0
F)
Where,
ᵨ =0.075 lb/cfm, C=0.24)
3.6 Heat Gain
3.6.1 Internal Heat Gain
a) People-Heat is generated wit in the human body by the process of
metabolism. The metabolism rate varies with the type of activity of the individual.
b) Lights: Loads due to the following is calculated below
Fluorescent light = total watts X 1.25 X 0.86 Kcal/hr
Total watts X 1.25 X 3.4 BTU/h
Or
Incandescent light = Total watts X 0.86 kcal./hr
Total Watts X 3.4 BTU/h
c) Fan Horse power in the draw-through systems: For Calculating this value, multiply
the brake horse of the fan by 641 to get the value in kcal/hr or by 2545 for BTU/h
d) Appliances: Some appliances give off both sensible and latent heats, the latent
heat being given off directly or as a result of their function such as cooking, drying,
etc. Refer to the data provided by manufacturers of the appliances.
Safety Factor: An additional 5% on RSH is taken as a safety factor and this also covers
items such as heat gain by the supply ducts, leak etc

42
Effective Room Sensible Heat (ERSH): It is obtained by adding up the items to A1 to
A6.
3.7 Effective Room Latent Heat Gain (ERLH)
3.7.1 Room Latent Heat
a) Infiltration: This is generally ignored.
b) Outside air (By Passed air): The latent heat portion (i.e. Latent heat removal for
condensation of the excess moisture from the bypassed fresh air) forms a part of the
room latent heat gain. The heat gain depends on the amount of moisture to be
condensed and drained off. And this is equal to the difference between the
surrounding and inside moisture contents and can be found out from the
psychometric Chart.
The heat gain is thus
= Qv X 60 X ᵨX h X BF X (W0- Wi)
Where,
Qv=Fresh Air Quantity is m3
/min or CFM
h=Latent Heat of condensation of water vapour in kcal/kg or BTU/lb of moisture
BF=By Bass Factor of the cooling coil.
(W0- Wi ): Difference of moisture content at outside and inside condition respectively
in the grams of water vapour/kg dry air or grains of water vapour/lb dry air.
Thus the heat gain is:
In Kcal/hr =m3
/min X 60 X 1.2 X (588.3/1000) X BF (W0- Wi)

43
= m3
/min X 60 X 1.2 X 0.1059 X BF X (W0- Wi)/7000
= CFM X BF X (W0- Wi) X 0.68
(W0 and Wi in g water vapour/kg dry air)
In BTU/h = CFM X 60 X 0.075 X 1059 X BF (W0- Wi)/7000
= CFM X BF (Wo-Wi) X 0.68
(Wo and Wi in Gr moisture/lb dry air)
(lb = 7000Gr)
(Density of air : 1.2 kg/m3
or 0.075 lbs/ft3
)
c) People (Occupancy): The moisture released by the occupants has to be
condensed. For the moisture released per person for various activities
Heat Gain= n X w X h
Where,
n=No. Of persons
w=Moisture released/person in g/person or in gr/person
h=Latent heat of condensation of moisture in kcal/g or BTU/gr.
I.e. Kcal/h = n X W X (588.3/1000)
d) Steam: In comfort application steam as such, or steam appliances are seldom used
in the conditioned space. In industrial or special application, where steam is used,
the corresponding heat gain has to be taken into account.
e) Appliance: If moisture generating appliances are used in the conditioned space,
find the moisture generation from the appliance rating.
Heat gain: W X h

44
Where,
W= moisture generated per hour in g or gr.
h= latent heat of moisture in kcal/g
f) Vapour transmission: As the water vapour content in the air increases, its water
vapour pressure increases. Thus the surrounding air will have higher water vapour
pressure than in the air conditioned space. This difference in pressure results in the
flow of the water vapour into the conditioned space from the surroundings.
g) Supply duct leakage : This loss is generally ignored.
Adding up the values of 4.8.2.a to 4.8.2.g and adding up 5% as safety factor,
effective room latent heat (ERLH) is obtained
3.7.2. Effective Room Total Heat Gain (ERTH):
It is the sum of ERSH and ERLH
3.7.3 Outside Air Heat
The sensible and latent heat removed from the fresh air taken in for ventilation
purposes (i.e. bringing the fresh air from outside conditions to the conditions in the
conditioned space) causes a load on the air conditioning apparatus.
The heat gain due to the by passed outside air has already been accounted for under
the ERTH gain.So the heat removed from the balance(1-BF) fresh air intake is to be
accounted for under this caption. The heat gain on account of fresh air intake is :
Sensible Heat = Qv X 60 X C X (1-BF) X Density X TD
Latent Heat = 42.36 X m3
/min. X (1- BF) (W0-Wi)
3.7.4 Return Duct Heat Gain

45
a) Return air heat gain, are generally ignored, as return air passage/ducts are
insulated, if running through non conditioned area and aspect of leakage is taken
care of ,by proper sealing of the return air passage/duct against leaks
b) The heat gain due to the fan horsepower is to be taken under the grand total heat
if the system is of blow through type
c) In chilled water system, the heat gain due to the chilled water pump horsepower
and gain through the chilled water piping (even if insulated) gets added to the load
on the refrigeration plant.
3.8 Determination of air Quantity (m3/min or cfm)
The air quantity selected should offset the room sensible and latent heat load. It
should also handle the total sensible and latent heat loads, i.e. including the outdoor
(fresh) air loads, etc.
Qda = ERSH/ (trm-tadp) (1-BF)60X ᵨX C
Where,
Qda = Dehumidified air quantity
ERSH = Effective Room sensible heat
Trm = Room Design temp
Tadp = Apparatus dew point.
C = Specific Heat of air
BF = By- Pass Factor of the coil.
Air Quantity in m3
/min. =ERSH/17.28 (trm-tadp) (1-BF)
ADP can be determined from the psychometric chart. By plotting the air-conditioning
process in the chart. Referring to the above chart, Point S Represent the condition of

46
the temperature and RH of the Supply air delivered to the room from the air
conditioning apparatus. The supply air in the mixing with the room air picks up the
room sensible and latent heat loads, resulting in the required comfortable
conditions. The point A is the condition of the air after picking up the sensible heat,
So the line SA represent the temperature rise of the supply air due to picking up the
sensible heat and line AM represent its latent heat pick up.
If 300 m3/min of supply air rises in temperature from S to A by picking up the
sensible heat in the room, then half the quantity of supply air will rise to double this
temperature range in picking up the same sensible heat. Likewise it latent heat of
pick up will also be doubled.
Points S1 and A1 represent the case of 150 m3
/min. Thus the temperature rise (line
SA) and latent heat pick up line (AM) (sensible and latent heat pick up) does not
change and the combined effect of sensible and latent heat pick will move in the
same line SM. Only the line S1M will be double than SM in the second case. In the
triangle SAM,SA and AM are proportional to the sensible and latent heat pick up
from the conditioned room and SM is the direction which the supply air conditions
will move, as it picks up the room sensible and latent heat loads in the proportion
and attains the room Condition. Thus the slope of the line SM is governed purely by
the ratio of the room sensible and latent heat loads, that is by the room sensible heat
factor. So the room sensible heat factor line has to be drawn on the psychometric
chart as the first step to plot the process. For drawing this line, the psychometric
chart has the sensible heat factor scale and an alignment point p, usually at
80Degree F and 50% RH.
a) Plot the room condition R on the Chart
b) Draw a base line connecting the alignment point (P) and the RSHF value on the
sensible heat factor scale.
c) The RSHF line is obtained by drawing a line through the room condition R, Parallel
to the baseline to intersect the saturation curve of the psychometric chart at X.

47
One can select any point on this line RX as the supply air conditions, as this air can
pick up the sensible and latent heat in the room. As explained earlier with reference
to above figure.
Once the supply air condition is selected (on the RSHF Line), the air quantity to be
handled to meet the load is determined by the difference in temperature between
room and supply air temperature selected.
There is however, no practical means to ensure that the condition of temperature
and RH of the leaving air from cooling coil is at exact condition selected on the
sensible heat factor line.
But there is no one condition of the temperature and humidity level, which can be
controlled in cooling plant. This is the point where RSHF line meets the saturation
curve in the psychometric chart, and this is the apparatus dew point (ADP).
So far we have considered the room sensible heat and latent load. Due to by-pass
effect in the cooling coil, sensible and latent heat load due to by-pass fresh air also
added to room load. So it is for effected room sensible heat and latent heat loads,
The ADP has to be established.
The By-pass fresh air comes out of the cooling coil at the same condition as it enters
the coil and mix with the balance (1-BF) air which gets treated by the coil. This
mixture of untreated fresh air and the (1- BF) treated air is supplied to the air
conditioned space. Therefore, to care of sensible and latent heat load of by-pass
fresh air, the balance (1- BF) air has to be cool down to a level lower than what
would otherwise have been done. Had there been no fresh air intake. To arrive at the
ADP for handling effective room load and in the psychometric chart, the effective
room sensible heat factor (ERSH) line has to be drawn as shown in above figure. The
point D, where ERSHF line crosses the saturation curve is the effective ADP. In our fig
point O,R and M represent the condition of Fresh air, return air and mixture of fresh
and return air respectively. The (1-BF) of the mixture of condition M gets cool to the
effective ADP at D.

48
Since ADP lies on the Saturation Curve, Moisture from the air will condense when
cool down to this temperature. Point M1 which is the intersecting point of the line
MD and RX (The room sensible heat factor line),is the condition of supply air, which is
the mixture of by-pass untreated fresh air and the cold and dehumidified (1- BF) air.
Line DM1 Represent the By Pass factor of the cooling coil and Line MM1, the (1- BF)
portion.
The supply air condition M1, deliver to the air condition space. Picks up the room
sensible and room latent heat, Moving Along the line M1R to give the final room
condition at R. Summarizing the portion of treated air, i.e. (1- BF) which comes out of
the coil at DM1R, The path DM1 being the shift in condition because of the by-pass
effect of the coil.M1R is the path due to pick up of RSH and RLH.
If Q is the air quantity (M3/m) or CFM, then (1- BF) X Q is the air quantity that gets
cool and dehumidified in passing through coil. It is this air which pickup the effective
room sensible latent heat loads. Thus
Q X (1-BF) X (trm - ADP) X 60 X ᵨ X 0.24 = ERSH
Where
ᵨ=density of air (1.2 Kg/m3
or 0.075 lb/cubic Feet)
And 0.24 is the specific heat of air,
Q (In m3
/m) =ERSH (kcal/h) / (1-BF) X (Trm-ADP) X17.28
Once the ADP is determined, The evaporator temperature can be selected from
manufacturers rating chart for cooling coil/Chiller.
These charts provide the refrigerant temperature to be maintained to handling the
load at the ADP determined the coil face air velocity for different Rows deep of coil.

49
Some assumption of the number of rows (Deep) and coil face air velocity has to be
made for selecting the coil size.
In some cases, the different combination of coil depth and Face (air) velocity will be
found more economical and suitable for the compressor selection. Then By Pass
Factor must be used to verify the resultant room sensible and latent heat and ADP.
Another Method adopted for rating of coil gives the (Sensible and latent heat
removal) capacity of unit area of coil face in heat units/Hr for Different
 Coils enter the air-conditioning ( Entering Air wet bulb)
 Coils face (Air) Velocities
 Refrigerant (chilled water /Brine) temperature
 Numbers of rows deep
 Fin Pitch

50
4.1 COMPRESSOR
Compressor is the heart of Air-condition system. It pumps and circulates refrigerant
through the system just as the heart pumps and circulate blood through the body.
Thus it supplies the necessary force to keep the system operating. While in operation
it lowers the pressure in the cooling coil. It results in low temperature in the space to
be cooled down by allowing the liquid refrigerant to evaporate (vaporize) the heat
latent vapour then flow towards compressor where they are compressed and thus
the temperature is raised. These high temperature vapours are discharge to the
condenser where heat flows from hot refrigerant vapour into the air or water passing
through the condenser. In sort the refrigerant absorbed heat from the cooling coil,
its temperature and pressure raised by the compressor and then same refrigerant is
discharge towards the condenser. The function of compressor can be summarized as:
a) To remove low temperature and low pressure vapours from the cooling coil though
the line called suction line.
b) To compress these vapours by increasing the pressure and temperature resulting in
an increase of saturation point of the refrigerant.
c) To discharge the vapours in high temperature and pressure to condenser through the
line called discharge line.
4.1.1 Types of compressor
1. Open type
2. Reciprocating
3. Rotary
4. Screw Compressor
5. Centrifugal
1. Open Type Compressor
An open type compressor is driven with electric motor with the help of pulley and
belt system. The compressor and motor are rotary type with mounted on the same

51
base plate on which is also mounted the condenser. This combination of component
put together is known as condensing unit
2. Reciprocating Compressor
The reciprocating compressor are slowly being phased out, because power
consumption of 0.94 KW
Their maintenance cost is high due to a large of number of moving parts.
3. Scroll Compressor
The scroll compressor are slowly replacing the reciprocating compressor in capacity
up to 40 Tons.
These compressors are more efficient requiring approx. 0.75 KW/Tons
They are maintenance free since they are rotary type with minimum moving parts.
4.Screw Compressor
These compressors are replacement for larger size reciprocating compressor and are
available in much larger capacity then reciprocating unit.
They have high efficiency between 68 to 0.75 IKW/ton.
They require minimum maintenance since there are moving parts.
5.Centrifugal compressor
These compressor are meant for large capacity between 200-2500 Tons
They have highest efficiency which can be as high as 0.62 KW/tons.
They require more maintenance then screw compressor but are preferred for their
larger capacity range.

52
4.2 Cooling Towers
The cooling towers generally being used are of following types. All cooling towers
currently being used are of Fiber Reinforced Plastic (FRP) Construction.
4.2.1 Induced draft
The induced draft cooling tower has the fans located at top of the tower, sucking or
drawing air up the tower. the induced draft is usually favored. In both these the air
and water are in counter flow. The cross flow induced draft towers are also used for
small capacity, in which air moves horizontally through the tower while water flows
down.
4.2.2 Forced Draft
The forced draft cooling tower has the fans fixed at bottom of the tower forcing air
up the tower.
4.2.3 Selection of cooling towers
 The types of cooling tower used depend upon the capacity, space and consideration
of noise and height.
 Force draft cooling towers suitable, where low height units are required.
 In other cases induced draft type are generally used.
 Circular cooling towers are more suitable when the space available is less, as the
clearance required between towers is less for this type.
 Cooling towers are also used to cool the water of engines for D.G sets.
 The generally required details of cooling tower i.e. size, weight, motor rating, based
on certain main brands are given in chart-I for cooling tower up to 300 TR and chart-I
for larger size cooling towers.
 The chart-I also gives equivalent capacity of cooling towers for D.G. sets.
 Specific details for other makes and type may vary by 10%.
 However, the information in this chart is adequate for basic planning.

53
4.2.4 Basis for Selection
 The a/c capacity of a cooling tower is depend on several factors i.e. Ambient wet
bulb temperature.
 Approach i.e. the difference in temperature between ambient wet bulb and water
after cooling.
 Normally the approach is 7o
F (-13.9o
C).e.g. if ambient is 83 o
F (28.3o
C), than water
should be at 90o
F (32.2o
C).
 The minimum approach can be 5o
F(-15 o
C).In such case the cooling tower capacity
reduce by 20%.
 Water flow rate per ton in USGPM or LPM. This is called cooling range which is 10 o
F
at 3 USGPM and 7.5o
F for 4 USGPM.
4.2.5 Importance of wet bulb
 The maximum ambient wet bulb differs from region to region in India
 Hence while specify the cooling tower and chiller parameters it is important to select
the correct wet bulb.
 This may also effect the water leaving temperature at cooing tower for example if
the wet bulb is 800
F then water leaving cannot be 900
F but it will be 920
F etc.
 This also effect the water inlet temperature at condenser of chiller.
 In most places the wet bulb temperature is higher in monsoon, but in many coastal
cities in south India the wet bulb is higher in summer. Hence select the wet bulb
whichever is higher and choose cooling tower and water entering condenser
according to wet bulb temperature
36
4.2.6 Water Consumption
Water is consumed in the cooling tower for the following reason:
 Some water is evaporated in the process of cooling the air, which in turned cools the
condenser water.
 Some water is also carried away by the cooling tower fans due to high air velocity.
In addition when the water is quite hard the salt of the water which has evaporated,
accumulate in the remaining water thus increasing the concentration of salts.

54
It is there for necessaries to bleed some water from the system in other words some
water carries these salts are put into the drain.
Normally the amount of water which has to be added to compensate for the above
mention losses is approx. 3 GPH or 12Litres per hour per tons of plant capacity.
Provision has to be made to add this water to the system.
4.3 Air Handling Unit
4.3.1 Types of AHUs
The various types of AHUs which are commonly used are
 Unitary type: These are used wherever space is to be saved.
 Unicom type: These are used, wherever the return air is carried in duct or where it is
necessary to kept more than one AHU. In a common room, but serving different
areas the arrangement will prevent mixing of return air.
 Sectional type: These are used for higher capacity or wherever space is available for
such AHUs and for mixing of duct return.
 High static sectional type
4.3.2 Sub classification
 Single skin body with M.S angle/G.I Sheet frame work
 Single Skin body with extruded aluminium frame work
 Double Skin body with extruded aluminium frame work

55
4.3.3 Source of centrifugal fan
The centrifugal fan in AHU may be either
 Imported fan (Nicotra, Comefri, Lau, Kruger etc.)
 Indian fans (M.S.I similar in performance to imported fans)
Indian fans (various AHU manufactures make in fans with out
Dated copied technology)
 The fans in items above mention are not in used accept on the insistence of the user
The fan outlet velocity should be as follows
 Upto 1600 FPM (8m/sec. ) for indigenous fans
 Upto 1800 FPM (9m/sec.) for indigenous form with imported design
 2000 FPM (10m/sec) for imported fan and housing assembly.
Imported fan may be excepted upto the outlet velocity of 2400 FPM (12m/sec,)
provided their computer selection indicated the fan noise level to be below 88db
4.3.4 Cooling Coils
The cooling coils in AHUs are
 Normally made of copper tubes and Aluminum fins. The copper tubes are expanded
on the tubes for a tight bonding and efficient heat transfer
 The coil generally are 3,4,6 and 8 rows deep for cooling application
 The number of fins on the coil can be 8, 10 or 12 fins per inch or approx. 3,4 or 5 fins
per centimeter

56
 The heat transfer improves with increase in the number of fins per inch per
centimeter. Hence more fins means higher heat transfer and increase capacity.
 The cooling coil is selected for air velocity through the coil ranging from 400 FPM
(2m/sec) to 600FPM (3m/sec.)
 However most cooling coils are generally selected for air velocity of 500FPM that is
2.5 m/sec.
 The area of coils are calculated accordingly e.g. coil area for 8000 CFM (13600
CMH)unit will be
8000/500 =16 sq.ft. = 1.48m2
4.3.5 Air handling unit selection
The air handling unit should be selected on the following basis.
 Capacity in TR should be at least 15 % more than the calculated capacity.
 Air quantity should be equal to the adjusted CFM. But air quantity within (-)10% of
adjusted CFM can be used if used of adjusted air quantity feasible due to space
constrained of if the next available size of AHUs as excess capacity of more than 10%.
4.3.6 Selection of Air-Handling Unit (Table)
AHU capacity based on:
 Water temperature of 12.8 0
C leaving and 7.20
C entering
 Air entering temperature between 250
C to 26.70
C and air leaving temperature
between 12.80
C to 13.90
C for 4 row and 11.10
C to 12.20
C for 6 row cooling coil
 High static AHUs are to be used when HE and Hepa filters are to be used

57
4.4 Air Filter
The various types air filters used in AHUs are
 Ordinary washable filters with efficiency of 80% by weight of dust particle
 Washable air filters having a efficiency of 95% down to 10 micron particle size.
 High efficiency air filters having an efficiency 99% down to 5 micron these are not
cleanable
 Bag filters having an efficiency of 95% by weight (these have high dust holding
capacity)

 Hepa filter having an efficiency of 99.99% down to particle size 0.3 micron.
 Normally used washable air filters of 95% efficiency for comfort application
 High efficiency (HE) filters are used where specific requirement for class 2000000
cleanliness is given.
 Hepa filters along with (HE) filters are to be used for operation theaters and
whenever clean room application in the drug industries and medical assembling
industries, where clean room required class 100000 or better, level of cleanliness.

58
5.1 DUCT SYSTEM
The function of the duct is to cover the air between two points, such as between the
air handling unit or air washer or room to be conditioned. It also carries the room air
back to the air conditioning apparatus. There are two types of air transmission
system adopted for air conditioning application. Low velocity and high velocity
systems. If the initial velocity of the main supply duct is within 760 m/min. (2500
fpm), it is classified as low velocity system. The high velocity system is employs
velocity above 760m/min. the low velocity system is adopted for comfort air
conditioning systems with the initial air velocity supply normally ranging from360 to
600 m/mint (1200 to 2000 fpm). Return air duct, whenever used, normally have low
velocity, air distribution system are divided into three categories depending upon
pressure.
 Class I fan- low pressure: up to 95mm (3.75”) water gauge
 Class II fan- medium pressure: up to 95mm to 170 (3.75” to 6.75”) water gauge
 Class III fan- high pressure: up to 170mm to 310mm (6.75 to 12.25”) water gauge
The duct has to be so sized that it is accommodated within the available space. Like
any other fluid passing through a pipe, air in passing though a duct suffers a pressure
drop due to friction. Larger the quantity of air passing through a given cross sectional
area of the duct, greater will be friction loss and pressure drop. The fan selected has
to deliver the required quantity of air overcoming the resistance offer by the various
components in the air distribution system, such as a cooling air washer, filter, supply
and return air outlets, damper etc. plus resistance offer by the duct system that is
fan has to work against the head.
As this head increase the fan will need higher power to deliver the required quantity
of air against the system head. Further the friction increase the noise level due to air
in motion also increase. Thus, on these two counts, the velocity of the air in the duct
has to be kept at a reasonable low. With the lower air velocity the size of the duct
increases. so duct system should be balance between the initial cost and the
operating cost for given rate of flow of air. The initial cost depends upon the size of

59
the duct. a small size of duct have low cost. But with smaller duct, the velocity will
increase so pressure losses due to the friction also increase. The fan will then have to
use more power to overcome the head, thereby increasing the operating cost.
Instead of evaluating the balance of initial costs in every case, designers depend
upon past experience and employ generally accepted recommended air velocities in
the ducts.
5.2 GAIN OR LOSS IN DUCT
The supply and return air ducts can gain or loss heat; the transfer the heat from the
surroundings to the air in the duct during the cooling cycle and from the air to the
surroundings during the heating cycle. This happens even the duct is passing through
the conditioned space and the air in the duct. The heat gain or loss, as the case may
be can be considerable when the duct are passing through non conditioned space. In
such cases, the duct have to be insulated. Duct having larger aspect ratios-will
gain/lose more heat than ducts having smaller aspect.
5.3 ASPECT RATIO
This is the ratio of the long side to the short side of the duct. As the aspect ratio
increases, more metal surface is required for the duct, for the same cross sectional
area.
Thus the higher aspect ratio increases not only the initial cost but also the operating
cost by way of increased heat gain or loss. Therefore this ratio is an important factor
to be considered in the duct design.
Ducts carrying small air quantities at low velocity have the greatest heat gain. This is
because the quantity of air by weight will be less and the heat transmission will be

60
higher because of the comparatively bigger surface area of the duct. For the same
cross sectional area, a circular duct required less surface area than a rectangular
duct. Also the circular duct being less in surface area than the rectangular type and
being free of corners offer less resistance to airflow. However it is usually becomes
difficult to accommodate the round duct in the available space hence the rectangular
duct are generally adopted. Rectangular duct select should be as nearly a square as
possible. This is because the squire duct needs less surface area than a rectangular
one for the same cross sectional area. The rectangular section should be so sized as
to have low aspect ratio- the aspect ratio in any case should not exceed 4:1. Since
the aspect ratio of a square section is one, a square duct or a duct very near square
section is best suited for minimizing the initial and operating cost.
1. Duct route should be as direct as possible for limiting the length of the duct to minimize the frictional losses.
2. Air velocity in the duct must be kept reasonably low to minimize the frictional losses and the noise
level.
3. Long radius smooth bends must be used for changing the direction.
4. Turning vanes must be provided, in case a sharp right angle turn is inevitable.
5. Properly fabricated transformation pieces must be used for construction and
expansion in the duct sizes.
6. Since the smooth surface offers less resistance to the air flow than a rough surface,
galvanized iron or an aluminium sheet metal should be used for the fabrication of the
ducts, if other materials are used, pressure loss can increase due to the roughness of
the surface
7. Aspect ratio in the rectangular duct should be kept within 4:1. Higher aspect ratio
can also create turbulence.
5.4 Methods of DUCT DESIGN
The methods can be adopted for sizing the duct systems.
1. Equal friction method

61
2. Velocity reduction method
3. Static regain method
Equal friction method:-
In this method, the friction loss per unit length of duct is kept constant throughout
the system. From the recommended velocity table, a suitable velocity is selected for
the main duct from noise level considerations. Knowing the total air flow rate and
velocity having been selected, the cross sectional area of the main duct and friction
losses per unit length of the duct is arrived at from the “air friction chart”. The air
friction chart corresponds to round ducts. The frictional loss in rectangular duct is
arrived at by first finding the equivalent round duct for the rectangular duct, and
then reading the frictional loss from the air frictional chart from the round ducts. For
using the equivalent table, first one side of the rectangular duct has to be decided,
where there is sufficient space and no restriction for selecting the duct sizes, it is
advantageous to decide on square duct.
Air flow is found from the heat load and corresponds to the standard air. If the
specific volume of the air of the application is substantially different from that of
standard air, then the friction loss arrived at by reading them from the friction chart.
This chart provides the friction per unit length established for the main duct. The
velocity of the air flowing in the different sections and the branch of the duct system
is automatically reduced since the air quantity are reduced in the branches, but the
friction loss per unit length of the duct is kept constant. The total system of the
friction loss, which are fan must overcome, is arrived at by calculating the loss in the
duct that has the highest resistance.
In case both short and long run of branches are present, dampers will have to be
provided at the entrance of the short run ducts for balancing purposes.

62
Velocity reduction methods
The size of the main duct is established by selecting a velocity from the
recommended range, as is done in equal friction method. Thereafter, for every
branch take off, arbitrarily reduced velocities are selected. With the selected velocity
and the air quantity to be handled, duct size is determined. If rectangular types of
duct are to be used, an equivalent round diameter is first determined for arriving at
the pressure loss. Using the longest run of duct, the total pressure loss is computed
including the elbow or fittings to arrive at the fan static pressure required for
supplying the air quantity. This method trough simple is normally not used as this
does not take into account relative pressure loss in various branches. So for
branching the system damper have to be provided at each branch take off.
Static regains methods
When the velocity of the air in the duct reduced, the resulting velocity pressure
different gets converted into static pressure or a static regain is obtained. The
principle of the static regain method is to size the duct such that the increase in the
static pressure at each branch or terminal offsets the pressure loss in the succeeding
section of the duct. Thus the static pressure for each terminal is the same.
This method involves lengthy and complicated calculations, and hence is generally
not used.

63
E-20 SHEET of heat load calculation.
JMI Engineers
HEAT LOAD CALCULATIONS
JOB NAME: Hospital Building ( Bokaro)
SPACE FOR: Blood Bank( G.F Block - C)
SIZE: 3050 Sq ft 36600.00 Cu ft
Estima
te for:
SUM
MER
SOLAR GAIN GLASS
HEA
T
GAI
N
CON
DITIO
N
DB
(°F)
WB
(°F)
%
R
H DP (°F)
G
R/
LB
ITEM Area Sun Gain Factor
Btu/
hour
OUTS
IDE 100 82 46
13
8
(Sq ft)
(Btu/h.sq
ft)
ROO
M 75 64 55 71
N - Glass 0.00 14 0.2 0
DIFF
EREN
CE 25
XXX
X
X
X
X
X 67
NE - Glass 105.00 12 0.2 252
E - Glass 0.00 12 0.2 0
OUTSIDE AIR
(VENTILATION)
SE - Glass 180.00 12 0.2 432
People
X 10
CFM/Per
son
S - Glass 0.00 12 0.2 0 Sq ft X
CFM/Sq
ft
SW - Glass 0.00 100 0.2 0
CFM
VENTILA
TION=
54
6
W - Glass 0.00 164 0.2 0
NW - Glass 0.00 123 0.2 0
EFF. SENSIBLE HEAT
FACTOR (ESHF) = 0.91
Indicat
ed
ADP = 56 °F
SOLAR & TRANS. GAIN WALLS &
ROOF
Select
ed
ADP = 54.0 °F
ITEM Area
Eq. temp.
diff. U
Dehum. temp
rise =
18.9
0 °F
(Sq ft) (°F) (Btu/h.sq ft)
DEHUMIDIFIED
CFM = 5585
N - Wall 0.00 22 0.36 0
NE - Wall 535.00 28 0.36 5393
E - Wall 0.00 36 0.36 0
SE - Wall 600.00 36 0.36 7776
S - Wall 0.00 34 0.36 0
SW - Wall 0.00 32 0.36 0
W - Wall 0.00 30 0.36 0
NW - Wall 150.00 24 0.36 1296

64
Roof Sun 0.00 53 0.12 0
NOTE
S
TRANS. GAIN EXCEPT WALLS & ROOF
Occu
panc
y = 73
Sqft/
pers
on
ITEM Area Temp. diff. U
Light
= 1.4
W/Sq
ft
Eq.
Load
= 9.15 KW
All Glass 285.00 25 0.58 4133
Air Change per
hour = 1.00
Glass Partition 0.00 20 1.1 0
Partition wall 630.00 20 0.4 5040
Ceiling 0 10 0.4 0
Floor 3050 15 0.4
1830
0
INTERNAL HEAT GAIN
People 42 Nos X 245
1029
0
Light 4270.00 W X 1.25 3.4
1814
8
Eq. Load 9150.00 W 3.4
3111
0
ROOM
SENSIBLE HEAT
(RSH)
1021
69
Supply duct
heat gain +
Supply
duct leak.
loss +
Safety
factor (%) 10.0
1036
4
Outside & Infiltered Air
CFM °F BF FACTOR
0 25 1 1.08 0
546 25 0.1 1.08 1474
EFFECTIVE ROOM
SENSIBLE HEAT
(ERSH)
1140
07
LATENT HEAT
People 42 Nos X 205 8610
Permeation Load 0
Safety
factor % 5.0 555
CFM GR/LB BF FACTOR
0 67 1 0.68 0
546 67 0.1 0.68 2488
EFFECTIVE ROOM
LATENT HEAT
(ERLH)
1165
2
EFFECTIVE ROOM
TOTAL HEAT
(ERTH)
1256
60
OUTSIDE AIR HEAT (SENSIBLE)
CHECK
FIGURES
CFM °F 1 - BF FACTOR
Btu/h/
Sq ft 53.9

65
=
546 25 0.9 1.08
1326
8
CFM /
Sq ft
= 1.83
OUTSIDE AIR HEAT (LATENT)
Sq ft /
TR = 222
CFM GR/LB 1 - BF FACTOR
CFM/
TR
= 407
546 67 0.9 0.68
2238
8
HEAT SUB TOTAL
1613
16
Return duct
heat gain &
leak. loss +
HP Pump
+
Dehum. &
Pipe loss
(%)
2.0 3226
TONS 13.71
GRAND TOTAL
HEAT
1645
42
JMI Engineers
SPACE FOR: ICU ( G.F Block - B)
SIZE: 630 Sq ft 7560.00 Cu ft
Estima
te for:
SUM
MER
SOLAR GAIN GLASS
HEA
T
GAI
N
CON
DITIO
N
DB
(°F)
WB
(°F)
%
R
H DP (°F)
G
R/
LB
Btu/
hour
OUTS
IDE 100 82 46
13
8
(Sq ft)
(Btu/h.sq
ft)
ROO
M 72 60 50 58
N - Glass 0.00 14 0.2 0
DIFF
EREN
CE 28
XXX
X
X
X
X
X 80
NE - Glass 0.00 12 0.2 0
E - Glass 0.00 12 0.2 0
OUTSIDE AIR
(VENTILATION)

66
SE - Glass 0.00 12 0.2 0
People
X 0
CFM/Per
son
S - Glass 0.00 12 0.2 0 Sq ft X
CFM/Sq
ft
SW - Glass 0.00 100 0.2 0
CFM
VENTILA
TION=
32
8
W - Glass 0.00 164 0.2 0
NW - Glass 0.00 123 0.2 0
EFF. SENSIBLE HEAT
Indicat
ed
ADP = 56 °F
ROOF
Select
ed
ADP = 54.0 °F
ITEM Area
Eq. temp.
diff. U
Dehum. temp
rise =
16.2
0 °F
DEHUMIDIFIED
CFM = 977
N - Wall 0.00 22 0.36 0
NE - Wall 0.00 28 0.36 0
E - Wall 0.00 36 0.36 0
SE - Wall 0.00 36 0.36 0
S - Wall 0.00 34 0.36 0
SW - Wall 0.00 32 0.36 0
W - Wall 0.00 30 0.36 0
NW - Wall 0.00 24 0.36 0
Roof Sun 0.00 53 0.12 0
NOTE
S
Occu
panc
y = 90
Sqft/
pers
on
ITEM Area
Temp.
diff. U
Light
= 1.5
W/S
qft
Eq.
Load
= 1.26 KW
All Glass 0.00 28 1.1 0
Air Change per
hour = 2.00
Partition wall 0.00 23 0.4 0
Ceiling 0 13 0.4 0
Floor 630 18 0.4 4536
INTERNAL HEAT GAIN
Light 945.00 W X 1.25 3.4 4016
Eq. Load 1260.00 W 3.4 4284
ROOM
SENSIBLE HEAT
(RSH)
1455
1
Supply duct
heat gain +
Supply
duct leak.
loss +
Safety
factor (%) 10.0 1554

67
CFM °F BF FACTOR
0 28 1 1.08 0
328 28 0.1 1.08 991
EFFECTIVE ROOM
SENSIBLE HEAT
(ERSH)
1709
6
LATENT HEAT
Permeation Load 0
Safety
factor % 5.0 161
CFM GR/LB BF FACTOR
0 80 1 0.68 0
328 80 0.1 0.68 1782
EFFECTIVE ROOM
LATENT HEAT
(ERLH) 3378
EFFECTIVE ROOM
TOTAL HEAT
(ERTH)
2047
4
CHECK
FIGURES
CFM °F 1 - BF FACTOR
Btu/h/
Sq ft
= 73.6
328 28 0.9 1.08 8916
CFM /
Sq ft
= 1.55
Sq ft /
TR = 163
CFM/
TR
= 253
328 80 0.9 0.68
1603
9
HEAT SUB TOTAL
4542
9
Return duct
heat gain &
leak. loss +
HP Pump
+
Dehum. &
Pipe loss
(%)
2.0 909
TONS 3.86
GRAND TOTAL
HEAT
4633
8

68
JMI Engineers
SPACE FOR: Surgeon (M) ( F.F Block - B)
SIZE: 155 Sq ft 1860.00 Cu ft
Estima
te for:
SUM
MER
SOLAR GAIN GLASS
HEA
T
GAI
N
CON
DITIO
N
DB
(°F)
WB
(°F)
%
R
H DP (°F)
G
R/
LB
Btu/
hour
OUTS
IDE 100 82 46
13
8
(Sq ft)
(Btu/h.sq
ft)
ROO
M 75 64 55 71
N - Glass 0.00 14 0.2 0
DIFF
EREN
CE 25
XXX
X
X
X
X
X 67
NE - Glass 0.00 12 0.2 0
E - Glass 0.00 12 0.2 0
OUTSIDE AIR
(VENTILATION)
SE - Glass 0.00 12 0.2 0
People
X 5
CFM/Per
son
S - Glass 0.00 12 0.2 0 Sq ft X
CFM/Sq
ft
SW - Glass 0.00 100 0.2 0
CFM
VENTILA
TION= 20
W - Glass 0.00 164 0.2 0
NW - Glass 30.00 123 0.2 738
EFF. SENSIBLE HEAT
Indicat
ed
ADP = 56 °F
ROOF
Select
ed
ADP = 54.0 °F
ITEM Area
Eq. temp.
diff. U
Dehum. temp
rise =
18.9
0 °F
DEHUMIDIFIED
CFM = 271
N - Wall 0.00 22 0.36 0
NE - Wall 0.00 28 0.36 0
E - Wall 0.00 36 0.36 0
SE - Wall 0.00 36 0.36 0
S - Wall 0.00 34 0.36 0

69
SW - Wall 0.00 32 0.36 0
W - Wall 0.00 30 0.36 0
NW - Wall 102.00 24 0.36 881
Roof Sun 0.00 53 0.12 0
NOTE
S
Occu
panc
y = 52
Sqft/
pers
on
ITEM Area
Temp.
diff. U
Light
= 1.1
W/S
qft
Eq.
Load
= 0.16 KW
All Glass 30.00 25 0.58 435
Air Change per
hour = 1.00
Partition wall 0.00 20 0.4 0
Ceiling 0 10 0.4 0
Floor 155 15 0.4 930
INTERNAL HEAT GAIN
Light 170.50 W X 1.25 3.4 725
Eq. Load 155.00 W 3.4 527
ROOM
SENSIBLE HEAT
(RSH) 4971
Supply duct
heat gain +
Supply
duct leak.
loss +
Safety
factor (%) 10.0 502
CFM °F BF FACTOR
0 25 1 1.08 0
20 25 0.1 1.08 53
EFFECTIVE ROOM
SENSIBLE HEAT
(ERSH) 5526
LATENT HEAT
Permeation Load 0
Safety
factor % 5.0 35
CFM GR/LB BF FACTOR
0 67 1 0.68 0
20 67 0.1 0.68 89
EFFECTIVE ROOM
LATENT HEAT
(ERLH) 739
EFFECTIVE ROOM
TOTAL HEAT
(ERTH) 6265
CHECK
FIGURES
CFM °F 1 - BF FACTOR Btu/h/ 49.6

70
Sq ft
=
20 25 0.9 1.08 474
CFM /
Sq ft
= 1.75
Sq ft /
TR = 242
CFM/
TR
= 422
20 67 0.9 0.68 800
HEAT SUB TOTAL 7538
Return duct
heat gain &
leak. loss +
HP Pump
+
Dehum. &
Pipe loss
(%)
2.0 151
TONS 0.64
GRAND TOTAL
HEAT 7689

71
RESULT & DISCUSSION
For building of GOVERNMENT MEDICAL COLLEGE we have calculated heat load floor
wise and have found that total cooling load comes out to be 520 TR. For equipment
selection we have selected rotary screw compressor chilling machine, air handling
unit, cooling tower, split casing pumps, MS and GI pipes with Y-strainers, and valves
G.I Duct sheet.
TOTAL AREA OF BUILDING IN Sq Feet
Ground
floor
First floor Second
floor
Third
floor
Fourth
floor
Hospital Building 31190 14670 16680 15990 16770
Total cooling
Area Of
building
95200
HEAT LOAD SUMMARY IN TONNS
Ground
floor
First
floor
Second
floor
Third
floor
Fourth
floor
Load 203.7 81.3 72.3 56.7 107
Total 520

72
TOTAL AREA OF BUILDING IN Sq Feet
Ground
floor
First floor Second
floor
Third
floor
Medical College
Building
4075 7530 7450 3245
Total cooling
Area Of
building
22300
HEAT LOAD SUMMARY IN TONNS
Heat Load Calculation for Building
Ground
floor
First
floor
Second
floor
Third
floor
Load 31.24 46.9 50.8 21.06
Total 150

73
CONCLUSION
From the above study we conclude that this design can be used for
calculation of cooling load for any residential, government building and
commercial building with proper selection of equipment, duct lay outing
and piping.

74
BIBLIOGRAPHY
S.No. Authors Books
1 C.P. ARORA Refrigeration air conditioning
2 Carrier Mannual Carrier Air-conditioning Hand Book
3 Blue Star Charts and Specification Blue Star Manual Book
4 Gupta and Associate Gupta Hand Book Manual
5 P.L Ballaney Refrigeration and Air-conditioning
6 R.C Patel Refrigeration and Air-conditioning
7 Sarao & Ghambhir Refrigeration and Air-conditioning
8 Ananthanarayanan
Basic Refrigeration and Air-
Conditioning

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