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8/18/2019 Design Of Air Conditioning System For Auditorium
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Table Of Contents
1.
Objective…………………………………………………………………
2.
Introduction………………………………………………………………
3. Building Design & Floor Plan…………………………………………
4. Heat Load Calculations ………………………………………….........
5. System Selection based on Energy Efficiency and life Cycle
Analysis……………………………………………………………………….
Conclusion………………………………………………………………….
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Introduction
Building energy can be saved and pollution decreasedwhile utility expenditures are minimized if energy
conservation measures are incorporated into the design,
maintenance and operation of a facility. Building cooling
load components are; direct solar radiation, transmission
load, ventilation/infiltration load and internal load.
Calculating all these loads individually and adding them
up gives the estimate of total cooling load. The load, thus
calculated, constitutes total sensible load.
Normal practice is that depending on the building type
certain percent of it is added to take care of latent load.
Applying the laws of heat transfer and solar radiation
makes load estimations. Step by step calculation procedure has been adequately reported in the literature.
Principles of solar energy calculation are applied to
determine the direct and indirect solar heating component
of the building. The requisite data of building material
properties, climate conditions and ventilation standard are
also established as per the ISHRAE standards.
The one dimensional heat conduction equation in
rectangular, spherical and cylindrical coordinates is
solved using finite difference technique. The
variation of auditorium building temperature with time is
obtained in terms of wet bulb temperature of cooling air
and initial building temperature. Factors directly affecting
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thermal comfort of the human are air temperature,
moisture content of the air, radiant exchange and air
movement.
Location and Environmental Conditions
Our College Auditorium building which is to be designed
is located on the outskirts of Nagpur having coordinates
21.094796N, 78.980848E . Being located in tropical
region , Nagpur experiences harsh summers with
temperature rising as high as 48°C and dry winters with
temperature droping down to 4°C .
The ambient design temperatures for Nagpur as per
ISHRAE guidelines are tabulated below:
Summer
(2% Accept.)
Monsoon
(2% Accept.)
Winter
(99% Accept.)
Dry Bulb Temp. – 41.4 C Dry Bulb Temp – 26.2C Dry Bulb Temp. – 11.5C
Mean wet bulb temp –23.6 C Mean Wet Bulb Temp - 31.9C Mean Wet Bulb Temp – 9.4C
Design temperatures for summer and monsoon are
selected for 2% acceptance conditions to achieve higher
accuracy in calculations and that for winter are selected
for 99% acceptance conditions.
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Building Design & Floor Plan
Our college auditorium building being constructed on ahilly contour poses a unique task in design of its HVAC
system. It is somewhat covered by a hill from the North-
West side which allow a very little or almost no solar
radiation to enter from this direction.
Being built on the first floor and located on a hillycontour, this building has been constructed by making the
floors offset to each other. Most of the windows are
located on the north-west wall so that almost no heat
enters through these windows. Also, the area surrounding
the auditorium especially on the north-west side is
covered with trees which also entraps some of the
radiation.
The auditorium is built with a height of 4.572 meters or
15 feet with concrete steps for seating from the inside
which also adds to insulation. Our college auditorium
encompasses a total of 800 people which is fairly justified
with the NSDC guidelines. As the auditorium is built on a
basement but due to hilly contour the complete area of basement roof covers only half of the area of the
auditorium floor. This auditorium has a peaked roof with
a false ceiling with attic ventilation. The walls of
auditorium are cladded with plywood from inside.
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Rest of the features can be seen from the floor plan shown
below:
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Heat Load Calculations
The net heat load from a building is a combined effect ofthe following factors:-
Solar heat gain through walls and roof (fabric heat
gain); sensible in nature.
Heat gain through fenestration (transmitted andradiated heat through glass windows); sensible in
nature.
Load due to occupants inside the building; sensible
and latent in nature.
Load due to ventilation and infiltration; sensible and
latent in nature.
Load due to lighting; sensible in nature and due to
electrical appliances; sensible as well as latent in
nature.
Before beginning with the heat load calculations, we need
to define the inside design conditions which are to be met
by HVAC system. The inside dry bulb temperature of theunconditioned building can be predicted for the given
ambient temperature using Humphrey’s Thermal
Neutrality correlation for tropical regions:
Ti=0.534T0+12.9
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This correlation gives the optimum temperature at which
the occupants would feel comfortable or at least would
not feel uncomfortable. The following table enlists inside
temperature for different seasons obtained using abovecorrelation:
Summer Monsoon Winter
Dry Bulb Temp – 35.0076C Dry Bulb Temp – 26.8908C Dry Bulb Temp – 19.041C
The inside design conditions for the building space by
considering ASHRAE comfort chart , the most suitable
conditions for the building have been selected as follows:-
Dry Bulb Temperature = 24˚C
Wet Bulb Temperature = 15.52˚C
Relative Humidity = 40%
Humidity Ratio = 0.00742 kg/kgDA
Dew Point Temperature = 9.57˚C
Specific Volume = 0.8510 m3 /kg
Specific Enthalpy = 43 KJ/KgDA
The detailed calculation procedure is elaborated below
considering all the above parameters-
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FABRIC HEAT GAIN:-
Considering the scope of this competition we are here
adopting the ASHRAE recommended CLTD-CLF
method, which gives considerably accurate results, for theestimation of the solar heat gain through walls, doors and
roof.
1. Through walls
Solar heat gain through walls is given by the equation-
= × ×
Where, U is the overall heat transfer coefficient through
the wall and is given as-
= 1
+
+(×)
Where
R is the thermal resistance of the wall
hi is the inside film coefficient = 8.347 W/m2-K (still air)
ho is the outside film coefficient = 23.3 W/m2-K(3.7 m/s)
A is the cross-sectional area for the heat flow
CLTD value for different walls facing a particular
direction at different solar times is obtained from
ISHRAE handbook. The maximum of these values is
selected for calculations.
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Thermal resistance of the wall is calculated for the
composition of wall shown below:-
For cement plaster , L=0.0127m and k = 56.782 W/m-K
For face brick , L =0.1016m and k = 12.886 W/m-K
For concrete block, L =0.1016m and k = 7.994 W/m-K
For plywood, L =0.1363m and k = 6.018 W/m-K
Note: Due to the presence of concrete steps to occupy the
audience, the resistance due to the area of the wall with
steps and the resistance due to rest of the wall area are to
be considered in parallel combination with each other.
This arrangement is represented by thermal circuit shown
below:
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Having known the values of thermal resistances, overall
heat transfer coefficient following which corresponding
heat load for different walls can be calculated.
2. Through Doors
The heat gain through the doors is calculated by taking
into account the design temperature difference instead of
cooling load temperature difference as no time lag in
radiative heat transfer occurs though the doors.
= × × ( − )
The doors are made of wood of 1 inch thickness with
conductivity k = 6.234 W/m-k and have area of 1.44m2
each. The no. of doors on each wall are tabulated below :
Direction of Wall No. of Doors
North-East 1 – Double Door
2 – Single Doors
South-West 1 – Double Door
2 – Single Door
South-East 2 – Double Doors
3. Through Roof
As the auditorium has peaked roof which is attic
ventilated with a false ceiling below it, the CLTD values
from Table 10 of ISHRAE Handbook are reduced by 25%
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and the roof area is taken as the projected area of the
peaked roof.
The calculations are performed based on this knowledge.
HEAT GAIN THROUGH
FENESTRATION :- The transfer of heat is accounted by two modes viz.
conduction and radiation. The governing equations for
each of these mode are :-
= × × ×
= × × (0 − )
Where
SHGFmax = maximum solar heat gain factor through glass
based on table 7 of ISHRAE Handbook
SC = shading coefficient based on table 5 of ISHRAE
Handbook (selecting double pane ordinary glass for
horizontal window and regular plate glass for vertical
window)CLF = cooling load factor for glass without interior
shading (based on direction and solar time)
U = Overall heat transfer coefficient based on Table 6 of
ISHRAE Handbook = 3.12 W/m2-K
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direction orientation SC SHGFmax CLF
NE horizontal 0.9 154 0.45
NE vertical 0.94 154 0.45SW Horizontal 0.9 167 0.59
SW Vertical 0.94 167 0.59
LOAD DUE TO OCCUPANTSInternal heat load due to occupants consists of both
sensible and latent components which can be calculatedas:-
Qu = (No. of people) ×sensible heat gain
person × CLF
= (. ) ×
Since the latent heat gain from the occupants is
instantaneous, the CLF for latent heat gain is 1 and the
value of CLF for sensible heat gain is taken as 0.5.
LOAD DUE TO INFILTRATION ANDVENTILATION
The heat load due to infiltration is calculated using ACH
method by taking ACH = 0.5 air changes/hr for a well-
sealed building. This heat load is in the form of sensible
as well as latent load which are given as:-
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=(To-Ti)
= ℎ( − )
Where
Vo is the volumetric flow rate of the infiltrated air
C pm is the average specific heat of moist air
hfg is the latent heat of vaporization of water
To and Ti are the outdoor and indoor dry bulb temperatures
Wo and Wi are the outdoor and indoor humidity ratios.
is the density of moist air at outsidetemperature(calculated using perfect gas equation)
= ×
3600 =0.64925 m3/sec
Where gross volume = total volume of conditioned space
= 4674.64m3
The heat load due to ventilation is calculated in similar
fashion as :
= ( − )
= ℎ( − )
Where
is the volumetric flow rate for ventilated air which istaken as 15 cfm per person as per ASHRAE guidelines but
to maintain the indoor air quality for comfort as per
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ISHRAE standards this quantity is reduced to 5 cfm
/person
BPF is the bypass factor of the cooling coil which isselected based on the applicatons from table 14 of ISHRAE
Handbook
LOAD DUE TO LIGHTINGThe heat load for lighting is calculated for two types of
lights viz. spotlights (incandescent) and fluorescent lights.
Basically the heat load due to lighting is calculated usingthe following equation:
Q = (installed wattage)(Usage Factor)(Ballast Factor)CLF
Where
Installed wattage is the total input power to the lights in the
conditioned space
Usage Factor accounts for any lights that are installed but
are not switched on at the time at which load calculations
are performed
Ballast factor takes into account the load imposed by
ballasts used in fluorescent lights(ballast factor value of
1.25 is taken for fluorescent lights, while it is equal to 1.0
for incandescent lamps)
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CLF is function of the number of hours after the lights are
turned on, type of lighting fixtures and the hours of
operation of the lights(CLF value of 0.73 is selected for
fluorescent lights whereas CLF=0.1 is selected forspotlights)
LOAD DUE TO APPLIANCES
The only running appliances inside the auditorium are
fans present on each of the columns which are 10 in
number and the heat load consists of two parts viz.sensible and latent load which are calculated as:
= ( ) × ( ) ×
= ( ) × ( ℎ )
Each fan has an installed wattage of 100W and the usage
factor is assumed to be 0.8 based on hours of operation
while CLF is selected as 0.58 .
Latent heat fraction of the fan is taken as 0.07.
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Job Name Design of HVAC System Estimated for Local Solar Peak
Address Wanadongri,Nagpur Time Load
Space Used for Auditorium Summer 3 P.M. 44.4 deg. C
Size 24.830m × 43.09m = 1069.9247 sq. m × 4.572m = 4891.695cu. M
Watts CONDITIONS DB(deg. C) WB(deg.C) %RH HUMIDITY RATI O(kg/kgDA)
Item Area or Quantity Sun Gain or Temp. Diff. Factor Sensible Latent Total Outdoor 41.4 23.6 21.6 0.01075
Indoor 35 19.44 21.6 0.00775
SOLAR GLASS GAIN Selected 24 15.52 40 0.00742
Window(NE) 5.05 17.4 14.47 1271.84 Room Conditions
Window(SW) 4.24 17.4 19.52 1440.61 VENTILATION
1000 People .0023595 cu. m/sec/person= 2.3595 cu. m/sec
SOLAR &TRANS. GAIN ‐ WALLS & ROOF
Wall(NE) 174.96 17.95 5.0709 15925.32 INFILTRATION
Wall(SW) 175.77 20.18 5.071 17987.03 Gross Air Changes
Wall(SE) 99.45 19.07 5.576 10574.95 Volume 4674.64 cu. M per sec. 0.00013 .64925cu. m/sec
Roof 1022.45 16.26 4.799 79783.55 SENSIBLE HEAT FACTOR &
TRANS. GAIN EXCEPT WALLS AND ROOF APPARATUS DEW POINT
Floor 639.17 6 2.9 11121.55
Door(NE) 11.52 17.4 5.994 1201.485 ESHF = 214986.4/274573.2 0.7829
Door(SW) 11.52 17.4 5.994 1201.485
Door(SE) 11.52 17.4 5.994 1201.485 Indicated adp = 4.44 deg. C Selected adp = 9.57 deg. C
INFILTRATION AND OUTSIDE AIR
Volume Density Specific Heat (1‐.075)(24‐9.57)= 13.34 deg. C Dehumidified rise
Infiltration 0.64925 1.0902 1021.6 17.4 12581.95
Outside Air 2.3597 1 .0902 1 021.6 .075(BPF) 17.4 3429.68 195442.2/(1.0902*1021.6*13.34) =13.15 cu. m/sec Dehumidified flow rate
INTERNAL HEAT
People 1000 People 70.337W/person 0.5(CLF) 35168.5
Lights
Fluorescent Lights 50 Nos. 0.7(Usage Factor) 60W/light 1.25(Ballast Factor) 0.73(CLF) 1916.25
Sp ot li ght s 10 Nos . 0 .3 (U sag e Fac tor ) 5 75 W/ li ght 1(B al la st F ac tor ) 0 .1 (CL F) 17 2. 5
Appliance (Fan) 10 Nos. 0.8(Usage Factor) 100W/fan 0.58(CLF) 464
ROOM SENSIBLE HEAT 195442.2
Supply Duct Supply Duct Fan Safety
Heat Gain% 3% Leakage Loss% 2% H.P.% 5% Factor 1.1
EFFECTIVE ROOM SENSIBLE HEAT 214986.4
ROOM LATENT HEAT
Volume Density Humidity Diff. Enthalpy
Infiltration 0.64925 1.0902 0.00333 2403340 5664.7
Outside Air 2.3597 1.0902 .075(BPF) 0.00333 2403340 1544.129
People 1000People 46.891W/person 46891
Steam
Appliance (Fan) 10 Nos. 100W/fan 0.07(Latent Heat Fraction) 70
Room Latent Heat Subtotal 54169.83
Supply Duct Safety
Leakage Loss% 2% Factor% 8%
EFFECTIVE ROOM LATENT HEAT 59586.81
EFFECTIVE ROOM TOTAL HEAT 274573.2
OUTSIDE AIR HEAT
Sensible 2.3597 1.0902 1021.6 0.05(BPF) 17.4(Temp. Diff.) 43442.68
Latent 2.3597 1.0902 0.05(BPF) 0.00333 2403340 19558.97
Grand Heat Sub‐total 337574.9
Return Duct Return Duct Pump
Heat Gain% 5% Leakage Loss% 2% H.P.% 5%
Grand Heat Total 378083.8
TONNAGE OF REFRIGERATION = 107.50 TR
HEAT LOAD ESTIMATION SHEET
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System Selection based on Energy Efficiency
And Life Cycle Analysis
Selection of a suitable air conditioning system depends
on:
1. Capacity, performance and spatial requirements
2. Initial and running costs3. Required system reliability and flexibility
4. Maintainability
5. Architectural constraints
The relative importance of the above factors varies from building owner to owner and may vary from project to
project. The typical space requirement for large air
conditioning systems may vary from about 4 percent to
about 9 percent of the gross building area, depending
upon the type of the system.
Considering a system capacity of 108 TR and a single
zone system for auditorium, we provide a comparative
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analysis of available HVAC systems known to us which
are :-
All Water Systems
All Air Systems
Unitary Refrigerant Systems
Storage Cooling Systems
1. All Water Systems
In all water systems the fluid used in the thermaldistribution system is water, i.e., water transports energy
between the conditioned space and the air conditioning
plant. When cooling is required in the conditioned space
then cold water is circulated between the conditioned
space and the plant, while hot water is circulated through
the distribution system when heating is required. Since
only water is transported to the conditioned space, provision must be there for supplying required amount
of treated, outdoor air to the conditioned space for
ventilation purposes. Depending upon the number of
pipes used, the all water systems can be classified into a
2-pipe system or a 4-pipe system.
A type of all water system which is generally
commercially used is the Central Chilled Water
System which consists of a chilled water plant which is
remotely located with only AHUs being close to the
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conditioned space.The chilled water system may be air
cooled or water cooled .
Advantages of All Water Systems
1. The thermal distribution system requires very less
space compared to all air systems. Thus there is no
penalty in terms of conditioned floor space. Also the
plant size will be small due to the absence of large
supply air fans.
2. Individual room control is possible, and at the sametime the system offers all the benefits of a large central
system.
3. Since the temperature of hot water required for space
heating is small, it is possible to use solar or waste heat
for winter heating.
4. It can be used for new as well existing buildings
(retrofitting).
5. Simultaneous cooling and heating is possible with 4-
pipe systems.
Disadvantages of All Water System
1. Requires higher maintenance compared to all air
systems, particularly in the conditioned space.
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2. Draining of condensate water can be messy and may
also create health problems if water stagnates in the
drain tray. This problem can be eliminated, if
dehumidification is provided by a central ventilation
system, and the cooling coil is used only for sensible
cooling of room air.
3. Generally involves high initial costs.
4. Control of humidity, particularly during summer is
difficult using chilled water control valves.Prime candidates for using such systems would be large
convention centres with less external walling when
compared to internal floor space.Such structures have
internal service cores which tend to use only small
areas.
2.All Air Systems
As the name implies, in an all air system air is used as
the media that transports energy from the conditioned
space to the A/C plant. In these systems air is processed
in the A/C plant and this processed air is then conveyed
to the conditioned space through insulated ducts using blowers and fans. This air extracts (or supplies in case
of winter) the required amount of sensible and latent
heat from the conditioned space. The return air from the
conditioned space is conveyed back to the plant, where
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it again undergoes the required processing thus
completing the cycle. No additional processing of air is
required in the conditioned space. All air systems can
be further classified into:
1. Single duct systems, or
2. Dual duct systems
One of the all air systems is the Central DX System
which is well suited for single zone applications by
locating the equipment properly and providing for the
usual acoustic attenuation, the noise of the plant can be
kept within limits .Generally these systems may have to
be water cooled so that the heat rejection equipment
like cooling towers can be remote located from the
plant.
Advantages of All Air Systems are:
a) Relatively small space requirement
b) Excellent temperature and humidity control over a
wide range of zone loads
c) Proper ventilation and air quality in each zone is
maintained as the supply air amount is kept constant
under all conditions
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Disadvantages of All Air Systems are :
a) High energy consumption for cooling, as the air is
first cooled to a very low temperature and is then heated
in the reheat coils. Thus energy is required first for
cooling and then for reheating. The energy consumption
can partly be reduced by increasing the supply air
temperature, such that at least one reheat coil can be
switched-off all the time. The energy consumption can
also be reduced by using waste heat (such as heat
rejected in the condensers) in the reheat coil.
b) Simultaneous cooling and heating is not possible.
Prime candidates for such applications are very large
auditoriums, when built in exclusive buildings .Large
indoor auditoriums calling for,say,1500 tons of cooling
could be economically cooled with 10 × 150 ton plants
3.Unitary Refrigerant Systems
Unitary refrigerant based systems consist of several
separate air conditioning units with individual
refrigeration systems. These systems are factory
assembled and tested as per standard specifications, and
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are available in the form of package units of varying
capacity and type. Each package consists of refrigeration
and/or heating units with fans, filters, controls etc.
Depending upon the requirement these are available in the
form of window air conditioners, split air conditioners,
heat pumps, ductable systems with air cooled or water
cooled condensing units etc. The capacities may range
from fraction of TR to about 100 TR for cooling.
Depending upon the capacity, unitary refrigerant based
systems are available as single units which cater to asingle conditioned space, or multiple units for several
conditioned spaces. Figure 36.9 shows the schematic of a
typical window type, room air conditioner, which is
available in cooling capacities varying from about 0.3 TR
to about 3.0 TR. As the name implies, these units are
normally mounted either in the window sill or through thewall.
One of the unitary refrigerant systems that is
commercially used for conditioning is Packaged
Equipment System . With large capacity ,reliable ,
factory-made equipment being available at unmatchable
costs , one can use such equipment also for auditoriums.
Multiple package units/ duct able splits can be used well.
Factory made comfort equipment with cooling coils whih
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2. Power consumption per TR could be higher compared
to central systems.
3. Close control of space humidity is generally difficult.
4. Noise level in the conditioned space could be higher.
5. Limited ventilation capabilities.
6. Systems are generally designed to meet the appliance
standards, rather than the building standards.
7. May not be appealing aesthetically.8. The space temperature may experience a swing if on-
off control is used as in room air conditioners.
9. Limited options for controlling room air distribution.
Prime candidates for using such systems are smallcapacity halls used by educational institutions.This,of
coure, gets stretched, to systems being used for large
assembly areas like marriage halls,community centers
,etc.
4. Storage Cooling Systems
On specific applications,such as temple
halls,churches,etc. where one needs cooling only for
,say,three hours a day and even that,only once a
week,storage systems can be used.Thermal storage
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systems can be as simple as the “ice storage ” ones, or as
sophisticated as “eutectic salt in custom containers”.
Costs will dictate the use of low end systems, but with ice
systems using direct ice melt, one may need to have an
AHU with a greater than normal coil bypass area.
As we can see from above that storage cooling system is
not a good choice for auditoriums as these systems can
efficiently work if it is operated for only 3 or 4 years aweek as these systems primarily run on ice and are not
capable to provide conditioning for long durations and
also require a considerable maintenance cost if stretched
for large capacities .So storage cooling systems are not
used in air conditioning purpose primarily.
From the economic as well as service point of view ,these
systems are not efficient for large capacities even though
the initial costs are low but the maintenance costs turn out
to be considerably high enough
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Conclusion
So, we have selected Packaged Equipment System out
of other alternatives for air conditioning of auditoriums as
this system being a compact alternative is quite efficient
in operation. Though the installation cost being high
comparative to other alternatives the maintenance cost islow for such systems with a fair enough service life.
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