Design Of Air Conditioning System For Auditorium

National Student Design Competition 2016

Final Report (On Design of HVAC System for College Auditorium)

Submitted ByYeshwantrao Chavan College Of Engineering Nagpur Team Members: Pushkar A. Pandit Shriwardhan Gaurihar Sarang Siras

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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Objective As posed by ISHRAE this year, we are tasked with the design of HVAC system for a College Auditorium with a design capacity of nearly 1000 people. Moreover, as the problem suggests to design the building based on the climate zone as well as the location of local Student Chapter. So we hereby select ‘Shri Datta Meghe Auditorium’ located in our college campus which best suits the requirements of this project.

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Introduction Building energy can be saved and pollution decreased while 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 a hilly contour poses a unique task in design of its HVAC system. It is somewhat covered by a hill from the NorthWest 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 hilly contour, 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 of the following factors:Solar heat gain through walls and roof (fabric heat gain); sensible in nature. Heat gain through fenestration (transmitted and radiated 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 the unconditioned 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 above correlation: Summer

Dry Bulb Temp – 35.0076C

Monsoon

Dry Bulb Temp – 26.8908C

Winter

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 the estimation of the solar heat gain through walls, doors and roof. 1. Through walls Solar heat gain through walls is given by the equation𝑄 = 𝑈 × 𝐴 × 𝐶𝐿𝑇𝐷 Where, U is the overall heat transfer coefficient through the wall and is given as-

𝑈=

1 1 1 + +(𝑅×𝐴) ℎ𝑖 ℎ𝑜

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. 𝑄 = 𝑈 × 𝐴 × (𝑇𝑜 − 𝑇𝑖 ) 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 North-East South-West South-East

No. of Doors 1 – Double Door 2 – Single Doors 1 – Double Door 2 – Single Door 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 :𝑄𝑤𝑖𝑛𝑑𝑜𝑤𝑟 = 𝐴𝑢𝑛𝑠ℎ𝑎𝑑𝑒𝑑 × 𝑆𝐻𝐺𝐹𝑚𝑎𝑥 × 𝑆𝐶 × 𝐶𝐿𝐹 𝑄𝑤𝑖𝑛𝑑𝑜𝑤𝑐 = 𝑈 × 𝐴 × (𝑇0 − 𝑇𝑖 ) 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 NE NE SW SW

orientation horizontal vertical Horizontal Vertical

SC 0.9 0.94 0.9 0.94

SHGFmax 154 154 167 167

CLF 0.45 0.45 0.59 0.59

LOAD DUE TO OCCUPANTS Internal heat load due to occupants consists of both sensible and latent components which can be calculated as:Q soccupants = (No. of people) ×

sensible heat gain × CLF person

𝑄𝑙𝑜𝑐𝑐𝑢𝑝𝑎𝑛𝑡𝑠 = (𝑁𝑜. 𝑜𝑓 𝑝𝑒𝑜𝑝𝑙𝑒 ) ×

𝐿𝑎𝑡𝑒𝑛𝑡 𝐻𝑒𝑎𝑡 𝐺𝑎𝑖𝑛 𝑝𝑒𝑟𝑠𝑜𝑛

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 AND VENTILATION The heat load due to infiltration is calculated using ACH method by taking ACH = 0.5 air changes/hr for a wellsealed building. This heat load is in the form of sensible as well as latent load which are given as:-

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𝑄𝑠𝑖𝑛𝑓𝑖𝑙𝑡𝑟𝑎𝑡𝑒𝑑 =𝜌𝑜 𝑉𝑜 𝐶𝑝𝑚 (To-Ti) 𝑄𝑙𝑖𝑛𝑓𝑖𝑙𝑡𝑟𝑎𝑡𝑒𝑑 = 𝜌𝑜 𝑉𝑜 ℎ𝑓𝑔 (𝑊𝑜 − 𝑊𝑖 ) Where Vo is the volumetric flow rate of the infiltrated air Cpm 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. 𝜌𝑜 is the density of moist air at outside temperature(calculated using perfect gas equation) 𝑉𝑜 = 𝐴𝐶𝐻 ×

𝐺𝑟𝑜𝑠𝑠 𝑉𝑜𝑙𝑢𝑚𝑒 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 : 𝑄𝑠𝑣𝑒𝑛𝑡 = 𝑉𝑣𝑒𝑛𝑡 𝐵𝑃𝐹𝜌𝑜 𝐶𝑝𝑚 (𝑇𝑜 − 𝑇𝑖 ) 𝑄𝑙𝑣𝑒𝑛𝑡 = 𝑉𝑣𝑒𝑛𝑡 𝐵𝑃𝐹𝜌𝑜 ℎ𝑓𝑔 (𝑊𝑜 − 𝑊𝑖 ) Where 𝑉𝑣𝑒𝑛𝑡 is the volumetric flow rate for ventilated air which is taken 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 is selected based on the applicatons from table 14 of ISHRAE Handbook

LOAD DUE TO LIGHTING The 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 using the following equation: Q slighting = (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 for spotlights)

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: 𝑄𝑠𝑎𝑝𝑝𝑙𝑖𝑎𝑛𝑐𝑒𝑠 = (𝑖𝑛𝑠𝑡𝑎𝑙𝑙𝑒𝑑 𝑤𝑎𝑡𝑡𝑎𝑔𝑒) × (𝑢𝑠𝑎𝑔𝑒 𝑓𝑎𝑐𝑡𝑜𝑟) × 𝐶𝐿𝐹 𝑄𝑙𝑎𝑝𝑝𝑙𝑖𝑎𝑛𝑐𝑒𝑠 = (𝐼𝑛𝑠𝑡𝑎𝑙𝑙𝑒𝑑 𝑤𝑎𝑡𝑡𝑎𝑔𝑒) × (𝐿𝑎𝑡𝑒𝑛𝑡 ℎ𝑒𝑎𝑡 𝑓𝑟𝑎𝑐𝑡𝑖𝑜𝑛)

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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HEAT LOAD ESTIMATION SHEET Job Name Design of HVAC System Address Wanadongri,Nagpur Space Used for Auditorium Size 24.830m × 43.09m = 1069.9247 sq. m × 4.572m = 4891.695cu. M

Item

Area or Quantity

Estimated for Summer

Sun Gain or Temp. Diff.

Watts Sensible

Factor

Latent

Total

SOLAR GLASS GAIN Window(NE) Window(SW)

5.05 4.24

17.4 17.4

14.47 19.52

1271.84 1440.61

CONDITIONS Outdoor Indoor Selected  Room Conditions 1000 People

Local Solar Time 3 P.M.

Peak  Load 44.4 deg. C

DB(deg. C) WB(deg.C) %RH HUMIDITY RATIO(kg/kgDA) 41.4 23.6 21.6 0.01075 35 19.44 21.6 0.00775 24 15.52 40 0.00742 VENTILATION .0023595 cu. m/sec/person=

2.3595 cu. m/sec

SOLAR &TRANS. GAIN ‐ WALLS & ROOF Wall(NE) Wall(SW) Wall(SE)        Roof 

174.96 175.77 99.45 1022.45

17.95 20.18 19.07 16.26

5.0709 5.071 5.576 4.799

15925.32 17987.03 10574.95 79783.55

6 17.4 17.4 17.4

2.9 5.994 5.994 5.994

11121.55 1201.485 1201.485 1201.485

Gross  Volume

TRANS. GAIN EXCEPT WALLS AND ROOF Floor Door(NE) Door(SW) Door(SE)

Infiltration Outside Air People  Lights Fluorescent Lights Spotlights Appliance (Fan)

Supply Duct Heat Gain%

Infiltration Outside Air People Steam Appliance (Fan)

639.17 11.52 11.52 11.52 INFILTRATION AND OUTSIDE AIR Volume Density Specific Heat 0.64925 1.0902 1021.6 2.3597 1.0902 1021.6 .075(BPF) INTERNAL HEAT 1000 People 70.337W/person 50 Nos. 10 Nos. 10 Nos.

 0.7(Usage Factor) 0.3(Usage Factor) 0.8(Usage Factor)

17.4 17.4

60W/light  1.25(Ballast Factor) 575W/light 1(Ballast Factor) 100W/fan

0.5(CLF)

35168.5

0.73(CLF) 0.1(CLF) 0.58(CLF)

1916.25 172.5 464

1.1 214986.4

Enthalpy 2403340 2403340

5664.7 1544.129 46891

0.07(Latent Heat Fraction)

Sensible Latent

Return Duct Heat Gain%

Return Duct 5% Leakage Loss%

70 54169.83

59586.81 274573.2 17.4(Temp. Diff.) 0.00333

Pump  2% H.P.%

43442.68 2403340

0.7829 Selected adp = 9.57 deg. C Dehumidified rise

195442.2/(1.0902*1021.6*13.34) =13.15 cu. m/sec

195442.2 Safety  5% Factor

100W/fan

Safety 2% Factor% 8% EFFECTIVE ROOM LATENT HEAT EFFECTIVE ROOM TOTAL HEAT OUTSIDE AIR HEAT 2.3597 1.0902 1021.6 0.05(BPF) 2.3597 1.0902 0.05(BPF)

Indicated adp = 4.44 deg. C

12581.95 3429.68

Room Latent Heat Subtotal Supply Duct  Leakage Loss%

214986.4/274573.2

(1‐.075)(24‐9.57)= 13.34 deg. C

ROOM SENSIBLE HEAT Supply Duct  Fan  3% Leakage Loss% 2% H.P.% EFFECTIVE ROOM SENSIBLE HEAT ROOM LATENT HEAT Volume Density Humidity Diff. 0.64925 1.0902 0.00333 2.3597 1.0902 .075(BPF) 0.00333 1000People 46.891W/person 10 Nos.

ESHF = 

INFILTRATION Air Changes 4674.64 cu. M per sec. 0.00013 .64925cu. m/sec SENSIBLE HEAT FACTOR & APPARATUS DEW POINT

19558.97

Grand Heat Sub‐total

337574.9

Grand Heat Total

378083.8

5%

TONNAGE OF REFRIGERATION = 107.50 TR

Dehumidified flow rate

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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 costs 3. 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 thermal distribution 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 same time 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 4pipe 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 a single 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 the wall. 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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are only 3 rows deep – theoretical does not meet the ‘adp’ needs of the application, but in practice such equipment has been used with the ensuing, higher relative humidity never posing any serious problem. Advantages of unitary refrigerant systems are : 1. Individual room control is simple and inexpensive. 2. Each conditioned space has individual air distribution with simple adjustment by the occupants. 3. Performance of the system is guaranteed by the manufacturer. 4. System installation is simple and takes very less time. 5. Operation of the system is simple and there is no need for a trained operator. 6. Initial cost is normally low compared to central systems. 7. Retrofitting is easy as the required floor space is small. Disadvantages of Unitary refrigerant systems are: 1. As the components are selected and matched by the manufacturer, the system is less flexible in terms of air flow rate, condenser and evaporator sizes.

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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 onoff control is used as in room air conditioners. 9. Limited options for controlling room air distribution. Prime candidates for using such systems are small capacity 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 a week 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 is low for such systems with a fair enough service life.

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