Design Of Air Conditioning System For Auditorium

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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Design Of Air Conditioning System For Auditorium

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