Pumping Station Desing - Second Edition by Robert L. Sanks, George Tchobahoglous, Garr M. Jones

against which both

the

supply

and

exhaust

fans

must

operate. When used

in air

conditioning systems, such

energy-recovery devices

can

also precool

the hot

makeup

air by

thermal exchange with cooler exhaust

air. This dual

use

makes them more economical

in

applications requiring heating

in

cold weather

and

refrigerated

air

conditioning

in

summer.

23-6. Corrosion Protection

Metallic equipment, especially ductwork,

is

suscepti-

ble to

deterioration

in

varying degrees

due to

mois-

ture,

acid

(from

hydrogen

sulfide

and

moisture),

and

salt

often

contained

in the air in wet

wells.

Raw

edges

of cut

metal sheets

and

surface

scratches pro-

vide

a

starting point

for

oxidation

to

spread under

coatings

of

metals such

as the

zinc

on

galvanized

steel. Type

316

stainless steel

is

generally corrosion

resistant,

but its

discoloration

may be

unacceptable

aesthetically. Copper

and

copper-bearing

materials

are

highly susceptible

to

attack

by

hydrogen

sulfide

and

should

be

avoided

or

protected

in

wastewater sta-

tions.

Aluminum

should

not be

used

in

damp chlorine

rooms. Where metals

are the

logical choice

for

items

such

as

louvers, dampers,

and

grilles, they should

be

type

316

stainless steel, aluminum,

or

galvanized

steel,

or be

protected with

an

epoxy

or

baked phe-

nolic corrosion-resistant coating applied

by the

man-

ufacturer.

Polyvinyl

chloride (PVC)

or fiberglass-reinforced

plastic (FRP)

are

more corrosion resistant (and more

expensive) than type

316

stainless steel,

but

their

detracting

characteristics

—

such

as fire and

smoke rat-

ings, structural strength, greater weight

and

support

requirements,

and the

greater

difficulty

of

on-site

modification

—

must

be

given

due

consideration.

PVC

and

FRP

ductwork

are

usually factory-fabricated

from

sheets

of

material

4.8 mm

(3/16 in.)

or

more

in

thick-

ness. Joints

in

either type

of

duct

can be

made

by

sol-

vent

welding.

FRP

joints also

can be

formed with

multiple

layers

of

glass cloth laid

up by

hand

in a

resin

binder. Duct accessories used, such

as

dampers, turn-

ing

vanes, hangers,

and

fasteners, should

be

equally

corrosion resistant.

Fans,

tanks,

and

piping

may

also

be

fabricated

from

PVC and

FRP. Heat exchangers (which must

be

metal

for

good heat

transfer)

should

be

protected

by

corro-

sion-resistant coatings, such

as

thermosetting

phenolics.

The

metal

is

dipped

and

then oven baked

for

several

hours.

Two or

more coats

may be

used

for

severe appli-

cations. Epoxy compounds

and

some phenolics

can be

applied cold

by

painting

or

spraying.

The

resistance

of

the

coatings

to the

various causes

of

corrosion

can be

obtained

from

their manufacturers.

The

heat-transfer

capability

of

corrosion-protected heat exchangers

may

be

decreased

by 15% or

more,

so

compare cost, energy,

and

permanence

in

making

a

selection.

The

types

and

severity

of

corrosion

to be

expected

should

be

ascertained

from

previous installations

at

the

same

or at a

similar

site,

if

possible.

Also consult

experienced coating vendors.

The

difficulty

and

cost

of

replacing unprotected components should

be

weighed against

the

cost

for

protection.

Note that cor-

rosion protection increases

first

cost

from

20 to

400%

but

replacement costs

can be

even higher.

Electrical switches

and

control components should

also

be

protected

from

corrosion. Such protection

can

be

provided

in a

number

of

different

ways.

•

Control hydrogen

sulfide

at the

source

to the

great-

est

extent

possible.

A

complete odor-control system

mitigates

the

problem except

for

electrical equip-

ment

within

wet

wells

or

under tank covers.

•

Isolate control equipment

in a

separate room

and

design

the

ventilation

to

exclude

sulfides

or, if

that

is

impossible, treat

the air

supply

for the

entire

room.

If

electronic equipment

is

installed, pressur-

ize the

room slightly

and

eliminate chlorine

and

other corrosive gases

as

well.

•

Hermetically seal relays

and

switches

in

plastic

boxes attached

to

plastic conduit, seal

the

boxes

as

completely

as

possible,

and put

potassium perman-

ganate pellets

in the

bottom

of the

boxes.

•

Pressurize

the

boxes

or

panels

to

maintain

25

kPa

(0.1

in. WC)

with either uncontaminated

air or

with

air filtered

through potassium permanganate

and

activated carbon. Such pressurizing

in

accordance

with

NEC 500

also reduces enclosure explosion

hazard

ratings

from

NEC

Class

1,

Division

1 to

Class

1,

Division

2.

Adequate

differential

pressure

can

be

monitored

by an

attached pressure switch

and

alarm.

To

guard against excessive compressed

air

consumption

if the

bolts

are

loose,

the

gaskets

are

old,

or the

panel

is

warped,

a

limited amount

of

air can be

supplied

by a

small, inexpensive, dia-

phragm-type

fish

tank

air

pump.

23-7. Sequence

of

Design Steps

Pumping station design

is a

cooperative

effort

among

the

various design disciplines that must

be

coordi-

nated

as the

design progresses. Each design step

affecting

other disciplines should have

confirmed

agreement before proceeding

to

avoid later redesign.

Step

1

:

Write

and

distribute

a

design memorandum

defining

the

applicable codes,

the

indoor

and

out-

door design temperatures,

the

building materials

and

orientation,

the

sources

and

amounts

of

heat

gain

(especially

from

motors

and

engines),

and the

design

air

change rate

for

each

space.

State

the

energy source

to be

used

and

describe

the

proposed

ventilating,

heating,

and

cooling systems.

Step

2:

Calculate

the

heating

and

cooling loads

for

each space. Measure

and

record wall, glass,

and

roof

areas

from

the

architectural

plans.

List

expo-

sure

directions

and

calculate space volumes. Obtain

or

calculate heat transmission

coefficients

(U

val-

ues)

for

wall, roof,

and

window construction.

Use

the

methods

in the

ASHRAE

Handbook

of

Funda-

mentals

[12]

to

calculate

the

heat loss

and the

heat

gain

for

each space.

Sum the

heating

or

cooling

loads

in

watts (British thermal units

per

hour)

for

each

air

handling

unit

and

obtain

the

grand totals.

Step

3:

From

the

appropriate designer, obtain

the

locations

of the

main pumps

and

piping, cranes

or

hoists, columns, stairs, beams,

and

other obstacles

to

air

distribution. Also obtain

the

heat emissions

from

simultaneously operating equipment. Calcu-

late

the

initial ventilation rates starting with

12 air

changes

per

hour

for wet

wells (more

in

some states

or

for

severe conditions)

and 6 (or

more)

air

changes

per

hour

for dry

wells. Make preliminary

equipment selections. Estimate

the

outside

air

infil-

tration

to

each space

by

either

the

crack

or the air

change method,

and

calculate

the

expected outdoor

air

heating

and

cooling loads.

See the

ASHRAE

Handbook

of

Fundamentals [12]

for

details

of the

crack

and air

change methods.

Step

4:

From space volumes, heat gains,

and

design

air

change rates, calculate

the

normal

and

required

purge

air

supply quantities

for

each space.

If the

summer

airflow

for

heat removal exceeds

30 air

changes

per

hour, consider evaporative

or

refriger-

ated cooling. Choose supply

and

return/exhaust reg-

ister locations, assign

the

proper

air

quantity

to

each, and,

from

catalog data, select register sizes

to

give

desired

air

throw (for supply)

or

velocity

(return/exhaust).

Note

the

corresponding pressure

drops.

Step

5:

Make

an

initial estimate

of

HVAC

motor num-

ber and

sizes

for the

electrical engineer.

Fan

motor

horsepower

can be

approximated

in SI

units

by

W=^

(23-2a)

where

W is

brake watts,

Q is the

airflow

in

cubic

meters

per

second,

R is the

static resistance

in

pas-

cals,

and E is the fan

static

efficiency

as a

decimal.

In

U.S.

customary units,

bhp

=

6?5lE

(23

-

2b)

where

bhp is

brake horsepower,

Q is the

airflow

in

cubic

feet

per

minute,

R is the

static resistance

in

inches

of

water column,

and E is the fan

static

effi-

ciency

as a

decimal.

Lacking more precise information,

use the fan

resistance

and

efficiency

assumptions given

in

Table 23-5

for the

initial

motor size calculations.

Then,

for

safety,

select

the

next larger standard

motor size. Size

all

fans

for at

least

60 Pa

(0.25

in.

WC)

static resistance.

For

propeller

fans

more than

0.6 m (2 ft) in

diameter,

use a

motorized shutter

to

reduce resistance

and to

give more positive closure.

Step

6:

Calculate

the final

heating

and

cooling loads

and

modify

the

initial

air

distribution rates

as

required

to

meet

the

allowable

difference

between

the

temperatures

of the

supply

air and the

room.

Maximum

temperature

differences

of

6O

0

F

for

heat-

ing

and

25

0

F

for

cooling

are

recommended.

Step

7:

Determine feasible types

and

locations

for

air-handling equipment

and lay out

tentative duct

runs.

Avoid

obstructions

and

consider

how the

ducts

should

be

supported.

Step

8:

Beginning with

the

outlets

farthest

from

the

fan,

assign cumulative

air

quantities

to

each duct

section. Using

air

friction

charts

or

calculators, size

the

ducts

(1) to fit the

available space,

(2) to

give

a

streamlined

airflow,

(3) to

look presentable,

and (4)

to

simplify

duct fabrication.

Step

9:

Obtain gross

face

areas using

velocities

of

(1)

1.3 m/s

(250

ft/min)

for

wall intake

and

relief

louvers,

(2) 2.5 m/s

(500

ft/min)

for

wall exhaust

louvers,

and (3) 5 m/s

(1000

ft/min)

for

openings

to

roof exhaust

fans.

Confirm

opening

sizes

and

loca-

tions with

the

architect

and

structural engineer.

Table

23-5.

Fan

Static

Pressures

and

Efficiency

Static

Resistance

3

Without

heat

recovery

With

heat

recovery

Pa in. WC Pa in. WC

Directed

supply

250 1.0 620 2.5

Directed

exhaust

125

0.5 500 2.0

a

Unless

more

accurate

information

is

available,

use in

Equation

23-

2(a)

or (b)

with

an

assumed

fan

static

efficiency

of 65% as

initial

assumptions

for

estimating

the

size

of

motor

required.

Step

10:

Draw

the

duct systems

on the

plan. From

the

resistances

for

registers, straight lengths

of

ducts,

and

fittings,

cumulatively

sum the

actual friction drop

of

the

longest duct

run

from

the

farthest

point back

to

the

fan. Duct

friction

can be

calculated

by the

Darcy-

Weisbach

equation, which

for air

takes

the

form

p

=

f

D

(cfxL/D)p

v

(23-3)

where

p is the

total pressure

friction

loss

in

pascals

(inches

of

water column),

/

D

is the

friction

factor

(dimensionless),

c/is

the

conversion

factor,

L is the

duct

length

in

meters (inches),

D is the

duct diame-

ter

in

meters (inches),

and

p

v

is the

velocity pres-

sure

in

pascals (inches

of

water column).

The

friction

factor

is a

function

of

Reynolds number,

however,

and can be

calculated only

by

iteration.

Air

resistance

for

straight ducts

is

more easily cal-

culated

with

a

slide rule made

especially

for

this

purpose,

with

a

programmable calculator,

or

with

a

computer. Duct

friction

tables

and

charts

and a

list-

ing of fitting

friction

losses

are

available

in the

ASHRAE

Handbook

of

Fundamentals

[12].

Step

11

:

Make

final

best selections

of the

heat trans-

fer

and

air-handling equipment based

on the

design

flowrate

and

the

calculated total

air

resistance,

including

the

friction

drops

for

louvers, dampers,

filters,

coils,

ducts,

and

registers.

The

friction drop

for

clean, pleated glass

fiber filters is

about

25 Pa

(0.1

in.

WC),

but the fan

motor must

be

sized

to

supply

the

required

air at a

dirty

filter

resistance

of

120

to 160 Pa

(0.5

to

0.65

in.

WC).

Filter

and

coil

resistances

are

usually considered

to be

part

of the

external

resistance

—

separate

from

the

internal

resistance

of the

air-handling unit given

by the

unit

manufacturer.

Step

12:

Show

the

selected

air

supply

and

exhaust

equipment

on the

plans. Provide

(1)

raised house-

keeping pads

for floor-mounted

units,

(2)

vibration-

isolated supports

for

suspended

and

wall-mounted

units,

(3) flexible

connectors

for

duct connections,

and

(4)

ample access

for filter,

motor, belt,

and

bearing maintenance. Allow space

for

replacing

heating

and

cooling coils

and filters.

Provide direct access

to the

equipment. Floor-

mounted

units are, therefore, preferable

to

ceiling-

hung

equipment. Rooftop equipment

is

more likely

to be

maintained

if it is

accessible

by

stairs.

Step

13:

Select

the

additional air-handling equip-

ment

for

purging

and

summer heat removal.

The

equipment

required

may be

only strategically

placed high wall

or

roof exhaust

fans

and low

dam-

pered weatherproof wall intakes. Nonducted intakes

and

exhaust

fans

should

be

selected

to

result

in no

more than

60 Pa

(0.25

in. WC)

negative pressure

within

the

space.

A

force

of

about

120 N (27

Ib)

is

required

to

open

a

standard outwardly swinging

door against such

a

vacuum.

Step

14:

Recheck

the

actual motor kilowatts

(or

horsepower) needed

for the

designed systems

and

correct

the

motor list

as

required. Make sure motors

are

selected

for the

electrical

characteristics avail-

able. Coordinate motor, starter,

and

control loca-

tions with

the

electrical engineer

so

that

all of the

necessary wiring

is

provided. Motors

and

controls

should

be

listed

by

Underwriters Laboratories

(UL).

High-efficiency

motors

are

cost

effective

for

continuously operating

fans

and

pumps.

A

specifi-

cation requirement that air-handling equipment

shall

be

rated

in

accordance with

the

AMCA

No. 99

Standards gives

increased

assurance that

it

will

meet

the

rated capacity.

Step

15:

Specify

each component

of the

system,

stating each important criterion

of

construction

and

performance

necessary

to

determine

the

acceptabil-

ity

of

items subsequently submitted

by the

success-

ful

bidder. Equipment schedules

are a

simple

way

of

summarizing required equipment designations,

capacities,

and

special

features. Write

a

perfor-

mance description

of (1) the

intended sequence

of

operation

and (2) the

operating temperatures.

Require

a

submittal

of

control components

and

schematic

and

wiring diagrams. Require

accep-

tance tests

to

ensure that

the

installed system

func-

tions

as

intended

and

meets specification

requirements.

23-8.

Ventilating

System

Design

Pump rooms located where

the

outdoor temperature

range

is

between

freezing

and

35

0

C

(95

0

F)

and in

areas where

noise

need

not be

suppressed

can

often

be

ventilated

without

ductwork. Those rooms located

below grade, however, require ductwork.

For

pump

rooms

at or

above grade, louvered

air

intakes (with

dampers

and filters) and

roof exhaust

fans

(with grav-

ity

shutters

and

sequencing thermostats)

are

sufficient.

For

lower outdoor temperatures, heating

is

usually

required.

For

higher temperatures, some

form

of

cool-

ing

is

normally necessary.

Heating

and

cooling calculations

are

always neces-

sary

to

determine

the

required extent

of

heat removal

or

heat addition.

The air

quantities

to be

handled,

as

derived

from

the

heat load calculations, must

be

com-

pared

to

those determined

by

required

air

change rates

or

other

criteria,

and the

larger quantity must

be

used

as

the

design basis.

Heating

Load

Calculations

Structure

heat load

is

composed

of (1)

transmission

loss through walls, windows,

and

roof

and (2)

infiltra-

tion

of

cold outside air, which must

be

heated

to the

design space temperature.

The

transmission loss through

any

portion

of the

structure

envelope

is

H

= U x A x

A7

(23-4)

where,

in SI

units,

H is the

heat loss

in

watts

per

hour,

U

is the

overall heat transmission coefficient

in W/

(m

2

• h •

0

C),

A is the

surface area

in

square meters,

and

AT

is the

inside-to-outside temperature difference

in

degrees Celsius.

In

U.S. customary units,

H is in

Brit-

ish

thermal units

per

hour,

U is in

Btu/(ft

2

• h •

0

F),

A is

in

square

feet,

and

A7

is in

degrees Fahrenheit. Trans-

mission

coefficients

for

typical walls, windows,

and

roofs

as

well

as

methods

of

calculating

the

coefficient

for

the

combinations

of

materials

are

given

in

Chapter

25 of the

ASHRAE Handbook

of

Fundamentals

[12].

Cooling

Calculations

Heat gain throughout

the

structure, which makes

up the

external cooling load,

is

more

difficult

to

calculate than

heat

loss because

of the

effects

of

radiated heat

from

the

sun.

This additional heat source must

be

considered

in

addition

to the

heat gain

by

conduction

due to the

differ-

ence between outdoor

and

indoor temperatures. Sunlight

does

not

heat space

air

directly,

but it

raises

the

tempera-

ture

of

sunlit

surfaces

as a

result

of

absorbed radiation.

Many

factors

influence

the

amount

and

timing

of

solar heat reaching

the

space.

The

intensity

of

incident

solar radiation depends

on

latitude, time

of

year, time

of

day, cloud cover,

and

atmospheric pollution.

The

time

lag

between solar heat input

and

interior

air

tem-

perature rise

may

vary

from

a few

minutes

to

several

hours.

The lag is

affected

by the

surface color

of the

exterior,

the

heat storage capacity

and the

insulating

value

of the

construction,

and the

daily outdoor tem-

perature

range, among other

factors.

The

combined heat

from

both

the

higher outdoor

temperature

and

solar radiation moves progressively

through

walls

and

roof,

finally

raising

the

interior sur-

face

temperatures. Loss

from

those surfaces occurs

by

radiation

to

cooler

surfaces

and by

convection

to the

adjacent

air.

Air

warmed

by

convection expands,

becomes lighter, rises,

and is

replaced

by

cooler

air,

which

continues

the

convective process.

Except

for

minor reflected losses, sunlight passes

directly

through clear glass

and is

absorbed

by the

interior

surfaces

it

strikes. Their increased tempera-

ture

transfers heat

to

other

cooler

surfaces

by

reradia-

tion

and to the air by

convection.

Simply stated, solar heat gain

is

taken into account

by

using

a

higher outdoor temperature than actually would

exist

at the

time considered

for

each heat gain calcula-

tion.

The

ASHRAE

Handbook

of

Fundamentals

[12]

contains tables

of

cooling load temperature

differential

(CLTD)

data (for

use in

calculating conduction heat gain

through

sunlit walls

and

roofs)

and

cooling load

factors

(CLF) (for calculating

the

solar radiation through glass).

Both

sets

of

data include

the

effect

of

time delay

due to

thermal storage.

The

total resistance

to

heat

transfer

for

each type

of

construction

is

found

by

adding

the

resis-

tances

of its

components,

as

illustrated

in

Example

23-1.

Peak

Heat

Gain

The

peak heat gain

to a

space

is the

largest

sum of

external

and

internal

heat gains that occur simulta-

neously.

Heat removal capacity,

in the

form

of

ventila-

tion

or

cooling, equal

to the

peak heat gain must

be

available

to

maintain

the

design space temperature.

External heat gain results

from

an

outdoor tempera-

ture that

is

higher than

the

indoor temperature

as

well

as

from

solar radiation. Internal heat gain comes pri-

marily

from

operating motors

and

engines.

Air

Intake

and

Exhaust

Openings

Intakes

for

ventilation

air

should

be

through

screened

louvers that exclude rain, snow, birds,

and

insects.

Ducted intakes that

are

connected

to

air-handling units

with

filters

that exclude insects need only bird screens.

Nonducted intakes should include insect screens

or

filters.

All

screens

and filters

must

be

readily

accessi-

ble for

cleaning

or

replacement. Bird screens

do not

need mesh openings smaller than

25 mm

(

1

I

2

in.).

The

air

resistance

of filters

used

at

unducted wall intakes

should

not

exceed

24 Pa

(0.1

in. WC)

when clean

to

limit negative pressure

in the

space

(created

by

exhaust

fans)

to 63 Pa

(0.25

in.

WC);

low

negative

pressure makes doors easier

to

open

and

close.

Both

the

louver

and

screen should

be

constructed

of a

cor-

rosion-resistant material

or

should have

a

corrosion-

protective coating. Provide

an

intake damper

at

each

louver

in

cold climates, preferably with positive clo-

sure

on

failure

of its

pneumatic

or

electric actuator.

Hurricane-prone locations

may

warrant

an

additional

manually

operated damper

(on the

outdoor

air

intake)

that

can be

locked closed when

a

storm approaches.

Air

exhaust openings should

be

protected

from

the

weather

by a

wall louver, hood, penthouse,

or

weather

cap.

The

discharge openings should

be

located

so as to

minimize

unwanted recirculation into

air

intakes.

A

high-velocity vertical discharge

of

untreated

wet

well

exhaust

air

—

about

20

m/s

(4000

ft/min),

which

requires

250 Pa (1 in. WC)

static

pressure

—

may

reduce odor near

the

station (but with attendant

noise).

A

thorough model study

of the

likely results (which

are not

readily predictable) should

precede

any use of

this approach

for

odor control.

Ducted

Ventilation

A

ducted

air

supply

is

usually necessary

to

achieve

the

desired

air

distribution pattern. Once delivered

to the

right

location,

the air can be

directed upward

or

down-

ward

by

registers (grilles with directional vanes

and

attached balancing dampers),

or it can be

more thor-

oughly

mixed with room

air by

diffusers

that give high

entrainment

of

room air.

The

distance

an

airstream

travels

after

leaving

the

outlet (the

"throw")

depends

on

its

velocity, temperature,

and any

obstructions

to

straight-line

flow.

Catalog data

on

grilles, registers,

and

diffusers

give

the

throw

and

pressure drop

for a

given

flowrate

of

standard air.

The

throw

is

longer

for an

out-

let

discharging horizontally within about

0.6 m (2 ft) of

a

smooth ceiling

due to the low

induction

of

room air.

Install

a

manual volume damper

in

each main duct

branch

and at

each inlet

or

outlet grille

to

allocate

or

"balance"

the

total

air

quantity among

the

various

points

of

supply

and

exhaust. Such dampers

are

fur-

nished

as a

standard part

of

registers,

but

must

be

sep-

arately

specified

for

grilles

and

diffusers.

Complete

testing

and

balancing

of

ducted systems should

be

specified

to

deliver

the

design

air

quantity within

10%

at

each register

and

diffuser.

Wherever

an air

duct penetrates

a

rated

fire

separa-

tion

required

by

code, install

an

accessible

fire

damper

of

equal

fire

rating. Codes generally require rated

fire

separations between

floors,

around vertical

shafts

and

stairs,

at the

envelope enclosing

life

safety

exitways,

and,

in

some instances, around mechanical equipment

rooms

and

storage rooms.

The fire

damper

is

held

open

by a

fusible

link that melts

and

allows

the fire

damper

to

close automatically

if the air

temperature

rises

above

its

fusion

point

52 or

74

0

C

(125

or

165

0

F).

The

fusible

link

is

selected

on the

basis

of

location

and

normal

air

temperature. Such dampers

are

required

by

NFPA Standard

9OA

(which

is

included

by

reference

in

many other codes)

to

prevent

fire and

smoke

from

spreading through

the air

ducts. High-

temperature thermostats

or

smoke detectors

(as

required

by

applicable codes) should

be

provided

and

interlocked

to

stop both

the

supply

and the

exhaust

fans

if fire is

detected. Some codes require tight-clos-

ing, remotely operable smoke dampers (which

are

actuated

by

smoke detectors)

at

certain locations.

Detailed duct design methods

are

given

in

Chapter

33 of the

ASHRAE

Handbook

of

Fundamentals

[12].

Ventilation

Controls

Excessive

and

intricate controls

for

ventilation sys-

tems

(or any

HVAC

system)

are

likely

to

result

in

more operational problems than they were intended

to

overcome.

The

rule

is to

keep controls

as

simple

as

feasible

while

still

maintaining

the

temperature within

an

appropriate range.

A

temperature variation

from

above

freezing

to

43

0

C

(UO

0

F)

would

not be

serious

in

an

unmanned pumping station,

but in an

occupied

office

or

control room

the

temperature variation

should

be

held

to a

range

of

6

0

C

(1O

0

F)

or

less.

The

following

are

suggested ventilation control

functions:

•

Energize controls when

the

supply

fan

starts.

•

Control

outside-return

air

mixing dampers

to use at

least

10%

outside

air

initially

and

modulate them

to

increase

the

proportion

of

outdoor

air as the

space

temperature rises, subject

to a

discharge low-limit

control

set at

7

0

C

(45

0

F)

minimum.

•

Interlock

the

exhaust

fan to run

with

the

supply fan,

and

the

exhaust damper

to

operate

in

conjunction

with

the

outside

air

damper.

•

Interlock

the

supply

fan to

stop

if the

supply

air

high-limit

thermostat

or

smoke detector functions,

or if the

supply

air

low-limit thermostat senses

a

temperature below

3

0

C

(37

0

F).

• For a wet

well, start

a

supplementary exhaust

fan and

open

an air

intake damper

to

purge

the

space

if a gas

detector senses

20% of the LEL of a

combustible

gas.

• For a dry

well, start

an

additional exhaust

fan and

open

an air

intake damper when

a

space high-tem-

perature thermostat senses

a

temperature above

3O

0

C

(85

0

F).

Example

23-1

Design

of a

Ventilating

System

Problem:

Plans

for a

water pumping station

are

shown

in

Figures

23-1

through 23-4. There

are

currently

four

200-hp pumps (three

duty,

one

standby),

but six

350-hp pumps

(five

duty,

one

standby)

will

be

used

in the

future.

Although

the

office

is

currently occupied,

the

station will

Figure

23-1.

Basement

plan.

Figure

23-2. Ground-level floorplan.

Figure

23-3. Boiler

room

plan.

Figure

23-4. Section

A-A

from

Figure 23-2.

also

be

unattended

in the

future.

The

latitude

is

41.5° north,

the

elevation

is 660 ft, and the

daily

temperature

range

is

2O

0

F.

Other temperature data

are as

follows:

Design

conditions Winter

Summer

Time

equaled

or

exceeded

99%

1%

Outside temperature,

dry

bulb

-2O

0

F

94

0

F

Outside temperature,

wet

bulb

—

74

0

F

Inside temperature (occupied)

7O

0

F

78

0

F

Inside temperature (unoccupied)

6O

0

F

105

0

F

Calculate

(1) the

heat transmission

coefficients

for

superstructure walls, roof,

and

glass;

(2)

the

heat loss

for the

pump room;

(3) the

present heat gain

for the

pump room;

(4) the

future

heat

gain

for the

pump room;

and (5) the

requirements

for

heat

and

ventilation.

Solution:

(1)

Heat transmission

coefficients.

Refer

to the

ASHRAE

Handbook

of

Fundamentals

[12],

pp.

23.12-23.20

and

Tables

1,

3A,

and 3B. Sum the

component thermal resistances

(R) for

each

surface

of the

enclosure

and

invert

(1

/R)

to

obtain

the

heat transmission

coefficients

(LO

values

(i.e.,

U -

l/R).

An

example follows.

Lower

walls

Roof Glass

Double

solar

bronze,

Outside

surface:

0.17

Outside

surface:

0.17

nonreflective

4-in.

face

brick:

0.44

Built-up

roof:

0.33

1-in.

styrofoam:

4.0

3-in.

lightweight

fill

4.69

1-in.

air

space:

0.94 1-in.

polyurethane

6.25

6-in.

concrete

block:

0.91 6-in.

flexicore

0.66

2-in.

glazed

tile:

0.60

Inside

surface:

1.35

Inside

surface:

0.61

Winter

R =

2.05

fl

=

8.41

/?=

12.71

Summer/?

=1.79

U

=

1/8.41

=0.12

U =

1/12.71

=

0.079

U =

0.49

a

U

=

0.56

a

Factors

for

heat

transfer

through glass

differ

for

winter

and

summer conditions (see ASHRAE [12],

p.

27.10, Table

13).

(2)

Heat loss

for

pump room. Calculate

the

heat losses using Equation 23-4

(Q=AxUx

AT).

(Also

see

the

ASHRAE

Han

dbook

of

Fundamentals

[12],

p. 25.

Iff.)

For

algebraic consis-

tency,

heat gains

are

shown

as

positive

and

heat losses

as

negative.

See

Figure 23-4

for

high

and

low

walls.

In

tabular format, heat losses

are as

follows:

Low

walls above grade:

(88 + 17 +

27)12.7

x

0.12(60

+ 20) =

-16,094

High

walls: 2(31.5

+ 88) x 3.3 x

0.28

x 80 =

-17,667

Glass: 2(85

+

30)3

x

0.49

x 80 =

-27,048

Doors:

[(12

x 10) + (3 x

7)](0.59

-

0.28)

x 80 =

-3,497

Roof/ceiling:

30 x 88 x

0.079

x 80 =

-16,685

Infiltration

(based

on

0.75

AC/h

at

8O

0

F

AT):

(20.7

x 88 x 30 x

0.75

x

1.08

x

80)/60

=

-59,020

Winter

pump room heat loss

(Btu/h):

-140,011

(3)

Pump room present summer heat gain. Refer

to

ASHRAE [12],

pp.

26.8, 26.19, 26.29,

26.32,

and

Tables

5

A,

1IA,

and 24.

Note that heat gain

due to

motors

is

Heat gain

(Btu/h)

=

bph(^-^\2545

(23-5)

V^y

where

bhp is the

brake horsepower,

E is the

overall motor

efficiency

as a

decimal,

and

2545

is

the

British thermal units

per

hour

•

brake horsepower.

For the

cooling load temperature difference (CLTD) adjustment

for

roof

and

walls

in the

fol-

lowing

example,

see

ASHRAE [12],

pp.

26.8

and

26.10,

Tables

5 and 7,

Note

2. For

glass trans-

mission

factors,

see

Tables

10,

11,

and 13, pp.

26.14-26.17,

and

Table

29, p.

27.29 together with

their

accompanying explanations

and

example.

For our

example, maximum heat transfer occurs

at

4

P.M.

on

July

21.

The

pump room wall

is

Group

No. C.

Directions

are

identified

using capital

letters

(N =

north, etc.).

Structure

solar

and

transmission heat gain

The

heat gain

is

determined

as

follows:

Roof

with

a

suspended ceiling

(see

ASHRAE

[12],

p.

26.8,

Table

5,

Roof

12):

0.079

x 30 x

88[(30

+ 1) + (78 -

104)

+ (84 -

85)]

=

+834

N

wall:

0.12

x

15.3

x

88[(12

-

0)0.83

+

(78

-

104)

+

(84

-

85)]

=

-2,753

E

wall:

0.12

x

15.3

x

28[(29

-

0)0.83

+

(78

-

104)

+

(84

-

85)]

=

-151

W

wall:

0.12

x

15.3

x

18.6[(16

-

0)0.83

+ (78 -

104)

+

(84

-

85)]

=

-469

S

wall (upper):

0.12

x 85 x

1.5[(20

-

1)0.83

+

(78

-

104)

+

(84

-

85)]

=

-172

N

glass, solar:

3 x 86 x 38 x

0.70

x

0.55

=

+3,775

E

glass, solar:

3 x 30 x

216

x

0.24

x

0.55

=

+2,566

W

glass, solar:

3

x 30 x 216 x

0.49

x

0.55

=

+5,239

S

glass, solar:

3

x 86 x 109 x

0.43

x

0.55

=

+6,651

Glass conduction:

3(86

+

30)2

x

0.56(-12)

=

^,677

Total summer structure external heat gain

(Btu/h)

=

+10,843

Internal

heat gain

Pump motor heat gain

in

Btu/h

(from

Equation 23-5):

3

x

200[(1

-

0.9)/0.9]

2545

=

+169,667

Estimated auxiliary motor heat:

20[(1

-

0.67)/0.67]

2545

=

+25,070

Heat

from

lights:

30x88x2x3.413=

+18,021

Heat added

by

people:

2

x 350 =

+700

Total internal peak heat gain

in

Btu/h:

+213,458

Present

total internal heat gain: Adding

the

structure heat gain

to the

internal peak heat gain

yields

the

pump

room's

present maximum heat gain:

+213,458

+

10,843

=

+224,301

(4)

Pump room future internal heat gain.

Pump

motor heat gain:

5 x 350

(10/90) 2545

=

+494,861

Pumping Station Desing - Second Edition by Robert L. Sanks, George Tchobahoglous, Garr M. Jones

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