Tài liệu Handbook of Mechanical Engineering Calculations P16 doc

16.1
SECTION 16
HEATING, VENTILATING, AND
AIR CONDITIONING
ECONOMICS OF INTERIOR CLIMATE
CONTROL
16.2
Equations for Heating, Ventilation,
and Air-Conditioning Calculations
16.2
Determining Cooling-Tower Fan
Horsepower Requirements
16.12
Choosing an Ice Storage System for
Facility Cooling
16.13
Annual Heating and Cooling Energy
Loads and Costs
16.22
Heat Recovery Using a Run-Around
System of Energy Transfer
16.24
Rotary Heat Exchanger Energy
Savings
16.26
Savings from ‘‘Hot-Deck’’
Temperature Reset
16.28
Air-to-Air Heat Exchanger
Performance
16.29

Steam and Hot-Water Heating
Capacity Requirements for Buildings
16.32
Heating Steam Required for
Specialized Rooms
16.32
Determining Carbon Dioxide Buildup
in Occupied Spaces
16.33
Computing Bypass-Air Quantity and
Dehumidifier Exit Conditions
16.34
Determination of Excessive Vibration
Potential in Motor-Driven Fan
16.36
Power Input Required by Centrifugal
Compressor
16.37
Evaporation of Moisture from Open
Tanks
16.38
Checking Fan and Pump Performance
from Motor Data
16.40
Choice of Air-Bubble Enclosure for
Known Usage
16.41
Sizing Hydronic-System Expansion
Tanks
16.45

SYSTEM ANALYSIS AND EQUIPMENT
SELECTION
16.53
Building or Structure Heat-Loss
Determination
16.53
Heating-System Selection and
Analysis
16.55
Required Capacity of a Unit Heater
16.58
Steam Consumption of Heating
Apparatus
16.65
Selection of Air Heating Coils
16.67
Radiant-Heating-Panel Choice and
Sizing
16.72
Snow-Melting Heating-Panel Choice
and Sizing
16.75
Heat Recovery from Lighting Systems
for Space Heating
16.77
Air-Conditioning-System Heat-Load
Determination—General Method
16.78
Air-Conditioning-System Heat-Load
Determination—Numerical

Computation
16.85
Air-Conditioning System Cooling-Coil
Selection
16.90
Mixing of Two Airstreams
16.97
Selection of an Air-Conditioning
System for a Known Load
16.99
Sizing Low-Velocity Air-Conditioning-
Systems Ducts—Equal Friction
Method
16.102
Sizing Low-Velocity Air-Conditioning
Ducts—Static-Regain Method
16.111
Humidifier Selection for Desire
Atmospheric Conditions
16.114
Use of the Psychrometric Chart in Air-
Conditioning Calculations
16.119
Designing High-Velocity Air-
Conditioning Ducts
16.122
Air-Conditioning-System Outlet- and
Return-Grille Selection
16.125
Selecting Roof Ventilators for

Buildings
16.130
Vibration-Isolator Selection for an Air
Conditioner
16.134
Selection of Noise-Reduction
Materials
16.136
Choosing Door and Window Air
Curtains for Various Applications
16.139
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Source: HANDBOOK OF MECHANICAL ENGINEERING CALCULATIONS
16.2
ENVIRONMENTAL CONTROL
Economics of Interior Climate Control
EQUATIONS FOR HEATING, VENTILATION, AND
AIR-CONDITIONING CALCULATIONS
A variety of calculation procedures are used in designing heating, ventilating, and
air-conditioning systems. To help save time for design and application engineers,
technicians, and consulting engineers, some 75 design equations are presented at
the start of this section of the handbook. These equations are used in both manual
and computer-aided design (CAD) applications. And since this handbook is de-
signed for worldwide use, the first group of equations presents USCS and SI ver-
sions to allow easy comparisons of the results. Abbreviations used in the equations
follow this presentation.
Btu
⅐

min
H
ϭ
1.08
ϫ
CFM
ϫ ⌬
T (1)
S
3
h
⅐
ft
⅐ Њ
F
kJ
⅐
min
H
ϭ
72.42
ϫ
CMM
ϫ ⌬
T (2)
SMM
3
h
⅐
m

⅐ Њ
C
Btu
⅐
min
⅐
lb DA
H
ϭ
0.68
ϫ
CFM
ϫ ⌬
W (3)
L
3
h
⅐
ft
⅐
grHO
2
kJ
⅐
min
⅐
kg DA
H
ϭ
177,734.8

ϫ
CMM
ϫ ⌬
W (4)
LMM
3
h
⅐
m
⅐
kgHO
2
lb
⅐
min
H
ϭ
4.5
ϫ
CFM
ϫ ⌬
h (5)
T
3
h
⅐
ft
kg
⅐
min

H
ϭ
72.09
ϫ
CMM
ϫ ⌬
h (6)
TMM
3
h
⅐
m
H
ϭ
H
ϩ
H (7)
TSL
H
ϭ
H
ϩ
H (8)
TM SM LM
Btu
⅐
min
H
ϭ
500

ϫ
GPM
ϫ ⌬
T (9)
h
⅐
gal
⅐ Њ
F
kJ
⅐
min
H
ϭ
250.8
ϫ
LPM
ϫ ⌬
T (10)
MM
h
⅐
L
⅐ Њ
C
AC CFM
ϫ
60 min /h
ϭ
(11)

HR VOLUME
AC CMM
ϫ
60 min /h
ϭ
(12)
HR VOLUME
MM
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HEATING, VENTILATING, AND AIR CONDITIONING
HEATING, VENTILATING, AND AIR CONDITIONING
16.3
Њ
F
Ϫ
32
Њ
C
ϭ
(13)
1.8
Њ
F
ϭ
1.8
Њ
C
ϩ

32 (14)
where H
S
ϭ
sensible heat, Btu /h
H
SM
ϭ
sensible heat, kJ /h
H
L
ϭ
latent heat, Btu/h
H
LM
ϭ
latent heat, kJ/h
H
T
ϭ
total heat, Btu/h
H
TM
ϭ
total heat, kJ/h
H
ϭ
total heat, Btu/h
H
M

ϭ
total heat, kJ/h
⌬
T
ϭ
temperature difference,
Њ
F
⌬
T
M
ϭ
temperature difference,
Њ
C
⌬
W
ϭ
humidity ratio difference, gr H
2
O/lb DA
⌬
W
M
ϭ
humidity ratio difference, kg H
2
O/kg DA
⌬
h

ϭ
enthalpy difference, Btu /lb DA
⌬
h
ϭ
enthalpy difference, kJ /lb DA
CFM
ϭ
airflow rate, ft
3
/min
CMM
ϭ
airflow rate, m
3
/min
GPM
ϭ
water flow rate, gal/min
LPM
ϭ
water flow rate, L/min
AC/HR
ϭ
air change rate per hour, English
AC/HR
M
ϭ
air change rate per hour, SI
AC/HR

ϭ
AC/HR
M
VOLUME
ϭ
space volume, ft
3
VOLUME
M
ϭ
space volume, m
3
kJ/h
ϭ
Btu/h
ϫ
1.055
CMM
ϭ
CFM
ϫ
0.02832
LPM
ϭ
GPM
ϫ
3.785
kJ/lb
ϭ
Btu/lb

ϫ
2.326
m
ϭ
ft
ϫ
0.3048
m
2
ϭ
ft
2
ϫ
0.0929
m
3
ϭ
ft
3
ϫ
0.02832
kg
ϭ
lb
ϫ
0.4536
1.0 GPM
ϭ
500 lb steam/h
1.0 lb steam/h

ϭ
0.002 GPM
1.0 lb H
2
/h
ϭ
1.0 lb steam/h
kg/m
3
ϭ
lb/ft
3
ϫ
16.017 (density)
m
3
/kg
ϭ
ft
3
/lb
ϫ
0.0624 specific volume
kg H
2
O/kg DA
ϭ
gr H
2
O/ lb DA/ 7000

ϭ
lb H
2
O/lb DA
Steam Pipe Pressure Drop and Flow Rate Equations
2
0.01306W (1
ϩ
3.6/ ID)
⌬
P
ϭ
(15)
5
3600
ϫ
D
ϫ
ID
5
⌬
P
ϫ
D
ϫ
ID
W
ϭ
60 (16)
Ί

0.01306
ϫ
(1
ϩ
3.6/ ID)
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HEATING, VENTILATING, AND AIR CONDITIONING
16.4
ENVIRONMENTAL CONTROL
W
ϭ
0.41667VA D
ϭ
60VA D (17)
INCHES FEET
2.4WW
V
ϭϭ
(18)
AD60AD
INCHES FEET
where
⌬
P
ϭ
pressure drop per 100 ft of pipe (psig/100 ft)
W
ϭ

steam flow rate, lb/h
ID
ϭ
actual inside diameter of pipe, in
D
ϭ
average density of steam at system pressure, lb/ft
3
V
ϭ
velocity of steam in pipe, ft/min
A
INCHES
ϭ
actual cross-sectional area of pipe, in
2
A
FEET
ϭ
actual cross-sectional area of pipe, ft
2
Condensate Piping Equations
H
Ϫ
H
SSS SCR
FS
ϭϫ
100 (19)
H

LCR
FS
W
ϭϫ
W (20)
CR
100
where FS
ϭ
flash steam, %
H
SSS
ϭ
sensible heat at steam supply pressure, Btu/lb
H
SCR
ϭ
sensible heat at condensate return pressure, Btu/ lb
H
LCR
ϭ
latent heat at condensate return pressure, Btu/lb
W
ϭ
steam flow rate, lb/h
W
CR
ϭ
condensate flow based on percentage of flash steam created during
condensing process, lb/h. Use this flow rate in steam equations above

to determine condensate return pipe size.
HVAC Efficiency Equations
BTU OUTPUT EER
COP
ϭϭ
(21)
BTU INPUT 3.413
BTU OUTPUT
EER
ϭ
(22)
WATTS INPUT
Turndown ratio
ϭ
maximum firing rate
Ϻ
minimum firing rate (that is 5
Ϻ
1, 10
Ϻ
1,
25
Ϻ
1)
GROSS BTU OUTPUT
OVERALL THERMAL EFF
ϭ
GROSS BTU INPUT
ϫ
100% (23)

BTU INPUT
Ϫ
BTU STACK LOSS
COMBUSTION EFF
ϭ
BTU INPUT
ϫ
100% (24)
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HEATING, VENTILATING, AND AIR CONDITIONING
HEATING, VENTILATING, AND AIR CONDITIONING
16.5
Overall thermal efficiency range 75%–90%
Combustion efficiency range 85%–95%
Equations for HVAC Equipment Room Ventilation
For completely enclosed equipment rooms:
0.5
CFM
ϭ
100
ϫ
G (25)
where CFM
ϭ
exhaust airflow rate required, ft
3
/min
G

ϭ
mass of refrigerant of largest system, lb
For partially enclosed equipment rooms:
0.5
FA
ϭ
G (26)
where FA
ϭ
ventilation-free opening area, ft
2
G
ϭ
mass of refrigerant of largest system, lb
Psychrometric Equations
The following equations are from Carrier Corporation publications.* These equa-
tions cover air mixing, cooling loads, sensible heat factor, bypass factor, temperature
at the apparatus, supply air temperature, air quantity, and determination of air con-
stants. Abbreviations and symbols for the equations are given below.
Abbreviations
adp apparatus dew point
BF bypass factor
(BF) (OALH) bypassed outdoor air latent heat
(BF) (OASH) bypassed outdoor air sensible heat
(BF) (OATH) bypassed outdoor air total heat
Btu/ h British thermal units per hour
cfm, ft
3
/min cubic feet per minute
db dry-bulb

dp dew point
ERLH effective room latent heat
ERSH effective room sensible heat
ERTH effective room total heat
ESHF effective sensible heat factor
Њ
F degrees Fahrenheit
fpm, ft/min feet per minute
gpm, gal/ min gallons per minute
*Handbook of Air-Conditioning System Design, McGraw-Hill, New York, various dates.
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HEATING, VENTILATING, AND AIR CONDITIONING
16.6
ENVIRONMENTAL CONTROL
gr/ lb grains per pound
GSHF grand sensible heat factor
GTH grand total heat
GTHS grand total heat supplement
OALH outdoor air latent heat
OASH outdoor air sensible heat
OATH outdoor air total heat
rh relative humidity
RLH room latent heat
RLHS room latent heat supplement
RSH room sensible heat
RSHF room sensible heat factor
RSHS room sensible heat supplement
RTH room total heat

Sat Eff saturation efficiency of sprays
SHF sensible heat factor
TLH total latent heat
TSH total sensible heat
wb wet-bulb
Symbols
cfm
ba
bypassed air quantity around apparatus
cfm
da
dehumidified air quantity
cfm
oa
outdoor air quantity
cfm
ra
return air quantity
cfm
sa
supply air quantity
h specific enthalpy
h
adp
apparatus dew point enthalpy
h
cs
effective surface temperature enthalpy
h
ea

entering air enthalpy
h
la
leaving air enthalpy
h
m
mixture of outdoor and return air en-
thalpy
h
oa
outdoor air enthalpy
hr
m
room air enthalpy
h
sa
supply air enthalpy
t temperature
t
adp
apparatus dew point temperature
t
edb
entering dry-bulb temperature
t
es
effective surface temperature
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HEATING, VENTILATING, AND AIR CONDITIONING
HEATING, VENTILATING, AND AIR CONDITIONING
16.7
t
ew
entering water temperature
t
ewb
entering wet-bulb temperature
t
ldb
leaving dry-bulb temperature
t
lw
leaving water temperature
t
lwb
leaving wet-bulb temperature
t
m
mixture of outdoor and return air dry-
bulb temperature
t
oa
outdoor air dry-bulb temperature
t
rm
room dry-bulb temperature
t
sa

supply air dry-bulb temperature
W moisture content or specific humidity
W
adp
apparatus dew point moisture content
W
ea
entering air moisture content
W
es
effective surface temperature moisture
content
W
la
leaving air moisture content
W
m
mixture of outdoor and return air mois-
ture content
W
oa
outdoor air moisture content
W
rm
room moisture content
W
sa
supply air moisture content
Air Mixing Equations (Outdoor and Return Air)
cfm

ϫ
t
ϩ
cfm
ϫ
t
oa oa ra rm
t
ϭ
(27)
m
cfm
sa
(cfm
ϫ
h )
ϩ
(cfm
ϫ
h )
oa oa ra rm
h
ϭ
(28)
m
cfm
sa
(cfm
ϫ
W )

ϩ
(cfm
ϫ
W )
oa oa ra rm
W
ϭ
(29)
m
cfm
sa
Cooling Load Equations
ERSH
ϭ
RSH
ϩ
(BF)(OASH)
ϩ
RSHS* (30)
ERLH
ϭ
RLH
ϩ
(BF)(OALH)
ϩ
RLHS* (31)
*RSHS, RLHS, and GTHS are supplementary loads due to duct heat gain, duct leakage loss, fan and
pump horsepower gains, etc.
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HEATING, VENTILATING, AND AIR CONDITIONING
16.8
ENVIRONMENTAL CONTROL
ERTH
ϭ
ERLH
ϩ
ERSH (32)
TSH
ϭ
RSH
ϩ
OASH
ϩ
RSHS* (33)
TLH
ϭ
RLH
ϩ
OALH
ϩ
RLHS* (34)
GTH
ϭ
TSH
ϩ
TLH
ϩ
GTHS* (35)

RSH
ϭ
1.08†
ϫ
cfm
ϫ
(t
Ϫ
t ) (36)
sa rm sa
RLH
ϭ
0.68†
ϫ
cfm
ϫ
(W
Ϫ
W ) (37)
sa rm sa
RTH
ϭ
4.45†
ϫ
cfm
ϫ
(h
Ϫ
h ) (38)
sa rm sa

RTH
ϭ
RSH
ϩ
RLH (39)
OASH
ϭ
1.08
ϫ
cfm (t
Ϫ
t ) (40)
oa oa rm
OALH
ϭ
0.68
ϫ
cfm (W
Ϫ
W ) (41)
oa oa rm
OATH
ϭ
4.45
ϫ
cfm (h
Ϫ
h ) (42)
oa oa rm
OATH

ϭ
OASH
ϩ
OALH (43)
(BF)(OATH)
ϭ
(BF)(OASH)
ϩ
(BF)(OALH) (44)
ERSH
ϭ
1.08
ϫ
cfm ‡
ϫ
(t
Ϫ
t )(1
Ϫ
BF) (45)
da rm adp
ERLH
ϭ
0.68
ϫ
cfm ‡
ϫ
(W
Ϫ
W )(1

Ϫ
BF) (46)
da rm adp
ERTH
ϭ
4.45
ϫ
cfm ‡
ϫ
(h
Ϫ
h )(1
Ϫ
BF) (47)
da rm adp
TSH
ϭ
1.08
ϫ
cfm ‡
ϫ
(t
Ϫ
t )* (48)
da edb ldb
TLH
ϭ
0.68
ϫ
cfm ‡

ϫ
(W
Ϫ
W )* (49)
da ea la
GTH
ϭ
4.45
ϫ
cfm ‡
ϫ
(h
Ϫ
h )* (50)
da ea la
Sensible Heat Factor Equations
RSH RSH
RSHF
ϭϭ
(51)
RSH
ϩ
RLH RTH
ERSH ERSH
ESHF
ϭϭ
(52)
ERSH
ϩ
ERLH ERTH

TSH TSH
GSHF
ϭϭ
(53)
TSH
ϩ
TLH GTH
*RSHS, RLHS, and GTHS are supplementary loads due to duct heat gain, duct leakage loss, fan and
pump horsepower gains, etc.
†See below for the derivation of these air constants.
‡When no air is to be physically bypassed around the conditioning apparatus, cfm
da
ϭ
cfm
sa
.
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HEATING, VENTILATING, AND AIR CONDITIONING
HEATING, VENTILATING, AND AIR CONDITIONING
16.9
Bypass Factor Equations
t
Ϫ
tt
Ϫ
t
ldb adp edb ldb
BF

ϭ
1
Ϫ
BF
ϭ
(54)
t
Ϫ
tt
Ϫ
t
edb adp edb adp
W
Ϫ
WW
Ϫ
W
la adp ea la
BF
ϭ
1
Ϫ
BF
ϭ
(55)
W
Ϫ
WW
Ϫ
W

ea adp ea adp
h
Ϫ
hh
Ϫ
h
la adp ea la
BF
ϭ
1
Ϫ
BF
ϭ
(56)
h
Ϫ
hh
Ϫ
h
ea adp ea adp
Temperature Equations at the Apparatus
(cfm
ϫ
t )
ϩ
(cfm
ϫ
t )
oa oa ra rm
t *

ϭ
(57)
edb
cfm †
sa
t
ϭ
t
ϩ
BF(t
Ϫ
t ) (58)
ldbadp edbadp
Both t
ewb
and t
lwb
correspond to the calculated values of h
ea
and h
la
on the psychro-
metric chart.
(cfm
ϫ
h )
ϩ
(cfm
ϫ
h )

oa oa ra rm
h *
ϭ
(59)
ea
cfm †
sa
h
ϭ
h
ϩ
BF(h
Ϫ
h ) (60)
la adp ea adp
Temperature Equations for Supply Air
RSH
t
ϭ
t
Ϫ
(61)
sa rm
1.08cfm †
sa
Air Quantity Equations
ERSH
cfm
ϭ
(62)

da
1.08(1
Ϫ
BF)(t
Ϫ
t )
rm adp
ERLH
cfm
ϭ
(63)
da
0.68(1
Ϫ
BF)(W
Ϫ
W )
rm adp
ERTH
cfm
ϭ
(64)
da
4.45(1
Ϫ
BF)(h
Ϫ
h )
rm adp
*When t

m
, W
m
, and h
m
are equal to the entering conditions at the cooling apparatus, they may be
substituted for t
edb
, W
ea
, and h
ea
, respectively.
†See footnote on page 16.9.
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HEATING, VENTILATING, AND AIR CONDITIONING
16.10
ENVIRONMENTAL CONTROL
TSH
cfm *
ϭ
(65)
da
1.08(t
Ϫ
t )
edb ldb
TLH

cfm *
ϭ
(66)
da
0.68(W
Ϫ
W )
ea la
GTH
cfm *
ϭ
(67)
da
4.45(h
Ϫ
h )
ea la
RSH
cfm
ϭ
(68)
sa
1.08(t
Ϫ
t )
rm sa
RLH
cfm
ϭ
(69)

sa
0.68(W
Ϫ
W )
rm sa
RTH
cfm
ϭ
(70)
sa
4.45(h
Ϫ
h )
rm sa
cfm
ϭ
cfm
Ϫ
cfm (71)
ba sa da
Note: cfm
da
will be less than cfm
sa
only when air is physically bypassed around the
conditioning apparatus.
cfm
ϭ
cfm
ϩ

cfm (72)
sa oa ra
Derivation of Air Constants
60
1.08
ϭ
0.244
ϫ
(73)
13.5
where 0.244
ϭ
specific heat of moist air at 70
Њ
F db and 50% rh, Btu/ (
Њ
F
⅐
lb DA)
60
ϭ
min/h
13.5
ϭ
specific volume of moist air at 70
Њ
F db and 50% rh
60 1076
0.68
ϭϫ

13.5 7000
where 60
ϭ
min/h
13.5
ϭ
specific volume of moist air at 70
Њ
F db and 50% rh
1076
ϭ
average heat removal required to condense 1 db water vapor from the
room air
7000
ϭ
gr/lb
60
4.45
ϭ
13.5
*See footnote on page 16.9.
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HEATING, VENTILATING, AND AIR CONDITIONING
HEATING, VENTILATING, AND AIR CONDITIONING
16.11
where 60
ϭ
min/h

13.5
ϭ
specific volume of moist air at 70
Њ
F db and 50% rh
Equations for Steam Trap Selection
The selection of the trap for the steam mains or risers is dependent on the pipe
warm-up load and the radiation load from the pipe. Warm-up load is the condensate
which is formed by heating the pipe surface when the steam is first turned on. For
practical purposes, the final temperature of the pipe is the steam temperature. Warm-
up load is determined from
W(t
Ϫ
t )(0.114)
ƒi
C
ϭ
(74)
1
hT
l
where C
1
ϭ
warm-up condensate, lb/h
W
ϭ
total weight of pipe, lb (from tables in engineering handbooks)
t
ƒ

ϭ
final pipe temperature,
Њ
F (steam temp.)
t
i
ϭ
initial pipe temperature,
Њ
F (usually room temp.)
0.114
ϭ
specific heat constant for wrought iron or steel pipe (0.092 for copper
tubing)
h
l
ϭ
latent heat of steam, Btu/ lb (from steam tables)
T
ϭ
time for warm-up, h
The radiation load is the condensate formed by unavoidable radiation loss from
a bare pipe. This load is determined from the following equation and is based on
still air surrounding the steam main or riser:
LK(t
Ϫ
t )
ƒi
C
ϭ

(75)
2
h
l
where C
2
ϭ
radiation condensate, lb /h
L
ϭ
linear length of pipe, ft
K
ϭ
heat transmission coefficient, Btu/(h
⅐
lin ft
⅐ Њ
F)
The radiation load builds up as the warm-up load drops off under normal op-
erating conditions. The peak occurs at the midpoint of the warm-up cycle. There-
fore, one-half of the radiation load is added to the warm-up load to determine the
amount of condensate that the trap handles.
Safety Factor
Good design practice dictates the use of safety factors in steam trap selection. Safety
factors from 2 to 1 to as high as 8 to 1 may be required, and for the following
reasons:
1. The steam pressure at the trap inlet or the back-pressure at the trap discharge
may vary. This changes the steam trap capacity.
2. If the trap is sized for normal operating load, condensate may back up into the
steam lines or apparatus during start-up or warm-up operation.

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HEATING, VENTILATING, AND AIR CONDITIONING
16.12
ENVIRONMENTAL CONTROL
3. If the steam trap is selected to discharge a full and continuous stream of water,
the air could not be vented from the system.
The following guide is used to determine the safety factor:
Design Safety factor
Draining steam main 3 to 1
Draining steam riser 2 to 1
Between boiler and end of main 2 to 1
Before reducing valve 3 to 1
Before shutoff valve (closed part of time) 3 to 1
Draining coils 3 to 1
Draining apparatus 3 to 1
When the steam trap is to be used in a high-pressure system, determine whether
the system is to operate under low-pressure conditions at certain intervals such as
nighttime or weekends. If this condition is likely to occur, then an additional safety
factor should be considered to account for the lower pressure drop available during
nighttime operation.
DETERMINING COOLING-TOWER FAN
HORSEPOWER REQUIREMENTS
A cooling tower serving an air-conditioning installation is designed for these con-
ditions: Water flowrate, L
ϭ
75,000 gal /min (4733 L/s); inlet water temperature,
T
i

ϭ
110
Њ
F (43.3
Њ
C); outlet water temperature, T
o
ϭ
90
Њ
F (32.2
Њ
C); atmospheric
wet-bulb temperature, T
w
ϭ
82
Њ
F (27.8
Њ
C); total fan efficiency as given by tower
manufacturer, E
T
ϭ
75%; recirculation of air in tower, given by tower manufacturer,
R
c
ϭ
8.5%; total air pressure drop through the tower, as given by manufacturer,
⌬

P
ϭ
0.477 in (1.21 cm) H
2
O. What is the required fan horsepower input under these
conditions? If the weather changes and the air outlet temperature becomes 102
Њ
F
(38.9
Њ
C) with a wet-bulb temperature of 84
Њ
F (28.9
Њ
C)?
Calculation Procedure:
1. Determine the fan horsepower for the given atmospheric and flow
conditions
The fan brake horsepower input for a cooling tower is given by the relation: BHP
ϭ
L
ϫ ⌬
T
ϫ
V
sp
ϫ
R
c
ϫ ⌬

P/
⌬
H
ϫ
6356
ϫ
E
T
, where
⌬
T
ϭ
T
o
Ϫ
T
i
; V
sp
ϭ
specific volume of outlet air, ft
3
/lb (m
3
/kg);
⌬
H
ϭ
difference between enthalpy of
outlet air and inlet air, Btu/lb (kJ/kg); other symbols as given earlier. Determine

the enthalpy difference between the outlet air and inlet air by referring to a psy-
chrometric chart where you will find that the enthalpy of the outlet air for the first
case above, H
o
ϭ
72 Btu/ lb (167.5 kJ/kg); from the same source the enthalpy of
the inlet air, H
i
ϭ
46 Btu/ lb (107.0 kJ/kg); likewise, V
sp
ϭ
15.1 ft
3
/lb (93.7 m
3
/
kg) from the chart. Substituting in the equation above, BHP
ϭ
(75,000
ϫ
8.337
ϫ
20
ϫ
1.085
ϫ
15.1
ϫ
0.477)/ (72

Ϫ
24)
ϫ
6356
ϫ
0.75
ϭ
788.51; say 789 hp
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HEATING, VENTILATING, AND AIR CONDITIONING
HEATING, VENTILATING, AND AIR CONDITIONING
16.13
(588.3 kW). In the above equation the constants 8.337 and 1.085 are used to convert
gal/ min to lb/min and air flow to ft
3
/lb, respectively.
2. Determine the power input required for the second set of conditions
For the second set of conditions the air outlet temperature, T
o
ϭ
102
Њ
F (38.9
Њ
C)
and the wet-bulb temperature is 84
Њ
F (28.9

Њ
C). Using the psychometric chart again,
⌬
H
ϭ
75
Ϫ
48
ϭ
27 Btu /lb (62.8 kJ/ kg), and the specific volume of the air at
this temperature—from the chart—15.2 ft
3
/lb (0.95 m
3
/kg). Substituting as before,
BHP
ϭ
764 hp (569.6 kW).
Related Calculations. This procedure can be used with any type of cooling
tower employed in air conditioning, steam power plants, internal combustion en-
gines, or gas turbines. The method is based on knowing the tower’s air outlet
temperature. Use the psychometric chart to determine volumes and temperatures
for various air states. As presented here, this method is the work of Ashfaq Noor,
Dawood Hercules Chemicals Ltd., as reported in Chemical Engineering magazine.
With the greater environmental interest in reducing stream pollution of all types,
including thermal, cooling towers are receiving more attention as a viable way to
eliminate thermal problems in streams and shore waters. The cooling tower is a
nonpolluting device whose only environmental impact is the residue left in its bot-
tom pans. Such residue is minor in amount and easily disposed of in an environ-
mentally acceptable manner.

CHOOSING AN ICE STORAGE SYSTEM FOR
FACILITY COOLING
Select an ice storage cooling system for a 100-ton (350-kW) peak cooling load,
10-h cooling day, 75 percent diversity factor, $8.00/ month kW demand charge, 12-
month ratchet—i.e., the utility term for a monthly electrical bill surcharge based
on a previous month’s higher peak demand. Analyze the costs for a partial-storage
and for a full-storage system.
Calculation Procedure:
1. Analyze partial-storage and full-storage alternatives
Stored cooling systems use the term ton-h instead of tons of refrigeration, which
is the popular usage for air-conditioning loads. Figure 1 shows a theoretical cooling
load of 100 tons (350 kW) maintained for 100 h, or a 1000 ton-h (3500 kWh)
cooling load. Each of the squares in the diagram represents 10 ton-h (35 kWh).
No building air-conditioning system operates at 100 percent capacity for the
entire daily cooling cycle. Air-conditioning loads peak in the afternoon, generally
from 2:00 to 4:00 pm, when ambient temperatures are highest. Figure 2 shows a
typical building air-conditioning load profile during a design day.
As Fig. 2 shows, the full 100-ton chiller capacity (350 kW) is needed for only
2 of the 10 h in the cooling cycle. For the other 8 h, less than the total chiller
capacity is required. Counting the tinted squares shows only 75, each representing
10 ton-h (35 kWh). The building, therefore, has a true cooling load of 750 ton-h
(2625 kWh). A 100-ton (350 kW) chiller must however, be specified to handle the
peak 100-ton (250 kW) cooling load.
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HEATING, VENTILATING, AND AIR CONDITIONING
16.14
ENVIRONMENTAL CONTROL
12

Tons
Hours
100
90
80
70
60
50
40
30
20
10
123456789101112123456789101112
FIGURE 1 Cooling load of 100 tons (351.7 kW) maintained for 10 h, or a
1000 ton-h cooling load. (Calmac Manufacturing Corporation.)
12
Tons
Hours
100
90
80
70
60
50
40
30
20
10
123456789101112123456789101112
FIGURE 2 Typical building air-conditioning load profile during a design day.

(Calmac Manufacturing Corporation.)
The diversity factor, defined as the ratio of the actual cooling load to the total
potential chiller capacity, or diversity factor, percent
ϭ
100 (Actual ton-hours) /total
potential ton-hours. For this installation, diversity factor
ϭ
100(750)/ 1000
ϭ
75
percent. If a system’s diversity factor is low, its cost efficiency is also low.
Dividing the total ton-hours of the building by the number of hours the chiller
is in operation gives the building’s average load throughout the cooling period. If
the air-conditioning load can be shifted to off-peak hours or leveled to the average
load, 100 percent diversity can be achieved, and better cost efficiency obtained.
When electrical rates call for complete load shifting, i.e., are excessively high,
a conventionally sized chiller can be used with enough energy storage to shift the
entire load into off-peak hours. This is called a full-storage system and is used most
often in retrofit applications using existing chiller capacity. Figure 3 shows the same
building air-conditioning load profile but with the cooling load completely shifted
into 14 off-peak hours. The chiller is used to build and store ice during the night.
The 32
Њ
F(0
Њ
C) energy stored in the ice then provides the required 750 ton-h (2625
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HEATING, VENTILATING, AND AIR CONDITIONING

HEATING, VENTILATING, AND AIR CONDITIONING
16.15
12
Tons
Hours
100
90
80
70
60
50
40
30
20
10
123456789101112123456789101112
FIGURE 3 Building air-conditioning load profile of Fig. 2 with the cooling
load shifted into 14 off-peak hours. (Calmac Manufacturing Corporation.)
12
Tons
Hours
100
90
80
70
60
50
40
30
20

10
123456789101112123456789101112
FIGURE 4 Extending the hours of operation for 14 to 24 results in the lowest
possible average load
ϭ
750 ton-h / 24
ϭ
31.25. Demand charges are greatly
reduced and chiller capacity can often be decreased 50 to 60 percent, or more.
(Calmac Manufacturing Corporation.)
kWh) of cooling during the day. The average load is lowered to (750 ton-h) /14 h
ϭ
53.6 tons (187.6 kW), which results in significantly reduced demand charges.
In new construction, a partial-storage system is the most practical and cost ef-
fective load-management strategy. In this load-leveling method, the chiller runs
continuously. It charges the ice storage at night and cools the load directly during
the day with help from stored cooling. Extending the hours of operation from 14
to 24 h results in the lowest possible Average Load, (750 ton-h) /24 hours
ϭ
31.25
tons (109.4 kW, as shown by the plot in Fig. 4). Demand charges are greatly reduced
and chiller capacity can often be decreased by 50 to 60 percent, or more.
2. Compute partial-storage demand savings
Cost estimates for a conventional chilled-water air-conditioning system comprised
of a 100-ton (350 kW) chiller with all accessories such as cooling tower, fan coils,
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HEATING, VENTILATING, AND AIR CONDITIONING
16.16

ENVIRONMENTAL CONTROL
pumps, blowers, piping, controls, etc., show a price of $600/ton, or 100 tons
ϫ
$600/ ton
ϭ
$60,000. The distribution system for this 100-ton (350-kW) plant will
cost about the same, or $60,000. Total cost therefore
ϭ
$60,000
ϩ
$60,000
ϭ
$120,000.
With partial-storage using a 40 percent size chiller with ice storage at a 75
percent diversity factor, the true cooling load translates into 750 ton-h (2626 kWh)
with the chiller providing 400 ton-h (1400-kWh) and stored ice the balance, or 750
Ϫ
400
ϭ
350 ton-h (1225 kWh). Hence, cost of 40-ton chiller at $600/ton
ϭ
$24,000. From the manufacturer of the stored cooling unit, the installed cost is
estimated to be $60 /ton-h, or $60
ϫ
350 ton-h
ϭ
$21,000. The distribution system,
as before, costs $60,000. Hence, the total cost of the partial-storage system will be
$24,000
ϩ

$21,000
ϩ
$60,000
ϭ
$105,000. Therefore, the purchase savings of the
partial-storage system over the conventional chilled-water air-conditioning system
ϭ
$120,000
Ϫ
$105,000
ϭ
$15,000.
The electrical demand savings, which continue for the life of the installation,
are: (100 tons
Ϫ
40 tons chiller capacity)(1.5-kW/ ton at peak summer demand
conditions, including all accessories)($8.00/ mo/ kW demand charge)(12 mo/yr)
ϭ
$8640.
3. Determine full-storage savings
With full-storage, 100-ton (350-kW) peak cooling load, 10-h cooling day, 75 per-
cent diversity factor, 1000-h cooling season, $8.00/mo/kW demand charge, 12-
month ratchet, $0.03 /kWh off-peak differential, the chiller cost will be (10 h)(100
tons)(75 percent)($60 /ton-h, installed)
ϭ
$45,000. The demand savings will be, as
before, (100 tons)(1.5 kW/ton)(12 mo)($8.00 /mo /kW demand charge)($8.00)
ϭ
$14,400/ year. Energy savings are computed using the electric company’s off-peak
kWh off-peak differential, or ($0.03/ kWh)(1000 h)(100 tons)(1.2 average kW /ton)

ϭ
$3,600/ year. The simple payback time for this project
ϭ
(equipment cost, $)/
(demand savings, $
ϩ
energy savings, $)
ϭ
$45,000/ $18,000
ϭ
2.5 yr. After the
end of the payback time there is an annual energy savings of $18,000/ year. And
as rates increase, which they usually do, the annual savings will probably increase
above this amount.
Related Calculations. Ice storage systems are becoming more popular for a
variety of structures: office buildings, computer data centers, churches, nursing
homes, police stations, public libraries, theaters, banks, medical centers, hospitals,
hotels, convention centers, schools, colleges, universities, industrial training centers,
cathedrals, medical clinics, manufacturing plants, warehouses, museums, country
clubs, stock exchanges, government buildings, and courthouses.
There are several reasons for this growing popularity: (1) utility power costs can
be reduced by shifting electric power demand to off hours by avoiding peak-demand
charges; (2) lower overall electric rates can be obtained for the facility if the kil-
owatt demand is reduced, thereby eliminating the need for the local utility to build
new generating facilities; (3) ice storage can provide uninterrupted cooling in times
of loss of outside, or inside, electric generating capability during natural disasters,
storms, or line failures—the ice storage system acts like a battery, giving the cooling
required until the regular coolant supply can be reactivated; (4) environmental reg-
ulations are more readily met because less power input is required, reducing the
total energy usage; (5) by making ice at night, the chiller operates when the facil-

ities’ electrical demands are lowest and when a utility’s generating capacity is un-
derutilized; (6) provision can be made to use more environmentally friendly HCFC-
123, thereby complying with current regulations of federal and state agencies; (7)
facility design can be planned to include better control of indoor air quality, another
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HEATING, VENTILATING, AND AIR CONDITIONING
16.17
environmentally challenging task faced by designers today; (8) new regulations,
specifically The Energy Policy Act of 1992 (EPACT), curtails the use of and elim-
inates certain fluorescent and incandescent lamps (40-W T12, cool white, warm
white, daylight white, and warm white deluxe), which will change both electrical
demand and replacement bulb costs in facilities, making cooling costs more im-
portant in total operating charges.
Designers now talk of ‘‘greening’’ a building or facility, i.e., making it more
environmentally acceptable to regulators and owners. An ice storage system is one
positive step to greening a facility while reducing the investment required for cool-
ing equipment. The procedure given here clearly shows the savings possible with
a typical well-designed ice storage system.
There are three common designs used for ice storage systems today: (1) direct-
expansion ice storage where ice is frozen directly on metal refrigerant tubes sub-
merged in a water tank; cooled water in the tank is pumped to the cooling load
when needed; (2) ice harvester system where a thin coat of ice is frozen on refrig-
erated metal plates and periodically harvested into a bin or water tank by melting
the bond of the ice to the metal plates; the chilled water surrounding the ice is
pumped to the cooling load when needed; (3) patented ice bank system uses a
modular, insulated polyethylene tank containing a spiral-wound plastic tube heat
exchanger surrounded with water; at night a 26

Њ
F(
Ϫ
3.33
Њ
C) 75 percent water/25
percent glycol solution from a standard packaged air-conditioning chiller circulates
through the heat exchanger, freezing solid all the water in the tank; during the day
the ice cools the solution to 44
Њ
F (6.66
Њ
C) for use in the air-cooling coils where it
cools the air from 75
Њ
F (23.9
Њ
C) to 55
Њ
F (12.8
Њ
C).
The patented system has several advantages over the first two, namely: (1) ice
is the storage medium, rather than water. One pound (0.45 kg) of ice can store 144
Btu (152 J) of energy, while one pound of water in a stratified tank stores only 12
to 15 Btu (12.7 J to 15.8 J). Hence, such an ice storage system needs only about
one-tenth the space for energy storage. This small space requirement is important
in retrofit applications where space is often scarce.
(2) Patented systems are closed; there is no need for water treatment or filtration;
pumping power requirements are small; (3) power requirements are minimal; (4)

installation of the insulated modular tanks is fast and inexpensive since there are
no moving parts; the tanks can be installed indoors or outdoors, stacked or buried
to save space. Currently these tanks are available in three sizes: 115, 190, and 570
ton-h (402.5, 665, and 1995 kWh). (5) A low-temperature duct system can be used
with 45
Њ
F (7.2
Њ
C) air instead of the conventional 55
Њ
F (12.8
Њ
C) air in the air-
conditioning system. This can permit further large savings in initial and operating
costs. The 45
Њ
F (7.2
Њ
C) primary air requires much lower air flow [ft
3
/min (m
3
/m)]
than 55
Њ
F (12.8
Њ
C) air. This reduces the needed size of both the air handler and
duct system; both may be halved. Energy savings from the smaller air-handler
motors may total 20 percent, even after figuring the additional energy required for

the small mixing-box motors.
This procedure is based on data provided by the Calmac Manufacturing Cor-
poration, Englewood, NJ. The economic analysis was provided by Calmac, as were
the illustrations in this procedure. Calmac manufactures the Levload modular in-
sulated storage tanks mentioned above that are used in their Ice Bank Stored Cool-
ing System. Their system, when designed with a low-temperature heat-recovery
loop, can also make the chiller into a water-source heat pump for winter heating.
Thus, office and similar buildings often require heating warm-up in the morning on
winter days, but these same buildings may likewise require cooling in the afternoon
because of lights, people, computers, etc. Ice made in the morning to provide heat-
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HEATING, VENTILATING, AND AIR CONDITIONING
16.18
ENVIRONMENTAL CONTROL
FIGURE 5 Counterflow heat-
exchanger tubes used in the Ice
Bank. (Calmac Manufacturing
Corporation.)
ing supplies free afternoon cooling and melts to be ready for the next day’s warm-
up. Even on coldest days, low-temperature waste heat (such as cooling water or
exhaust air), or off-peak electric heat can be used to melt the ice. Oil or gas con-
nections to the building can thus be eliminated.
Nontoxic eutectic salts are available to lower the freezing point of the water in
Calmac Ice Banks to either 28
Њ
F(
Ϫ
2.2

Њ
C) or 12
Њ
F(
Ϫ
11.1
Њ
C) and, consequently,
the temperature of the resulting ice. Twenty-eight-degree ice, for example, can
provide cold, dry primary air for many uses, including extra-low temperature airside
applications. Twelve-degree ice can be used for on-ground aircraft cooling, off-peak
freezing of ice rinks, and for industrial process applications requiring colder liquids.
Other temperatures can be provided for specialized applications, such as refrigerated
warehouses.
Figure 5 shows the charge cycle using a partial storage system for an air-
conditioning installation. At night a water-glycol solution is circulated through a
standard packaged air-conditioning chiller and the Ice Bank heat exchanger, by-
passing the air-handler coil. The cooling fluid is at 26
Њ
F(
Ϫ
3.3
Њ
C) and freezes the
water surrounding the heat exchanger.
During the day, Fig. 6, the water-glycol solution is cooled by the Ice Bank from
52
Њ
F (11.1
Њ

C) to 34
Њ
F (1.1
Њ
C). A temperature-modulating valve, set at 44
Њ
F (6.7
Њ
C)
in a bypass loop around the Ice Bank, allows a sufficient quantity of 52
Њ
F (11.1
Њ
C)
fluid to bypass the Ice Bank, mix with 34
Њ
F (1.1
Њ
C) fluid, and achieve the desired
44
Њ
F (6.7
Њ
C) temperature. The 44
Њ
F (6.7
Њ
C) fluid enters the coil, where it cools the
air passing over the coil from 75
Њ

F (23.9
Њ
C) to 55
Њ
F (12.8
Њ
C). Fluid leaves the coil
at 60
Њ
F (15.6
Њ
C), enters the chiller and is cooled to 52
Њ
F (11.1
Њ
C).
Note that, while making ice at night, the chiller must cool the water-glycol to
26
Њ
F(
Ϫ
3.3
Њ
C), rather than produce 44
Њ
F (6.7
Њ
C) or 45
Њ
F (7.2

Њ
C) water temperatures
required for conventional air-conditioning systems. This has the effect of ‘‘derating’’
the nominal chiller capacity by about 30 percent. Compressor efficiency, however,
is only slightly reduced because lower nighttime temperatures result in cooler con-
denser water from the cooling tower (if used) and help keep the unit operating
efficiently. Similarly, air-cooled chillers benefit from cooler condenser entering air-
temperatures at night.
The temperature-modulating valve in the bypass loop has the added advantage
of providing unlimited capacity control. During many mild-temperature days in the
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HEATING, VENTILATING, AND AIR CONDITIONING
16.19
SI Values
FC
44 6.7
55 12.8
60 15.6
75 23.9
FIGURE 6 Charge cycle. (Calmac Manufacturing Cor-
poration.)
SI Values
FC
52 11.1
60 15.6
34 1.1
44 6.7

55 12.8
60 15.6
75 23.9
FIGURE 7 Discharge cycle. (Calmac Manufacturing Cor-
poration.)
spring and fall, the chiller will be capable of providing all the cooling needed for
the building without assistance from stored cooling. When the building’s actual
cooling load is equal to or lower than the chiller capacity, all of the system coolant
flows through the bypass loop, as in Fig. 8.
Using 45
Њ
F (7.2
Њ
C) rather than 55
Њ
F (12.8
Њ
C) system air in the air-conditioning
system permits further large savings in initial and operating costs. The 45
Њ
F (7.2
Њ
C)
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HEATING, VENTILATING, AND AIR CONDITIONING
16.20
ENVIRONMENTAL CONTROL
SI Values

FC
26 –3.3
32 0.0
FIGURE 8 When cooling load equals, or is lower than
chiller capacity, all coolant flows through bypass loop. (Calmac
Manufacturing Corporation.)
low-temperature air is achieved by piping 38
Њ
F (3.3
Њ
C) water-glycol solution from
the stored cooling Ice Bank to the air handler coil instead of mixing it with bypassed
solution, as in Fig. 6. The 45
Њ
F (7.2
Њ
C) air is used as primary air and is distributed
to motorized fan-powered mixing boxes where it is blended with room air to obtain
the desired room temperature. Primary 45
Њ
F (7.2
Њ
C) air requires much lower ft
3
/
min (m
3
/min) than 55
Њ
F (12.8

Њ
C) air. Consequently, the size and cost of the air
handler and duct system may be cut in about half. Energy savings of the smaller
air handler motors total 20 percent, even counting the additional energy required
for the small mixing-box motors.
The recommended coolant solution for these installations is an ethylene glycol-
based industrial coolant such as Union Carbide Corporation’s UCARTHERM
௡
or
Dow Chemical Company’s DOWTHERM
௡
SR-1. Both are specially formulated for
low viscosity and superior heat-transfer properties, and both contain a multi-
component corrosion inhibitor system effective with most materials of construction,
including aluminum, copper, solder, and plastics. Standard system pumps, seals,
and air-handler coils can be sued with these coolants. However, because of the
slight difference in the heat-transfer coefficient between water-glycol and plain wa-
ter, air-handler coil capacity should be increased by about 5 percent. Further, the
water and glycol must be thoroughly mixed before the solution enters the system.
Another advantage of ice storage systems for cooling and heating is provision
of an uninterrupted power supply (UPS) in the event of the loss of a building’s
cooling or heating facilities. Such an UPS can be important in data centers, hos-
pitals, research laboratories, and other installations where cooling or heating are
critical.
Figure 9a shows the conventional ‘‘ice builder’’ and Fig. 9b the LEVLOAD Ice
Bank. When ice is stored remote from the refrigerating system evaporator, as in
Fig. 9b, the evaporator is left free to aid the cooling during the occupied hours of
a building or other structure. Figure 10 compares the chiller performance of a Partial
Storage Ice Bank, Fig. 9b (upper curve), with an ice-builder system, Fig. 9a (lower
curve), on a typical design day. Note that when compressor cooling is done through

ice on the evaporator, suction temperatures are low and kW/ ton is increased.
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HEATING, VENTILATING, AND AIR CONDITIONING
HEATING, VENTILATING, AND AIR CONDITIONING
16.21
Ice Builder
Refrigerant
System Coolant
(Water)
LEVLOAD Ice Bank
Solid Ice
Melted Ice Area
System Coolant
(Water-Glycol)
FIGURE 9 (a) Typical ice-builder arrangement. (b) LEVLOAD Ice
Bank method of ice burn-off (ice melting). (Calmac Manufacturing
Corporation.)
SI Values
FC
18 –7.8
35 1.7
45 7.2
FIGURE 10 Comparison of chiller performance of a Partial
Storage Ice Bank (upper curve) and conventional ice-builder
system (lower curve). (Calmac Manufacturing Corporation.)
With discussions still taking place about chlorofluorocarbon refrigerants suitable
for environmental compliance, designers have to seek the best choice for the system
chiller. One approach finding popularity today as an interim solution is to choose

a chiller which can use an energy-efficient refrigerant today and the most environ-
mentally friendly refrigerant in the future. Thus, for some chillers, CFC-11 is the
most energy-efficient refrigerant today. The future most environmentally friendly
refrigerant is currently thought to be HCFC-123. By choosing and sizing a chiller
that can run on HCFC-123 in the future, energy savings can be obtained today,
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HEATING, VENTILATING, AND AIR CONDITIONING
16.22
ENVIRONMENTAL CONTROL
and, if environmental requirements deem a switch in the future, the same chiller
can be used for CFC-11 today and HCFC-123 in the future. It is also possible that
the same chiller can be retrofitted to use a non-CFC refrigerant in the future. A
number of ice storage systems have adopted this design strategy. Many firms that
installed ice storage systems in recent years are so pleased with the cost savings
(energy, equipment, ducting, UPS, etc.) that they plan to expand such systems in
the future.
Chillers using CFC-11 normally produce ice at 0.64 to 0.75 kW/ton, depending
on the amount of ice produced. Power consumption is lower when larger quantities
of ice are produced. When HCHC-123 is used, the power input ranges between 0.7
and 0.8 kW/ ton, again depending on the number of tons produced. As before,
power consumption is lower when larger tonnages are produced. New centrifugal
chillers produce cooling at power input ranges close to 0.5 kW/ ton.
In all air-conditioning systems, designers must recognize that there are three
courses of action open to them when CFC refrigerants are no longer available: (1)
continue to use existing CFC-based equipment, taking every precaution possible to
stop leaks and conserve available CFC supplies; (2) retrofit existing chilling equip-
ment to use non-CFC refrigerants; this step requires added investment and changed
operating procedures; (3) replace existing chillers with new chillers specifically

designed for non-CFC refrigerants; again, added investment and changed operating
procedures will be necessary.
To avoid CFC problems, new high-efficiency chlorine-free screw chillers are
being used. And there are packaged ammonia screw chiller available also. Likewise,
a variety of alternative refrigerants are now being produced for new and retrofit
refrigeration and air-conditioning uses.
Table 1 shows an economic analysis of typical partial-storage and full-storage
installations. A conventional chilled-water air-conditioning system is compared with
a partial-storage 40 percent size chiller with Ice Banks in the partial-storage anal-
ysis. Full-storage produces the simple payback time of 2.5 years for the investment.
Data in this analyze are from Calmac Manufacturing Corporation. When using a
similar analysis, be certain to obtain current prices for components, demand charges,
and electricity. Values given here are for illustration purposes only.
ANNUAL HEATING AND COOLING ENERGY
LOADS AND COSTS
A 2000 ft
2
(185.8 m
2
) building has a 100,000 Btu/h (29.3 kW) heat loss in an area
where the heating season is 264 days’ duration. Average winter outdoor temperature
is 42
Њ
F (5.6
Њ
C); design conditions are 70
Њ
F (21.1
Њ
C) indoors and 0

Њ
F(
Ϫ
17.8
Њ
C)
outdoors. The building also has a summer cooling load of 7.5 tons (26.4 kW) with
an estimated full-load cooling time of 800 operating hours. What are the total winter
and summer estimated loads in Btu /h (kW)? If oil is 90 cents /gallon and electricity
is 7 cents /kWh, what are the winter and summer energy costs? Use a 24-h heating
day for winter loads and a boiler efficiency of 75 percent when burning oil with a
higher heating value (HHV) of 140,000 Btu/ gal (39,018 MJ /L).
Calculation Procedure:
1. Compute the winter operating costs
The winter seasonal heating load, WL
ϭ
N
ϫ
24 h/day
ϫ
Btu/ h heat loss
ϫ
(average indoor temperature,
Њ
F
Ϫ
average outdoor temperature,
Њ
F)/ (average in-
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HEATING, VENTILATING, AND AIR CONDITIONING
16.23
TABLE 1
Economic Analysis of Typical Partial Storage and Full Storage Installations*
Partial storage
Assume: 100-ton peak cooling load, 10-h cooling day, 75 percent diversity
factor, $8.00 /mo /kW demand charge, 12-mo ratchet.*
Conventional Chilled Water Air Conditioning System:
100-ton chiller at $600 / ton, installed** $ 60,000
Distribution system 60,000
Total $120,000
Partial storage (40% size chiller with Ice Banks):
At 75 percent diversity factor, the true cooling load translates into 750 ton-
h with the chiller providing 400 ton-h and stored cooling the balance, or 350
ton-h. Therefore:
40-ton chiller at $600 / ton, installed $ 24,000
Stored cooling at $60 / ton-h, installed 21,000
Distribution system 60,000†
Total $105,000
Purchase savings: $ 15,000
Demand savings:
60 tons
ϫ
1.5 kW/ ton‡
ϫ
12 mo
ϫ
$8.00

ϭ
$8,640/ yr
*Utility term for a monthly electrical bill surcharge based on a previous
month’s higher peak demand.
Full storage
Assume: 100-ton peak cooling load, 10-hour cooling day, 75% diversity fac-
tor, 1000-hour cooling season, $8.00/ mo / kW demand charge, 12-mo ratchet,
$0.03/ kWh off-peak differential.
Full storage:
10 h
ϫ
100 tons
ϫ
75%
ϫ
$60/ ton-h, installed $45,000
Demand savings:
100 tons
ϫ
1.5 kW/ ton
ϫ
12 mo
ϫ
$8.00 $14,400/ yr
Energy savings:
$0.03/ kWh
ϫ
1000 h
ϫ
100 tons

ϫ
1.2 Avg. kW /ton $ 3600/yr
Total savings: $18,000/ yr
Simple payback: $45,000
Ϭ
$18,000
ϭ
2.5 yr.
**The $600 /ton includes all accessories, such as cooling tower, fan coils,
pumps, blowers, piping, controls, etc.
†Figure shown is for conventional temperature system. This cost could be
reduced by 50 percent by using a low temperature duct system.
‡The 1.5 kW/ ton is figured at the peak summer demand conditions and also
includes all accessories.
Model Model Model
Specifications 1098 1190 1500
Total ton-hour capacity 115 190 570
Tube surface /ton-h, ft
2
12.0 12.0 12.0
Nominal discharge time, h 6-12 6-12 6-12
Latent storage cap., ton-h 98 162 486
Sensible storage cap., ton-h 17 28 84
Maximum operating temp.,
Њ
F 100 100 100
Maximum operating press., lb /
in
2
90 90 90

Outside diameter, in 89 89 —
Length
ϫ
width, in — — 268
ϫ
96
Height, in 68 101 102
Weight, unfilled, lb 1,060 1,550 4,850
Weight, filled, lb 9.940 16,750 50,450
Model Model Model
Specifications 1098 1190 1500
Volume of water/ ice, gals. 980 1620 4860
Volume of solution in HX, gals. 90 148 555
Press. drop (25% glycol, 28
Њ
F),
PSI
20 gal / min 4.0 — —
40 gal / min 9.7 5.0 —
60 gal / min 18.8 9.0 —
80 gal / min — 13.0 2.8
160 gal / min — — 7.0
240 gal / min — — 13.0
The outlet temperatures from the tanks vary with the rate at which the tanks
are discharged. See LEVLOAD Performance Manual for details.
LEVLOAD and CALMAC are registered trademarks of Calmac Manufac-
turing Corporation. The described product and its applications are protected
by United States Patents 4,294,078; 4,403,645; 4,565,069; 4,608,836;
4,616,390; 4,671,347 and 4,687,588.
*Calmac Manufacturing Corporation

SI values in procedure text.
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HEATING, VENTILATING, AND AIR CONDITIONING
16.24
ENVIRONMENTAL CONTROL
door temperature,
Њ
F
Ϫ
outside design temperature,
Њ
F), where N
ϭ
number of days
in the heating season. For this building, WL
ϭ
(264
ϫ
24
ϫ
100,000)(70
Ϫ
42)/
(70
Ϫ
0)
ϭ
253,440,000 Btu (267.4 MJ).

The cost of the heating oil, CO
ϭ
Btu seasonal heating load
ϫ
oil cost $/gal/
boiler efficiency
ϫ
HHV. Or, for this building, CO
ϭ
(253,440,000)(0.90)/ 0.75
ϫ
140,000
ϭ
$2,172.34 for the winter heating season.
2. Calculate the summer cooling cost
The summer seasonal electric consumption in kWh, SC
ϭ
tons of air conditioning
ϫ
kW/ ton of air conditioning
ϫ
number of operating hours of the system. The
kW/ design ton factor is based on both judgment and experience. General consensus
amongst engineers is that the kW/ton varies from 1.8 for small window-type sys-
tems to 1.0 for large central-plant systems. The average value of 1.4 kW /design
ton is frequently used and will be used here. Thus, the summer electric consump-
tion, SC
ϭ
7.5
ϫ

1.4
ϫ
800
ϭ
8400 kWh. This energy will cost 8400 kWh
ϫ
$0.07
ϭ
$588.00.
The total annual energy cost is the sum of the winter and summer energy costs,
$2172.34
ϩ
$588.00
ϭ
$2760.34.
Related Calculations. Use this procedure to compute the energy costs for any
type of structure, industrial, office, residential, medical, educational, etc., having
heating or cooling loads, or both. Any type of fuel, oil, gas, coal, etc., can be used
for the structure.
This procedure is the work of Jerome F. Mueller, P.E. of Mueller Engineering
Corp.
HEAT RECOVERY USING A RUN-AROUND
SYSTEM OF ENERGY TRANSFER
A hospital operating room suite requires 6000 ft
3
/min (169.8 m
3
/s) of air in the
supply system with 100 percent exhaust and 100 percent compensating makeup air.
Winter outdoor design temperature is 0

Њ
F(
Ϫ
17.8
Њ
C); operating-room temperature
is 80
Њ
F (26.7
Њ
C) with 50 percent relative humidity year-round. How much energy
can be saved by installing coils in both the supply and exhaust air ducts with a
pump circulating a non-freeze liquid between the two coils, absorbing heat from
the exhaust air and transferring this heat to the makeup air being introduced?
Calculation Procedure:
1. Choose the coils to use
In the winter the exhaust air is at 80
Њ
F (26.7
Њ
C) while the supply air is at 0
Њ
F
(
Ϫ
17.8
Њ
C). Hence, the coil in the exhaust air duct will transfer heat to the nonfreeze
liquid. When this liquid is pumped through the coil in the intake-air duct it will
release heat to the incoming air. This transfer of otherwise wasted heat will reduce

the energy requirements of the system in the winter.
As a first choice, select a coil area of 12 ft
2
(1.1 m
2
) with a flow of 6000 ft
3
/
min (169.8 m
3
/s). While a number of coil arrangements are possible, the listing
below shows typical coil conditions at face velocities of 500 ft/ min (152.4 m /min)
and 600 ft /min (182.8 m/min) with a coil having 8-fins/in coil. Entering coolant
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HEATING, VENTILATING, AND AIR CONDITIONING
HEATING, VENTILATING, AND AIR CONDITIONING
16.25
temperature
ϭ
45
Њ
F (7.2
Њ
C); entering air temperature
ϭ
80
Њ
F (26.7

Њ
C) dry bulb,
67
Њ
F (19.4
Њ
C) wet bulb. Various manufacturers’ values may vary slightly from these
values.
Ft/ min (m/ min)
face velocity
Temp rise,
Њ
F(
Њ
C)
No. of
rows
Total MBtu /h
(kWh)
Leaving
dry bulb,
Њ
F(
Њ
C)
Leaving wet
bulb,
Њ
F
(

Њ
C)
500 (152.4) 12 (21.6) 6 16.6 (4.86) 57.3 (14.1) 56.5 (13.6)
500 (152.4) 12 (21.6) 8 20.1 (5.89) 54.2 (12.3) 53.9 (12.2)
600 (182.9) 10 (18) 4 14.3 (4.19) 62.1 (16.7) 59.7 (15.4)
The middle coil listed above, if placed in the exhaust duct, would produce 20.1
MBtu/ h (5.89 kWh) with a 12
Њ
F (21.6
Њ
C) temperature rise with a leaving air tem-
perature of 53.9
Њ
F (12.2
Њ
C) when the liquid coolant enters the coil at 45
Њ
F (7.2
Њ
C)
and leaves 57
Њ
F (13.9
Њ
C).
2. Determine the coil heating capacity
The heating capacity of a coil is the product of (coil face area, ft
2
)[heat release,
Btu/(h

⅐
ft
2
) of coil face area]. For this coil, heating capacity
ϭ
12
ϫ
20,100
ϭ
241,200 Btu/ h (70.7 kWh). The incoming makeup air can be heated to a temper-
ature of: (heating capacity, Btu) /(makeup air flow, ft
3
/min)(1.08)
ϭ
241,200/ 6000
ϫ
1.08
ϭ
37.2
Њ
F (2.9
Њ
C). Hence, the makeup air is heated from 0
Њ
F(
Ϫ
17.8
Њ
C) to
37.2

Њ
F (2.9
Њ
C).
The energy saved is—assuming 1000 Btu /lb of steam (2330 kJ/kg)—241,200
/1000
ϭ
241.2 lb/h (109.5 kg /h). With a 200-day heating season and 10-h
operation/ day, the saving will be 200
ϫ
10
ϫ
241.2
ϭ
482,400 lb /yr (219,010
kg/ yr). And if steam costs $20/ thousand pounds ($20/454 kg), the saving will be
(482,400/ 1000)($20)
ϭ
$9,648,00/ year. Such a saving could easily pay for the
heating coil in one year.
Related Calculations. Use this general approach to choose heating coils for
any air-heating application where waste heat can be utilized to increase the tem-
perature of incoming air, thereby reducing the amount of another heating medium
that might be required. Most engineers use the 1000 Btu /lb (2300 kJ /kg) latent
heat of steam as a safe number to convert quickly from hourly heat savings in Btu
(kg) to pounds of steam. This procedure can be used for industrial, commercial,
residential, and marine applications.
The procedure is the work of Jerome F. Mueller, P.E., of Mueller Engineering
Corp.
Figure 11 is a typical run-around coil detail that is very commonly used in

energy recovery systems in which the purpose is to exrract heat from air that must
be exhausted. Normally about 40 to 60 percent of the heat being wasted can be
recovered. This seemingly simple detail has two points that should be carefully
noted. The difference in fluid temperatures in this closed system creates small ex-
pansion and contraction problems and an expansion tank is required. Most impor-
tantly the temperature of the incoming supply air can create a coil temperature so
low that the coil in the exhaust air stream begins to ice up. Normally this betins at
some 35
Њ
F (1.67
Њ
C). This is when the three-way bypass valve comes into play and
the glycol is not circulated through the outside air coil. Obviously if the outside air
temperature is low enough, the system will go into full bypass and the ciruclating
pump should be stopped.
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