Tài liệu Handbook of Mechanical Engineering Calculations P9 pdf

9.1
SECTION 9
AIR AND GAS COMPRESSORS
AND VACUUM SYSTEMS
Estimating the Cost of Air Leaks in
Compressed-Air Systems
9.1
Selecting an Air Motor for a Known
Application
9.4
Air-Compressor Cooling-System Choice
for Maximum Coolant Economy
9.10
Economics of Air-Compressor Inlet
Location
9.14
Power Input Required by Centrifugal
Compressor
9.16
Compressor Selection for Compressed-
Air Systems
9.18
Sizing Compressed-Air System
Components
9.24
Compressed-Air Receiver Size and
Pump-Up Time
9.26
Vacuum-System Pump-Down Time
9.27
Vacuum-Pump Selection for High-

Vacuum Systems
9.30
Vacuum-System Pumping Speed and
Pipe Size
9.33
Determining Air Leakage in Vacuum
Systems by Calculation
9.34
Checking the Vacuum Rating of a
Storage Vessel
9.36
Sizing Rupture Disks for Gases and
Liquids
9.39
Determining Airflow in Pipes, Valves,
and Fittings
9.40
System Economics and Design Strategies
ESTIMATING THE COST OF AIR LEAKS IN
COMPRESSED-AIR SYSTEMS
Find the cost of compressed air leaking through a 0.125-in (0.3175-cm) diameter
hole in a pipe main of a typical industrial air piping system, Fig. 1, to the atmo-
sphere at sea level when the air pressure in the pipe is 10 lb/in
2
(gage) (68.9 kPa),
the plant, Fig. 2, operates 7500 h/yr, air temperature is 70
Њ
F (21.1
Њ
C), and the cost

of compressed air is $1.25 per 1000 ft
3
(28.3 m
3
). What is the cost of the leaking
air when the pipe pressure is 50 lb/in
2
(gage) (344.5 kPa) and the other variables
are the same as given above?
Calculation Procedure:
1. Find the volume of air discharged to the atmosphere
Air flowing through an orifice or nozzle attains a critical pressure of 0.53 times the
inlet or initial pressure. This reduced pressure occurs at the throat or vena contracta,
which is the point of minimum stream diameter on the outlet side of the air flow.
If the outlet or back pressure exceeds the critical pressure then the vena contracta
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Source: HANDBOOK OF MECHANICAL ENGINEERING CALCULATIONS
9.2
PLANT AND FACILITIES ENGINEERING
FIGURE 1 Typical compressed-air system main and branch pipes (Factory Manage-
ment and Maintenance).
or throat pressure rises to equal the backpressure. Air flow through a hole in a pipe
or tank replicates the flow through an orifice or nozzle.
When an inlet air pressure of 10 lb/ in
2
(gage)
ϩ
14.7

ϭ
24.7 lb/in
2
(abs) (170.2
kPa), the critical pressure is 0.53
ϫ
24.7
ϭ
13.09 lb/ in
2
(abs) (90.2 kPa). Since
13.09 lb /in
2
(abs) is less than the atmospheric backpressure of 14.7 lb/ in
2
(abs)
(101.3 kPa), the throat pressure equals the backpressure, or 14.7 lb/in
2
(abs) (101.3
kPa). Knowing this, we can compute weight of the escaping air from W
ϭ
1.06
A(P
1
[P
Ϫ
P
1
]/T)
0.5

, where W
ϭ
leakage rate, lb /s (kg /s); A
ϭ
area of leakage
hole, in
2
(cm
2
); P
ϭ
pipeline or initial air pressure, lb/ in
2
(abs) (kPa); P
1
ϭ
outlet
or backpressure, lb/ in
2
(abs); T
ϭ
absolute temperature of the air before leakage
ϭ Њ
F
ϩ
460.
Substituting, using the values given above, W
ϭ
1.06
ϫ

0.012272(14.7[24.7
Ϫ
14.7]/ 530)
0.5
ϭ
0.006851 lb/s (0.0031 kg/ s). Converting this air leakage rate to
lb/ h (kg/h), multiply by 3600 s/h, or 0.006851
ϫ
3600
ϭ
24.66 lb/ h (11.19 kg/
h). Since the cost of compressed air is expressed in $/ft
3
, the flow rate of the
leaking air must be converted. Since air at 14.7 lb/in
2
(abs) (101.3 kPa) weighs
0.075 lb per ft
3
(1.2 kg/ m
3
), the rate of leakage is 24.66/0.075
ϭ
328.8 ft
3
/h (9.31
m
3
/h).
2. Determine the annual cost of the air leakage

This compressed-air plant operates 7500 h/ yr. Since the leakage rate is 328.8
ft
3
/h, the annual leakage through this opening is 7500
ϫ
328.8
ϭ
2,466,000 ft
3
(58,691 m
3
). At a cost of $1.25 per 1000 ft
3
, the annual total cost of this leak is
$1.25
ϫ
2,466,000/ 1000
ϭ
$3,082.50. This is a sizeable charge, especially if there
are several leaks of this size, or larger, in the system.
3. Find the rate of leakage at the higher line pressure
When the backpressure is less than the critical pressure, a different flow equation
must be used. In the second instance, the critical pressure is 0.53 (50 lb/in
2
(gage)
ϩ
14.7)
ϭ
34.29 lb /in
2

(abs) (236.26 kPa). Since 34.29 lb/in
2
(abs) (236.26 kPa)
is greater than the atmospheric backpressure of 14.7 lb/ in
2
(abs) (101.3 kPa), the
critical pressure is greater than the backpressure. Air leakage through the hole is
now given by W
ϭ
0.5303(ACP)/(T)
0.5
, where C
ϭ
flow coefficient
ϭ
1.0; other
symbols as before.
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AIR AND GAS COMPRESSORS AND VACUUM SYSTEMS
9.3
FIGURE 2 Typical compressed-air plant showing compressor and its associated piping and accessories (Power).
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AIR AND GAS COMPRESSORS AND VACUUM SYSTEMS
9.4
PLANT AND FACILITIES ENGINEERING
Substituting, W

ϭ
0.5303(0.012272)(1.0)(64.7)/ (530)
0.5
ϭ
0.01829 lb /s (0.0083
kg/ s). Converting to an hourly flow rate as earlier, 3600
ϫ
0.01828
ϭ
65.84 lb/ h
(29.89 kg/ h).
4. Compute the annual cost of air leakage at the higher pressure
Following the same steps as earlier, annual leakage cost
ϭ
65.84 lb /h (7500
h/ yr)(1.25/ 1000 ft
3
)/ 0.075 lb /ft
3
ϭ
$8,230.00 per year. Again, this is a significant
loss of revenue. Further, the loss at the higher pressure is $8230 /3082.50
ϭ
2.67
times as great. This points out the fact that higher pressures in a compressed-air
system can cause more expensive leaks.
Related Calculations. Compressing air requires a power input to raise the air
pressure from atmospheric to the level desired for the end use of the air. When
compressed air leaks from a pipe or storage tank, the power expended in compres-
sion is wasted because the air does no useful work when it leaks into the atmo-

sphere.
In today’s environment-conscious world, compressed-air leaks are considered to
be especially wasteful because they increase pollution without producing any ben-
eficial results. The reason for this is that the fuel burned to generate the power to
compress the wasted air pollutes the atmosphere unnecessarily because the air pro-
duces only a hissing sound as it escapes through the hole in the pipe or vessel.
SELECTING AN AIR MOTOR FOR A KNOWN
APPLICATION
Show how to select a suitable air motor for a reversible application requiring 2 hp
(1.5 kW) at 1000 rpm for an industrial crane. Determine the probable weight of
the motor, its torque output, and air consumption for this intermittent duty appli-
cation. An adequate supply of air at a wide pressure range is available at the
installation.
Calculation Procedure:
1. Assemble data on possible choices for the air motor
There are four basic types of air motors in use today: (1) radial-piston type; (2)
axial-piston type; (3) multi-vaned type; (4) turbine type. Each type of air motor has
advantages and disadvantages for various applications. Characteristics of these air
motors are as follows:
(1) Radial-piston air motors, Fig. 3, have four or five cylinders mounted around
a central crankshaft similar to a radial gasoline engine. Five cylinders are preferred
to supply more horsepower with evenly distributed power pulses. In such a unit
there are always two cylinders having a power stroke at the same time. The radial-
piston motor is usually a slow-speed unit, ranging from 85 to 1500 rpm. It is suited
for heavy-duty service up to 20 hp (15 kW) where good lugging characteristics are
needed. Normally they are not reversible, though reversible models are available at
extra cost.
(2) Axial-piston air motors, Fig. 4, are more compact in design and require less
space than a four- or five-cylinder radial-piston motor. Air drives the pistons in
translation; a diaphragm-type converter changes the translation into rotation. This

arrangement supplies high horsepower per unit weight. Axial-piston motors are
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AIR AND GAS COMPRESSORS AND VACUUM SYSTEMS
AIR AND GAS COMPRESSORS AND VACUUM SYSTEMS
9.5
FIGURE 3 (a) Five-cylinder piston-type radial air motor used in sizes from about 2 hp (1.5
kW) to 22 hp (16.4 kW) and at speeds from 85 to 1500 rpm. (b) How five-cylinder air motor
distributes power. Two cylinders are always on power stroke at any instant (Gardner-Denver
Company).
available in sizes from 0.5 to 2.75 hp (0.37 to 2.1 kW). They run equally well in
either direction. To make the motor reversible, a four-way air valve is inserted in
the line.
(3) Multi-vaned motors, Fig. 5, are suitable for loads from fractional hp (kW)
to 10 hp (7.5 kW). They are relatively high-speed units which must be geared down
for usable speeds. The major advantages of multi-vaned motors is light weight and
small size. However, if used at slow speed, the gearing may add significantly to
the weight of the motor.
(4) Air-turbine motors deliver fractional horsepowers at exceptionally high
speeds, from 10,000 to 150,000 rpm, and are an economical source of power. They
are tiny impulse-reaction turbines in which air at 100 psi (689 kPa) impinges on
buckets for the driving force. Force-feed automatic lubrication sprays a fine film of
oil on to bearings continuously, minimizing maintenance.
Based on the load requirements, 2 hp (1.5 kW) at 1000 rpm, a reversible radial-
piston air motor, Table 1, would be a suitable choice because it delivers up to 2.8
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AIR AND GAS COMPRESSORS AND VACUUM SYSTEMS

9.6
PLANT AND FACILITIES ENGINEERING
Diaphragm converts
piston translation
to rotation
Pistons
(four or five)
Output
shaft
FIGURE 4 Axial-piston air motor available in various output sizes (Keller Tool Com-
pany).
FIGURE 5 Typical multi-vane type air motor, available in fractional hp sizes and up to
some 10 hp (7.5 kW) (Gast Manufacturing Company).
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AIR AND GAS COMPRESSORS AND VACUUM SYSTEMS
AIR AND GAS COMPRESSORS AND VACUUM SYSTEMS
9.7
TABLE 1
Specifications of Typical Air Motors
Rated hp*
(kW)
Speed at rated
hp—rpm
Free speed
rpm
Weight
lb (kg)
Stall torque

ft—lbs
Air consumption
at rated hp ft
3
free air/min
Radial piston motors (non-reversible)*
2.9 (2.2) 1,500 3,200 130 (59) .......... ..............
3.3 (2.5) 1,300 3.000 130 (59) .......... ..............
3.8 (2.8) 1,200 2,700 130 (59) .......... ..............
Radial piston motors (reversible)*
2.5 (1.7) 1,200 2,200 135 (61.3) .......... ..............
2.8 (2.1) 1,000 1,950 135 (61.3) .......... ..............
3.2 (2.4) 900 1,600 135 (61.3) .......... ..............
5.2 (3.9) 750 1,600 200 (90.8) .......... ..............
*at 90 lb / in
2
(620 kPa).
Ingersoll-Rand.
hp (2.1 kW) at 1000 rpm with air delivered to the motor at 90 lb/ in
2
(620 kPa).
The weight of this motor, Table 1, is 135 lb (61.3 kg).
2. Compare the advantages of air motors to other types of motive power
Air motors have a number of advantages over their usual competitors—electric
motors. These advantages are: (1) In explosive or gaseous environments, air motors
are lower in cost than larger, heavier, explosion-proof electric motors. Air motors
operate relatively trouble-free in moist, humid environments where the electric mo-
tor may suffer from a buildup of fungus and corrosion. And since the air motor
requires little maintenance, it can be mounted in inaccessible locations. (2) With
an air motor, the output speed can be varied from zero to free-speed no-load rotation

by merely changing the volume of air supplied to the motor. Controls are simple
in design and use. (3) Air motors can weigh as little as one-quarter that of electri-
cally-powered units; their physical dimensions are about 50 percent those of elec-
trical devices. Further, air motors do not spark; they cannot burn out from over-
loading; the air motor is not injured by stalling. Air motors start and stop positively;
they have a consistent output torque which can be changed by varying the inlet air
pressure.
Air motors do, however, have limitations. Thus: (1) Compared to electric motors,
air motors are inefficient. An air motor requires about 5 hp (3.7 kW) input to the
air compressor to produce one horsepower (0.7 5kW) at the motor outlet. (2) Air
motors are rarely practical in sizes greater than 20 hp (15 kW). Their most efficient
range is 1 /20 to 20 hp (0.04 to 15 kW). (3) The initial cost of an air motor is high;
in larger sizes, above 1 hp (0.75 kW), air motors cost up to five times that of
equivalent electric motors.
3. Check the motor duty cycle and load against the unit’s characteristics
When selecting an air motor, the first factor to be considered is the type of duty
cycle, intermittent or continuous. A crane, for which this motor will be used, does
have an intermittent duty cycle because it is not normally used continuously. There
is a rest period while the crane load is being put on the crane and again while
being off-loaded from the crane.
The great majority of air-motor applications have a low-load cycle; the air mo-
tors are used for only a few seconds continuously and have long off-duty periods.
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AIR AND GAS COMPRESSORS AND VACUUM SYSTEMS
9.8
PLANT AND FACILITIES ENGINEERING
Torque
Horsepower

Governor
controlled
curves
Rated performance
Horsepower
Torque
Speed RPM
1
Free speed
0
Stall speed
FIGURE 6 Performance curve of a typical air motor. Note how a built-in governor can change
the shape of the curve by limiting the maximum speed of the air motor (Product Engineering).
The duty cycle will usually determine the type of motor and the size of compressor
that must be used.
4. Check the horsepower and speed required
Performance curves, Fig. 6, show an air motor’s torque and horsepower (kW) output
at various rpm. Such curves can be varied somewhat by using governors or by
modifying the air intake or exhaust ports. However, the basic shape of the perform-
ance curve depends on the fundamental design of the air motor. It is common
practice to rate an air motor at its maximum output, i.e., at the top of the dome-
shaped performance curve. The reversible radial-piston motor chosen here has ad-
equate horsepower and speed for the anticipated load.
5. Determine the effect of air pressure and quantity on the air motor output
Table 2 shows how the air pressure available at the motor inlet affects both the
power output and rpm of typical air motors. For the motor being considered here,
the output would be sufficient at the lowest air pressure listed. Thus, the motor
choice is acceptable.
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AIR AND GAS COMPRESSORS AND VACUUM SYSTEMS
9.9
TABLE 2
Effect of Air Pressure on Motor Performance
Motor style
no.
Rated hp (kW)
1234 5
rpm at rated hp
1 2 345
Free speed—rpm
12345
At 60 psi lb
/in
2
1.9 2.5 3.2 6.4 0.4 1,200 1,100 850 800 900 2,650 2,250 2,030 1,900 1,600
(413 kPa) (1.4) (1.9) (2.4) (4.8) (0.29)
Gardner-Denver Company.
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AIR AND GAS COMPRESSORS AND VACUUM SYSTEMS
9.10
PLANT AND FACILITIES ENGINEERING
Related Calculations. When using tables of air-motor performance, it is im-
portant to keep in mind that the stall torque and air consumption vary for each
motor. Hence, these values are not listed in the usual performance tables. There are
so many variables in air-motor choice that stall torque and air consumption are
unique for each application and are supplied by the motor manufacturer when the

motor choice is made.
As a general rule of thumb, stall torque ranges between 2 and 2.5 times the
torque developed when operating at maximum horsepower output.
In small motors, up to 2.5 hp (1.9 kW), air consumption varies from 35 to 40
ft
3
/min (0.99 to 1.1 m
3
/min) of free air per hp (0.746 kW). Larger air motors
consume 20 to 25 ft
3
/min (0.57 to 0.70 m
3
/min) of free air per hp. These con-
sumption rates apply to non-reversible motors. Reversible air motors consume 30
to 35 percent more air.
The data, tables and illustration in this procedure are from Product Engineering
magazine.
AIR-COMPRESSOR COOLING-SYSTEM CHOICE
FOR MAXIMUM COOLANT ECONOMY
Select a suitable cooling system for a two-stage 5000-hp (3730-kW) engine-driven
air compressor, Fig. 7, installed in a known arid hard-water area when the rated
output of the compressor is 25,000 ft
3
/min (708 m
3
/min) at 100 lb/ in
2
(abs) (689
kPa). Water conservation is an important requirement for this compressor because

of the arid nature of the area in which the unit is installed. Use standard cooling-
water requirements in estimating the capacity of the cooling system.
Calculation Procedure:
1. Assess the types of cooling systems that might be used
Several types of cooling systems can be used for air compressors such as this.
Because the air compressor is used in an arid area subject to water shortages, a
recirculating system of some type is immediately indicated. Since both the engine
and air-compressor cooling water require temperature reduction in such an instal-
lation, the two requirements are usually combined in one cooling system.
The first arrangement that might be chosen, Fig. 8, combines a heat exchanger
for engine power-cylinder cooling and a cooling tower for raw-water cooling for
the compressor. Either a natural-draft cooling tower, such as that shown, or a me-
chanical-draft cooling tower might be used. The cooling-tower choice depends on
a number of factors. In an arid area, however, natural-draft towers are known to
perform well in dry climates. Further, they require much less piping and electric
wiring than mechanical-draft towers.
Another possible cooling-system arrangement uses a closed coil in the cooling
tower for both the power and air cylinders, Fig. 9. This totally closed system does
not allow contact between the compressor and engine cooling water with the at-
mosphere. This means that the compressor and engine cooling water can be treated
to reduce scale formation. Raw water recirculated through the cooling tower does
not contact the compressor coolant.
Where installation costs are critical, raw water can be used to cool the air-
compressor cylinders, Fig. 10. The engine power cylinders, which usually operate
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AIR AND GAS COMPRESSORS AND VACUUM SYSTEMS
9.11
FIGURE 7 Gas-engine driven compressor has oil-cooled power pistons. Compressor, left, uses metallic piston-rod packing (Cooper-

Bessemer Corp.).
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AIR AND GAS COMPRESSORS AND VACUUM SYSTEMS
9.12
PLANT AND FACILITIES ENGINEERING
FIGURE 8 Heat exchanger, center, for power-cylinder cooling and raw-water cooling
for the compressor (Ingersoll-Rand Co.).
FIGURE 9 Closed cooling system for power and air cylinders utilizing pipe coil in
the cooling tower (Ingersoll-Rand Co.).
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AIR AND GAS COMPRESSORS AND VACUUM SYSTEMS
AIR AND GAS COMPRESSORS AND VACUUM SYSTEMS
9.13
FIGURE 10 Raw water cools the air cylinders; power cylinders use
closed system protected by thermostatic valve (Ingersoll-Rand Co.).
110°F (43.3°C) 130°F (54.4°C)
FIGURE 11 Cooling-tower recirculating system is not recommended because of the
possibility of scale and impurities buildup (Ingersoll-Rand Co.).
at a higher temperature, are cooled by a closed system protected by a thermostatic
valve.
An open cooling-tower system, Fig. 11, is not recommended for installations
such as this because of the possible heavy scale buildup. However, such an open
cooling system might be used where the economics of the installation permit it and
scale buildup is unlikely to occur.
2. Determine the air-compressor cooling load
Use flow rates given in Table 6, page 9.24 to estimate the cooling water flow rate

for this air compressor. Thus, with the intercooler and jacket in series, using the
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AIR AND GAS COMPRESSORS AND VACUUM SYSTEMS
9.14
PLANT AND FACILITIES ENGINEERING
higher flow rate of 2.8 gal/ min (364.3 L /s) per 100 ft
3
/min (100 m
3
/s), the cooling-
water flow required for this 25,000 ft
3
/min (708 m
3
/s) air compressor is (25,000/
100)(2.8)
ϭ
700 gal/ min (44.2 L/s).
3. Compute the engine jacket-water cooling load
The engine jacket water cooling load is computed separately using a cooling-water
temperature rise of 10 to 20
Њ
F (5.6 to 11.1
Њ
C) during passage through the engine,
as given in the Internal-Combustion Engine section of this handbook. Further, the
usual jacket-water flow rate is 0.25 to 0.60 gal /(min bhp) (0.02 to 0.05 kg /kW).
For a 5000-hp (3730-kW) engine, using the maximum flow rate, (5000)(0.6)

ϭ
3000 gal/ min (189.3 L /s).
Additional cooling water may be used for the turbocharger, if fitted, and for
aftercooling. Steps for calculating these cooling-water flows are given in the section
cited above. Such cooling-water flows are usually additive to the jacket-water flow,
depending on the cooling arrangement used.
Related Calculations. Cooling systems for air and gas compressors are im-
portant for reliable and safe operation of these units. Hence, great care must be
exercised in choosing the most reliable and economic cooling system.
Today, both mechanical-draft and natural-draft cooling towers are popular
choices. An economic study is needed to determine the best choice when the cool-
ing effectiveness of both types of towers are about equal. Data given on cooling
towers elsewhere in this handbook can be helpful to the designer in choosing the
best type of tower to use for a given installation of air or gas compressors.
Straight flow-through cooling of small compressors and their drive engines is
often used where adequate water supplies are available. Thus, in large cities the
cooling water may be taken from the water main and discharged to the sewer after
passage through the compressor and engine. Cost of the water may be small com-
pared to the investment in a cooling tower. But with increased environmental con-
cerns, this scheme of cooling may soon be extinct.
ECONOMICS OF AIR-COMPRESSOR INLET
LOCATION
A plant designer has the option of locating an air-compressor inlet pipe either inside
the compressor building or outside the structure. The prevailing average indoor
temperature is 90
Њ
F (32.2
Њ
C) while the average outdoor temperature is 50
Њ

F (10
Њ
C).
Air requirements for this plant from the compressor are: 1000 ft
3
/min (28.3 m
3
/
min) of free air at 70
Њ
F (21.1
Њ
C) at 100 lb /in
2
(gage) (689 kPa) for 7500 h/yr; the
200-hp (149.1 kW) compressor drive motor operates at full load throughout the
7500 hr load year. Determine which is the best location for the compressor intake
based on power savings with an electric power cost of $0.04/kWh.
Calculation Procedure:
1. Determine the power savings possible with cooler intake air
Compute the intake volume, I, required to deliver 1000 ft
3
/min of free air at each
of the possible intake temperatures from I
ϭ
1000(density of air at 70
Њ
F, lb/ft
3
/

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AIR AND GAS COMPRESSORS AND VACUUM SYSTEMS
AIR AND GAS COMPRESSORS AND VACUUM SYSTEMS
9.15
density of air at inlet temperature, lb/ft
3
), where I
ϭ
intake volume, ft
3
(m
3
), re-
quired at the air inlet temperature.
For inlet air at 50
Њ
F (21.1
Њ
C), the outside intake air temperature, using a table
of air properties, I
ϭ
1000(0.07493/ 0.7785)
ϭ
962.49 ft
3
(27.2 m
3
). With an inlet

temperature of 90
Њ
F (32.2
Њ
C), using the inside-of-the-building air intake, I
ϭ
1000(0.07493/ 0.07219)
ϭ
1037.95 ft
3
(29.37 m
3
), say 1038 ft
3
(29.37) cu m).
The power saving from using lower-temperature intake air is then hp saving
ϭ
100(intake volume required at the higher intake temperature
Ϫ
intake volume re-
quired at the lower intake temperature)/ intake volume required at the higher intake
temperature. Substituting, hp saving
ϭ
100(1038
Ϫ
962)/ 1038
ϭ
7.32 percent.
2. Find the annual power saving with the lower intake temperature
The annual power saving, P kWh, can be found from: P

ϭ
(hp saving/100)(motor
hp)(0.746 kW /hp)(annual operating hours). Since the compressor operates at full
load 7500 h/ yr using 200 hp, the annual power saving is P
ϭ
(7.32/ 100)(200
hp)(0.746)(7500)
ϭ
81,910.8 kWh.
The annual cost saving, A
ϭ
(kWh/ h per yr saved)(power cost, $/ kWh). With
a power cost of $0.04 /kWh, the annual cost saving, A
ϭ
(81,910.8)(0.04)
ϭ
$3276.43. If an outside inlet were more expensive than an indoor inlet, this saving
could be used to offset the increased cost.
Related Calculations. As a general rule, an outside air intake, Fig. 12, is more
economical than an inside air intake when the air in the building is at a higher
temperature than the outside air. The only time an outside air intake might be less
desirable than an indoor air intake is when the outside air is polluted with corrosive
vapors, excessive dust, abrasive sand, etc., which would be injurious to people or
machines. Under these circumstances the designer might elect an indoor air intake.
However, before choosing an indoor intake, review the efficacy of outdoor air filters
of various types, Fig. 13.
Air in industrial districts may contain from 1 to 4 grains of dirt per 1000 ft
3
(28.3 m
3

). If the intake air for a compressor contains only 1 gram per 1000 ft
3
(28.3 m
3
), 7200 grains of dirt will pass into the air compressor in 1 week’s oper-
ation. With the higher level of 4 grains per 1000 ft
3
(28.3 m
3
), 28,800 grains, or
over 4 lb (1.8 kg) will be carried into the compressor during 1 week’s operation.
Frequently, much of the dirt carried in the air is abrasive. If this dirt is allowed to
get into the compressor cylinders it will mix with the lubricating oil and cause rapid
wear of piston rings, cylinder walls, valves, and other parts.
Intake-air filters, Fig. 13, can reduce much of the danger of abrasive particles
in the supply air. Each type has its favorable features. Viscous coated wire filters,
Fig. 13a, are often used for small- and medium-size compressors. Centrifugal air-
flow units, Fig. 13b, and traveling-curtain oil-bath filters, Fig. 13c, are popular for
larger air compressors. The final choice of an intake filter is a function of com-
pressor capacity, intake-air quality, annual operating hours, and expected life of the
compressor installation. Filter manufacturers can be most helpful to the plant de-
signer in evaluating these factors.
The general procedure given here is valid for air compressors of all types: cen-
trifugal, reciprocating, vane, rotary, etc., and the procedure can be used for air
compressors in plants of all types—chemical, petroleum, manufacturing, marine,
industrial, etc. This procedure has universal application because it is based on the
properties of air, the compressor power input, the annual operating hours, and the
cost of power. These values can be found in any application of air compressors in
industry today.
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AIR AND GAS COMPRESSORS AND VACUUM SYSTEMS
9.16
PLANT AND FACILITIES ENGINEERING
(a)
(b)
10-ft (3-m)
FIGURE 12 (a) Outside intake-air filter for air compressor should have intake pipe as
short as possible and be fitted with long-radius elbows (Ingersoll-Rand Co.). (b) Glazed-
tile tunnel for outdoor-air intake.
Compressed-Air and -Gas System
Components and Layouts
POWER INPUT REQUIRED BY CENTRIFUGAL
COMPRESSOR
A centrifugal compressor handling air draws in 12,000 ft
3
/min (339.6 m
3
/min) of
air at a pressure of 14 lb/in
2
(abs) (96.46 kPa) and a temperature of 60
Њ
F (15.6
Њ
C).
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AIR AND GAS COMPRESSORS AND VACUUM SYSTEMS
AIR AND GAS COMPRESSORS AND VACUUM SYSTEMS
9.17
(a)
(b) (c)
FIGURE 13 (a) Viscous coated-wire intake-air filter (Air-Maze Corp.). (b) Centrifugal
air-flow oil-bath intake-air filter also acts as a silencer. (c) Traveling-curtain oil-bath intake-
air filter cleans itself in the oil (American Air Filter Co., Inc.).
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AIR AND GAS COMPRESSORS AND VACUUM SYSTEMS