United States
Environmental Protection
Agency
Office of Air Quality
Planning and Standards
Research Triangle Park, NC
EPA 340/1-92-015e
September 1992
Revised March 1993
Stationary Source Compliance Training Series
J>EPA COURSE #345
EMISSION CAPTURE AND
GAS HANDLING SYSTEM
INSPECTION
Instructor Reference Material
Fans and Fan Systems
-------
EPA 340/1-92-015e
Revised March 1993
Course Module #345
Emission Capture And
Gas Handling System Inspection
Instructor Reference Material
Fans and Fan Systems
Prepared by:
Crowder Environmental Associates, Inc.
2905 Province Place
Piano, TX 75075
and
Entrophy Environmentalist, Inc.
PO Box 12291
Research Triangle Park, NC 27709
Contract No. 68-02-4462
Work Assignment No. 174
EPA Work Assignment Manager: Kirk Foster
EPA Project Officer: Aaron Martin
US. ENVIRONMENTAL PROTECTION AGENCY
Stationary Source Compliance Division
Office of Air Quality Planning and Standards
Washington, DC 20460
September 1992
Revised March 1993
-------
FOREWORD
This reference document is a compilation of selected published technical
papers, articles, and reports on industrial fans and fan systems selection,
design, and operation. These technical papers are suggested as a
source of background information and possible reference material for
persons serving as instructors for the inspector training course #345 -
Emission Capture And Gas Handling System Inspection. The document
is not being distributed as an EPA publication and is not available to the
public.
-------
CONTENTS
ITEM 1 - Fans - Special Report, John Reason, Power Journal,
September 1983
ITEM 2 - Chapter 7 - Fans, Handbook Of Ventilation Control. Henry
McDermott, Ann Arbor Science, 1977
ITEM 3 - Fans, Power Special Report, Robert Aberbach, Power
Journal, March 1968
ITEM 4 - Selecting Fans And Blowers, Robert Pollak, Chemical
Engineering Journal, January 22, 1973
ITEM 5 - Fans And Blowers, Paul Cheremisinoff, Pollution Engineering
Magazine, July 1974
ITEM 6 - Fans And Fan Systems, John Thompson, Chemical
Engineering Journal, March 21, 1983
ITEM 7 - Basic Courses In Fan Selection, Fan Density Correction, And
Fan Arrangements And Classes, Chicago Blower
Corporation, 1981
-------
ITEM 1
Special Report - Fans
John Reason
Power Journal
-------
Special Report
By John Reason, Associate Editor
Nir.e'.\-nine percent of ill air-mov-
ing applications are bandied b>
three basic fan types —centrifugal.
propeller, and axial. Each of these types
has different cap-abilities in terms of the
pressure generated, tie volume of gas
handled, the degree of control possible.
and resistance to wear and corrosion. In
se\erai cases, characteristics overlap and
a choice must be made between two
basic fan types for the same applica-
tion.
Centrifugal fans operate by forcing
the air to rotate in the fan housing. The
resulting centrifugal force acting on the
rotating air mass develops the pressure
to move the air stream.
Propeller fans move air. without
developing significant pressure, simply
by the angle of attack of the propeller
blades. The housing, if any. plays little or
no part in controlling the air
flow and flow control is
minimal.
Principal benefit is extreme simplicity.
Axial fans are essentially propeller
fans enclosed in a housing, which pro-
vides a degree of control that, in some
IN THIS REPORT
Where fans are used
S«2
Fan theory
Shapes and characteristics S • 6
Adjustable-pitch fans.
S -10
Fan systems.
S «12
Inlet/outlet essentials,
S '14
Economics of fan drives _ S • 16
Repair and rebuilding
S- 18
Noise and vibration
S-20
cases. ;s better than that of a centrifugal
fan. Air now can be straight through,
provided the fan drive can be mounted in
the gas stream, though frequently the
duct configuration forces the flow to go
through a right angle before entering the
fan as with the centrifugal.
Van&-axiai fans have guide vanes
before and, or after the wheel to stream-
line the air flow better for fan action.
Axial fans without guide vanes are
know-n as tube-axial.
Rising energy costs and pollution-
control regulations have had a significant
effect on the choice and flesign bi tarts
over the last 10 years. Fans must be
more efficient, and they must also have
the" capacity to -move air and flue gas
through the pollution-control equip-
ment— electrostatic precipitators. fabric
niters, scrubbers, heat exchangers, incin-
erators, etc. In many cases, this need
has mandated more fans as
well as more horse-
power to drive
them.
^Reprinted horn Power, September 1983
McGraw-Hill. Inc.. 1983. All right
-------
Where fans control energy
Material transport fan is some-
times arranged to have trans-
ported material pass through the
fan wheel. Single-stage paddle-
wheel axial fan is used with
replaceable blades of wear-resist-
ant material.
j5&) Blower used here to transport
^O^ shredded garbage into boiler con-
^^ sists of three to five stages of
^•884 centrifugal-fan wheels running on
vKf a single shaft, usually at 3600
rpm. Air flows through each stage
in series. Different wheel diame-
ters develop pressures up to
about 15 psig. Compressor wheels
have axial or backward-curved
blades. Transported material nev-
er flows through the blower, at
least by intention.
Roof ventilators are usually the
simplest form of propeller fan,
directly driven or through a belt by
a totally enclosed motor, with no
air-flow control. Individual ventila-
tors are merely switched off when
less ventilation is needed. If more
pressure is needed, propeller is
enclosed in tube to become a
tube-axial fan.
Flyash reinfection fan Is used on
stoker-fired boilers, especially
when firing wood chips, to ensure
complete combustion of carryover
char. Duty is extremely erosive at
high temperatures. Fans used are
centrifugal with open hard-faced
radial blades to resist erosion and
to make replacement easy.
Overfire-air fan used with stoker-
fired boilers, blows cool, clean,
ambient air into furnace to provide
turbulence and complete combus-
tion of gases from grate. Airfoil or
backward-curved centrifugal fan
or fixed-pitch axial fan may be
used. Control is with inlet vanes or
speed control.
Shredded
.waste fuel
'A
0
*•»»
o
Hot-primary-air fan, occasionally
used to handle high-moisture
coals, takes hot air from the pri-
mary-air heater and blows it
through the pulverizer to the burn-
ers. This air contains paniculate
matter picked up from the air-
heater elements. Fan erosion bur-
den is comparable to that of a
gas-recirculation fan. Centrifugal
fan with straight-radial blades is
normally needed.
Cold-primary-air fan blows am-
bient air through the pulverizers, a
preheater, and the burners. High
pressure at relatively low volume is
needed (low specific speed). Fan
has a narrow large-diameter cen-
trifugal wheel with backward-
curved or airfoil blades, or is of
the two-stage axial type. Air han-
dled is clean, so no special con-
struction is needed; flow rate is
controlled by the number of pul-
verizer dampers open. Axial fans,
if used, must be properly applied
to avoid" operation in stall region.
An inlet silencer is required on this
type of fan.
Office or control room
Air-conditioning and ventilation
fans must develop sufficient pres-
sure to drive air through cooling
coils and ductwork into office
space. Centrifugal-fan wheels with
forward-curved squirrel-cage con-
struction are used on small units
with on/off control. On large units,
backward-curved blades with inlet-
vane control are used.
Cooling coils
/—IT-'
Rue-gas recirculation fan must
operate in an extremely high-tem-
perature, erosive atmosphere.
Open-radial, straight-radial, or ra-
dial-tip centrifugal fans are used,
often of high-alloy construction.
Recirculated gas introduced into
the hopper area is known as "gas
recirculation" and reduces grate
temperature and furnace absorp-
tion. Recirculated gas introduced
near furnace exit is known as "gas
tempering-." Flue-gas recirculation
redistributes heat absorption
throughout boiler, but has no
effect on total absorption. It may
be used to reduce grate tempera-
ture, cut oxides of nitrogen, or
increase superheater temperature
during low-load operation.
Inlet
silencer
Forced-draft (FD) fan develops
medium pressure to blow clean
combustion air into the boiler (un-
derfire air in the case of a stoker-
fired boiler, secondary air for pul-
verized-coal-, oil-, and gas-fired
boilers). Fan may be centrifugal or
axial. Centrifugal fan is usually
double-inlet type with backward-
curved or airfoil blades, with inlet-
vane or speed control. Axial fan is
normally single-stage with adjust-
able-pitch blades or inlet-vane
control. On balanced-draft boilers,
fan must develop sufficient pres-
sure to equal resistance of air
ducts, air heater, and burner, or
fuel bed, making the furnace the
point of zero pressure. Air may be
prewarmed before entering fan.
but neither temperature nor par-
ticulate loading requires special
fan construction.
-------
flow in utility and industrial plants
Induced-draft (ID) fan handles a
large volume of hot flue gas and
must develop enough pressure to
drive the combustion products
through pollution-control equip-
ment and air heaters while main-
taining zero pressure in furnace.
This fan is the largest power con-
sumer in a modern powerplant. If
paniculate-control equipment is
located upstream of the ID fan,
erosion is not a severe problem,
but fan must handle some flyash
let through by failed or inefficient
collection equipment. Before par-
ticulate-control equipment was
mandated, most ID fans were cen-
trifugal, with radial or radial-tip
blades. Today, with particulate-
removal efficiencies typically in
excess of 98%. all types of centrif-
ugal and axial fans are seen on
induced-draft service. Corrosion is
seldom a problem except on boil-
ers firing black liquor or municipal
garbage. ID fans are usually fitted
with an exhaust silencer of the
resonant type.
Indirect-reheat fan adds heat to
stack gases for better dispersion
by blowing ambient air through an
indirect-reheat coil into the stack.
This arrangement eliminates the
need to place heater coils in the
wet, cool flue gas with an attend-
ant corrosion and pressure loss.
Reheat fan blows ambient air and
needs no special construction.
Fabnc filter or electrostatic precipitator
Resonant outlet
silencer
—(/////////A—I
Wet scrubber booster fan is nor- g
mally the only fan in a steam- f
generation system that must resist
the action o< corrosive gases 6
because it is located in the wet ^
flue gas downstream of the scrub- .
ber. Centrifugal fan wheel must be f.
used, usually of open radial con-
struction for ease of blade repair
and replacement, though single-
thickness, backward-curved
blades are used. Construction is of
stainless steel or special alloys.
making this a high-cost fan. Simi-
lar duty ID fans are sometimes
needed in the paper industry on
recovery boilers burning black
liquor or on condensation heat-
recovery systems.
^
* — 1
Bypass \_
X
/
To stack
Flyash-transport blower con-
sists of multiple stages of centrifu-
gaMan wheels on a single shaft.
Flyash is usually transported by
suction. Flyash-laden air passes
through a centrifugal separator to
drop out flyash and then through a
bag separator to ensure that no
flyash passes through the blower.
Backward-curved
fj Backward-inclined
Squirrel-cage
Dry scrubber booster fan is
located upstream of the scrub-
bers. This is the preferable loca-
tion for a retrofit fan installed to
provide added pressure to move
air through the flue-gas-^esulfurt-
zation system. Rue gas handled is
hot, well above dewpoint, so cor-
rosion is not a problem. Note that
part of flue gas may bypass FGD
system when low-sulfur fuel is
burned, so scrubber booster fan
doesn't handle full flue-gas vol-
ume.
Hollow airfoil
Open-radial
• 2 Radial-tip
Propeller
Radial
Axial
Induced-draft cooling-tower
fans are large-diameter propeller
fans driven by gear drive or belt.
Because of the large size of these
fans and their outdoor location,
careful attention is paid to blade
shape and inlet and outlet condi-
tions to hold power consumption
and noise to a minimum. Speed
control or adjustable-pitch blades
may be used. Fans handle wet
ambient air with no corrosion or
erosion problems. Blade materials
may be extruded aluminum or
glass-reinforced plastic.
Forced-draft cooling-tower fans
are usually axial type, but centrifu-
gal-fan wheels may be used on
some towers, especially small,
packaged units where space sav-
ing is paramount. Centrifugal fans
have forward-curved squirrel-cage
wheels. Axial-fan blades and
housing are often of fiberglass.
Mechanical-draft cooling tower
-------
Duct walls
Pilot tube
Gas tlow
Velocity pressure = total - static pressure
Total pressure
1. Pitot tube measures velocity pressure as impact of moving air
on tip. It can be traversed across duct to find velocity profile
Area = 3 sq ft
Velocity = 2333 ft/sec
VP = 0.3 in. H,O
SP - 4.6 in. H20
Static pressure
Area - 1 sq ft
Velocity = 7000 ft /sec
VP = 3 m. H20 Velocity-to-static-pressure
SP - ^ in. H2u rega|n effjc|ency. 96o/o
2. Static-pressure regain occurs whenever air stream slows
down. Velocity pressure is converted to static pressure
The pressure developed by a fan is mea-
sured not in pounds per square inch but
in inches of water. For instance, a typical
forced-draft fan feeding secondary air to
a pulverized-coal boiler develops about
30 in. H2O, which is equal to about 1.0
psia. This is measured as the difference
in pressure between the fan's inlet and
outlet.
But it is important to understand the
difference between static pressure and
velocity pressure. When the fan's outlet
is completely closed off by dampers so
that there is no flow, all of the pressure
developed is static pressure and can be
measured with a manometer or sensitive
pressure gage. When the dampers are
opened and gas flows, both static and
velocity pressure are developed. Velocity
pressure is really a measure of the kinet-
ic energy in the moving gas stream. It is
measured by a pilot tube pointing
upstream and is actually recorded as the
difference between the static pressure
and the pressure created by the gas
stream impacting the end of the pilot
tube (Fig 1).
The sum of static pressure and veloci-
ty pressure is known as total pressure.
Stalic pressure can be converted to
velocity pressure and vice versa, and this
is an important part of fan-system
design. For instance, if a certain volume
of gas is moving along a duct and comes
to a point at which the duct cross section
increases (Fig 2), the velocity of the
moving gas decreases and the drop in
velocity pressure is accompanied by a
corresponding rise in. static pressure.
This is known as static-pressure regain.
(Note that the conversion of velocity to
static pressure is always less than 100%
efficient because of turbulence at the
duct transition.)
s • 4
All you need to know
Static-pressure regain is routinely
used at the outlet of a fan. As the gas
leaves the trailing edge of the fan blades,
virtually all its energy is in the form of
velocity pressure. This is converted to
static pressure as it slows down in the
outlet duct.
Fan output characteristics are often
published in terms of static rather than
total pressure, although total pressure
would appear to be more useful. How-
ever, the published static-pressure levels
are measured under standard test condi-
tions defined by the Air Movement Con-
trol Assn (AMCA). These standards are
virtually universal and provide a stan-
dard comparison for fan performance.
At the same time, it must be realized
that the fans are tested under ideal con-
ditions, which can never be attained in
an aclual fan system.
Note that most fan characteristics
show the fan's output pressure (static or
total) rising to a maximum point at some
level of output flow roughly correspond-
ing to peak efficiency. The essence of
good fan-system design is to select a fan
that operates at that point for most of its
duty cycle. The fan static pressure dur-
ing operalion equals ihe sum of pressure
drops through the fan system—duct
bends, coil bundles, heaters, eic. This
total pressure can be calculated approxi-
mately from AMCA data, but if an
accurate estimale is needed, • it may be
necessary to build a tesl model.
Output volume vs pressure
In addition to outpul pressure, a key
consideration when selecting a fan is the
volume of gas that it must handle. For
instance a propeller fan for ventilation
handles a large volume of air, but only
needs to develop a fraction of an inch of
water gage pressure. Conversely, a pri-
mary-air fan must develop enough pres-
sure to blow the pulverized coal through
the burners (typically about 50 in. H2O),
but it handles only a relatively small
volume of air. An ID fan must handle a
much greater volume of air than an FD
fan on the same boiler, because il musi
blow the hot expanded flue gas.
Specific speed is a term used to
describe a fan's pressure/volume-han-
dling characteristics. It is defined as the
speed of a hypothetical fan of the same
shape and design, but of unknown size,
that delivers 1.0 cfm at 1.0 in. H2O.
Specific speed of a fan is different at
each operating point, but the specific
speed at the peak-efficiency point is
unique to each fan design. Thus, it is a
very useful figure for comparing the per-
formance of various fan designs. Fig 3 is
a plot of efficiency vs specific speed for a
number of different fan types. Note that
fans of high specific speed, such as those
with backward-inclined airfoil fan, have
better efficiency than fans of low specific
speed, such as a high-pressure radial-tip
blower.
How efficiency is defined
Efficiency, which 30 years ago was
hardly a consideration in fan design, is
now of vital importance. The cost of
fan-drive energy can easily run to
$250/hp-yr, and the present value of the
energy cost over the life of the fan must
be balanced against the capital cost of a
more efficient fan.
There are two ways of specifying fan
efficiency. Total (or mechanical) effi-
ciency takes into account the total ener-
gy in the gas stream (static plus velocity
pressure) as a percent of energy input to
the fan. Static efficiency takes into
FANS • a special report
-------
80
7°
6°
50
Xf
w\\
3 4 5 6 8 10 15 20
Specif c speed, based on static pressure, •
30 40
1000 rpm
Typ* Width Typical duty
Radial blades
Narrow Pressure blower
Narrow Pressure blower
Narrow Pressure blower
Medium Pressure blower
Medium Pressure blower
Medium Pressure blower
Wide Industrial exhauster
Typ« Width Typical duty
Airfoil blade*
H Narrow Mechanical draft
I Medium Mechanical draft
J Wide Ventilation
Backward curved
K Narrow Mechanical draft
L Medium Mechanical Graft
M Wide Ventilation
Forward curved— —
N Medium Mechanical draft
•s
j 80
| 70
i 60
03
o 50
2
Type
Vane-
CD
p
0
R
S
Prope
T
n/*
fi
•^s
0 30 40 5C
Specific speed, b
Hub- to- Typical
tip ratio duty
High
High
Medium
Medium
Low
Low
Ventilatio
Vent latio
Ventilatio
Ventilatio
Ventilatio
N
60
ased c
*
T
T
1
1
Vent ation
\o±-^—^^
x, K X
p" V' ^^.
\
\ *
80 100 150 200
n total pressure, + 1
Mechanical efficiency *
_S
"N
300
000 rpm
flow X total pressure
BMP X 63 56
x static pressure
-.„.., . , BHp ^ 63 56
3. Each fan shape has a unique value of specific speed at peak efficiency. High-specific-speed fans are more efficient
about fan theory
account only the static-pressure output
of the fan (see formulas. Fig 3). Static
efficiency is generally used by manufac-
turers to specify the performance of their
fans, since total efficiency depends on
the total fan system.
Fan efficiency at any operating point
can be estimated from the performance
curve. Static pressure can be read from
the curve, and velocity pressure can be
found as a function of air flow. Note that
performance curves are determined for
the fan handling air at standard temper-
ature and pressure and at a density of
0.075 lb/ft3. These figures must be cor-
rected for the actual density of the gas,
using the fan laws (see box at right).
Streamlined flow is the key
Good fan efficiency depends on
streamlined flow of the gas stream over
Separation at blade
trailing edge
the fan blades. Any turbulence in the
stream or separation of smooth flow
from the blades increases friction losses
and distorts the even distribution of flow
between the biadtS; generating shock
losses (Fig 4).
Fig 5 shows the component vectors of
gas flow at the leading and trailing edges
of a centrifugal-fan blade. Clearly, in the
ideal fan design, Vectors R, and R:
should be parallel to the fan-blade sur-
face at both leading and trailing edges of
blade.
The static pressure developed by the
fan is given by the equation shown in Fig
5. The first term in this equation depends
solely on fan design, but both of the
other terms depend on streamlined orien-
tation to and from the fan blades. For
instance, if the entry vector A, points in
the direction of blade rotation, the pres-
Separation at
blade leading edge
4. Efficient fan operation depends on the
streamlined flow of air through blade
passages without separation, turbulence
5. Basic equation for fan-output pressure
(right) shows importance of correct blade
shape in relation to direction of air flow
Vector T—Peripheral velocity of tan blade
Vector R—Relative velocity of gas to blade
Vector A—Absolute velocity of gas stream
Subscnpt 1—Inlet conditions
Subscript 2—Outlet conditions
p—Gas density
P—Pressure head
Five simple fan laws
guide design., application
When working with a fan wheel of
fixed size and shape, the following
relationships apply:
Volume (cfm) is proportional to
fan speed (rpm).
Static pressure (in. H2O) is pro-
portional to speed squared.
Horsepower is proportional to
speed cubed.
Horsepower is proportional to
gas density (lb/ft3).
Static pressure. developed is
proportional to gas density.
sure developed by the fan is lessened.
If the relative velocity of the gas
stream departs significantly from the
blade angle at either leading or trailing
edges, friction and shock losses increase.
Fan laws aid design, application
Compressible gases flow in extremely
complex patterns and it is almost impos-
sible to predict the performance of a fan
theoretically from its design and shape.
But fortunately, fan design and applica-
tion is simplified by a series of very
straightforward fan laws that predict
how the parameters of a specific fan
design are related to each other. Thus, if
a mode! fan wheel is built and tested in
the laboratory, the fan laws can be used
to predict precisely the characteristics of
a scaled-up fan, built with the same
shape as the model. These same laws can
be used to predict the performance of a
complete system from a scale model.
For instance, as the fan's wheel diame-
ter is increased, the pressure increases as
the square of the diameter; as the fan's
speed is increased, the volume delivered
varies in proportion. The basic fan laws
are defined in the box above.
Power. September 1983
S • 5
-------
Shape and pitch of
Most fans consist of a rigid wheel
designed to rotate at a constant speed in
a fixed housing. The output characteris-
tic of each fan is determined by width,
depth, curvature, and pitch of the blades;
by the speed and diameter of the wheel;
and, to some extent, by the housing. The
different characteristic curves determine
each fan's suitability for a particular
application in industry. Different wheel
geometries also affect the fan's ability to
resist erosion and the buildup of deposits
on the blades.
The major wheel shapes used in indus-
trial fans and their corresponding char-
acteristics are shown in this section.
Note, however, that there is often a
subtle variation between characteristics
of similar fans produced by different
manufacturers and, sometimes, different
terms are used to describe virtually the
same blade shape. Only the fan's charac-
teristic curve and the manner in which
the fan reacts with the complete system
are of ultimate importance.
Typically, manufacturers offer fans
with a range of characteristic curves.
These are presented as plots of static
pressure (in. H:O) against output in
cubic feet per minute (cfm). As a fan's
output changes over the range of the
curve, the horsepower needed to drive it
changes, as does the fan's efficiency.
These variables may also be plotted on
the curve.
How centrifugal fans vary
In general, a wide centrifugal-fan
wheel of small diameter produces a large
volume of gas at low static pressure
(high specific speed). A narrow, large-
diameter wheel drives a small volume of
gas at relatively high static pressure (low
specific speed). The simplest type of fan
blade is fiat radial and, while this is used
in some applications, its shape is usually
modified to improve efficiency. Single-
thickness blades may be backward
inclined; backward curved; or backward
inclined and forward curved to produce a
radial tip. Even better efficiency can be
obtained with a backward-inclined air-
foil-shaped blade, which promotes
smooth, nonturbulent gas flow. Finally,
forward-curved blades are used in some
clean-air-handling applications.
Backward-inclined blades are used
to handle noncontaminated air, as in a
boiler forced-draft service. Over the
operating range, the static pressure falls
off as the gas flow increases (caused by
opening of control dampers or decrease
in system resistance). However, at flow
rates below the designed operating
range, the air flow may break away from
the blade surface, causing a region of
instability. This is shown by the dip in
the curve to the left of the peak pressure
(below). The fan should not be operated
in this region because of the low efficien-
cy and possible pulsations in the output
pressure.
Limiting power
ievef
/ I Unstable
/ \ region
cfm.
Observe that the horsepower drawn by
this fan increases to a maximum at the
peak pressure and then falls off. This
gives the backward-inclined fan a useful
nonoverloading characteristic —if the
drive motor is sized for the maximum
horsepower, it can never be overloaded
by unexpected changes in system resist-
ance. This is particularly valuable if the
motor must be sized and ordered for the
fan before the system, resistance is known
accurately.
Backward-curved, single-thickness
blades (sometimes known as single-
thickness airfoil blades) are used more
frequently than flat backward-inclined
blades because they promote smoother
6. Single-inlet centrifugal, induced-draft fan for utility application, has radial-tip blades
protected by wearplates bolted in position. Speed is 710 rpm at 6642 hp
cfm
s . e
FANS . a special report
-------
blades determine characteristics
air flow and have somewhat stronger
construction. Characteristic curve is sim-
ilar to the backward-inclined blade, but
the instability region may be less pro-
nounced so that the fan can be operated
over-the full range of air flow from wide
open to shutoff.
A fan with backward-inclined blades
has a high specific speed for its size and
overall shape. This means a high speed is
needed to generate the required pressure
which, in turn, means the wheel has low
resistance to erosion from particles in the
air stream. However, in the many situa-
tions where particulates are not a prob-
lem, this wheel has wide application.
Static efficiency is relatively high —
between 77% and 80%.
Hollow airfoil blades are now used in
many applications to increase the fan
efficiency. These blades act like an air-
plane wing and promote extremely
smooth flow and efficiencies as high as
91%. Needless to say. airfoil blades are
costly to construct. They are also more
expensive to repair or rebuild, and so
they are used only with clean, nonabra-
7. Hollow airfoil fan blades must be built
up from steel sheet. Construction is costly
as is the addition of wear plates, but
efficiency more than compensates
Static Pr
Horse,-
cfm
sive
full
can
gases. Operation is stable over the
range, noise level is low, and the fan
be operated at high speed.
Radial blades make no attempt to
produce smooth air flow, with the result
that particles in the gas stream are
deflected away from the blade surface
and give this type of fan maximum
resistance to abrasion. Also, because the
flat, radially mounted blades project the
gas stream straight out from the fan hub,
the fan has a low specific speed. This
means it runs slower than the backward-
inclined fan to generate the same pres-
sure, which also improves its resistance
to abrasion.
When energy costs were low, the radi-
al-blade fan was ideal for induced-draft
cfm
service on a coal-fired boiler where the
particulate loading was very high. But
the efficiency of the radial-blade fan is
low (70-72%). This fact, together with
the stringent particulate-control require-
ments imposed on today's boilers, has
virtually removed the radial-blade fan
from ID service.
Straight radial blades are still used in
several applications where heavy particu-
late loading is the governing factor. In
the boiler room, these applications
include flue-gas recirculation and hot
primary air.
Open radial blades are specified for
the extremely abrasive service needed in
flyash reinjection. These blades are vir-
tually like paddles with no centerplate
and often no side plates. Radial-blade
fans may also be used for material trans-
port—for example, when shredded re-
fuse is injected into the boiler for cofir-
ing. Blade wear is extremely heavy, but
blades can be very easily replaced or
lined. Efficiency is as low as 65%.
cfm
s • 7
-------
Radial-tip blades (sometimes known
as backward inclined, forward curved)
are today's answer to the deficiencies of
straight radial blades. Radial-tip blades
present a low angle of attack at the inner
or leading edge, which allows the gas
stream to follow the blade shape with a
minimum of turbulence. The trailing or
outer edge of the blade is curved up
almost to an axial direction, giving the
fan a low specific speed and hence, a
good resistance to abrasion. Operation is
stable throughout its range and efficien-
cy is high (78-83%).
cfm
This fan has many of the characteris-
tics of other backward-inclined types,
including high efficiency and self-limit-
ing horsepower. It is ideal for handling
gas streams with moderate dust loading,
especially induced-draft applications
where high specific pressure and low
specific speed are needed.
rotor, produce a fan of low-to-medium
efficiency, but high-volume handling
capacity for its size. Because of the for-
ward curve of the blades, they can very
easily accumulate deposits from dirty
gas streams and so can only be used for
handling clean air. The characteristic
curve has a significant unstable region
operating close to its point of maximum
efficiency (Fig 10). Today, when the
equivalent capital cost of one horsepower
may run to thousands of dollars, there is
intense interest in the development of
reliable, variable-speed drives for fans
(see fan drives, p S-16), but the fact that
most boiler fans today are still controlled
by dampers attests to the fact that vari-
able-speed drives are still far from fully
accepted.
Forward-curved blades, mounted
on what is often called a squirrel-cage
cfm
where the fan cannot be operated. For-
ward-curved blades have little use in
today's boiler house, but are used exten-
sively in small-size air-handling equip-
ment.
Output control is a problem
Centrifugal fans are essentially con-
stant-output devices and do not lend
themselves well to applications needing
variable gas flow. In most applications
where variable output is needed, the fan
wheel is run at full speed at all times,
and the flow is throttled by inlet vanes or
outlet dampers. Clearly, this approach
wastes energy, especially if the fan is
required to operate at less than design
output for long periods.
Outlet dampers are the most energy-
wasteful, since their effect on the fan is
merely to push its operating point back
down the curve away from the design
operating point. Power input remains
high, while efficiency drops off drastical-
ly. A. typical situation where outlet
dampers are the only practical means of
control occurs when a primary-air fan is
feeding air to a number of pulverizers.
As load increases and more pulverizers
are brought into operation by opening
their dampers, fan output is increased.
Inlet vanes, properly installed, are
somewhat more energy-efficient because
they can alter the fan's characteristic at
low output levels. If the damper blades
are oriented so that, when partially open,
they impart a pre-swirlto the air stream
in the direction of rotation of the fan,
horsepower needs of the fan are reduced
at the same time that output flow is
reduced (Fig 9).
Speed control of the fan wheel is the
ideal method of output control^ If the
speed of the fan can be varied, the fan's
output characteristic can be continuously
varied with the flow so that it is always
Propeller fans are essentially low-
pressure, high-volume devices used
mainly for driving clean ventilation air.
Most propeller fans are simple, low-cost
devices, but for cooling-tower applica-
tions, their size and power requirements
become very high and precise engineer-
ing is vital.
Axial fans are essentially propeller-
type fans enclosed in a housing or tube.
In the tube, some of the velocity pressure
generated by the impact of the blades on
the air stream is converted to static
8 Adjustable-pitch axial fan provides
good flow control at high efficiency
FANS • a special repon
-------
Net (100%)
operating load \
\ \
\
\
50 75
Flow, % test block
100
9. Fan efficiency drops off quickly as output is reduced by inlet-vane control at
constant speed. Proper orientation of vanes keeps efficiency loss to a minimum
50 75
Row, % test block
100
10. Variable-speed operation alters fan's characteristic curve to cross system-
resistance curve at optimum point. Higher efficiency is maintained at all levels
pressure. If straightening blades are
installed in the tube, the swirl velocity
pressure is also converted to static pres-
sure and efficiency is improved further.
cfm
Pressure developed by an axial fan
follows the same fan laws as apply to the
centrifugal wheel, but it also depends on
the ratio of hub-to-tip diameter. (The
higher the hub-to-tip-diameter ratio, the
higher the pressure.) On a simple propel-
ler fan, the parts of the blades close to
the hub are crowded together and move
at a lower velocity. Result is that gas
tends to recirculate through the center of
the fan. In fact, both static pressure and
efficiency of such a fan can be increased
by installing a hub cover over the fan
wheel. Axial fans for powerplant appli-
cations are designed with -high hub-to-
tip-diameter ratios. '
The significant feature of the axial-fan
11. Single-inlet radial-tip wheel has wear
plate on faces of blades
characteristic is the deep stall area to the
left of the peak pressure point. This is
caused by the fan blade stalling in much
the same way as an airplane wing stalls.
If the fan is operated in this region
(because of an accidental blockage in the
flow, for instance) it continues to pump
energy into the gas without developing
significant flow. The fan housing can
overheat rapidly under such conditions.
This stall region, together with the
relatively low pressure developed by a
single-stage axial fan (about 20 in. H2O
maximum), have largely kept axial fans
out of the power field until the 1970s. In
addition, the machined or die-cast blades
are more susceptible to wear from panic-
ulate in the gas stream and more diffi-
cult and costly to repair than centrifu-
gal-fan blades.
A typical problem with the stall region
would occur with two forced-draft fans
operating in parallel. If one fan were
operated first at low furnace load, it
would be impossible to bring the second
fan up to speed, through its stall region,
without reducing load on the first fan.
Variable-pitch blades, together with
the high cost of energy and the decreased
amount of particulates in flue-gas
streams, have increased interest in axial
fans for power applications (see next
page). Although centrifugal fans remain
the workhorse of utility and industrial
plants, variable-pitch axial fans are now
an important possibility for new power-
plants.
Specify static or total
It is important to note that fan-output
characteristics are universally specified
in terms of static pressure and static
efficiency (see page S-4 for definition).
.Clearly, when applying a fan, the engi-
neer is more interested in total pressure
and mechanical efficiency. Both these
values depend on the system to which the
fan is connected and cannot, therefore,
be specified for the fan per se.
Power. Saptambar 1983
S • B
-------
Adjustable-pitch axials
Impeller casing Straightening vanes
Diffuser Blade
Inlet box
Couplings
Main dnve motor
Control drive
Hub
Hub
Main bearing assembly
Floating shaft
12. Large two-stage axial fan allows complete access to variable-pitch control
mechanism without dismantling. Blades are removable through casing openings
0.09 0.07 0.05
Density, Ib/ft^
200 400 600
800 1000 t200 1400 1600 1800 2000
Flow, cfm + 1000
13. Axial-fan characteristics are chosen so that test-block condition is in area of
highest efficiency. Actual output can be raised or lowered from test-block level
0.09 0.07 0.05 200 400 600 800 1000 1200 1400 1600 1800 2000
Density, Ib/ft3 Flow, cfm •+• 1000
14. Power needed to drive fan depends on flow, blade angle, and gas density.
Nomograph to left of curves allows horsepower to be figured at any density
Efficiency at different loads is the hall-
mark of adjustable-pitch axial fans. Tra-
ditionally, axial fans have been consid-
ered to be less erosion-resistant, more
costly, and to develop less pressure than
centrifugal fans of equivalent size and
power. But with the increasing costs of
energy and the development of reliable
adjustable-pitch blades, axial fans have
become a serious competitor for the cen-
trifugal fan in many powerplant applica-
tions. Several studies have shown that,
over a range of loads, the adjustable-
pitch axial fan has lower lifetime costs
than the centrifugal fan with adjustable-
speed drive.
The simplest type of adjustable-pitch
axial fan is one in which the pitch of the
blade can be altered only when the fan is
stationary. The value of this feature is
that the fan characteristics can be
adjusted after the fan system is built, to
compensate for errors or changes in the
estimate of fan-system resistance.
Note, however, that an axial fan is
designed to operate at maximum effi-
ciency when the blades are at the design
pitch (sometimes indicated as 0 deg on
the performance curve). After the sys-
tem is built and tested, blade pitch can
be adjusted positively or negatively to
obtain the design performance. Manu-
facturer-supplied characteristic curves
show bands of efficiency, indicating the
loss of efficiency incurred by deviating
from the design pitch.
Flow-rate variations during operation
with this type of fan are made with
conventional inlet-vane controls. Adjust-
able-speed operation is rarely used with
an axial fan. However, belt-driven fans
lend themselves fairly easily to speed
adjustment by changing belt and pul-
lies.
The adjustable-pitch axial power-
plant fan consists of a large-diameter
hub mounted on a bearing assembly. On
the periphery of the hub are a number of
blade shafts on which the fan blades are
mounted. The inner ends of the blade
shaft carry a short crank and guide shoe,
which is located between two regulating
rings. A hydraulic cylinder, mounted
axially in the bearing assembly, moves
the guide rings back and forth to alter
the blade pitch.
Because the bearing assembly does not
rotate, hydraulic supply to operate the
variable-pitch mechanism can be a con-
ventional hydraulic tank mounted out-
side the fan housing. Pitch can be con-
trolled by a standard pneumatic signal,
which controls the hydraulic positioning
of the guide rings. Blade pitch varies
continuously during boiler operation to
FANS • a apacwJ report
-------
find increasing applications
accommodate not only changes in boiler
load, but also upset in boiler operation
that happen too quickly for adequate
response by conventional damper con-
trol.
The inherent low pressure developed
by an axial fan is handled for boiler-
draft applications by using two stages
with guide vanes between the stages and
following the second stage. Fig 12 shows
the arrangement for a typical two-stage
mechanical-draft fan. Primary-air and
ID axial boiler fans normally have two
stages. A single stage is usually adequate
for FD fans.
Fig 13 shows a series of axial-fan
characteristic curves developed for a
two-stage axial induced-draft fan on a
660-MW coal-fired unit. Test-block con-
dition is 1.220.000 cfm at 26 in. H,O
When the customer ordered this fan, he
selected three representative operating
points. Given these operating points, the
manufacturer selected a fan from 1500
possible combinations of speed, blade
length, hub diameter, and number of
blades so that each of the operating
points fell in the highest efficiency zone
possible. Another important considera-
tion was that all operating points were
kept wpell away from the fan's stall line.
Test-block condition was selected at
the point where the customer felt the
boiler would most frequently operate.
Fan coordinates were then selected to
place the test-block condition in the
region of peak fan efficiency. Note that
this still allows the other two design
operating points to fall in regions of
relatively high efficiency.
Unlike conventional fan performance
curves, which are derived under standard
conditions at standard gas density, the
curves in Figs 13 and 14 are design
curves used to select a fan for a real
application, where gas density changes
with temperature. Both horsepower and
output pressure vary with gas density
and, to simplify the selection process, the
manufacturer supplies a nomograph to
the left of the curves so that pressure
and/or horsepower at any operating
point can be read directly.
In this installation, the customer
elected to invest in the extra cost of a
two-speed motor to increase flexibility
still further. The fan may be used at one
speed or the other, depending on the
operating load or changes in boiler char-
acteristics. At each design operating
point, the same output can be achieved
at either speed by altering the blade
pitch angle. Using the curves, it's a sim-
ple matter to find which speed and blade
angle gives the highest efficiency.
15. Axial-fan blades can be removed for repair, replacement and/or balancing by plant
personnel or, as shown here, complete fan may be overhauled by manufacturer
16. Two-stage, adjustable-pitch axial fan has intermediate casing section removed for
inspection. Note large hub size and straightening vanes in outer annulus of housing
Power. September 1983
-------
System design is vital
Fan-wheel geometry determines the
shape of the fan's characteristic curve,
but that curve may be quite meaningless
if the fan is incorrectly sized for the
complete system or if the system is poor-
ly designed. Published fan characteris-
tics provided by the manufacturer to
describe a fan's performance are, of
necessity, determined under ideal test
conditions defined by the Air Movement
Control Assn. No practical fan system
can provide the ideal flow conditions of
the test setup, but it's vital that they be
approached as closely as possible. This
means that fan inlet and outlet must be
correctly sized and shaped, inlet and
outlet dampers must be correctly posi-
tioned, ductwork bend must be correctly
located, and the fan itself must be sized
to cover the range of gas flows through
the system at optimum efficiency.
System resistance, or static-pres-
sure drop throughout the fan system, is
the principal parameter that must be
determined. This can be found by using
handbook value of pressure drop through
ducts, branch piping, standard elbows,
etc, and by published value of pressure
drop through purchased components,
such as heat-exchanger bundles, burners.
air heaters, and particulate-removal
equipment.
Note that the pressure drop at any
point in the duct system is very depen-
dent on the turbulent or laminar state of
the gas flow. In a handbook, this turbu-
lence is estimated by specifying the num-
ber of duct diameters that the bend is
away from other components in the sys-
tem. The net effect is that all the pres-
sure-drop values are estimates that may
vary significantly from actual conditions,
especially if the fan system is poorly
designed.
Total system resistance is the sum
of the individual pressure drops in all the
fan-system components at the specified
flow rate, known as the test-block condi-
tion. To allow for the uncertainties
involved, it is normal to add 15% to the
estimated flow rate or 32% to the pres-
sure drop (pressure drop varies as the
square of the flow rate).
Once the system resistance at the test-
block condition is known, resistance at
other flow rates can be calculated using
the simple relationship that pressure
drop is proportional to flow rate squared.
This produces a system-resistance curve
of static-pressure drop against flow rate,
which may be plotted on the same scale
as the fan's characteristic curve. The
point at which the two curves intersect is
the operating point of the fan (Fig 17).
Inaccurate estimates of system re-
sistance can be costly. If system resist-
ance is overestimated, then the fan speci-
fied will be unnecessarily large and it
may be necessary to add throttling vanes
to the system to reduce the flow to the
required level (Fig 18). This involves a
continued energy wastage throughout
the life of the fan. For this reason, some
fan engineers now consider that the tra-
ditional 15% safety factor used on test-
block rating is too high. On the other
hand, allowance must be made for dete-
rioration of the system in between over-
hauls because of buildup of deposits,
wear, etc.
If the system resistance is underesti-
mated, the specified flow rate will not be
achieved and it may be necessary to
rebuild part of the system by enlarging
duct sizes or adding turning vanes to
reduce turbulence. Alternatively, the fan
speed must be increased to raise output
pressure. Neither of these modifications
comes cheap, and they are often quite
impractical. This dilemma has contrib-
uted to the interest in adjustable-pitch
axial fans whose characteristic curves
can be adjusted for optimum output
after the complete fan system has been
designed and built.
Note that most fans, especially boiler-
draft fans, do not operate in the test-
block condition for all of their duty
cycle. As the fan is throttled by closing
dampers, the point of operation moves
along the characteristic curve, and both
horsepower and efficiency change with
output flow.
Constant-speed centrifugal fans must
be selected for a test-block condition
equal to the air flow at maximum contin-
uous rating (MCR). If the system is
operated at part load for much of its
duty cycle, there will be a constant and
unavoidable loss of fan efficiency. Axial
fans with adjustable blades can be oper-
ated above the test-block condition by
increasing the pitch of the blades. So the
test-block condition for axial fans is
selected for the load level that the cus-
tomer estimates will be the median boiler
load.
Another important consideration in
the selection of the test-block point is
that at no time during the duty cycle will
the fan be operated in an unstable part
of its characteristic curve. This is partic-
ularly important with axial fans and also
with backward-inclined and forward-
curved centrifugal fans.
Gas density is of critical importance in
20 25 30
Row cfm -i- 1000
17. Fan system's resistance to gas flow (expressed in terms of
static pressure) varies with flow, as shown by curve
15 20 25 30 35
Flow, cfm -<- 1000
18. Overestimate of system resistance leads to an over sizing of
fan drive and an excess flow, which may need throttling
FANS . a special report
-------
to good performance
19. Twin induced-draft fans exhausting from stoker boiler. Note gradual increase in
duct cross section and generous windboxes to improve system efficiency
selecting fan size and drive horsepower.
A fan is a volumetric device, and at
constant speed delivers a constant num-
ber of cubic feet per minute, regardless
of gas density. Both fan output charac-
teristics and system-resistance curves are
calculated for air at 70F, at sea level,
and at a density of 0.075 lb/ftj. In
practice, the density of gas handled var-
ies with both temperature and height
above sea level. Reference to the fan
laws shows that as density increases,
static pressure and drive horsepower
increase in proportion.
In many applications, particularly
boiler-draft fans, it is the mass of air
entering the boiler that is important, not
the volume. Thus, if the air being blown
by the fan is at a temperature and pres-
sure other than standard conditions,
appropriate corrections must be made,
using standard density tables together
with the fan laws, before the fan size and
horsepower are selected.
System resistance also varies with
changes in temperature and density, and
it cannot always be assumed that all the
components of a fan system will change
their resistance according to the same
laws. This means that system resistance
must be calculated for each component
at the expected temperature and density,
to derive the system-resistance curve.
Note that estimated system resistance
is strongly influenced by design of the
fan's inlet and outlet connections (see
next page). Poor inlet and/or outlet
design cannot be adequately compen-
sated for merely by adding a safety
factor. AMCA publishes a manual of
factors to use in estimating system resist-
ance with poor inlet or outlet.
20. Double-inlet centrifugal fan (left) has inlet boxes on both
sides. Drive/bearing configuration is AMCA Arrangement 7
21. Radial inlet vanes (below) impart a swirl to gas in direction
of fan rotation. This reduces power as it reduces flow
Powor. September 1983
S . 13
-------
The integrity of a fan's performance
curve is vitally dependent on the design
of both the inlet and the outlet. This is
true whether the fan is drawing in
ambient air and discharging into a duct
system; ducted at both inlet and outlet;
or drawing air from a duct and discharg-
ing to atmosphere. Most important, air
flow must be smooth and free from tur-
bulence and, as far as possible, air flow
must be constant and must fill the pas-
sages between the fan blades as uniform-
ly as possible.
Turbulence and uneven flow of air
into the fan increases system resistance,
lowers fan capacity and efficiency, and
leads to damaging vibrations. As a
result, the fan's performance is less than
that shown on the published manufactur-
er's curve, and there is inaccurate coordi-
nation between the fan and the system
resistance curve.
Smooth flow at a centrifugal fan's
inlet is achieved with an inlet bell or
spun shape to promote streamlined flow
into the fan wheel. Different styles of fan
wheels have different shaped inlet
RADIAL-FAN INLET BELL
bells. The radial-blade paddle-wheel cen-
trifugal fan has a short cylindrical inlet
that directs gas axially into the heels of
the blades. The forward-curved fan,
because of its large blade width and
Inlet/outlet design is
shallow blade depth, has a wide inlet
mouth to ensure that the blades are
completely filled with air. The back-
ward-curved fan has an intermediate-
size inlet with a smoothly curved shape.
In this case, recirculation of flow
right angle bend to enter the center of
the fan. In such cases it is often neces-
sary to add guide vanes to the duct to
ensure smooth flow around the elbow.
An inlet box or plenum is one alter-
native. This provides sufficient space for
the gas flow to move into the fan inlet
BACKWARD-CURVED-FAN INLET
through the clearance between fan inlet
and wheel helps to distribute the flow
evenly across the blades.
A ducted inlet must be arranged so
that it is completely filled with air at a
uniform pressure and so that the direc-
tion of flow is parallel to the shaft axis.
Ideally, this requires a smooth, straight
duct length of up to eight duct diameters
before the air enters the fan. On many
systems this requirement is impractical,
especially with in-line centrifugal fans
where the gas stream must go through a
INLET BOX
evenly from all directions. The fan inlet
must have the same smoothly shaped bell
to ensure even flow into the blades.
A building wall located too close to a
fan inlet can cause undesirable turbu-
lence in the fan inlet, as can an inlet box
FORWARD-CURVED-FAN INLET
that is too shallow. If the situation is
unavoidable, the addition of a splitter
may reduce turbulence and allow the
wheel to fill properly.
Inlet dampers have a very significant
effect on fan performance and must be
•^v2
* V^&
^WSg'~-'
P^
TURNING VANES
BUTTERFLY INLET DAMPER)
carefully located. For instance, a butter-
fly damper located too close to the fan
FANS . a special report
-------
critical to system performance
inlet can have disastrous effects on fan
performance. Inlet dampers should be
located so that, when partially open, they
impart a swirl to the air stream in the
direction of fan rotation. This reduces
the power needed by the fan at the same
time that it reduces the air flow through
the fan.
Spinning flow at the fan inlet caused
by the shape of inlet duct or box, or by a
INLET SWIRL
mechanical cyclone air cleaner upstream
of the fan, can seriously detract from fan
performance. If the spin is in the same
direction as the fan-wheel rotation, static
pressure developed by the fan is reduced,
much as though the fan were permanent-
ly throttled with inlet vanes. Air swirl in
the opposite direction to fan rotation
produces a slight increase in output stat-
ic pressure, but a disproportionate
increase in power consumption and
noise.
The effect of spinning flow on fan
performance is difficult to determine
before the system is built. It should be
avoided whenever possible by the correct
placement of duct elbows or with egg-
crate-shaped straightening vanes.
Outlet is important too
The housing for a centrifugal fan is
usually of the single-outlet scroll type
exhausting into a duct, but if the fan is
exhausting to atmosphere or into a ple-
num, the multivane diffuser type may be
FAN HOUSINGS-
used. In either case, the function of the
housing is to collect the high-velocity air
as it leaves the fan wheel and convert the
velocity into static pressure.
Some types of centrifugal fans, when
exhausting to atmosphere, can operate
satisfactorily without a housing. This is
particularly true of the backward-curved
wheel, in which much of the conversion
from velocity to static pressure occurs
within the wheel. On the other hand,
forward-curved fan wheels must be used
with a housing to operate satisfactorily.
Discharge ductwork also has an
important influence on fan performance.
When the fan has a scroll-type housing,
the discharge duct usually has a much
larger cross-sectional area than the hous-
ing. Ideally, the discharge piece should
have a gradual taper to permit the veloc-
ity pressure to be efficiently converted to
static head.
Air discharge from a centrifugal fan
has a nonuniform velocity profile caused
by centrifugal forces tending to move
OUTLET VELOCITY PROFILE
air to the outside of the scroll. The
velocity pressure of this fast-moving air
is not fully converted to static pressure
until the flow has evened out. Usually a
long tapered outlet duct is impractical,
but there should be three to six diame-
ters of straight duct to fully develop
static pressure. Without this outlet duct,
there is a static pressure loss equal to
about half the velocity pressure.
Elbows in the outlet duct close to the
fan create a loss in static pressure. If an
OUTLET ELBOWS
elbow is unavoidable, it should be in the
direction of wheel rotation to minimize
turbulence created in the stream.
Axials also need inlet care
Axial-fan performance may also be
degraded by poor inlet conditions, espe-
cially propeller-type axials used on cool-
ing towers. Duct-mounted vane-axial
fans have more control over the air flow,
but their efficiency still suffers from
turbulent, uneven inlet flow.
Excessive tip clearance allows air leak-
age of the high-pressure discharge air
around the tips to the low-pressure inlet
side. The loss lowers both efficiency and
total pressure capability.
Inlet bells to ensure smooth flow of
air to the fan blades are particularly
important on forced-draft cooling tow-
ers. The loss in total pressure available
from such a fan may be as high as 26%
without the inlet bell. Inlet conditions on
induced-draft towers are equally impor-
tant, but more difficult to evaluate
because of the structural members that
interfere with streamlined air flow.
Excessive approach velocity reduces
the static pressure developed by an air-
cooler fan. It should not exceed 50% of
the fan discharge velocity. Excessive
approach velocity can occur if the fan is
located too close to the ground or to an
adjacent structure.
In a wet tower, the fan approach
velocity, as measured at the inlet louvers,
should be about 800-1200 ft/min for a
counterflow and 400-600 ft/min on a
crossflow tower. But the trend is to lower
flow rates as fan-horsepower cost contin-
ues to rise.
Excessive discharge velocity
wastes horsepower in the form of
unwanted velocity pressure. If the exit
velocity pressure for a cooling tower s ID
fan approaches 0.3 in. HjO, a velocity-
regain stack is probably in order.
VELOCITY REGAIN STACK
A velocity-regain stack recovers the
excessive velocity pressure, and lowers
total fan pressure and horsepower.
Recirculation through ID fans on wet
cooling towers can be caused by prevail-
ing winds. Exit velocity from velocity
regain stacks is generally low. and this
can be blown back through the louvers to
cause loss of efficiency and higher cold-
water temperatures. Allowance for recir-
culation must be made during the tow-
er's design phase.
Power. September 1983
15
-------
Drive technique is key to
Because most fans have a fixed design
and operate most effectively at a single
design speed, the single-speed squirrel-
cage induction motor has traditionally
been accepted as the universal drive sys-
tem. Squirrel-cage motors are inexpen-
sive, very reliable, require little mainte-
nance, and in most cases can be directly
coupled to the fan wheel, resulting in an
extremely simple drive mechanism. For
these reasons, squirrel-cage induction
motors still drive the majority of fans in
use today.
But unfortunately, the gas flow
required from a fan is not always con-
stant. While the output of an FD fan on
a base-load utility boiler may remain
unchanged for weeks on end, the same
fan on an industrial boiler used to gener-
ate steam for an industrial process may
be required to change output continuous-
ly, from shutoff to full load.
In the days when efficiency and power
consumption didn't matter, any variation
in a fan's output flow could be handled
with inlet or outlet dampers, while the
fan continued to operate at full speed.
While this system is inefficient, it retains
the basic simplicity of the squirrel-cage
drive and is still used on the vast majori-
ty of fans in service today.
But the search for an economical low-
cost drive for powerplant fans is now in
earnest. Described here are some of the
methods being used in utility and indus-
trial plants today. All of them involve
higher capital costs and lower reliability
than the tried-and-tested squirrel-cage
motor. But this higher cost must now be
balanced against the present worth of
power savings over the life of the fan.
Hydraulic coupling placed between
the fan and the motor drive allows the
fan to slow down when heavily throttled
by control dampers, while the motor
continues to run at close to full speed.
The power consumption of a fan is pro-
portional to the cube of the speed, while
the power loss in the hydraulic coupling
is directly proportional to the slip. Thus,
net power savings are possible, depend-
ing on the duty cycle of the fan. The
hydraulic coupling is a simple, rugged
piece of equipment that does little to
compromise fan reliability.
Two-speed motors with inlet vane
control offer an attractive low-cost way
of reducing energy consumption. Con-
ventional two-winding, two-speed motors
are generally too bulky and costly in the
horsepower sizes needed for boiler con-
trol. However, the advent of the pole-
amplitude-modulation (PAM) motor has
revived interest in the technique. This is
a single-winding motor designed to oper-
ate at two different, but adjacent, syn-
chronous speeds. The fan is chosen so
that the test-block rating can be met at
one synchronous speed, while the net
operating conditions are met at the adja-
cent speed. Flow control is with conven-
tional inlet dampers, but considerable
energy saving is possible at low boiler
loads.
There is some difficulty involved with
changing from one synchronous speed to
the other because the magnetic flux must
be allowed to decay in the low-speed
configuration before the motor is re-
energized for high-speed operation. To-
tal change may take as long as four
seconds. PAM motors are best applied
on boiler systems where there are a num-
ber of fans operating in parallel. System
upset can then be minimized by ensuring
that only one fan changes speed at a
time.
10.0
80
40 -
38 -
36 -
34
32 =
50
60 70
Capacity factor
60 70
Capacity factor
22. Present worth of fan drives (1981 dollars) includes capital 23. Adjustable-frequency and steam-turbine drives become
and 35-yr operating costs for FD fan on 500 MW unit favorable on this large 800-MW-unit, ID fan drive
80
S • 16
FANS . a «peoal report
-------
efficient operation
Adjustable-frequency drives have
been widely touted as the answer to
adjustable-speed drives for fans, but they
are presently only making slow inroads
into the mechanical-draft-fan market.
Principal drawback is the high initial
cost, and there is also a lingering resist-
ance on the part of powerplant engineers
to allow unfamiliar electronic equipment
onto the plant floor.
Packaged adjustable-frequency drives
up to a few hundred horsepower are now
commonplace in industry, and reliability
and cost problems have been largely
overcome. Higher-horsepower, higher-
voltage motor controls are limited by the
blocking voltage of available solid-state
power devices. To provide adequate
blocking voltage to generate an adjust-
able-frequency supply for a 4160-V
motor, for instance, requires five silicon-
controlled rectifiers (SCR) in series.
Additional SCRs and other devices
needed to control the power-carrying
devices multiply geometrically with the
voltage.
These problems are alleviated to a
large extent with the load-commutated
inverter driving a synchronous rather
than a squirrel-cage motor. A synchro-
nous motor provides its own commutat-
ing voltage to turn off the SCRs at the
end of each cycle. The result is a greatly
simplified and more reliable adjustable-
frequency power supply. However, it
cannot, of course, be retrofitted to an
existing fan motor, which must be
replaced with a more costly synchronous
motor before the load-commutated in-
verter can be applied.
While there is no doubt that the
adjustable-frequency drive offers the
greatest possibility for power savings
over the full range of boiler load, the
number of installations in the US, after
years of development, still numbers less
than 100. This number will rise rapidly
as solid-state power technology advances
and power costs rise still higher. Also
around the corner are economical adjust-
able-frequency drives for use with squir-
rel-cage motors.
Wound-rotor motors, although
widely considered an old technology, can
be economically attractive for medium-
size industrial fan drives, when rotor slip
energy, conventionally wasted in resistor
banks, is inverted and fed back into the
power line. The electronics needed are
relatively low in cost, and, in the event of
failure, the motor can be run at full
speed with damper control of air flow.
DC motors with variable-voltage sol-
id-state control are occasionally used for
industrial fan drives. While the cost of
the motor and its maintenance is higher
than for an ac motor, cost of the solid-
state drive is low and reliability is high.
Steam-turbines offer many possibili-
ties as economical fan drives, but they
are difficult to evaluate economically
against more conventional motor drives.
The obvious advantage of variable speed
to match fan output must be weighed
against the much higher installation and
maintenance costs of mechanical-drive
steam turbines when compared with
electric motors. Also, there is little ener-
gy to be saved in varying speed, unless
the fan drive can be integrated with the
overall plant cycle.
But there are many ways to integrate
turbine drives with the plant steam cycle,
given the basic fact that fan power is
directly related to boiler load and. there-
fore, to steam flow. For instance, if the
main turbines are limited by the flow
200
300
400 500 600
Unit output. MW
24. Economic choice of drive vanes significantly with unit size.
These curves are for ID fan at 70% service factor
25. Packaged adjustable-frequency drive and standard 480-V
squirrel-cage motor varies speed between 175 and 1925 rpm
Power. September 1983
S • 17
-------
through the exhaust, condensing turbines
can be used with the additional advan-
tage of increasing the capacity of the
main unit. Alternatively, backpressure
turbine drives can be used if there is
adequate use for the exhaust steam. Typ-
ical examples are feedwater heating or
flue-gas reheat.
Economic selection is complex
An economic comparison of boiler fan
drives, taking into account both the capi-
tal cost and the present-worth of fuel
costs over the life of the plant, involves a
number of complex factors that are dif-
ferent for every plant and cannot be
generalized.
One very important factor that is often
ignored is the plant capacity factor. A
base-load unit operating at full load at
all times, except for scheduled downtime,
presents very different economics to a
cycling unit. Capacity factor is the
weighted average load for one year of
operation. A base-loaded plant might
have a capacity factor of about 70%. A
cycling unit can easily have a capacity
factor as low as 50%. Clearly, to evaluate
fan-drive economics on the basis of a
hypothetical 100% load factor can pro-
duce some very erroneous results.
Unit size also has a significant effect
on the choice of drive. In a study con-
ducted by Steams-Roger Engineering
Corp and presented at the 1981 EPRI
symposium on powerplant fans, it was
found that for forced-draft fans below
150 MW, variable inlet vanes and two-
speed PAM motor drives are most eco-
nomical. At 500 MW and above, the
axial fan overtakes the two-speed motor,
and adjustable-frequency drives are bet-
ter than variable inlet vanes.
Steam turbines are almost never eco-
nomical for forced-draft fans because of
their high capital cost. However, for ID
fans on units above about 500 MW, the
steam-turbine drive becomes seriously
competitive, especially when combined
with exhaust-limited main steam tur-
bines. Turbine drives also tend to be
more competitive at low load factors
because of their high efficiency at low
loads.
Axial fans also appear to offer better
economics on large boiler systems, espe-
cially in cases where one axial fan
replaces two centrifugals.
Fig 22 shows the present worth of
total fan-drive costs for the forced-draft
fans on a 500-MW unit at different
capacity factors. Fig 23 shows the extent
to which these economics are turned
around when evaluating the much larger
drive for an induced-draft fan on an
800-MW unit. Fig 24 shows how the
economics of various drive options
change as unit size increases.
Repair/rebuild:
Because most fans are expected to oper-
ate continuously for long periods, often
around the clock, they must eventually
wear out. For this reason, fan manufac-
turers have developed extensive methods
of repair and rebuild, which they offer as
part of their overall customer service.
These services range all the way from
replacing a faulty bearing to dismantling
and reconstructing a fan wheel from the
shaft up.
Fan erosion is the principal concern
of powerplant operators, caused by the
impact of particulates in the gas stream
on the fan-wheel components. Corrosion,
which plagues operators in some chemi-
cal plants, is not a serious problem in
steam generation, provided flue gas pass-
ing through the fan is at a temperature
well above the acid dewpoint. Only when
induced-draft or scrubber booster fans
are located downstream of a wet scrub-
ber does corrosion become a problem.
Structural fatigue, caused by the cycling
of previously base-loaded fans, is occa-
sionally a problem.
Paniculate-collection equipment,
mandated on modern powerplants, has
significantly reduced the level of erosion
experienced by fans. But this advantage
has been partly nullified by the need for
more efficient fan wheels, with back-
ward-curved and airfoil blades. These
are less resistant to erosion and more
costly to repair. Flue-gas-recirculation
fans and flyash-reinjection fans (now
more frequently used to improve boiler
operation at variable load and with low-
grade fuels) must still bear the full brunt
of flyash erosion.
In-situ or in-shop
There is some difference of opinion
among powerplant operators as to
whether fans are most economically
repaired in the plant, or whether they
should be removed to a manufacturer's
repair shop, where they can be worked
on and rebalanced under controlled con-
ditions. The difference is largely in main-
tenance.
On a well-maintained fan, wear plates
are regularly inspected and replaced
before erosive wear begins to affect the
structural integrity of the wheel. But in
all too many cases, wear is allowed to
progress to the point where only a com-
plete rebuild is feasible.
What type of wear plate
Most common form of wear plate is
mild-steel floor plate, welded onto the
fan blades and onto the back and center
plates in the areas where erosive parti-
cles impinge on the fan wheel. The raised
pattern on the plate has the effect of
deflecting erosive particles away from
the blade surface so that wear may actu-
ally be less than with another harder
wear material.
One advantage of using a relatively
soft sacrificial material for wear protec-
tion is that wear patterns on a new fan
are often unknown. After a few months
or years of service, with frequent inspec-
tion, the areas of greatest wear can be
26. Fan-wheel (left) shows maximum erosion near center disc, where incoming
particulate strikes blade surface. Rebuilt wheel, with wear plates added, is at right
FANS . a special report
-------
A part of fan operation
seen. The mild-steel wear plates can be
removed and harder protective material
placed where it is really needed.
The most commonly used hard-faced
wear plate is chromium carbide weld
overlay, applied on soft steel sheet. The
weld overlay provides the hard erosion-
resistant surface, while the soft carbon
steel backing enables the material to be
cut, shaped, and welded onto the fan
blades.
The mild-steel base plate makes it
possible to remove the wear plate from
the fan if necessary, but this requires
skilled work with a gas torch. Most fan
wheels are built with regular carbon steel
plate (20-25,000 psi yield strength), and
present little problem. However, a fan
built with high-strength structural steel
such as SSS100 (about 100.000 psi)
should be treated with care to avoid
degrading its structural integrity. It is
always possible to add more weld overlay
onto the mild-steel base plate, provided
the erosive wear has not progressed
through to the structural material.
Note that the weld overlay is not
intended to provide any structural
strength to the fan and may actually be
cracked when delivered from the manu-
facturer. This cracking is due to the
method of manufacture and does not
compromise the strength of the fan
wheel.
Flame- or plasma-sprayed metal-
lizing of a fan wheel is frowned on by
most manufacturers because of the thin-
ness of the protective layer and because
it can peel off if not correctly applied.
However, if the wheel surface is properly
prepared beforehand, and the sprayed
coating is regularly inspected, this may
be a viable method of extending fan life.
But it should be understood that spray
coatings involve a certain amount of
alloying with the base material and may
interfere with future fan weldments.
Ceramic or tungsten carbide tiles
are receiving considerable attention be-
cause of their extreme hardness, but
their big problem is secure attachment
and cost. Ceramic tiles have been suc-
cessfully applied to flat fan blades, such
as straight radial, and are excellent for
housing linings, but epoxy grouting
methods limit operating temperatures.
To weld or bolt
Continuous arc or gas welding is the
universally accepted method of construc-
tion for centrifugal fans and, if properly
done, enables a fan wheel to be disman-
tled almost as easily as it is assembled.
However, wear plates are often bolted in
place in addition to bead welding around
VT/F
•JJJJ
27. Pattern on wear plates deflects
particles from surface, reducing wear
'/«-"/«in. thick
chromium carbide
weld overlay
28. Tungsten carbide wear plate is
welded onto areas of heavy erosion
Blade
Soft steel
29. Wear plate must be carefully welded
around edges to prevent entry of particles
30. Tungsten carbide or ceramic tiles
provide ultra-hard wear surface (right)
31. Final step of rebuild is static and dynamic balancing in shop. Wheel must again be
statically balanced in-situ to allow for resonant frequency of foundation
Power. September 1983
19
-------
Noise/vibration
32. Large wear plates may be bolted to
blade and bead-welded around edges
the edges. The original idea of bolting is
to allow easy removal of wear plates
when worn. However, welding is always
needed to prevent the entry of particu-
lates under the wear plate.
If the wear plate is wide, bead welding
around the edges may not be adequate,
and bolts provide additional fastening to
prevent the plate from lifting off the fan
blade. This is possible on a full wear
plate attached to backward-inclined
blades, especially as it wears thin.
Axials need different skills
The variable-pitch axial fan, although
a more precisely built piece of equipment
than a centrifugal fan, appears to lend
itself to more customer-oriented repair
than centrifugal fans. On most variable-
pitch fans, a complete set of blades can
be replaced by plant staff within a single
shift. The removed blades can then be
sent to the manufacturer or repaired
in-house.
Blades are usually protected from ero-
sion by a nose strip of stainless steel or
chrome, attached to the blade with stain-
less steel screws.
To balance an axial fan, it is necessary
to know the moment arm (WR2) of each
blade, and to locate the blades around
the hub so that the total WR2 is evenly
distributed. One manufacturer allows
the customer to control the WR2 by
altering the length of the screws that
attach the nose strip. Moment arm of
each blade is then measured by a manu-
facturer-supplied balancing machine at
the customer's facility. A computer pro-
gram then tells the customer where to
reattach each blade.
Plasma-spray coatings can also be
used to protect axial-fan blades. Because
welding of axial-fan blades is never
needed, there is no chance that the spray
coating will interfere with fan repair in
the future.
An efficient fan is a quiet fan. Here's one
case where the need to cut energy costs
and the mandate to reduce pollution go
hand in hand. And, as in so many other
mechanical systems, the cheapest and
easiest way to reduce both noise and
vibration is to tackle them at the design
stage.
Vibrations in a fan system range all
the way from the high-pitched aerody-
namic noise caused by vortices as air
leaves the trailing edge of the fan blade,
to low-frequency pulsations in the duct-
work. They may result in heavy cost
penalties because of the noise irritation
they cause, because of the energy loss,
the mechanical damage and wear on the
fan. or a combination of these effects.
Three major sources of noise
Fan noise may be caused by the blade-
passing frequency, by air turbulence
around the blades, or by mechanical
vibration of the fan housing or drive
resulting from inadequate design or
maintenance.
Blade-passing frequency is the ma-
jor and most common noise source. It is
caused by pressure pulsations as the
blades pass a stationary object such as
the cutoff sheet in a centrifugal fan or a
straightening vane on an axial fan.
Unfortunately, this frequency often falls
in the speech-interference range-
between 150 and 1200 Hz.
This noise can be minimized at the
design stage by selecting a large, slow-
moving fan, rather than a small fan, to
move the same volume of air. Since the
blade-passing frequency of the larger fan
is slower, the noise output is less objec-
tionable.
Blade-passing-frequency noise also
can be reduced by modifying the shape
of the fan cutoff or by moving it away
from the fan blades. Another technique
Resonator
Plug
33. Adjustable resonator at fan cutoff is
one way to stop blade-passing noise
34. Inlet silencers attenuate noise from two forced-draft fans feeding air to waste-fuel
boiler. Good drive mounting and alignment is also vital for noise reduction
S . 20
FANS . a special report
-------
control: now more vital than ever
is a resonant silencer mounted at the
cutoff, which can be adjusted to absorb
vibrations at the blade-passing frequency
(Fig 33).
Turbulent noise is caused by vortices
breaking away from the trailing edge of
a fan blade. Blades that are designed to
promote the smooth flow of air with a
minimum of turbulence, such as hollow
airfoil blades, generate less noise, rein-
forcing the general principle that effi-
cient fans are the quietest.
Mechanical vibration is caused most
frequently by an unbalanced fan wheel,
but it can also be driven by misaligned or
out-of-balance couplings and by turbu-
lence in the gas flow. Vibration also can
be amplified by inadequate foundations.
Here again, correction at the source is by
far the preferred remedy, though una-
voidable noise can be reduced by lagging
around the fan housing and ductwork, or
by complete enclosure of the housing.
Turbulence in the gas stream is usu-
ally caused by incorrect design of ducts
and fan inlets and outlets. It can also be
caused by a fan operating in its unstable
region. This might happen because of a
miscalculation in the fan-system resist-
ance, which results in the fan's operating
Disc-wobble resonance may confuse fan-balancing task
Deflection
Fan balancing is usually a fairly straightforward job, but it
can be made virtually impossible by a problem known as
disc-wobble resonance. This resonance is caused by
flexing of the central support disc (see sketch) which
makes it almost impossible to locate balance weights by
the usual single-plane balancing method.
The problem of disc-wobble resonance has been
extensively investigated by Southwest Research Institute,
San Antonio, Tex. It has been found on centrifugal fans
ranging from small single-inlet process fans to large
double-inlet, induced-draft fans. According to Senior
Research Engineer, Harold Simmons, "Whenever exces-
sive unexplained vibrations occur, the fan impeller should
be checked for disc-wobble resonance by impact or
shaker testing."
One severe case of disc reso-
nance occurred at a plant in which
dual ID fans were installed to con-
vert pressurized boilers to balanced
draft. Typically, with only one fan
running, bearing vibrations would
remain at 2-3 mils for a few days and
would then begin to shift in phase
and amplitude. Then, within a few
hours, the amplitude would exceed
the 5-mil alarm point, and the phase
would shift over 180 deg. The fan
would then have to be shut down
and the other fan started to keep the
unit operating at half load.
Bearing-vibration amplitude plot-
ted against speed during startup
and shutdown tests (see curve)
showed that the fan was operating
near a resonance. Normally, vibra-
tion caused by unbalance alone
increases as the speed squared, but,
in this case, it increased much faster
than that, especially over 900 rpm.
The dramatic 180-deg shift did not
appear in the bearing-housing vibra-
tions, indicating that the resonance
was probably not a shaft critical
speed.
To locate the cause of the
unknown resonance, a large vari-
able-speed shaker was bolted to the
concrete foundations so as to apply
horizontal vibrations. Shaker speed
Blade
Distortion of Center disc may cause fan
wheel to vibrate as shown here
5 •
tr
t CD O
S °2
CO *-
200
Vibration amplitude increases much
more rapidly than bearing vibration
was varied from well below, to 50% above, fan speed,
and vibration data were taken from bearing housings,
shaft, and foundations.
Shaft critical speed was found to be well above the fan
running speed of 900 rpm, and the foundation response
was found to be highly damped, eliminating these two
components as sources of the resonance.
The disc resonance was identified by measurements
from inside the wheel. When the shaker speed reached
930 rpm (15.5 Hz), the whole wheel vibrated at 20 mils or
more, while the shaft remained relatively still. Measure-
ment of the vibration at selected locations around the
wheel showed that the center disc was wobbling, as
shown in the sketch.
Calculations estimated that 50%
increased stiffness of the center disc
was needed. These calculations
were confirmed with a scale model,
and the stiffening plate was installed
on one of the two fans. Shaker tests
were repeated and the wobble fre-
quency was found to have increased
by 25%.
When the fan was restarted, the
vibration level was significantly re-
duced, as shown below. The modi-
fied fans have now been operating
for several years without a repeat of
the problem.
Modifications to the center disc
are not always necessary. In another
case, several fans suffering from
disc-wobble resonance were suc-
cessfully field-balanced. These fans
were in ventilation service, and were
not exposed to the large thermal
excursions such as ID fans experi-
ence. Even when the disc-wobble
resonance was within 1% of fan
speed, the fans could be balanced
for both horizontal and axial bear-
ing-housing vibrations using a con-
trolled balancing procedure.
Accurate amplitude and phase
data must be obtainable to balance
a fan near resonance. Also, the
effects of extraneous excitations,
such as shaft misalignments and fan
interaction, must be minimized.
After center-disc
modification
400 600
Speed, rpm
800
Power. September 1983
21
-------
point being misplaced from its design point
(Fig 35).
Gas-stream turbulence not only causes
mechanical vibration in fan housing and
ductwork, but also reduces overall sys-
tem efficiency.
Silencers should be used on a fan
system only after all reasonable attempts
have been made to reduce fan noise by
proper design. This is because any silenc-
er produces a pressure drop in the system
and thus consumes power. And the
greater the noise attenuation, the greater
the pressure drop. Nevertheless, inlet
silencers are often unavoidable on any
large fan drawing in ambient air, and
outlet silencers are frequently needed on
Q.
0
35. Inaccurate estimate
of fan-system-resistance
curve may result in oper-
ating point falling in un-
stable part of fan-charac-
teristic curve
Flow
fans exhausting to atmosphere.
Absorptive silencers are generally
used on fan intakes when the fan is
drawing in clean ambient air. These
silencers consist of perforated rectangu-
lar plates packed with mineral wool or
fiberglass (Fig 36). The width of the
plates and the pass width between them
Vibration can accompany staggering inefficiency and energy losses
Serious vibration problems repeatedly broke outlet
dampers and cracked inlet boxes on two 7000-hp forced-
draft centrifugal fans feeding a balanced-draft boiler at
one utility company. But that was not the only problem.
Robert Perry of Process Equipment Inc. Birmingham,
Ala, a supplier who also consulted on the problem,
reports that there was a significant temperature gradient
between inlet and outlet ducts. In fact, the outlet duct
was too hot to touch, whereas the temperature should
have been close to ambient.
Clearly, the vibration was caused not by unbalanced
rotors or loose foundations, but simply by serious fan
inefficiency. Uneven distribution of air to the fan wheel
was causing the air passages be-
tween rotor blades to fill up intermit-
tently, instead of smoothly and con-
tinuously. Result was shock losses
that were causing the vibration.
The duct layout to the fans is
shown in the sketch. Velocity tra-
verses made at various points in the
duct showed that the air flow was
quite smooth and continuous, and
the velocity profile quite flat up to
the 90-deg turning elbow, but at and
after the elbow. Perry found serious
velocity-pressure pulsations. It was
clear that serious vortex shedding
and turbulence was occurring from
this point to the fan inlet.
According to Perry, "The problem
could easily have been solved with
the installation of turning vanes at
the 90-deg elbow and in the inlet
boxes, but, to my dismay, the oper-
ating engineers were reluctant to
accept the fact that inlet-duct design
was the cause of both the vibration
problem and the serious and costly
inefficiency.
"The original specifications called
for a guaranteed fan efficiency of
92%, and there was no reason to
suppose that figure was not attainable on these back-
ward-inclined airfoil-bladed fans. When we checked the
actual efficiency with the boiler at full load, we found the
figure to be closer to 43%. At the 50% load, the fan
efficiency checked out at 17%."
Utility management could not be convinced that the
investment of $200,000 in turning vanes was justified.
They were still unconvinced when it was shown that, at an
energy cost of $300/hp-yr, the cost of the fan inefficiency
was $546,000/yr per fan.
Engineers familiar with fan systems will confirm that
this type of problem is quite common, but this utility still
has not installed turning vanes.
Air flow i* smoothand uniform up to the
90-deg turning elbow, after which serious
turbulence and vortex-shedding occurs
To boiler
S . 22
FANS
i special report
-------
Improved fan performance eliminates need for silencers
Noise emanating from the No. 12 stack of the Michigan
City generating station, Northern Indiana Public Service
Co, was causing serious complaints from local residents
in 1979. Most of the irritating noise was at the blade-
passing frequency of 145 Hz and its second harmonic.
Originally, the utility planned to install silencers on the
ID fans. However, a complete fan-system study, con-
ducted with the help of Stone & Webster Engineering
Corp, Boston, Mass, showed that the problem was
directly related to poor fan performance and could be
solved without the use of silencers.
Unit 12, rated 500 MW, has twin ID fans with airfoil
blades, originally driven by constant-speed 890-rpm
7000-hp motors and discharging into a 500-ft stack.
The fans are oriented north and south of the stack, and
each has an east and west inlet duct. The inlet boxes
had opposed-blade dampers for flow control. Hydraulic
coupling between fan and drive motor was not in use
because of its slow response to system transients.
Extensive performance evaluations of the complete
system showed that static pressures ranged from 1.0 in.
H2O above the manufacturer's curve to 7.0 in. H20
below, and all calculated horsepower points were above
the manufacturer's curve. Misalignment between
dampers appeared to be the probable cause of flow
unbalance, ranging from 11% to 28% of flow. Damper
problems also seemed to be the cause of flow unbal-
ance between the east and west inlet ducts of the south
ID fan. Measured fan efficiencies were all below manu-
facturer's predictions, because of the unbalance prob-
lems and gas counterspin developed by dampers. Yet
another problem caused by the dampers was that at
low loads, they caused the fans to operate in the
unstable portion of the characteristic curve.
The cutoff of a centrifugal fan is the part of the
housing closest to the rotor blade tips at the discharge.
It normally has a small radius and is parallel to the blade
tips. Modifications to this shape can reduce pressure
pulsations as the blades pass it.
Several methods of modifying the cutoff had already
been tried on the fans of boiler No. 12 before Stone &
Webster became involved. The distances between
blade tips and cutoff and the cutoff radius had both
been increased. These modifications had apparently
produced a 5-dB noise reduction.
A sloped cutoff can also be used to reduce fan noise.
This has the effect of smoothing out the pressure
pulsations. For a double-inlet fan, a V-notch is used
(see sketch). Aerodynamic-model tests of the ID fan
showed that a V-notch cutoff would not produce a
significant throttling effect, and so a new cutoff was
made for each fan and installed during an outage.
Control modifications and their effect on noise output
were compared, and are shown in the curve below. The
upper curve shows the original configuration using
opposed-blade dampers with the fans at full speed. The
second configuration is a temporary combination of
flow control with opposed-blade dampers and fan-
speed control. The third curve is a combination of
parallel-blade inlet louvers and variable-speed control.
This is the actual final configuration. (It is noisier than
configuration (2) at lower load because fan speeds are
higher.)
Parallel-blade inlet louvers were installed, replacing
the original opposed-blade dampers. Aerodynamic-
model tests showed that relocating the new inlet louvers
three feet closer to the inlet flange would improve fan
performance by inducing a swirl flow into the inlet, allow
better flow balance and fan control, and increase reli-
ability by reducing damper-blade actuator failures.
Speed control produced a significant reduction in fan
noise, as well as a 1.4-MW reduction in motor horse-
power. Variable-speed control always minimizes fan
noise because the fan is able to operate near peak
efficiency throughout its load range. In addition, the
low-frequency inlet-duct noise and vibration were
reduced because of the lower pressure drop across the
inlet dampers.
Sound-level reductions produced by the modifica-
tions to the inlet louvers and cutoff were about 6-7 dB
at peak load, while reductions due to speed control
were about 8 dB. Modifications to Unit 12 that were
made concurrently with the noise-abatement investiga-
tion increased the output power level from 468 MW to
500 MW But despite this increase in power, the
required noise reduction was achieved without the
installation of silencers.
Fan wheel with
10 backward-
inclined airtoil
blades
Inlet
V-*haped cutoff has the effect of smoothing out pressure
fluctuations as the fan blades pass the cutoff edge
130
120-
115-
1 110
G.
105
100
(1) Opposed-blade
damper control
(full speed!
810 rpm
795 rpm
726 rp,
(3) Inlet louvers and
variable-speed
control with
V-notch cutoff
Opposed-blade damper
control with variable speed
610 rpm
350 400 450 500
Boiler load. MW
How control modifications reduced the noise level at different
outputs. Configuration (3) is the method finally used
Power. September 1983
S • 23
-------
Absorptive material
Duct
Perforated sheet
Mineral wool or fiberglass
36. Pressure loss in an absorptive-type silencer is a function of
the number of decibels of sound attenuation needed
Outlet evase
Acoustic chambers
Gas flow v ^ --Outlet screen
37. Resonant silencer is used to cut noise in dirty-gas streams.
It must be sized to attenuate a particular sound frequency
Adaptor plate
determines the level of attenuation.
Resonant silencers are used to
attenuate noise at the outlet of fans
handling contaminated gas streams, such
as a boiler ID fan. The acoustic length of
this silencer consists of a series of reso-
nating chambers angled away from the
direction of gas flow. The frequency of
these resonating chambers is selected
according to the blade-passing frequency
of the fan. such that the reflected sound
waves exiting the chambers are 180 deg
out of phase with the incoming sound-
waves, effectively cancelling them (Fig
37). Because of the open nature of the
resonant chambers, the acoustic silencer
is far less liable to plugging.
Disc silencers, which are also used
to attenuate noise at fan inlets, consist of
two panels, one surrounding the fan inlet
and the other opposing it. Additional
disc panels may be used to provide more
inlet area. The panels form an inlet-gas
path that is tapered from the entering
edge to the fan inlet (Fig 38).
Fan-wheel balance is critical
A major source of fan-system vibra-
tion is unbalance of the fan wheel itself.
Fans that are perfectly balanced when
delivered may become unbalanced dur-
ing service as a result of the accumula-
tion of deposits on the fan blades or
uneven erosion of fan material. Accumu-
lation of material on the fan blades can
be minimized by the use of sootblowers
or by sonic horns (Fig 39).
In powerplant applications, unbalance
in a fan wheel is more likely to accumu-
late from the erosion of.fan material. To
prevent this unbalance from reaching
damaging levels, the fan wheel should be
checked periodically.
A fan wheel should always be bal-
anced in-situ after any work has been
done on it and before it is returned to
S • 24
Circumferential inlet
38. Disc silencer, used on fan intake,
doubles as inlet bell to smooth flow
service. In-situ balancing is essential
because shaft alignment, foundation
stiffness, and resonant frequencies of the
supporting structure have a critical
effect on the vibration amplitude.
Single-plane balancing is almost
always adequate for in-situ balancing,
provided the fan has been dynamically
balanced by the manufacturer after con-
struction. Single-plane balancing cor-
rects for radial vibratory forces caused
by uneven distribution of fan mass
around the shaft axis. It does not correct
for longitudinal vibration caused by
varying degrees of unbalance along the
shaft length. For this, dynamic or two-
plane balancing is needed.
^Fan housing
33. Sonic horn provides a low-cost way
to remove paniculate buildup from blades
Single-plane balancing is relatively
simple and can be performed in-situ with
a simple vibration analyzer. The tech-
nique involves first measuring the mag-
nitude and phase of vibrations caused by
the wheel unbalance. A trail weight is
then attached to the fan wheel at a
known point, and the amplitude and
phase of vibrations are again measured.
Knowing the position and mass of the
trial weight, the size and position of the
correct balancing weight can then be
determined by a simple vector analysis.
Technicians familiar with balancing can
often dispense with the vector analysis
and correctly place the balance weight
by estimation. (For a more detailed
description of balancing techniques, see
Power special report. Balancing rotating
machinery. October 1983.)
Reprints of this Special Report are
available at nominal cost. For a
complete price list covering this and
other reports, write to:
POWER Reprint Department
1221 Ave of the Americas
New York, NY 10020
FANS . i special report
-------
ITEM 2
Fans - Chapter 7
Patrick Dolan
Handbook Of Ventilation For Contaminant Control
-------
HANDBOOK OF VENTILATION
FOR CONTAMINANT CONTROL
(Including OSHA Requirements)
by
Henry J. McDermott
Regional Industrial Hygienist
Shell Oil Company
Walnut Creek, California
ANN ARBOR SCIENCE
PUBLISHERS INC
PO. BOX 1425 • ANN ARBOR, MICH. 48106
7
FANS
If hoods are the most important component in the ventilation sys-
tem, the fan along with the ducts leading into and out of the fan
rank second in importance. The fan, of course, generates the suction
in the system that draws contaminated air in through the hoods. If
the fan is too small the airflow will be too low. Fortunately, fan
selection does not always have to be perfectly accurate; fans have
some built-in flexibility since their capacity increases with higher
fan speeds although this also increases the fan's power consumption.
Speeding up the fan is the standard remedy for systems with inade-
quate airflow.
The ducts before and after the fan can almost be considered part
of the fan itself. These ducts establish smooth airflow into and out
of the fan so the fan can do the maximum work rhoying air. Poor
design of these ducts can lead to turbulence and uneven flow pat-
terns at the fan inlet and outlet, and the fan's capacity will be lower
than you expect from the fan size and speed.
This chapter covers fan selection criteria and the different types
of fans that are available. It should be valuable both as general back-
ground and when selecting a new fan for a new or existing ventila-
tion system. In many plants the problem is an existing system that
does not work properly. Chapter 11 on "Solving Ventilation System
Problems" reviews steps to improve system performance. Chapter 9
contains tips on lowering the cost of ventilation systems.
185
-------
186 Handbook of Ventilation
FAN AND SYSTEM CURVES
Fan selection involves choosing a fan to match the requirements
of the exhaust ventilation system. The fan must move the correct
quantity of air against the resistance to airflow caused by friction
and turbulence in the system. The relationship between flow rate
and resistance for both the exhaust system and the fan can be
plotted to help select the proper fan. The plots are called "rating
curves."
Exhaust System Curves
The pressure loss or resistance to airflow through a ventilation
system is proportional to the square of air velocity through the hood
and ducts. Once a system is designed, and the duct diameters and
lengths chosen, the amount of static pressure (suction) that the fan
must develop to pull different quantities of air can be estimated.
Figure 7-1 illustrates a system curve for the welding bench hood
system in Figure 6-5. At the design flow rate of 1050 fts/min the
2.0
t.
-------
188 Handbook of Ventilation
i 2.0
1.5
1.0
0.5
(b)
Brake horsepower
N
\
\
\
2.0
1.5
1.0
0.5
Static pressure
Brake horsepower
\
0.4
0.3
0.2 o
0.1
0
0.4
0.3
500 1000 1500 2000
Airflow - ft3/m1n
10
0.1
i
Outlet static
pressure
^n
[W
r 0.8 in.
of water
Inlet static
pressure
2.3 1n. of water
Figure 7-3 Fan static pressure is calculated from the static pressures at the
fan inlet and outlet, and the velocity pressure at the fan inlet.
let is positive. Both represent energy needed to overcome resistance
to airflow so the signs are not important.
Example: What is the fan static pressure for a system with
a static pressure reading of 2.3 in. of water suction on the
inlet side of the fan and 0.8 in. of water positive pressure on
the discharge side of the average fan inlet duct velocity is
3000 ft/min (Figure 7-3)?
Answer: The velocity pressure in the duct at 3000 ft/min
velocity is 0.56 in. of water (either from Equation 4-3 or
Table 6-3).
FSP=|SPlnlet +|SPoullel| -VPlnlel
= 2.3 + 0.8 — 0.56
= 2.5 in. of water
Brake Horsepower Curve
The amount of electrical power needed to spin the fan depends
on the fan's output and the system resistance. It can be plotted as
the "brake horsepower curve" on the fan rating diagram (Figure
7-2b). Brake horsepower is the amount of energy needed to run
the fan neglecting the drive losses between the fan and motor. Brake
-------
190 Handbook of Ventilation
horsepower data are based on manufacturers' tests of their fans fol-
lowing standardized procedures.1 The actual power consumption
will be higher than the brake horsepower rating because of drive
losses.
The shape of the brake horsepower curve shows the effect of oper-
ating the fan at different points along its static pressure curve. The
curve in Figure 7-2 is typical for one type of fan but other fans
have different shaped curves.
Mechanical Efficiency Curve
Mechanical efficiency is a measure of how much energy the fan
uses at different points on the static pressure curve. The goal is to
choose a fan that is operating near its peak efficiency. The key is
to find the correct fan size so the peak efficiency coincides with the
exhaust system design flow rate and static pressure.
Figure 7-2c shows all three fan curves plotted together as a fan
rating curve might appear in the fan manufacturer's literature.
Since the mechanical efficiency is a relative measurement there are
no units for mechanical efficiency; the shape of the curve indicates
its efficiency.
Operating Point
When the fan curves and exhaust system curve are plotted to-
gether, the point of intersection of the fan static pressure curve and
the system curve indicates the airflow through the system with that
fan. The intersection is called the "operating point." Figure 7-4
shows the system curve for the welding hood system from Figure
7-1 and the fan curves from Figure 7-2. The operating point shows
that 1050 ft3/min of air will be drawn through the system with
that fan. The brake horsepower can be determined by moving up
from the operating point to the brake horsepower curve, then read-
ing the brake horsepower on the right vertical axis. From Figure
7-4 about 0.25 horsepower is the power needed to rotate the fan
neglecting drive losses between fan and motor.
Fan Rating Tables
Every fan has a separate rating curve for each fan rotating speed.
Increasing the fan speed moves the curve upward while decreasing
the fan speed moves the curve down. Figure 7-5 shows that dif-
ferent operating points for a ventilation system can be achieved by
changing the fan rotating speed. Most fan manufacturers make
Fans 191
2.0
0.4
Static pressure
Brake
horsepower
I
t-
o
-C
2000
Airflow — ft /min
Figure 7-4 Plotting the system curve from Figure 7-1 with the fan curves
from Figure 7-2 shows the operating point for the system and fan.
Figure 7-5 Each fan has a separate static pressure curve for
rotating
-------
CFM
687
773
859
945
1031
1117
1203
1289
1375
1461
1547
1633
1719
1805
1891
1977
2063
2149
2235
2407
V SP
RPM BMP
889 0 04
957 005
1028 007
1101 008
1176 Oil
1253 012
1330 015
1409 017
1488 02,1
1568 023
1648 027
1730 031
1811 036
1894 041
1974 046
2059 0 52
2141 059
2222 065
2306 0 73
2474 091
V SP
RPM BMP
«1 CC6
1038 O.67
1103 009
1172 Oil
1241 012
1313 0 15
1389 017
1464 02!
1539 023
1618 027
1696 031
1774 035
1853 039
1934 044
2014 051
2097 0 56
?176 063
2258 071
2340 077
2505 0 94
'/•SP
RPM BMP
1061 OOfc
1115 0.09
1175 0.11
1237 fl.U
1304 015
1373 0 17
1445 021
1518 023
1591 026
1666 031
1742 034
1820 0 3R
1896 043
1975 048
2053 0 51
2134 061
2213 067
7293 0 74
2375 082
25V 099
V SP
RPM BMP
1218 012
1261 0 1 3
1310 tie,
1364 017
1423 021
I486 0.22
1550 025
1617 029
1688 l>37
1758 036
1829 041
1904 045
1979 051
2052 056
2130 C6?
2205 0 68
2285 0 76
2167 OR3
2441 u 91
260(i 1 09
r SP
RPM BMP
1365 016
1397 018
1439 021
I486 0 22
1537 0 25
1593 02*
1653 031
1714 0.35
1781 039
1847 043
1914 047
1984 05?
2058 0 58
?128 064
?204 071
2278 0 77
2352 084
24 ?8 09?
2504 1 01
?66l> 1 19
I1," SP
RPM BHP
1649 028
1677 031
1712 033
1750 036
1796 041
1845 0.44
1900 0.48
1955 OK
2015 0.57
2077 062
2141 068
2207 0 74
2272 081
2343 088
2414 095
2483 1 03
2555 1 1 1
2627 1 21
?775 1 41
2 SP
RPM BHP
1896 042
1920 045
1950 049
1988 053
2027 057
20.73 0 62
2122 0.67
2175 072
2231 0 7«
»89 0.84
2349 091
2411 098
2477 1 05
2542 1 13
2610 1 22
2677 1 31
2747 1 41
2888 1 62
2V SP
RPM BHP
2114 058
2139 062
2168 066
2203 071
2240 0 76
2282 082
2329 0 88
2378 0 94
2433 1JOI
2485 1.06
2545 1.16
2604 1,24
2666 1 32
2728 i<)
2797 1 51
2861 1 6?
2995 1 83
3" SP
RPM BHP
2314 076
2341 081
2368 086
2399" 0 92
2436 098
2476 1 04
2522 1 11
2569 1 18
2621 1.26
2674 1J4
2779 1.43
rm !.»
2847 1«
2909 1 72
2971 1 83
3100 206
3V SP
RPM BHP
2504 0 96
2527 1 02
2555 1 08
2584 1 14
2620 1 21
2662 1 29
2704 1 36
2751 1 45
2800 1 54
2150 1 63
1 If
3080 ?«fj
3205 2 3T
O"
o
O
?r
o
s
1289
1375
1461
1547
1633
1719
1805
1891
1977
2063
2149
• 2235
2407
2579
2751
2923
3095
3267
3439
4 SP
RPM BHP
2699 1 24
2730 1 32
2762 1 39
2797 1 47
2836 1 55
2878 1 64
2822 1 73
2969 1JU
3023 1.94
3075 2.04
3131 218
3187 2.28
3306 254
3432 212
3559 3 13
3692 3 46
3830 383
3969 4.23
4110 466
4V S
RPM e
2841
2868
2895
2927
2963
2999
3042
J0£7
3137 ;
3185 ;
3238 i
too ;
3494 J
3524 .
3651 .
3779 :
3912 '
4047 '
4191 4
p
HP
42
49
57
66
75
81
94
> 04
MS
t.26
.39
LSI
in
07
41
74
12
53
99
5 SP
RPM BHP
3001 1 68
3024 1 '6
3052 1 85
3086 1 94
3119 204
3160 215
3201 2 25
3245 2 37
J790 2.48
U40 241
33*4 179
»01 9.02
1*17 J 33
3738 S«7
3866 4 03
3994 4 42
4130 485
4267 5 31
5': SP
RPM BHP
3150 1 95
3180 2 05
3206 2 15
3238 2 25
3273 236
3312 247
3355 2 59
M98 2.72
3441 2JM
3493 2.91
3S99 3.21
3708 359
3825 JJ4
3947 431
4075 4 72
4206 5 16
6 SP
RPM BHP
3277 2 16
3299 2.26
3323 2 35
3356 2 47
3386 2 58
3424 271
3462 282
3503 2 95
3S4J JJ»
3990 3.22
3WO 3.53
379S XK
3910 4J1
4032 441
4154 SOI
6V SP
RPM BHP
3417 247
3438 2 57
3465 2 68
3498 2 81
3530 2 93
3567 3 06
3604 3 19
3644 3 33
M»3 3.49
37W 3J1
3*81 4.13
39*8 4J1
411* 4J1
4232 5.32
7 SP
RPM BHP
3551 2 79
3576 2 91
3603 3 03
3635 316
3671 331
3707 3 44
3746 3 58
3788 374
an 4.07
1975 4.41
4079 4.7*
4191 119
7VSP
RPM BHP
3660 3 01
3684 3 14
3710 3.26
3740 341
3771 3.54
3807 369
3842 383
3881 3 99
MM 4.31
40*9 440
4144 tfi
8 SP
RPM BHP
3788 3.37
3812 3.51
3843 364
3868 3 78
3903 3.94
3938 409
3973 4 ?5
4059 461
4147 4.97
4245 5.J7
8V SP
RPM BHP
3892 3.61
3916 3.75
3940 3.89
3970 404
4000 4.21
4030 4.35
4065 4.51
4143 4.S7
4232 5.26
~; *.•£
.1.
Selection within limed area renders most efficient quietest operation BHP shown does not include belt drive losses
CFM = Flow rate, fr/min
RPM Rotating, speefl, rev/min
SP = Fan static pressure, in. of water
BHP = Brake horsepower
Figure 7-fl Fan rating table for a backward inclined blade fan from a fan catalog. (Source: Reference 2)
CD
CO
-------
194 Handbook of Ventilation
"families" of the same fan design in different sizes with operating
curves that cover a range of efficient fan operating conditions. Fan
catalogs contain fan rating tables (Figure 7-6) for each size fan
constructed from the rating curves. Most fan tables have a shaded
portion to indicate the fan to select for maximum mechanical effi-
ciency.
Caution: Fan Rating Data
You can get in trouble using fan rating tables and curves if the
air density at the fan is different from standard conditions since
standard density air was used to construct the table; or if the fan
inlet connection does not match the ideal conditions used during the
fan tests.
Air Density
Fan ratings are developed from tests using "standard air" with
a density of 0.075 lb/ft3. This is the density of air at 70°F, 50 per-
cent relative humidity and a barometric pressure of 29.92 in. of mer-
cury. When air density varies significantly from this value, correc-
tions to fan ratings are needed. The primary factors affecting density
are air temperature and the plant's altitude above sea level. In
either case the volume capacity of the fan is not changed but the
static pressure developed by the fan varies with the air density. If
the air temperature or altitude increases, the air becomes less dense
and so, for example, a fan moving 1000 ft'/min of this less dense
air is moving less mass than a fan moving the same quantity of
"standard air." Consequently, the fan does not have to develop as
much static pressure when it is moving less dense air. Also the horse-
power requirement is lower since less air mass is being moved. Cor-
rections are made to the pressure and horsepower ratings using the
factors in Table 7-1.
Standard air is usually assumed during system design because
the duct friction and other pressure loss design data in Chapter 6
are based on standard air. After the system has been designed and
the fan capacity and static pressure calculated, corrections are made
to reflect actual operating conditions. Generally no density correc-
tions are needed for temperatures from 40°F to 100°F or for alti-
tudes from —1000 to +1000 ft (relative to sea level) because density
differences are not large enough to affect fan performance in indus-
trial exhaust systems.4
TV\e most coimmon density correction is -wHere a system is designed
Fans
195
Table 7-1 Air Density Correction Factors (Altitude and Temperature)
Air
Temp.,
°F
0
70
100
120
140
160
180
200
250
300
350
400
450
500
550
600
650
700
750
800
Altitude Above Sea
0
.87
1.00
1.06
1.09
1.13
1.17
1.21
1.25
1.34
1.43
1.53
1.62
1.72
1.81
1.91
2.00
2.10
2.19
2.28
2.38
1000
.91
1.04
1.10
1.14
1.18
1.22
1.26
1.29
1.39
1.49
1.59
1.69
1.79
1.88
1.98
2.08
2.18
2.27
2.37
2.48
1500
.92
1.06
1.12
1.16
1.20
1.24
1.28
1.32
1.42
1.52
1.62
1.72
1.82
1.92
2.02
2.12
2.22
2.32
2.42
2.52
2000
.94
1.08
1.14
1.18
1.22
1.26
1.30
1.34
1.45
1.55
1.65
1.75
1.86
1.96
2.06
2.16
2.26
2.36
2.47
2.57
2500
.96
1.10
1.16
1.20
1.25
1.29
1.33
1.37
1.47
1.58
1.68
1.79
1.89
1.99
2.10
2.20
2.31
2.41
2.51
2.62
3000
.98
1.12
1.19
1.23
1.27
1.31
1.36
1.40
1.50
1.61
1.72
1.82
1.93
2.03
2.14
2.24
2.35
2.46
2.56
2.66
Level
3500
.99
1.14
1.21
1.25
1.29
1.34
1.38
1.42
1.53
1.64
1.75
1.85
1.96
2.07
2.18
2.29
2.40
2.50
2.61
2.72
4000
1.01
1.16
1.23
1.28
1.32
1.36
1.41
1.45
1.56
1.67
1.78
1.89
2.00
2.11
2.22
2.33
2.44
2.55
2.66
2.76
4500
1.03
1.18
1.25
1.30
1.34
1.39
1.43
1.48
1.59
1.70
1.81
1.93
2.04
2.15
2.26
2.38
2.49
2.60
2.71
2.81
5000
1.05
1.20
128
1.32
1.37
1.42
1.46
1.51
1.62
1.74
1.85
1.96
2.08
2.19
2.30
2.42
2.54
2.65
2.76
2.86
Source: Reference 3.
the system will operate with different density air. In this case the
correct fan static pressure and horsepower to move the desired air
volume must be calculated. '
^
Example: The welding hood system in Figure 6-5 is to be
relocated to Denver, Colorado (altitude 5000 ft) although it
was designed to operate at sea level. What size fan will be
needed if the original specifications required a fan moving
1050 ft'/min at 0.86 in. of water fan static pressure and a
suitable fan had a 0.25 brake horsepower rating at those
conditions?
Answer: The fan must move 1050 ft'/min but, since the air
density is lower due to altitude, less mass of air will be
moved meaning less resistance and horsepower. From Table
7-1 the correction factor for 5000 ft and 70°F is 1.20.
•n*op
rsp.etu., = "*-••'»•
0.86
Density Factor 1.2
0.25
= 0.72 in. of water
= 0.21 brake h
-------
196 Handbook of Ventilation
Select a fan from standard fan tables rated at 1050 ftVmin
and 0.72 in. of water fan static pressure.
A second type of density correction is applied when you want a
system to move a specific volume of air and develop a specific fan
static pressure at nonstandard conditions.
Example: Select a fan to move 15,000 ft'/min at 30 in. of
water fan static pressure while operating at 200°F and 2500
ft altitude.
Answer: In order to use a fan rating table the static pressure
must be adjusted to the equivalent fan static pressure at
standard conditions. From Table 7-1 the correction factor for
200°F and 2500 ft is 1.37.
FSPequiv = FSPaciuai x Density Factor
= 30 in. x 1.37 = 41.1 in. of water
Select a fan rated at 15,000 ft'/min and 41 in. of water fan
static pressure from a standard rating table.
For the dilution ventilation systems discussed in Chapter 1 the
air volume itself must be corrected since the dilution effect is based
on air mass rather than simply air volume. In other words if calcu-
lations show that 1000 ft'/min Js needed for contaminant dilution
at standard conditions, a higher airflow rate will be needed at oper-
ating conditions where the air density is less than 0.075 pounds/
ft3. This is explained in Chapter 1.
Poor Fan Inlet Connections
The second thing to remember when using fan rating tables or
curves is that the tests used to develop the ratings were conducted
under ideal laboratory conditions. Often field conditions do not
equal the test conditions and so a fan will not perform as well as
the rating table predicts it will.
For the tests straight ducts are connected to both the fan inlet and
outlet (Figure 7-7). This helps assure that the air enters and leaves
the fan with minimum turbulence and nonuniform flow. Fan blades
are designed to be most efficient when air enters the fan in a straight
line. Elbows, fan inlet boxes or duct junctions near the fan can im-
part a spin to air entering the fan. If the spin is in the same direc-
tion as fan rotation, the amount of air moved will decrease along
with energy consumption since the fan blades have to "catch up" to
the air before acting on it. If the air spin is opposite to the fan rota-
tion the output will be reduced although the power consumption will
Vtp h i f*h or t n i n ovor»r>< ' ** i»-/-.
Fans 197
Inclined
manometer
Pitot tube
velocity traverse
Flow straightener
Bell-shaped inlet
Figure 7-7 The duct arrangement used to test fan performance provides ideal
airflow conditions that actual installations may not duplicate. This
can reduce fan performance below that expected from the manu-
facturer's rating table.
can be affected by an elbow or fan inlet box. The 'same figure also
shows that performance can be restored by adding turning vanes
inside the box to reestablish straight flow into the tan.5
Reduced fan performance due to poor inlet connections is insidious
in that it cannot be identified using the standard pressure measure-
ments used to test fan performance." Reduced performance caused
by a duct obstruction or a plugged filter can be identified using the
simple pressure tests discussed in Chapter 10. However, poor duct
inlet connections do not increase pressure losses in the system; they
just reduce the fan's ability to do useful work on the air. Special
testing techniques to measure spinning or uneven airflow into the
fan are covered in Chapter 10.
Poor Outlet Connections Reduce Static Regain
Poor fan outlet connections also have an adverse effect on fan
-------
198 Handbook of Ventilation
Vane
Inlet box without
turning vanes
10
Inlet box with turning
vanes to straighten
airflow entering fan
O)
4->
tO
O
Z 6
I
111
t-
3
I/I
I/I
01
I.
o.
u
0
I
T
Straight Inlet
(no Inlet box)
Vaned Inlet
box
0
20
30
40
50
Airflow - 1000 ftj/m1n
Figure 7-8 Illustration of the effect of a fan inlet box without turning vanes
on fan performance. Addition of turning vanes to straighten air-
flow restores fan performance. (Source: Reference 5)
with no elbows or other interference to smooth flow for 5 to 10 duct
diameters away from the fan outlet.
The reason for the adverse effect is that the air discharged from
a fan outlet does not have uniform velocity distribution (Figure
7-9) . Since air has weight it is thrown out by centrifugal force from
the spinning fan wheel, resulting in higher velocities at the outer
edge of the outlet than at the inner edge. Several duct diameters
downstream from the fan outlet the air velocity returns to near uni-
across tK
TKe velocity pressures
Fans
199
Figure 7-9 Uneven air velocity distribution at
the fan outlet results in a higher
velocity pressure there compared
to a location 5-10 duct diameters
downstream where air velocity dis-
tribution is more uniform.
energy) in the moving air is proportional to the square of the ve-
locity according to Equation 4-3:
vp-(-^—Y
v*~ [ 4005 )
where VP = velocity pressure, inches of water
V = air velocity, ft/min
Usually the velocity pressure is calculated from the average duct
velocity but for added accuracy the velocity pressure can be calcu-
lated using the individual velocity readings across the, duct as shown
in Figure 7-9:
/ V, + V 4- 4-V\2
,_ / Vi t- V2 + ... -i- Vn \
V N ; (7-2)
(4005)2
where Vj, V2... V,, = individual velocity readings, ft/min
N = number of individual velocity readings
Static Regain
Solving Equation 7-2 at both the fan outlet and a point several
duct diameters downstream from the fan outlet shows that the ve-
locity pressure in the system is higher at the fan outlet than do-wm-T
VP =
-------
200
Handbook of Ventilation
stream from the outlet. Since no energy was added to the system
between these locations the total energy is constant except for the
slight duct friction loss. So the drop in velocity pressure is balanced
by a corresponding increase in static pressure as velocity pressure
(kinetic energy) is converted into static pressure (potential energy)
according to Bernoulli's Law, as explained in Chapter 4. This phe-
nomena is known as "static regain" and is important in ventilation
work because the magnitude of friction, turbulent and other pres-
sure losses is directly proportional to the velocity pressure. These
losses can be minimized in the exhaust stack by converting as much
velocity pressure to static pressure as possible before the air reaches
elbows or other sources of pressure loss. Installing an elbow im-
mediately after the fan means you are causing turbulence pressure
losses in air with high velocity pressure so the losses will be greater
than if a short length of straight duct between elbow and fan outlet
permitted static regain to occur (Figure 7-10). When elbows can-
not be avoided, take advantage of the centrifugal motion of the air
at the fan outlet by using an elbow as in Figure 7-llb rather than
in 7-lla.8 Sometimes an elbow can be avoided by rotating the fan
housing during installation (Figure 7-llc).
Exhaust Stacks
Static regain can be carried one step further. In some systems a
gradual taper called an evas£ (Figure 7-12) is used to maximize
static pressure regain before the air is discharged from the stack.
At the stack discharge, the air decelerates from the duct velocity to
essentially zero velocity in the ambient environment. Thus one ve-
locity pressure of energy is lost through this deceleration, corre-
sponding to the one velocity pressure of acceleration energy added
as air entered the hoods and was accelerated to the duct velocity.
Slowing the air as much as possible in the stack reduces the magni-
tude of the deceleration loss although a high discharge velocity is
advantageous in some ventilation systems since it helps disperse the
contaminants in the exhausted air. Of course a system with no stack
at all on the fan outlet has a very high deceleration loss since the
velocity pressure is higher at the fan outlet due to the nonuniform
velocity distribution (Figure 7-9). Figure 7-12 shows that a
straight stack does not affect pressure loss in the system while no
stack at all on the fan outlet causes a loss of 0.5 VP. An evase per-
mits static regain to occur and so the fan size can be reduced in sys-
tems with an evase. Every system should have at least a short
straight stack on the fan out1' <
Fans 201
(a)
(b)
Figure 7-10 An elbow at the fan outlet (a) causes a higher pressure loss
than one located downstream (b) because the velocity pressure
is higher at the elbow in (a) than in (b).
Figure 7-11 When elbows cannot be avoided, their direction is important. In
(a) the elbow causes the air to change direction from the curv-
ing due to centrifugal action. The elbow in (b) takes advantage
of this curving airflow while (c) shows how rotating the fan
hou.sinp ran pliminpto »V.~ olK^,,.
-------
202 Handbook of Ventilation
Ol L.
=> OJ
ITJ 4-
Ol O
OJ
t- .
c
I. •!-
0 1
-C
OJ U1
1- 3
Q- O
O C
•r- (O
-2
-4
-6
Calculation based
on Evase outlet
velocity of 2000 ft/min
Evase stack causes
static regain
Straight stack
causes no loss
or gain
No outlet stack
causes 1/2 VP loss
0 2,000 4,000 6,000 8,000 10,000
Fan outlet velocity - ft/min
Figure 7-12 The type of fan discharge stack has an effect on static pressure
regain and pressure loss.
In summary, pressure losses on the discharge side of the fan can
be minimized by reducing the high velocity pressure at the fan out-
let through static pressure regain techniques.
CHOOSING THE RIGHT FAN
To choose the proper fan for a ventilation system you need to
know this information:"
• Air volume to be moved.
• Fan static pressure.
• Type and concentration of contaminants in the air since they
affect the fan type and materials of construction.
• Importance of noise levels as a limiting factor.
Fans
203
Once this information is available the type of fan best suited for
the system can be chosen. There is a variety of different fans avail-
able but they all fall into one of two classes: axial flow fans and
centrifugal fans.
Centrifugal Fans
Centrifugal fans move air by centrifugal action. Blades on a ro-
tating fan wheel throw air outward from the center inlet at a higher
velocity or pressure. Centrifugal fans are usually used in ventilating
systems rather than axial fans because for the volume flow rates
and pressures typical of industrial exhaust systems, centrifugal fans
are quieter and less expensive to install and operate. Centrifugals
cope better with uncertain or fluctuating airflow conditions- than do
axial fans but their efficiency is generally lower. They can be divided
into three classes depending on the shape and setting of the fan
wheel blades. Their applications and advantages overlap but there
are distinct differences (Table 7-2):
Table 7-2 Comparison of Centrifugal Fans
Factor
First cost
Efficiency
Operational stability
Tip speed
Abrasion resistance
Sticky material handling
Forward
Curved
Blade
Low
Low
Poor
Low
Poor
Poor
Backward
Inclined
Blade
High
High
Medium
High
Medium
Medium
Radial
Blade
Medium
Medium
Medium
Medium
Good
Good
Source: Reference 10.
Radial BZade Fans
Radial blade fans (Figure 7-13a) are used for dust systems since
the flat radial blades tend to be self-cleaning. They also have large
openings between blades and so are less likely to clog. They can be
built with thick blades to withstand erosion and impact damage from
airborne solids. Typical operating ranges are from small units up
to fans handling 100,000 ft3/min at 20 in. of water static pressure.
Their major disadvantage is that they are the least efficient fan for
local exhaust systems. The heavy construction adds to their cost.
They are seldom used for non-dust systems.
-------
204 Handbook of Ventilation
(a) RADIAL BLADE FAN
(b) FORWARD CURVED FAN
(c) BACKWARD INCLINED FAN
Figure 7-13 The shape of the fan blades for centrifugal fans. (Source: Ref-
erence 11)
The static pressure rating curve (Figure 7-14) shows that the
operating point for this fan should be selected well to the right of
the peak in the pressure curve. This will avoid pulsing flow since
the fan's pressure capacity varies little over a wide range or airflow
volumes to the left of the peak. The horsepower curve rises in an
almost straight line over the operating range of the fan.
Forward Curved Blade Fans
Forward curved blade fans (Figure 7-13b) are useful when
moving large volumes of air against moderate pressures (0 to 5 in.
of water) with low noise levels. These fans have many cup-shaped
blades that accelerate *r\n nir rind discharge it at a hiphpr
Fans
205
Airflow - ftJ/min—•-
Figure 7-14 Fan curves for a typical radial blade fan.
than the fan wheel tip is moving. The shape of the fan housing con-
verts the high air velocity into static pressure but this is an ineffi-
cient process and so the fan's overall efficiency is low. This poor
efficiency limits its application since some other types of fans are
more efficient in higher pressure systems. The fan's main advantage
is the high blade discharge velocity which means that high air speeds
are achieved with relatively low fan rotating speeds. Since fan noise
is related to fan speed, this feature makes the forward curved blade
fan quieter for some low and moderate pressure systems than other
fans. The high air velocity across the blades precludes its use when
erosive materials are in the airstream.
The static pressure rating curve for the forward curved fan (Fig-
ure 7-15) has a valley caused by blade inefficiency at low air vol-
umes and a peak where air velocity is the greatest that can follow
the contour of the blade surface. Beyond the peak the air starts to
break away from the fan blade. This peak is important in fan selec-
tion since if the fan is operating on the curve near the peak (Figure
7-16), minor changes in system pressure can cause severe fluctua-
tions in air volume through the system. This pulsing flow can occur
with other centrifugal fans but is more severe with forward curved
-------
206
Handbook of Ventilation
O L.
Ol
. s
c o
•r- O.
I <"
i- l-
O.-Q
U X)
•r- C
•!-> 10
C
ID
0
T 1 1 r
Static pressure
1 r
0
Airflow - ft3/min—-
Figure 7-15 Fan curves for a typical forward curved blade fan.
blades. The optimum operating point is well to the right of the peak
to avoid pulsing.
Note also from Figure 7-15 that the horsepower curve rises
sharply with increasing volume. If the system has less resistance
(a lower static pressure) than calculated, the fan will operate at a
higher flow rate through the system. As shown in Figure 7-17,
this can result in excessive power costs since the horsepower curve
rises so sharply. This feature is another disadvantage of the forward
curved fan.
Backward Inclined Blade Fans
Backward inclined blade fans (Figure 7-13c) are used more and
more for handling large volumes of air with little dust since this fan
is more efficient than the forward curved fan. The improved effi-
ciency occurs because the fan blades cause the pressure increase di-
rectly as the wheel rotates; the velocity of air leaving the wheel is
relatively low.11 This low blade discharge velocity is somewhat of
a disadvantage in big fans since bigb rotating speeds are needed to
-velocities.
Fans 207
"4-
O
-------
208
Handbook of Ventilation
Design
Airflow — ft /m1n—•-
o
Q.
01
o
-C
Ol
i.
01
2
O
Q.
Ol
I/)
t-
o
_ 0
0 Design Actual
Airflow - ft3/min—•-
Figure 7-17 The shape of the brake horsepower curve for forward curved
blade fans is important. In (a) the fan operating point and horse-
power based on design calculations is shown. If the airflow
through the installed system is higher than planned (b), the
power consumption is much higher than expected (AH) due to
the rising horsepower curve.
complete operating range. This is an improvement since the back-
ward inclined fan cannot operate smoothly to the left of its static
pressure curve peak. The shape of the blades adds some structural
strength to the fan wheel to help minimize one problem with the
Fans 209
Airflow — ft /min —»-
Figure 7-18 Fan curves for a typical backward inclined blade fan.
(a) Backward inclined
blades
(b) Airfoil\ blades
Figure 7-19 The difference between backward inclined blade fans and airfoil
fans is the shape of the blades.
backward inclined blade fan. Otherwise its application is similar to
the backward inclined fan.
Axial Fans
A screw or propeller action produces airflow in axial fans; the
air travels parallel to the fan shaft and leaves the fan in the same
direction as it entered. The three different types of axial fans share
the advantages of
-------
210 Handbook of Ventilation
efficiency. Their disadvantages are relatively high noise levels, low
pressure capability (less than 10 in. of water static pressure) and
they are not suitable for hot or contaminated air when the fan motor
is installed inside the duct.
Propeller Fans
Simple propeller fans (Figure 7-20) are not used in duct ventila-
tion systems because they do not produce pressure (either positive
pressure or suction). They are suitable for moving large volumes
of air as window fans or roof ventilators where there is no real re-
sistance to airflow.
Tube Axial Fans
Tube axial fans (Figure 7-21 a) are special propeller fans mounted
k
Motor
Figure 7-20 A propeller fan used for
dilution exhaust applications.
Impeller
Motor
-Impeller
Vanes
Motor
(a) Tube axial fan
(b) Vane axial fan
Figure 7-21 A tube axial fan (a) is a special propeller fan mounted in a duct
section. A- vane axial fan has vanes to straighten airflow and
Fans 211
inside a duct. The blades are specially shaped to enable the fan to
move air against low (0 to 3 in. of water static pressure) resistance.
Vane Axial Fans
Vane axial fans (Figure 7-21b) are similar to tube axial fans
but have vanes mounted in the duct to convert spinning air motion
into higher static pressure and also straighten out the moving air.
They are useful in systems with static pressures of about 2 to 10 in.
of water if noise is not a problem. Vane axial fans are available
with adjustable pitch blades for systems with changing static pres-
sure requirements. The fan rating curve can be changed by adjusting
the blade pitch so the fan can operate at different pressure-volume
relationships without sacrificing the fan's high efficiency.13
Rating curves for vane axial fans vary depending on blade shape,
straightening vanes and other factors. In general both the static pres-
sure and horsepower curves (Figure 7-22) exhibit a peak at 50
to 70 percent of the fan's wide open (no resistance) capacity. The
operating point is selected to the right of the peak to avoid pulsing
flow in the system due to small pressure changes similar to the puls-
ing illustrated in Figure 7-16. The horsepower curve also decb'nes
to the right of the peak, showing why axial fans are very efficient.
Airflow — ft3/min-
-------
212 Handbook of Ventilation
To the left of the peak the fan blades become quite inefficient in
moving the air as the volume drops off. This is due to disturbances
in airflow patterns over the blades and causes the steep rise in
horsepower at lower flows. The noise level in the system usually
increases as the horsepower increases.
A combination axial-centrifugal fan is also available from some
manufacturers. This fan has a backward inclined or airfoil blade
wheel mounted in-line in a duct section. The fan shaft is parallel to
the duct and splitter vanes and deflectors divert the air through the
fan wheel then straighten it out downstream to convert the spinning
air velocity into static pressure. The fan has a rating curve that is
similar to the backward inclined blade centrifugal fan (Figure 7-
18), but the unit is more compact and requires less installation space.
It is quieter than pure axial fans.
Throughout this section on the different fan types the shape of
their rating curve has been stressed. The shape of the static pressure
curve and the brake horsepower curve are important because they
illustrate to you what will happen if the pressure drop through the
system is different than the amount calculated during design. As
mentioned earlier, choosing a forward curved blade fan for a system
where the pressure loss could be significantly less than calculated is
a mistake since the power costs rise sharply with high airflow vol-
umes (Figure 7-17). For all fans the shape of the static pressure
curve shows the useful operating range of the fan in volume output
and pressure. Using fan rating tables rather than curves deprives
you of the graphic overview of fan performance but is more con-
venient since the fan manufacturer has selected the right fans from
his stock of different fans and motor sizes. Keep in mind that all
the data points on the tables are taken from individual fan curves
and each fan is capable of a wide range of pressure-volume output
combinations.
FAN NOISE
Fan noise can be a problem with some ventilation systems. Noise
complaints can occur in areas that are served by the system, in loca-
tions near the fan and occasionally in areas remote from any part
of the ventilation system. The best solution to most fan noise prob-
lems is selecting a quiet fan and providing the proper mounting to
minimize noise transmission. Attempts to quiet a noisy fan can be
expensive and ineffective if the noise source is turbulence in the air
moving through the fan.
Fans 213
Turbulent vs. Mechanical Noise
Fans make noise in two ways:"
• Turbulent noise from air moving over the fan blades, impact-
ing the housing and changing direction at the inlet and outlet. This
noise travels through the ducts and into the workrooms served by
the exhaust system. The magnitude of turbulent noise from these
different sources is the primary reason why some types of fans are
quieter than others. Table 7-3 lists the major noise sources for for-
ward curved and backward inclined fans. Minimize turbulent noise
problems by selecting a "quiet" fan from fan catalogs.
Table 7-3 Significant Noise Sources for Centrifugal Fans
Noise
Source
Inlet
Blades
Blade outer
edges
Housing
Cause
Blades cutting air
Air separating from blades
Air from top and bottom of
blades merging
Airstreams changing speed
and direction
Forward
Curved
Blade Fan
X
X
X
X
Backward
Inclined
Blade Fan
X
X
X
Source: Reference 14.
• Mechanical noise from the fan motor bearings and drive, and
also the noise radiated from the fan housing. Vibration noise from
unbalanced moving parts is also a form of mechanical noise. The key
design features to minimize mechanical noise problems are using
flexible connections between ducts and fan to reduce noise trans-
mission along the ducts, and providing an inertial base or vibration
isolators for mounting the fan (Figure 7-23). The area where the
fan is located should be as insensitive to noise as possible. Occa-
sionally a building structure is too weak or too flexible to withstand
the motion of the spinning fan (also called "live load") without ex-
cessive vibration. This problem should be identified before fan loca-
tion is decided.
Turbulent Noise Solutions
The best solution for turbulent noise problems is to select a quiet
fan using fan manufacturers' catalogs. On fan rating tables or dia-
-------
214 Handbook of Ventilation
Flexible wire
to motor
Inertial
base
Flexible duct
(both inlet and
outlet)
Fan inlet
Vibration
isolators
Structure adequate for
"live loads"
Figure 7-23 Illustration of ways to reduce noise and vibration transmitted by
the fan to the rest of the building.
grams (Figure 7-6) the best selections for quiet operations are
indicated. Since the zone of quiet operation usually matches the
maximum operating efficiency range (Figure 7-24), this is a good
way to select fans. Although this procedure helps you pick the quiet-
est fan of that make and model, it does not permit comparison of
different makes or even different type fans (for example a forward
curved blade with an airfoil fan) made by the same company.
Sound Power Level Ratings (noise ratings) for different fans can
be calculated from manufacturers' noise data collected according to
test procedures standardized by the Air Moving and Conditioning
Association. Fan catalogs usually show this information on diagrams
or illustrate the calculations needed to find the noise rating for com-
parison with other fans.
The quietest type of fan is the airfoil backward inclined blade fan
followed by the plain backward inclined blade fan. The noisiest fans
are the radial blade and axial flow fans. The forward curved blade
fan is also noisier than the backward inclined fan except that in
small sizes (generally about 18 in. wheel diameter or less) at low
static pressure the forward curved blade fan is quieter than a similar
size backward inclined fan since the rotating speed of the latter is
much greater. A 15-in. diameter forward curved blade fan operating
against 1.0 in. of water static pressure will pull over 2000 ft3/rnin.
This is enough for many small ventilation systems. However, in
general only airfoil fans should be considered for large installations
Fans 215
20 40 60 80
Airflow - % of maximum flow
100
Figure 7-24 The zone of quietest operation for a fan coincides with its most
efficient operating range.
where noise is a problem unless the airborne solids in the exhaust
stream make the radial blade fan a necessity.
Here are some additional guidelines for fan noise control:13
• Fan noise is proportional to fan static pressure. For critical
noise areas review the system design to reduce resistance so the fan
static pressure requirement is minimized.
-------
216 Handbook of Ventilation
• Use larger, slower fans rather than smaller, higher speed fans
when ether factors are equal.
• Pulsation noise can result from poor fan inlet conditions such
as inlet boxes (Figure 7-8) that load only one side of the fan wheel
with air. Hopefully these poor connections are eliminated as part
of the system design procedure.
• Some older fan design specifications may put a ceiling on fan
outlet velocities to control noise. This concept is no longer valid; fan
noise depends on fan efficiency, not on the fan outlet velocity.
• For supercritical noise applications the fan motor, drive and
bearing noise may have to be considered but if the isolation methods
in Figure 7-23 are used these mechanical noises probably will not
be a problem compared to turbulent noise.
FAN LAWS GOVERN OPERATION
There is a series of statements that explain how fans work. They
are called Fan Laws and are used to construct the fan rating tables
and curves. They also help you decide how to modify a fan's opera-
tion so it works properly in your system. These laws could have
been defined earlier in the chapter but they also make a good sum-
mary of fan selection principles.
Three Key Laws
The three most descriptive laws describe the relationship of vol-
ume, fan static pressure and brake horsepower to fan speed:
• Changes in volume (ft3/min) vary directly with changes in
fan speed. For a given fan doubling the speed will double the vol-
ume output.
• Changes in static pressure vary directly with the square of
changes in fan speed. If you double the fan speed, the static pressure
generated by the fan increases by a factor of four. A corollary to this
law is that static pressure also varies directly with the square of
changes in fan volume.
• Changes in brake horsepower vary directly with the cube of
changes in fan speed. Doubling the speed of a fan increases brake
horsepower by a factor of 8 (2 x 2 x 2 = 8).
All three of these fan laws act together so any change in fan speed
to increase volume output also increases fan static pressure and
brake horsepower. Especially as power costs increase due to rising
electric rates, the jump in brake horsepower should be considered
before increasing fan speed. If a fan is too small to do the job eco-
Fans 217
nomically perhaps it will be less expensive in the long run to replace
it with a larger fan using less power than would be consumed in
speeding up the small fan.
Example: A fan is rated in a manufacturer's catalog as
delivering 10,500 ft'/min of air at 3 in. of water fan static
pressure when running at 400 rev/min and requiring 6.2
horsepower. If the fan speed is increased to 500 rev/min,
determine the volume, static pressure and horsepower as-
suming standard conditions.
Answer: Capacity: Q = 10,500 ( __-- J = 13,125 ft'/min
Static Pressure: FSP = 3
Horsepower: HP = 6.2
500
400
500 V
\ '= 4.7 in. of
water
*
= 12.1 horsepower
There are additional fan laws besides the basic three. They deal
with the effect of different fan size and air density on fan perfor-
mance. They are summarized in Table 7-4.
Table 7^1 Fan Laws
Variables"
Fan Speed Fan Size
Air Density
Effect
Varies
Constant
Constant Constant
Varies0 Constant
Constant
Constant Varies
• Volume varies as fan speed
• Pressure varies as square of fan
speed '
• Power varies as cube of fan
speed \
• Volume varies as cube of wheel
diameter
• Pressure varies as square of
wheel diameter
• Tip speed varies as wheel
diameter
• Power varies as fifth power of
wheel diameter
• Volume constant
• Pressure varies as density
• Power varies as density
"Assumes
change).
"Assumes
same fan
constant system (hoods, and duct lengths and diameters do not
constant fan proportions as when selecting different wheel size of
type.
-------
218 Handbook of Ventilation
IS THE FAN WORKING PROPERLY?
Since the fan is the only moving part in most ventilating systems,
it often receives a lot of scrutiny when the system is not working
properly. This chapter has focused on providing background on the
different types of fans and how fans are selected for ventilation sys-
tems. If you have a fan in an existing system and the system is not
working properly, you have a different problem. Chapter 10 deals
with testing all ventilation system components including fans and
Chapter 11 tells you how to use test data in reviewing system per-
formance to correct ventilation system problems. Chapter 9 offers
guidelines for designing a system that will do the job at minimum
cost.
REFERENCES
1. AMCA Bulletin 110. "Standards, Definitions, Terms, and Test Codes for
Centrifugal, Axial and Propeller Fans" (Park Ridge, Illinois: Air Moving
and Air Conditioning Association, Inc., 1952).
2. Catalog, New York Blower Company, Chicago, Illinois.
3. Catalog, Chicago Blower Corporation, Glendale Heights, Illinois.
4. ACGIH Committee on Industrial Ventilation. Industrial Ventilation—A
Manual o/ Recommended Practice, 14th Ed. (Lansing, Michigan: American
Conference of Governmental Industrial Hygienists, 1976).
5. Geissler, H. "Purchased Fan Performance," Reprint No. 5483 (Pittsburgh,
Pennsylvania: Westinghouse Electric Corporation, 1959).
8. Trickier, C. J. "Field Testing of Fan Systems," Engineering Letter No.
E-3 (Chicago, Illinois: The N.Y. Blower Company).
7. Trickier, C. J. "Effect of System Design on the Fan," Engineering Letter
No. E-4 (Chicago, Illinois: The N.Y. Blower Company).
8. Tracy, W. E. "Fan Connections," Reprint No. 5100 (Pittsburgh, Pennsyl-
vania: Westinghouse Electric Corporation, 1955).
9. National Institute for Occupational Safety and Health. The Industrial En-
vironment—It* Evaluation and Control (Washington, D.C.: U.S. Govern-
ment Printing Office, 1973).
10. Cheremisinoff, P. N. and R. A. Young. "Fans and Blowers," Pollution
Engineering 6, No. 7 (1974).
11. Trickier, C. J. "Fundamental Characteristics of Centrifugal Fans," Engi-
neering Letter No. E-l (Chicago, Illinois: The N.Y. Blower Company).
12. Rogers, A. N. "Selection of Fan Types," Reprint No. 5312 (Pittsburgh,
Pennsylvania: Westinghouse Electric Corporation, 1957).
13. American National Standard Z 9.2-1971. "Fundamentals Governing the
Design and Operation of Local Exhaust Systems," (New York, New York:
American National Standards Institute, 1972).
14. Trickier, C. J. "How to Select Centrifugal Fans for Quiet Operation," Engi-
neering Letter No. E-13 (Chicago, Illinois: The N.Y. Blower Company).
-------
ITEMS
Fans - Special Report
Robert Aberbach
Power Journal
-------
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Fans move gas at relatively low differential pressures
V
/
13
v>
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O>
it
r.. _—-_-_.. Fan cycle
\ ^DiSChOfQf ^phOSf I
\ \-r
\ Sue/tor phase — "*" \
\
X. . . . _._.V
. Compressor cycle
Volume-^
Compression phase
Fan disch
FANS, BLOWERS, COMPRESSORS all move
air, but at greatly different pressures. Fan
pressure range is from a few inches of water
up to about 1 psi. Blowers work to about 50
psi; compressors span about 35 psi and up.
Illustration above shows how p-v diagrams for
fans and compressors differ. Fan overcomes
mainly static forces. Total pressure includes
velocity head, as shown at right.
Flow direction
How fans work
A fan can be defined as a volumetric machine which, pump-
like, moves quantities of air or gas from one place to
another. In doing this it overcomes resistance to flow by
supplying the fluid with the energy necessary for continued
motion. Physically, a fan's essential elements are a bladed
rotor (such as an impeller or propeller) and a housing to
collect the incoming air or gas and direct its flow.
Energy Factors. Since any device that makes something
else move is doing work, a fan demands energy itself to oper-
ate. The amount of energy required depends on the volume
of gas moved, the resistance against which the fan works,
and machine efficiency.
To better understand these energy relationships, look at
the pressure-volume diagram shown above for a typical fan
and a compressor. The work which the fan does is, of course,
represented by the area enclosed by the fan cycle. The
diagram also shows that the pressure changes involved are
relatively minor. Unlike blowers or compressors which work
against larger heads and where pressure increases are signif-
icant, the bulk of fan work goes into moving gas at relatively
low pressures.
Compression. Note too, from the diagram, that volume
remains virtually constant during the compression phase. This
happens because the change in specific weight (pounds per
cubic feet) of the gas between fan inlet and discharge is
negligible — less than 1.% fora 1-p'si pressure increase. Thus,
when analyzing fan power, the volume change during the
compression phase can be thought of as zero. Work accom-
plished b) the fan cycle is accurately expressed by this equa-
tion: Work = Ap x V; where Ap is the rise in pressure
across the fan and V is the volume of air or gas moved.
Pressure difference (either static or total, and often called
simply "head") is usually measured in inches of water. Static
pressure is the force per unit area exerted on walls, ducts,
piping, etc— just like the steam pressure on the inside of a
boiler drum. Static pressure is what overcomes the resistance
of ducts, fuel beds, filters, grates, etc.
Total pressure represents the combined effect of static
pressure and velocity pressure. Velocity pressure is that
head over and above static, caused by the movement of the
air or gas. The cutaway drawing above shows how mano-
meters inserted in a fan discharge duct measure static head,
velocity head, and total head.
Power. Starting with the equation for fan work and some
basic physical constants— and throwing in some simple math
—we can develop the equation that expresses air horsepower:
V x H
Ahp = 63<;g : where V is the volumetric flow through
the fan in cu ft per min and H is the head or pressure dif-
ference (in inches of water) across the fan. The air horsepower
may also be designated as either static or total. Since the
resistance to be overcome in fan applications is primarily
static pressure, the fan pressure developed is usually referred
to in terms of static head. On this basis, the calculated fan
power is known as static air horsepower (Ahps). When the
power calculations are based on total head, fan power is re-
ferred to as the total air horsepower (Ahp, ) and is equiva-
lent to the power oulpui.
FANS • A SPECIAL REPORT
-------
f"
t""
fe'*
-_?-.• "•.••tf.-rt-s-1.-. j^ijjT™-:? >v
- ~' "-virr-?- n^~ 5" - "--
Here are fan en
«r bortepowtr-requtred horsepower at
tOO% fan efficiency to movefltven quan-
- tfty of air against given pressure
.BrmJce ho«apoww*f any rnachbie fe *e
actual horsepower that It develops, Ifs
greater than the theoretJcaJ or air horse-
power by the tosses incurred vtaJhe iarr
* rough frtetion, teakage, «te.
Fan inl*t area—inside area of the inlet
collar on the fan
Fan outlet ana—inside area of fan out-
let (flange on outlet side)
Free demwy—theoretical operating con-
dition when static pressure and resist-
ance in the system equal zero
Mechanical efficiency of a fan Is the ratio
of power output to power input
et Input fe trie*KWBBJ»W« delivered
~
j>ow«r output, expressed ta horsepower,
varies Jwrth fen
\ dry. fit TOP i
t sea level conditions (barometric pres-
sure of 29.92 in. Ha) :^ , , .
Standard afr density te 0.075 tb percuft
Standard flue-gas density to aOTB ft par
cu ft at sea level (2S.92 to. *Hg baromet-
ric pressure) and 70 F
Static efficiency of a ten is the mechani-
cal efficiency multiplied by the ratio of
static pressure to the total pressure y
Yetocttjr JXBMUI* te fin Iftte&c
per vnR volume of flowtog ak. For atend-
ard »lr ft equals ;i,;-.
In similar fashion, static efficiency is associated with Ahps,
while mechanical efficiency (or total efficiency) is associated
with power output. Each is compared with shaft horsepower
—or the power input to the fan—to arrive at an actual effi-
ciency figure.
The equations expressing power relationships are sum-
marized below. Also, a sample problem is worked through
to show how the equations are applied. The typical fan
characteristic curve below shows how these physical proper-
ties are related to one another.
Density. Pressure and temperature of the air or gas also
influence power output, efficiency, etc. This comes about
because pressure and temperature affect gas density, and a
change in density changes total and static pressure and their
subsequent conversion into inches of water at standard con-
ditions. These relationships will be discussed in more detail
later; but for now just remember that head and horsepower
vary inversely as absolute fluid temperature and directly as
absolute fluid pressure (or directly with fluid density) — and
that adjustments often must be made for pressure and tem-
perature variations when calculating performance or select-
ing a fan for a particular application. Equations for pressure
and temperature corrections are also given below.
Other physical relationships—speed for instance—affect fan
performance too. These parameters are covered under the
fan laws, on p S • 7.
Basic equations help calculate fan performance
Fan efficiency, horsepower, pressure,
etc are shown on typical performance
curves at right. Here are the equations
used to calculate these variables:
Horsepower and efficiency
Total head = static head + velocity head
or, Ht = H. + H,
Static air horsepower (Ahp.) = -„ '
where V = flow in cfm and head is meas-
ured in terms of in. of water.
V x H
Power output (Ahp,) =
Static efficiency (E.) =
Ahp.
Mechanical efficiency (E,) =
Power input
Ahp,
power input
Pressure, temperature corrections
Head and horsepower vary inversely as
absolute fluid temperature and directly
as absolute fluid pressure. Or,
P T
Corrected head (H.) = Hb ^-^-
r b I •
P T
Corrected hp (hp.) = hpb D' T
r t I «
where P is absolute pressure and T is
absolute temperature in degrees Rankine.
Subscript "a" indicates condition after
correction, subscript "b" before.
Example. A fan develops 4.5 in. static
pressure and 0.7-in. velocity pressure
when flow is 8000 cfm of 100 F air; shaft
hp is 7.2. Calculate the static and me-
chanical efficiencies:
,_Jotol pressure
Volume -*•
8000 x 4.5
Static air hp = —^j^— = 5.7
6356
E. =|r= 79%;Ht
4.5+ 0.7 =5.2 in.
Power output =
6.55
Et = 7-5- = 86%
8000 x 5.2
6356
= 6.55 hp
POWER • MARCH 1968
-------
Axial-flow fan moves air or gas in a straight-through path
PROPELLER TYPE is the simplest axial-
flow fan. Besides direct-connected motor
shown, the fan may also be belt driven
HELICAL-FLOW pattern caused by
screw-like motion of rotating blades is
typical of the discharge from propeller
Guide vane
TUBE-AXIAL FAN mounts propeller in
cylinder. Vane-axial fan, bottom, straight-
ens the helical flow, also ups efficiency
Two basic designs are modified by blade shapes, attachments
There are two basic classes of fans: the axial-flow design
which moves gas or air parallel to the axis of rotation; and
the centrifugal-flow (or radial-flow) type which moves air or
gas perpendicular to the axis of rotation. Axial-flow fans
are better suited to low-resistance applications; centrifugal-
flow fans usually take care of the higher head jobs.
Propeller action. The axial-flow fan uses the screw-like
action of a multibladed rotating shaft, or propeller, to move
air or gas in a straight-through path.The leading edge of each
rotating propeller blade bites into the air or gas, which is
then propelled through the fan and discharged in a helical
pattern (shown above) by the blade's trailing edge.
Guide vanes. Two common varieties of axial-flow fans are
shown above. The tube-axial design is nothing more than a
propeller fan enclosed in a cylinder which collects and di-
rects air flow. In a vane-axial fan - an extension of the tube-
axial design — air-guide vanes on the discharge side of the
propeller straighten out the air-flow pattern and increase the
static pressure.
Again, there's overlapping, but the approximate pressure
range of a propeller fan is 0 to 1 in. of water; a tube-axial fan
moves air at pressures ranging from 1A to 2Vi in. of water.
A vane-axial fan can handle pressures ranging from a low of
Vi in. of water and reaching about 6 in.; special designs go
even higher.
Axial-flow fans can have widely differing characteristics,
depending on the design, blade type, ducting, etc. Generally,
they are applied where outlet velocities required are higher
than what centrifugal fans can produce, and pressure de-
mands are not above axial-flow limits. Usually vane-axial
fans are nonoverloading. Typical characteristic curves are
described and illustrated at bottom of facing page.
Centrifugal fans are advantageous when the air must be
moved in a system where the frictional resistance is relatively
high. The air drawn into the center of the revolving wheel
turns 90 deg, and enters the space between the wheel's blades.
The bladed wheel whirls air centrifugally between each pair
of blades and tosses it out peripherally at high velocity and
increased static pressure. As this happens, more air is sucked
in at the eye of the wheel. As the air leaves the revolving
Efficiency of propeller fan depends largely on inlet design
AIR RECIRCULATION around blade tips greatly reduces the
efficiency of a propeller type fan. Installing a curved, orifice-
like ring at the blade tips reduces the amount of recirculation
and improves efficiency. An angle-shaped ring just about elim-
inates recirculation. Optimum efficiency is achieved with a
combination curved,angle-shaped ring wh,ch also smooths flow
FANS • A SPECIAL REPORT
-------
Centrifugal-flow fan takes air in at eye, spins it out at right angles to the inflow
CENTRIFUGAL (or radial) fan turns air
90 deg between entry and discharge.
Housing directs flow, increases pressure
StrtomlmaiblrfoU)
BLADE TYPES in fan take many shapes.
Air foil blade is most efficient. Back-
ward curved blade picks up little dirt
AT SAME TIP VELOCITY (Vb) each type
blade produces different air velocity
(VJ. Vector Vab is V, relative to the blade
blade tips, part of its velocity is converted into additional
static pressure by divergence of the scroll-shaped housing.
Blade shapes employed are of three basic types: forward-
curved, straight, or backward-curved. Other configurations,
including an airfoil design, are merely variations of these
designs. Some commonly used shapes are shown in the cen-
ter illustration above.
In general, blade type limits top fan speed. Backward-
curved blade machines can operate at a relatively higher
speed than forward-curved designs. Type of blade employed
also depends on space limitations, allowable noise levels, effi-
ciency demanded by specified load conditions, and desired
fan performance characteristics. While mechanical efficien-
cies are much the same for each different blade type, differ-
ences in horsepower and pressure-volume relationships can
be significant.
Air velocity. The effect of different blade shapes on air ve-
locity is illustrated above. Note that a forward-curved blade
imparts a greater absolute velocity to air leaving the blade
than does a backward-curved blade at the same tip speed.
Thus, at the same operating speed, a backward-curved blade
develops less velocity head, and converts more energy into
static head.
A forward-curved fan produces a lower static head — but,
because it produces higher air velocity, is particularly suited
for handling large volumes of air in low-resistance systems.
Straight-blade designs have velocity characteristics that fall
between the two curved styles.
Comparison. Curves on the next page compare the effect
of different blade shapes on operating characteristics of cen-
trifugal-flow (radial-flow) fans. For each curve the fan is
considered running at a constant speed. Also; each curve
applies to the complete range of sizes for each fan type -
so long as different sizes are proportionally shaped with all
dimensions varying as a function of diameter.
Restrictions. To fully appreciate the significance of fan
characteristic curves, you must bear in mind that every fan
is restricted to that performance defined by its curve. The
fan must operate at a point that lies somewhere on the char-
acteristic plot.
For example, if the head required for a given volume is
less than that specified by the curves, additional resistance
must be placed in the system (a damper, serving as a throt-
tle, can handle this chore). Otherwise the fan will put out
increased capacity until it reaches a point on the character-
istic curve where the head matches system resistance. The
fan has no choice, it must operate at this point on the curve
where head, capacity and system resistance are in balance.
10 20 30 40 50 60 70 80 90 100
% of wide open volume
Performance varies with type of fan
Vane-axial fan performance characteristics are shown in graph
at left. Depending on the design, horsepower curve may in-
crease with flow (as shown at left), or may be flat and self
limiting (decreasing as capacity approaches 100%). Nonover-
loading (self-limiting) characteristic permits use of motor hp
size close to the actual hp required for operating speed. Guide
vanes improve horsepower and efficiency.
With axial-flow fans, maximum efficiency occurs at higher
percent output than with centrifugal fans. (Curves for centri-
fugal fans are shown on next page). A comparison of axial
and centrifugal types shows that the axial-flow propeller design
generally has lower horsepower and pressure curves; efficiency
curve is flatter and maximum efficiency is usually lower.
POWER • MARCH 1968
S • 5
-------
Blade shape affects radial fan efficiency
Characteristic curves
pinpoint fan operation
10 20 30 40 50 60 70 80 90 100
Flow volume, %
FORWARD-CURVED blade has peak efficiency near the
point of highest pressure. Since hp Increases rapidly as
capacity increases, there's danger of overloading motor
if system resistance has not been accurately calculated
._. ^. Total pressure
Si otic pressure ~~
10 20 30 40 50 60 70 80 90 100
Flow volume, %
BACKWARD-CURVED blade shows decreasing pressure
as fan capacity goes up; this expands range of stable
operation. Hp characteristic is nonoverloading. The fan
hits its peak efficiency at a point close to maximum hp
130
MO
S 90
O
O_
70
_ -— Jotol frttture
Static pressure —
Totalefficiency _
10 20 30 40 50 60 70 80 90 100
Flow volume, %
STRAIGHT BLADE characteristics are somewhere be-
tween the forward and the backward types. Maximum ef-
ficiency occurs near maximum pressure. An oversized
motor may be needed to prevent overload as hp Increases
A fan's characteristic curve, like those at left for a radial fan
or that for a vane-axial fan on the previous page, defines ex-
actly how the machine will perform. The fan has no alterna-
tive but to operate at a point on its characteristic curve where
operating factors (head, speed, output) are in equilibrium.
To better understand this concept, let's examine the graphs
on the facing page. A typical fan characteristic curve is
shown on the left. Points A, B, C, D represent fan system re-
quirements (the volume, pressure, etc) calculated at different
operating conditions for a particular application. The line
connecting these calculated points is the curve representing
system resistance. The point where it intersects the fan's
static-pressure characteristic curve at a given fan speed, is
the point where the fan operates.
Every fan operates only along its characteristic curve.
So if there is any error made in calculating point D (in vol-
ume, pressure, etc.), then this point will not fall on the fan's
characteristic curve and the fan will not be able to satisfy
the performance requirements of this application when oper-
ating at that particular speed.
For example, let's say that your calculation of required
system capacity is in error, and that you actually need 10%
more volume at the same pressure. To provide sufficient vol-
ume, point D is displaced to the right. But this shift along
the curve to obtain a 10% increase in volume, drops avail-
able pressure 14%. The fan is unable to meet both specified
system demands; volume or pressure must be sacrificed.
Similarly, if the system capacity is correct but the calcu-
lated pressure is 10% low, then you have to give up about
10% capacity to get the pressure actually required.
Speed variations. For a given fan, a family of characteris-
tic curves can be obtained by varying fan speed. The nature
of each curve remains the same since the change in operating
speed merely shifts the curve by a proportionate amount.
If system resistance is plotted on the same grid as the fam-
ily of curves for different fan speeds, we have a graph like
the one shown on the right side of the facing page. If the fan
runs at constant speed, any volumetric output less than that
indicated for the intersection of the system resistance and
speed curves, will be produced at higher pressure; this neces-
sitates throttling the output — a waste of energy. When the
fan is arranged for variable speed operation this squandering
of energy can be prevented by simply running the fan at a
lower rpm.
Oversizing. Frequently, in order to provide excess capac-
ity, an engineer will specify a volume and/or pressure that's
larger than the amount actually needed. Additional power,
of course, is needed to drive the larger fan; but the fan will
then be able to provide additional capacity without pressure
loss and without overloading the motor.
Fan laws. The way a fan will be affected by a change in
any operating condition can be predicted by a set of rules
known as the fan laws. These are summarized in the pane!
at right and apply to fans of the same geometric shape and
operating at the same point on the characteristic curve.
Also summarized are the mathematical concepts of speci-
fic speed and specific diameter. These relationships help ex-
plain how geometrically similar fans operate and are used
for fan design and fan selection (see page S* 12).
S • 6
FANS • A SPECIAL REPORT
-------
50 70
Flow, %
110
SYSTEM-RESISTANCE curve and fan static pressure charac-
teristic intersect at point where fan supply balances demand
Flow output -*•
FAN SPEED can be varied so that output pressure matches
system resistance for desired cfm of air; this conserves energy
How to calculate fan performance for different operating conditions
(1) Given fan size, system resistance,
and air density
When speed changes:
a. Capacity varies directly with speed
or,
Qz
b. Pressure varies as speed squared
°r'"P7 * \rpmj /
c. Horsepower varies as the speed
cubed or.
When pressure changes:
a. Capacity and speed vary as the
square root of the pressure or,
»pmi Q\_
rpm2f 02
b. Horsepower varies as the pressure
to the (3/2) power or,
(3) Constant pressure, density, rating
and a geometrically similar fan
When fan size (wheel diameter) changes:
a. Capacity and horsepower vary as
wheel diameter squared or,
Si hp, /dia,
~
b. Speed varies Inversely as wheel
When speed and wheel diameter change-'
a. Capacity varies as the product of
speed and wheel diameter cubed
Q, jPJDx ($!*
Wl & * ipm, * idia
b. Pressure varies as the product of
speed squared and wheel diameter
scared or. ^
c. Horsepower varies as the product
of speed cubed and the fifth power
of wheel diameter or,
{rpmi \* /dtaA*
rpmj / xdiflj /
d. Horsepower also vanes as the prod-
uct of capacity and pressure or,
ftp i _ Qi », PI
f3) Constant pressure
•a. Speed, capacity and horsepower
/vary Inversely with the square root
of density or,
rpirii jji hp|
And Oien speed, capacity and horse-
power vary inversely with square
TOot-of barometric pressure and di-
rectly as the square root of abso-
fafte temperature:
<4) Constant speed and capacity
a. Horsepower and pressure vary di-
mcUy wtth density and barometric
pressure and Inversely with abso-
lute temperature:
top i ^j te *ii _ 'El _ T*
»IPa' E <»2 bz T,
pS) Constant amount by weight
*. Capacity, speed and pressure vary
tevfiEsejy with density
-------
"'?rs%1
FORCED-DRAFT FANS at this generating station provide the
draft tor two boilers. Each fan puts out nearly 490,000 cfm and
runs at 890 rpm; wheel diameters top 98 in. Adjustable
inlet vanes (see p S-18) are used to control the fan volume
Fan selection and application
Selecting fans for an energy system is no easy task. In addi-
tion to choosing the right size and type, noise and vibration
problems must also be factored in. Fans are literally the
heart of air-moving systems for a wide range of applications
including the general areas of steam generation, heating,
ventilating and air conditioning.
Steam Generators use fans to mechanically provide the
draft necessary to maintain combustion in the furnace. The
system may be just forced draft (which pushes air into the
combustion chamber); or induced draft (which pulls excess
air and the products of combustion through the combustion
chamber) used in conjunction with forced draft.
Forced draft. Determining how much air a forced-draft
fan must supply involves more than just calculating the
fuel's excess air requirement. Sources of infiltration (which
reduce the amount of air needed) and leakage (which adds
to air demands) must be taken into account.
Overfire air. When coal is burned on stokers, some air is
supplied over the fire. To assure ample pressure at all loads,
overfire air is usually supplied by a separate fan, drawing
its supply from the room or from the preheated air supply
(choice depends on each plant's design, arrangement, etc).
Air-cooled walls may increase or decrease forced-draft
requirements, depending on how the air is moved. If the fan
discharges air through the walls, there can be considerable
leakage into the furnace and boiler room, while infiltration
from the boiler room to the furnace is reduced. If the
forced-draft fan exhausts air from an air-cooled wall,
there'll be considerable infiltration from the boiler room.
Induced-draft fan, which is located at the outlet of the
steam generator, handles gases produced by combustion of
the fuel and any air infiltration that occurs up to the fan
inlet — including any that leaks directly into the air heater.
Since the total amount of air infiltration can be significant
(as much as 20% of the gas weight handled), determining
ID fan requirements is an art.
From experience and a thorough knowledge of the plant,
come estimates of the total gas weight to be handled. Vol-
ume at fan inlet is calculated from the weight determina-
tion. Additional induced-draft capacity (about 20%) is usu-
ally provided so that the boiler can operate with high ex-
cess air if necessary and to compensate for any future drop
in fan efficiency as the blades wear and collect dirt.
Other uses. Coal-fired steam generators often use air from
the primary fans to mix air and coal in the proper ratio for
burning. Additional fans take the air used to dry coal at the
pulverizer, and vent it to atmosphere or to the furnace.
Overfire air fans, besides improving combustion, help
eliminate smoke by creating enough turbulence in the fur-
nace to keep the gas from arranging itself in layers. Fans
in pulverizer exhausters handle a mixture of coal and air
and move it on to the furnace.
Cooling towers, which provide for the evaporative cool-
ing of water by dropping it through an air stream, use fans
to keep the air moving. A large volume of air at a relatively
low velocity (less than 2000 fpm) is generally called for.
Pressure drops are also low —usually under Vi in. of water.
FANS • A SPECIAL REPORT
-------
ROOF VENTILATOR handles local air-removal jobs. Butterfly
dampers work automatically to keep weather outside. Remotely
controlled dampers are also available. Units of this general
design can handle up to 40,000 cfm with a 5-hp, 600-rpm fan
INTAKE UNIT warms air brought into the plant. This one has
steam-fired heater and thermostatic controls; single unit can
deliver up to 32,000 cfm and work at steam pressures to 150
pounds. The maximum output In Btu/hour is about 4Vi million
Cooling-tower designs may be either forced or induced draft.
Induced-draft towers generally employ propeller fans;
forced-draft towers use both radial- and axial-flow design.
Ventilation. As long as the weather remains uncertain,
natural ventilation will never be 100% reliable; mechanical
ventilation is a must. Basically, ventilation involves replac-
ing contaminated air with fresh air, and that means provid-
ing exhaust and make-up. It all boils down to a lot of air
being moved around, and that's where fans enter the picture.
Frequently, local units — like the roof ventilator shown
above —handle air exhaust. Other times it may be more
practical to exhaust through a central duct system. Some
intake units (above right) also heat the air as it's brought
into the structure. Heating load handled by the central heat-
ing plant is thus reduced.
Ventilation requirements in power plants are somewhat
unusual since to some extent draft fans can meet the air
change demands. However, additional mechanical ventila-
tion is usually needed for confined areas and hot spots.
Air conditioning systems, in the most complete sense, are
designed to control the temperature, humidity, and clean-
liness within a given space. The ultimate reason may be per-
sonnel comfort, product protection, process regulation, etc.
The equipment used ranges- from packaged air condi-
tioners to large custom-designed units integrated into a huge
central system serving an entire complex of buildings. What-
ever the size or however complex the air conditioning sys-
tem may be, a fan — or a combination of fans'— handles the
job of moving the air.
In the case of a packaged unit the fan may be included
as an integral part of the equipment. In large custom-de-
signed plants a centralized -Kn system is often used to dis-
tribute conditioned air. Depending on the type of system,
separate exhaust and return air fans are also used in addi-
tion to the normal supply fans.
AIR CONDITIONING SYSTEMS use fans for the circulation of
air. Centrifugal designs are generally employed In larger ca-
pacity systems. Fans like the one above can deliver up to
about 500,000 cfm of air at 7-in. (WG) static pressure. A
double inlet design is often used for increased total capacity
POWER • MARCH 1968
S • 9
-------
Fans I and 2 in series
System
resistance
Fans I and
-------
Overall efficiency varies with pressure drop through the ductwork
The resistance a fan encounters in a system is
largely a function of the duct work through which
the air or gas travels. Duct resistance is espe-
cially significant in air conditioning and ventila-
tion systems where the fans may have to move
air a considerable distance through circuitous
paths before it reaches tfie point of end use.
Resistance to air flow in a system, generally
measured in inches of water, is called static
pressure. It equals the sum of all the pressure
losses due to friction through the ductwork in-
cluding straight sections, restrictions and turns.
The two graphs at the right enable you to find
pressure loss when air flows through a straight
section of duct. The upper graph gives the pres-
sure drop in terms of inches of water per 100 ft
of duct. This is based on the use of standard
round sheet metal duct with air at 70 F. Varia-
tions in air temperatre of ±20 F do not meas-
urably affect the results.
If the duct is rectangular, the lower chart per-
mits conversion to the equivalent round size —
or lets you quickly determine what rectangular
size will match a given size round duct.
Example. Let's say that your air-moving re-
quirements are 10,000 cfm at a maximum veloc-
ity of 2000 fpm. Using the upper graph, follow
the 10,000 cfm line horizontally to the right until
rt intersects the diagonal line labeled 2000 fpm.
The point of intersection nearly falls on another
diagonal line representing a 30-in. diam duct.
This size meets the requirements of our example.
Dropping vertically from this point of intersec-
tion you also find that the pressure drop for this
size duct is 0.15 in. (WG) static pressure per 100
ft of straight length.
Turns in the ductwork produce further signifi-
cant losses. For Instance, a 90-deg round elbow
may cause a friction loss equal to as much as
a 60-diam length of straight duct.
The graph below gives you an idea of what
the equivalent straight line loss is, due to various
types and sizes of elbows.
« 60
ID 2.0 4.0
Radius ratio, R/D
6.0 6.0 10.0
100,000
80,000
60,000
40,000
30,000
20,000
0.01 0.02 0.04 0.06 O.I 0.2 0.4 0.6 I 2 4
Pressure loss, inches of water per 100 feet
DUCT DESIGN is a major factor in system resistance. Graph gives pressure
loss for round ducts as a function of air flow and velocity; text explains use
ELBOWS increase system pressure drop. Graph
gives ecj'Jvalent straight lengths for some sizes
3 4~5~~6 8~10 20 30 40 50 60 80100
One side of rectangular duct (B)
RECTANGULAR DUCT dimensions can be converted to the equivalent size
round duct with the above chart. Use round duct size to find pressure drop
POWER • MARCH 1968
S • 11
-------
Selecting size and type of fan comes next
In order to select the proper fan for
your application, you must know what
the air requirements are. The basic in-
formation needed is fan capacity (in
cfm) and pressure (in inches of water
at standard atmospheric conditions).
Operating temperature and barometric
pressure must be known so that density
corrections can be made. Armed with
this data, you can pick the proper size
and type of fan for your application.
Special considerations such as ex'reme
temperature, corrosive conditions, noise
limitations, etc. may then modify this
choice.
Capacity and density. Whether it's
air conditioning, air for combustion,
exhaust, ventilation or any other proc-
ess, the amount of air or gas demanded
by the system is the first parameter for
fan selection.
Next, corrections must be made for
actual operating conditions; tempera-
ture, barometric pressure, and altitude
all affect density which in turn affects
the system's static pressure. If the
chemical make-up of the gas varies
greatly from air, consult a chemical
handbook for the density data. In the
case of air the tables below give values
of barometric pressure and air density
at various altitudes and temperatures.
These tables contain all the data neces-
sary for density corrections. Mathe-
matically, remember that density varies
inversely with the absolute tempera-
ture, and directly with the barometric
pressure. Static pressure varies directly
with density.
Example 1. At standard conditions,
(70 F, 29.92 in. Hg) gas density is 0.080
Ib per cu ft. What is its density at 340 F
and at an elevation of 2500 ft? First,
from the tables, at 2500 ft elevation
the barometer reading is 27.31 in. Hg.
Thus, the gas density at the conditions
given in this example is:
0.080 X
460 + 70 ^ 27.31
X
29.92
460 + 340
or, 0.049 Ibpercuft.
Example 2. At 70 F, and at sea
level, air density is 0.0750 Ib per cu ft.
What's its density at 300 F and the
sanne elevation? Again from the tables,
the answer is 0.0523 Ib per cu ft.
Example 3. At standard conditions,
gas density is 0.076 Ib per cu ft. Oper-
ating conditions are 380 F, 2200-ft
elevation, and 10-in. (W.G.) static pres-
sure. First, correct static pressure to
standard conditions. From the tables,
at 2200-ft elevation barometric pres-
sure is 27.62-in. Hg. Then at operating
conditions the gas density is:
0.076 X
460 + 70 27.62
460 + 380 ^ 29.92
or, 0.0445 Ib per cu ft.Therefore, static
pressure at standard air conditions is:
10 X
0.0750
0.0445
= 16.8 in. (W.G.)
Pressure. To determine the actual
static pressure at which the fan oper-
ates, allowance must be made for pres-
sure drop through various parts of the
system and —as shown in the example
— correction must be made for density.
If granular or stringy material is han-
dled by the system, this must be con-
sidered in determining static pressure
Temp. F Density
Temp. T Density
Ttmp. F Density
0
5
10
15
20
25
30
35
40
45
50
55
60
65
70
75
80
85
90
95
100
105
110
115
120
125
130
135
140
145
150
155
160
165
170
175
180
185
190
195
200
.0864
.0855
.0846
.0836
.0828
.0819
.0811
.0803
.0795
.0787
.0779
.0772
.0764
.0757 i
.0750 E.
.0743 fe
.0736 f"
.0729 f-
.0723 fe
.0716 K
.0710 gr
.0704 F
.0698 F-
.0692 ?•
.0686 g
.0680 f
.0674 I1'.
.0669 t
.0663 f-
.0657 ?.
.0651 fr.
.0646 K
-0641 L
.0636 »
.0631 E
.0626 P
.0621 E
.0616 fe
.0611 K
.0607 5-
.0602 f
205
210
215
220
225
230
235
240
245
250
255
260
265
270
275
280
285
290
295
300
305
310
315
320
325
330
335
340
345
350
355
360
365
370
375
380
365
390
395
40P
405
.0598
.0593
.0589
.0584
.0580
.0576
.0572
.0568
.0564
.0560
.0556
.0552
.0548
.0545
.0541
.0537
.0534
.0530
.0527
.0523
.0520
.0517
.0513
.0510
.0507
.0504
.0500
.0497
.0494
.0491
.0483
.0485
.0482
.0479
.0477
.0474
.0470
.0467
.0464
.0462
^0459
410
415
420
425
430
435
440
445
450
455
460
465
470
475
480
485
490
495
500
505
510
515
520
525
530
535
540
545
550
555
560
565
570
575
580
585
590
595
600
605
610
.0456
.0454
.0451
.0449
.0446
.0444
.0441
.0439
.0437
.0434
.0432
.0429
.0427
.0425
.0423
.0420
.0418
.0416
.0414
.0412
.0410
.0408
.0405
.0403
.0401
.'0399
.0397
.0395
.0394
.0392
.0390
.0388
.0386
.0384
.0382
.0380
.0379
.0377
.0375
.0373
.0372
Elevation,
ft
0
100
200
300
400
600
600
700
800
900
1000
1100
1200
1300
1400
1500
1600
1700
1600
1900
2000
2100
2200
2300
2400
2500
2600
2700
2800
2900
3000
3100
3200
3300
3400
3500
3600
3700
3600
3900
Barometer, deration,
ta.-H£ ft
29.92
29.81
29.71
29.60
29.49
29.38
29.28
29.17
29.07
28.96
28.86
28.75
28.65
28.54
28.44
28.33
28.23
28.13
28.02
27.92
27.82
27.72
27.62
27.52
27.41
27.31
27.21
27.11
27.01
26.91
26.81
26.72
26.62
26.52
26.42
26.32
26.23
26.13
26.03
25.94
4000
4100
4200
4300
4400
4500
4600
4700
4800
4900
5000
5100
5200
5300
5400
5500
5600
5700
5800
5900
6000
6100
6200
6300
6400
6500
6600
6700
6800
6900
7000
7100
7200
7300
7400
7500
7600
7700
7800
7900
Barometer, Elevation,
In-Hg ft
25.84
25.74
25.65
25.55
25.46
25.36
25.27
25.17
25.08
24.99
24.89
24.80
24.71
24.61
24.52
24.43
24.34
24.25
24.16
24.07
23.98
23.89
23.80
23.71
23.62
23.53
23.44
23.35
23.26
23.17
23.09
23.00
22.91
22.82
22.74
22.65
22.56
22.48
22.39
22.31
8000
8100
8200
8300
8400
8500
8600
8700
8800
8900
9000
9100
9200
9300
9400
9500
9600
9700
9800
9900
10000
10100
10200
10300
10400
10500
10600
10700
10800
10900
11000
11100
11200
11300
11400
11500
11600
11700
11800
11900
B 8romftt.tr
In.-Hi
22.22
22.14
22.05
21.97
21.89
21.80
21.72
21.64
21.55
21.47
2138
21.30
21.22
21.14
21.06
20.98
20.90
20.82
20.74
20.66
20.58
20.50
20.42
20.34
20.26
20.18
20.1C
20.02
19.95
19.87
19.79
19.71
19.64
19.56
19.48
19.40
19.33
19.25
19.18
19.10
DENSITY VARIES directly with pressure, inversely with
absolute temperature. With these tables, based on air
density of 0.075 Ib per cu ft at sea level and 70 F, any
corrections for actual conditions can be made quickly
12
FANS • A SPECIAL REPORT
-------
Static
pressure (SP)
Volume
SPECIFIC SPEED, specific diameter added
to characteristic curve show that for every
value of D. there is a corresponding N«.
This relationship is used for fan selection
Static efficiency
(Es)
Specific
diameter
Specific speed (Ns)
SIMPLIFIED CURVE of static efficiency, D,
against N, eliminates parameters not needed
for selection. Composite curve on p S • 14
shows this data for a variety of fan types
(Cfm}"2
p — Solves Ns *(Rpm
SP
to
.
Operating _
speed, rpm
-
_
j
•
-
-
_
-
-
-i
-
-
j
—
-
~
-
-
-
tfO.OOO
7000
5000 ~.
4000 ~
Ns ~
3000200,000
150,000
2000 100,000
70,000
1500 50,000
lop-M600
*^ 20,000
TOO I5'°°D
10,000
500 700°
400 50°0
300 300°
2000
1500
200 ,000
-
• -
-.,— ';
~ —
•
-
__
I —
[ -
.
_ -
ISO
-
too -
—
hesHg
3
4
5
c
V
7
-------
1.0
f— forward (arvtd Mate,
L" ""1.'. ri'irir
ID" —^- i
Spea'fic speed (Ns) thousands
' 1kj""W I'bo
~200
STATIC EFFICIENCY and D. can be plotted against N, for one Enter with N. (from nomograph) and select the type of fan, Its
type fan (see previous page) or for all types on one grid. This efficiency and D.. Then return to the nomograph for a quick
results In composite curve—a tool for simplified fan selections. determination of actual diameter, and the fan is selected
obtained from table 1 and 2:
(1) Density of flue gas corrected to
operating conditions is:
27.82 460 + 70
0-078 X 2^92 X 460 + 362
or, 0.0468 Ib/cu ft
Equivalent SR at standard air condi-
tions = 10.0 X
= 16-° i°- sp
(2) Next, enter the nomograph with
the values of cfm, pressure and speed,
and solve for N,. For this example, the
specific speed is 31,800. (See dashed
line on nomograph).
(3) Enter the composite fan curve at
31,800 N. and read the specific diam-
eter and static efficiency — 0.48 and
73.5% respectively — for a radial-tip
fan, which the curve shows is most ap-
propriate for this application. A study
of the graph indicates that fprward-
curved blade could also be used, but
the efficiency is lower.
(4) Return to the nomograph with
the required SP and cfm (as dashed
line shows), where you can see that
with a D. of 0.48 the required fan
diameter in this case is 68 in.
(5) Finally calculate fan BHP:
cfm X SP
BHP - 6356 X SE
80,000 X 10.0 ,?1Tmp
or' 6356 X 0.735 ~ 1?1 BHP
Follow the same procedure no matter
at what speed the fan is to run.
Double-inlet fans. The composite
fan curve is plotted for single-inlet fans.
If you have a double-inlet fan in mind
just divide the required volume by two,
before determining specific speed and
specific diameter. Then apply the curve
in the normal manner. However, when
calculating BHP, use the full volume.
For example, using the previous
problem, suppose a double-inlet fan
were desired. Calculate specific speed
on the nomograph by considering the
volume to be 40,000 cfm (Vi the ac-
tual required volume). Then at 900
rpm the specific speed is 22,500. Enter
the composite fan curve where you can
see that a forward-curve fan will be
most efficient; read specific diameter as
0.43, static efficiency as 71.0%. From
the nomograph, with 16.0 SP, 40,000
cfm, and 0.43 D,, read the required
diameter as 43 in. BHP is then:
80,000 X 10.0 _ 177 c RHP
6356X0.71 -177'5BHP
Of course, the nature of the fan ap-
plication will have a bearing on the
type of fan considered. For instance,
it would be as unusual to consider a
vane-axial fan for a 386 F application,
as it would be to consider a speed of
3600 rpm. Special noise requirements
or vibration demands may further
modify the choice. But in any event,
this method permits quick selection of
a fan type and size for any application.
Knowing these values, it's an easy mat-
ter to select a specific model.
FANS • A SPECIAL REPORT
-------
Vibration may mean poor
selection or installation
Generally, there are two types of vibration — aerodynamic
and mechanical. The two can combine to cause trouble, or
may do their damage independently. Vibration has numer-
ous causes and many contributing factors. For instance, fan
vibration is directly related to speed. As operating rpm goes
up, vibration amplitude also goes up.
Aerodynamic vibration, also called pulsation, is caused
by turbulent flow of the air or gas. Characteristics of this
type of vibration are unsteady flow and a capacity that's
lower than calculated. If you substantially change the flow
through a fan system (such as by adjusting dampers, open-
ing an enclosure door, dropping an inspection plate in duct
work, etc) and vibration stops — or is reduced — you can be
certain that your vibration problem has aerodynamic origins.
Operating a fan on the unstable portion of its pressure
curve is one way of creating aerodynamic vibration. The
unstable portion is generally that to the left of maximum
pressure (see graph at right). Since an improper calculation
of system resistance kicks off the condition, the solution is
to somehow change the resistance by removing unnecessary
dampers, increasing outlet area, etc.
Poor inlet conditions reduce the air supply and cause
turbulence within the impeller. Some of the common trouble-
causing inlet arrangements, with illustrations of how the
problems can be solved, are shown at right.
Mechanical vibration, like the aerodynamic variety, has
many causes. Often, insufficient rigidity between fan and
motor is the culprit; poor support of the fan and motor is
another common offender. Lack of allowance for thermal
expansion can deform the fan blades. Loose connections,
poor support, and other such problems are hard to pin down.
Very often, the only way to determine the exact cause of
trouble is by trial and error.
Driving system is another source of mechanical vibration.
Possibilities include misaligned or unbalanced couplings;
loose or improperly fitting drive belts; belt sheaves which are
worn, out-of-round, loose, etc. Bent shafts or shafts running
at critical speeds can also cause vibration, as can bearings
that are close to failure.
If you suspect trouble in the driving system, you may
have to employ instrumentation to be sure. A stroboscope
spotlights faulty sheaves, couplings, shafts, etc; a vibration
analyzer points out unbalanced components.
Certain wheel conditions can cause severe vibrations. Even
a small amount of dirt on the fan wheel can cause unbal-
ance. Forward-curved blades are especially likely to pick
up foreign matter. Unbalance also results from corrosion
and abrasion, which pits or otherwise unevenly removes
material from the wheel. Distortion from other sources, such
as uneven stress, also produces vibration. Unbalance from
dirt or pitting is usually easy to spot visually. Stress-induced
distortion, unless severe, is almost impossible to spot by eye.
Solutions are available: for fans tending to pick up for-
eign matter, a regular cleaning program is the best bet. If
corrosion or abrasion causes unbalance, try a fan wheel
made of a different material (as a temporary cure, you may
be able to balance the wheel). If overstressing is causing dis-
tortion you've probably installed the wrong fan. This may
mean a new fan system — or at least a system modification.
Calculated
system
nsistonct
Copocity, cfm
PULSATION IS LIKELY when fan operates at a point to left of
peak static pressure. Usually, If you are operating In unstable
region, errors were made in calculating the system resistance
END CONNECTION too close to the fan causes turbulence
within wheel, producing pulsation. Cures include moving the
elbow and adding internal guide vanes to smooth out the flow
I
Splitter
WALL OF BUILDING close to the fan also causes turbulent
Inlet flow and pulsation. A splitter at fan inlet improves inlet
flow pattern and lets air supply fill the wheel properly
CONTROL AIR Injected close to Inlet may cause pulsation If
air stream temperatures differ greatly. Cure is to inject air
upstream using a mixing section to aid diffusion of streams
POWER • MARCH 1968
S • 15
-------
Allowable noise level varies with each installation
When selecting a fan for any applica-
tion, you must consider noise. How
much loudness can be tolerated from a
fan depends on the application.
Loudness varies with fan size and
static pressure; the problem with noise
has grown as fan sizes and operating
pressures have increased. Although the
fans and other system components have
themselves been made quieter through
improved designs and manufacturing
methods, today's applications teno to
produce more noise. For instance, it is
not unusual to have industrial ventila-
tion systems requiring fans that devel-
op over 35-in. static pressure. Not too
many years ago, a much quieter 5-in.
static pressure requirement would have
been considered high.
Noise sources. Fan noise is created
in several ways. Air turbulence creates
noise within the fan inlets, outlets and
blade passages. The resulting sound is
similar to a windy swish or a whistle.
Part of the sound is transmitted
through the fan outlet and pan through
the fan inlet. Outlet noise is largely
confined by ducts and therefore quick-
ly dissipated; inlet noise, on the other
hand, is not easily removed.
Mechanical sources are a second
cause of fan noise. Motors, bearings,
drive mechanism, etc start the ball roll-
ing and then transmit the noise through
ducts. Noise transmission is especially
significant if the plenum encloses the
fan drive.
Vibration is another source of un-
wanted noise. Also mechanical in na-
ture, it is readily transmitted through
ducts, supports, building structure, etc.
Harmonics. Proximity of the fan im-
peller to stationary objects such as mot-
or struts, wiring, etc may also produce
noise. As the impeller passes by the
stationary structural members at a con-
stant rate, a noise-accentuating, funda-
mental blade frequency (harmonic)
results. Changing either motor speed
or the number of blades changes the
noise level (number of blades should
not equal the number of motor struts).
Orifice design of axial-flow fans not
only affects efficiency (see p S-4), but
also has a sound effect. Surges of re-
circulating air between blacje tips and
orifice can cause a huffing and puffing
noise as pressure is rapidly built up
and relieved. Reducing clearance be-
tween tips and orifice ring, or changing
the axial distance between fan and ori-
fice, eliminates this noise.
Line frequency hum (120 cps) of the
fan motor becomes a source of noise
through resonance with some compon-
ent part of the fan. The hum may be
transmitted to the impeller through
motor shaft, or to the chassis through
motor support. Various types of pads,
grommets, resilient hubs, etc isolate res-
onating components from the motor.
Straight-bladed centrifugal fans,
commonly used in material handling,
are much noisier than other centrifugal
type fans (three to five times as loud).
This is because the blade passages do
not fill with evenly flowing air and
therefore pulses are generated at the
fan outlet.
Usually, a materials-handling instal-
lation doesn't allow much leeway in the
selection of blade shapes; but if low
noise level is a must, consider installing
a fan with blades having a slight curva-
ture. One plant, in 1966, installed a fan
with backward inclined blades for use
in a pelletizing process. The material
that the fan handles is extremely abr-
asive ; but the plant's engineers and the
fan manufacturer feel that what they
lose on initial cost, they more than
make up for with higher efficiency.
They believe that this fan should out-
last a straight-bladed fan and in the
long run be cheaper. Thus far their
operating experience tends to support
their belief.
The blades of many material-han-
dling fans are covered with a corrosion-
erosion resistant material. Then as the
fan wears, the blade covering can be
replaced rather than the entire fan.
For instance, the pelletizing plant fan
(paragraph above) has blades covered
with heavy diamond-stud plate, held in
place with bolts.
Noise control. If a fan installation is
too noisy, there are several ways to re-
duce sound level. One of the first things
to try is to change the direction of the
noise. If the fan has an open inlet,
point it away from people or towards
an acoustical shield. Careful position-
ing of the fan discharge also helps.
If the fan is attached directly to a
floor or structural member, the build-
ing may act as a sounding board. Sound
isolation mountings should remedy this
situation satisfactorily.
Ducting or plenums with a fan rig-
idly attached, transmit sound through
the system, thus creating a significant
amount of noise. Whenever possible,
flexible connections for fans and duct-
ing should be used.
Duct lining and silencing equipment
will also cut down on the amount of
the transmitted sound. Unfortunately,
neither lining nor silencers are entirely
practical in ducts carrying corrosive or
abrasive gases.
Fiber glass lining does an excellent
job if thick enough and long enough.
Generally, the lining should be at least
2 in. thick; as frequency (in cycles per
second) of the predominant sound
wave decreases, insulation must be
thicker. Duct size must often be in-
creased so that the added lining does
not restrict flow.
Absorption material does the best
job when it's placed across the sound
path. Thus lining is more effective in
an elbow than in a straight length of
duct. The same principle applies when
a baffle or wall is used to control noise
from outdoor fans. A concrete barrier
wall between the fan and personnel is
extremely effective. But don't place the
barrier so close to the fan that air in-
take is impeded.
Silencers aerodynamically designed
to match the fan's characteristics do a
good job of noise control, without re-
stricting flow. Choice of silencer or lin-
ing depends upon a careful analysis of
the particular system, but generally a
silencer does a better job on high-fre-
quency sound.
A fan enclosure, such as a separate
room acoustically surrounding the unit,
is effective when the only noise source
is the fan package itself. Such enclos-
ures are often used in power plants,
where the mechanical draft woOsh can
rival Hurricane Carolyn for decibels.
The enclosure should be large enough
so that inlet air flow is not restricted;
air intakes should be designed so that
the bulk of air is aimed at the axis of
the fan (see illustration); all cracks,
pipes, spaces, etc should be carefully
sealed since a lot of sound can seep
out of a small opening. Silencer or
muffler can be used in conjunction with
the enclosure (see illustration, right) for
additional noise reduction.
16
FANS • A SPECIAL REPORT
-------
BLADE DESIGN can affect fan noise, since using a different type blade may allow
slower operating speed — but not affect other performance characteristics. For
instance, a fan such as that at right—since it has more blades—gives desired out-
put even though it runs at slower speed than the fan at left; this means less noise
MATERIAL-HANDLING fan, by its nature,
tends to be noisy. Choice of blade Is
limited by the application, but using
fan blades with a slight curvature helps
Improved designs quiet fans and system components
SILENCERS, or mufflers, cut down noise significantly. One shown above is aero-
dynamically matched to the fan at left; thus, it has a minimal adverse effect on
the air-flow pattern and performance characteristics are preserved. Silencers
can be put at Inlet, outlet, or both. They are attached with a flexible coupling
ENCLOSURE, above, provides effective noise control. Walls
act as baffles, help reduce the Intensity of the sound waves
FIBER GLASS sound curb at left does excellent noise-reduction
job. The unit is designed for minimum added flow resistance
POWER • MARCH 196B
s- n
-------
Opposed Mate
f>arellei t>lo
-------
Fluid drive controls performance by varying fan speed
Fans can be controlled by varying the operating speed. The
method may be a variable speed motor, or a constant speed
motor driving through a magnetic or fluid coupling.
Fluid drives transmit power with no mechanical contact
between parts. Basically, input power drives an impeller,
which applies a force to a fluid. Power is then transmitted
from the fluid to an output shaft. On fan applications, output
speed can be varied down to about 20% of input speed.
With fluid drives, regulation is smooth and stepless. The
absence of mechanical connectors reduces the transmission
of vibration and shock. Necessary power input for starting is
less than with other control methods.
How ft works. Take a look at the mechanism shown in cross
section at the right. The unit illustrated is typical of a fluid-
drive mechanism used in applications ranging up to 800 hp.
The steel housing acts as an oil reservoir as well as an en-
closure for the mechanism. The input shaft, impeller and
casings rotate together at motor (input) speed. The runner
and output shaft rotate together at output speed. A vortex of
oil transmits power from impeller to runner.
Speed changes. Centrifugal force acts upon the working
fluid, pushing it into an annular shaped ring in the working
circuit and scoop tube chamber. The control mechanism
(manual or automatic) sets the position of the scoop tubes
for establishing the oil level in the scoop chamber. This de-
termines the amount of oil in the working circuit which in
turn determines the drive or output speed.
Pump. The circulating pump transfers oil from sump to
external oil cooler. After cooling, the oil returns and enters
-JHMrW-
iHfat shaft
Impelltr
Circulating
*0il from
cooler
Inner cosing
Offer cast/iff
the working circuit through the center of the input shaft. The
pump supplies a constant volume of oil to the system.
Torque limiting. The fluid-drive unit will stall when over-
loaded, and thereby limit the amount of torque transmitted
to the load. The maximum torque transmission can also be
regulated. Other adjustments permit governing the rate of
acceleration of the load.
last longer and are easy to maintain.
When the fan is run at reduced capac-
ity for a relatively long period, their
usually favorable horsepower alteration
saves money by cutting down on the
electricity consumed by the drive motor.
Keep in mind, too, that inlet and outlet
dampers can be used together for op-
timum characteristics.
Pitch control. One way to control
the output flow of an axial-flow fan
running at constant speed is to vary the
pitch of the fan blades. The change in
pitch can be made manually, remotely
or automatically. A variable-pitch fan
with automatic regulation is shown
in the photo above.
Sensing devices for automatic pitch
changes may be based on pressure read-
ings, temperature changes, humidity
indications, gas detection, etc. Signal
transmission can be through a pneu-
matic, hydraulic, mechanical or elec-
tric system.
Power savings at reduced _volumes
are substantial with variableipitch con-
trol. With automatic actuation, re-
sponse to change is quick and smooth.
Operation is also characterized by ex-
cellent sound level characteristics;
(since efficiency is maintained over a
wide range, sound level drops as pres-
sure and volume decrease).
Fail-safe features can be incorpo-
rated into the design of automatic-pitch
controls. For instance, in a pneumatic
system, mechanisms are available which
shift the pitch angle to either maximum
or minimum upon failure of control
air supply.
Manually adjusted blades are changed
when the fan is stopped by simply
loosening a lock nut at the base of each
blade. Index numbers on the fan hub
indicate the angle necessary for various
outputs. With remote control, the pitch
can be changed while the fan is run-
ning — thanks to a lever-linkage mech-
anism or a similar mechanical device.
Again, an index can indicate output.
Variable speed. Another way to con-
trol fan output is to regulate fan speed.
Common methods include variable
speed motors, magnetic couplings, and
fluid-drive units. A typical fluid-drive
unit, illustrated and described above,
uses hydraulic fluid as the power trans-
mission medium.
A magnetic coupling uses interacting
magnetic fields to transmit power. Con-
trol is obtained by adjusting the
strength of the excitation current which
produces the connecting field. As the
strength of the field varies, so does the
amount of slip between input drive and
output — thus regulating output speed.
Slip-ring (or wound-rotor) motors
provide step-wise change in speed by
connecting the motor windings to an
adjustable external resistance through
slip rings and brushes. As the resistance
is altered, so too is the motor speed.
One manufacturer now has a two-
speed motor on the market which
changes rpm by regulating the motor's
stator flux. This is done by modulating
the poles, not by adjusting the wind-
ings. Used with inlet vanes, precise
flow control is obtained.
Sometimes, when a supply of exhaust
steam is readily available, the fan will
be driven by a steam turbine. Such an
arrangement is frequently used in in-
dustrial plants to achieve an efficient
heat balance — as well as drive the fan.
Speed variation is feasible down to
about 35% of rated turbine speed.
POWER • MARCH 1968
S - 19
-------
BELT GUARD protects against person-
nel injury as it prevents drive damage
SHAFT SEAL keeps abrasive material
from leaving housing, entering bearing
LUBRICATED LABYRINTH seal design
(above) is effective, and has long life
HEAT FLINGER is mounted on shaft be-
tween fan and bearing to dissipate heat
IN PLACE, heat flinger also acts as a
seal, reducing leakage from the housing
COOLING WHEEL guard added to as-
sembly protects fingers and aids cooling
Fan accessories up efficiency and add protection
Auxiliary equipment usually associated
with fan installations includes items
such as external fittings, safety and con-
trol equipment, shaft fittings, inlet fit-
tings, etc. All these, often referred to
as accessories, are chosen as fan appli-
cation demands. Some, of course, are
so common as to be almost standard
items on any fan installation.
Access doors are one type of extern-
al fitting. If there's a possibility of dirt
collecting inside the fan, and if inlet
and outlet are inaccessible without re-
moving duct work, a door should be
installed, They are also invaluable for
fan inspection. If inspections are fre-
quent, a quick release access door is
preferable. This type, shown in the pho-
to above, has lever nuts -or lugs for fast
opening and securing. If only occasion-
al inspection is scheduled, a bolt-se-
cured door is satisfactory.
Inlet and outlet flanges can be sup-
plied when rigidly tight connections are
required. These flanges have pre-
punched bolt holes and can be made
of special alloys to allow for thermal
expansion of attached ductwork. Duct
fittings and mounting panels of all va-
rieties and shapes are available. Fiber
glass is available for applications that
demand corrosion resistance.
Stack caps are used with roof ven-
tilating units that discharge air vertical-
ly. When the fan is operating, the cap
lid automatically opens. (The lid can
also be remotely controlled). When the
fan is idle, the lid closes and the cap
provides a weatherproof closure. These
fittings are standard items for sizes
ranging up to about 6 ft diameter.
When corrosive gases are involved,
caps are available in corrosion-resistant
materials such as fiber glass.
Safety items should not be ignored.
Screens keep foreign matter from inlet,
also protect personnel. Belt guards (see
photo above) protect mechanism and
personnel. Vibration isolation mount-
ings can also be considered safety
items, since they help protect equip-
ment from serious damage while in-
sulating workers from an annoying
distraction.
Shutters and dampers were already
mentioned as a method of fan control.
It's well to keep in mind though, that
the damper installed should suit the ap-
plication, or else you're just wasting
money. For instance, outlet dampers
are generally less expensive than inlet
dampers; so when it's expected that a
damper will only be used infrequently
— and for a short time on each occa-
sion — use the outlet damper instead of
the more expensive inlet damper. How-
20
FANS • A SPECIAL REPORT
-------
A reliable drive motor is at the heart of effective fan performance
Fans have electric-motor drive in almost all installations. A
polyphase induction motor rs customary for ratings over 1 hp.
NEMA has adopted standards for motor dimensions, rat-
ings, insulation, enclosures, etc. Motors are thus easy to
specify and apply with assurance of interchangeability.
Type of enclosure, open or totally enclosed, is considered
when specifying a fan drive. Totally enclosed motors—which
prevent the interchange of inside-outside air—are recommend-
ed if ambient air contains anything harmful to motor insides.
Location of the motor depends on the gas being moved,
unit size, physical restrictions, etc. If gas is corrosive or ex-
cessively hot, the motor is placed outside the duct and con-
nected to the fan by a shaft or belts. One design mounts the
motor within a bifurcated duct. (See photo, right). When ap-
plication permits, the motor, either open or totally enclosed,
may be connected directly to the fan and located right in the
flow path. Air flow then helps cool the motor.
Starters, which provide overload protection as well as a
means of energizing the windings, may be of full- or reduced-
voltage types. Full-voltage starters are the least expensive
and are usually adequate since most fans can withstand the
acceleration forces developed when full voltage is applied at
standstill. In addition, the starting load of most fans consists
only of their own inertia; start-up is fast and the motor is not
subjected to high starting current for prolonged periods.
Very large fans may require reduced-voltage starters, how-
ever. They lower the inrush current on starting, then apply
full line voltage when the motor is near full speed.
Breakaway torque requirements for the motor are low,
amounting mainly to the bearing friction. After that, torque is
proportional to the square of the speed, as the fan starts to
move air. The motor must have a torque characteristic that
exceeds the fan requirements at every speed (curve, right).
The final operating point will be at the intersection of the
two curves, but it must be in the straight-line portion of the
motor curve at far right. Fortunately, the easy starting re-
quirements permit a low-slip motor to be used, having a steep
slope in the operating area. Speed won't vary much even if
some variation is encountered in the fan curve.
op ecu
ever, where long periods of reduced-
capacity operation are anticipated, use
the more efficient inlet dampers. Vari-
able inlet vanes (either manually or
automatically controlled) are also avail-
able for control purposes. (See p S • 18)
Bearings are another item which can
be thought of as fan system accessories.
They should, of course, be selected to
match the needs of the application.
When shaft and fan wheel are extra
heavy, use special-duty bearings.
A drain fitting is one inexpensive
item that should never be overlooked.
Welded to the low point of the scroll
case, it allows liquid to escape from-the
enclosure. A standard pipe^or a trap
can be connected to the drain fitting. If
there is even a possibility of rain, con-
densation, water, etc collecting in the
scroll case, be sure to include a drain
fitting in your specifications.
Shaft seals axe advisable when the
fan will be handling abrasive material
that might be blown out of the shaft
opening back towards the shaft bear-
ings. This may involve the use of noth-
ing more than a nonlubricated felt or
asbestos pad fixed over the shaft open-
ing. For added effectiveness try a lubri-
cated labyrinth seal.
A shaft cooler or cooling wheel is
used on high-temperature (above 300
F) applications to protect the inboard
bearing from heat radiated from the
fan housing and convected through the
shaft. The wheel — mounted between
inboard bearing and fan housing — not
only cools the bearing by drawing air
over it, but also acts as a shaft seal, re-
ducing leakage from the housing and
flinging abrasive material away from
the bearing. A cooling wheel guard
should be installed for protection and
for the added cooling it provides by
controlling the direction of air flow
over the cooling wheel.
Coolers should be made of material
having a relatively high heat transfer
coefficient; aluminum, brass, cast iron
and steel are all used. The wheel is
either clamped or shrunk-fit to the
shaft. It should not be secured to the
shaft with set screws, since these tend
to push the wheel away from the shaft.
When fan is working in an extreme-
ly high-temperature environment, pro-
vision should be made for an even
greater cooling requirement; water-
cooled end caps can handle this chore.
POWER • MARCH 1968
S • 21
-------
Y
-hr
£
_L
o e
o e
o e
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o o
• 1 e
0 0
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~~rr
r
T
5
Y
\ H -£ ^ Pilot tube positions
PITOT TUBE takes readings in rectang-
ular duct at those points indicated above
ROTATING ASSEMBLYis placed in bearings after the pedestals have been leveled
with shims. The shaft should be clean, oiled where it contacts bearing surfaces
IN ROUND DUCT keep tube within
±1/2% of R and ±1 deg of positions
Installation and maintenance
Installation starts with the unloading
of the fan. Move all parts carefully
taking care not to lift the fan wheel by
its blades or rim. All parts should be
throughly cleaned and inspected be-
fore actual assembly begins.
Foundations for heavy-duty indus-
trial fans should be reinforced poured
concrete. They should scale at least
three times the weight of the equip-
ment supported. When determining
top level of the foundation, allow at
least % in. for grouting.
Fan housing of large industrial
units is usually split for easy wheel and
shaft removal. Lower portion should
be set on the foundation in its approx-
imate position and leveled with shims.
Also put the bearing pedestals in ap-
proximate position. Then the wheel
and shaft can be carefully lifted and
set into the bearings (see photo
above). Align shaft by accurately
positioning and securing the bearings
and pedestals. Then align the drive
mechanism with the fan shaft. Finally,
the housing is adjusted by vertically
positioning its inlet to wheel; shaft
should also be centered in the hous-
ing's shaft opening, allowing for
thermal expansion when operating.
Duct work connections are made
after fan is ready for operation. Do
not force flanges that don't fit; you'll
either weaken the ducting or draw
the fan out of line. Insert gaskets be-
tween flanges and provide expansion
joints when hot gases will be handled.
Ducts should be independently sup-
ported. Avoid sudden changes in
duct size; keep 45 deg maximum an-
gle between duct and entering
branches. On turns, keep a centerline
radius of at least l'/i duct diameters.
Electrical connections. For electri-
cal hook-ups, use only wire specified
by insurance underwriters; include a
thermal overload switch. Before oper-
ation, carefully check wiring against
diagrams furnished by manufacturers.
Start-up. Make one final check of
bearings, wheel and shaft, electrical
connections, etc before starting the
fan. After a run-in period (eight hours
are sufficient for large industrial units)
during which bearings can be checked
and performance observed, the fan is
ready for full-time service.
Bearings must be carefully main-
tained. They should be oiled or
greased regularly, using a highly re-
fined mineral base lubricant. Do not
use too much grease, as overheating
and leakage can result. Use special
lubricants for applications such as
those where temperatures are extreme,
or the atmosphere is extremely wet.
When bearings are used at high
temperature or in a dusty atmosphere,
close attention to lubrication is even
more important. Dirty bearings need
thorough cleaning before greasing. If
the lubricant receptacle is not clean,
new grease will just carry foreign mat-
ter into the bearing.
Clean fans regularly. If parts are
kept free of dirt, grease and grime, the
fan will operate more efficiently and
last longer.When the fan is belt-driven,
check belt tension frequently. If the
belt is too tight, extra load is placed
on bearings, belts overheat, and life
of the drive mechanism is shortened.
Field testing. Very often after the
22
FANS • A SPECIAL REPORT
-------
Rotation and discharge symbols for radial, fans simplify identification
'Top
ROTATIONS AND DISCHARGE directions are noted from drive
side of fan. On single-inlet fans, drive side is always consid-
ered as that side opposite inlet—regardless of actual drive
location. If the mounted fan is inverted (suspended from the
ceiling, for example), the direction of rotation and discharge
is determined as if the fan were resting normally on the floor
Fan inlet box position is designated with reference to horizontal centerline
INLET BOX designations refer to horizontal line through cen-
ter of fan shaft as seen from drive side. If fan has drive con-
nection at each end of shaft, the drive side is that with higher
horsepower. For angular arrangements (Nos. 3 and 6 above)
give the angle between centerlines as shown in drawings.
Nos. 3* and 6* arrangements above interfere with floor
fan is installed, you'll want to check
its output. This means a field test.
Tests are also called for when the sys-
tem functions improperly, needs bal-
ancing, or when changes are planned
and performance data are needed.
If done carefully, field testing is
accurate. Volume measurements to
± 10% of actual flow can be made at
the fan's inlet or outlet.
Instruments. A pitot tube and man-
ometer offer the best way to determine
air velocity or pressure. Place the
pitot tube in a long run of straight
duct, about 10 diameters from fan. If
the test is made on the inlet side, the
test point can be closer to the fan, but
there should still be a long straightjrun
of duct on the outlet side. ,^~'
Procedure. With pitot tube and
manometer, record velocity pressures
(Hv) in inches of water at 20 points
in the duct. (See drawings on the fac-
ing page). Take the square root of
each velocity pressure reading and then
average these 20 values. Call the value
(vll^)avg. Check temperature in the
duct and local barometric pressure
(in. Hg) to find the actual air density
(D) in Ib per cu ft.
D = .075
530
460 + local F temp.
barometer reading \
29.92 /
Next, figure the ratio of standard to
actual air density (K). That is:
K=0.075 -r- D
The average air velocity, V, in the
duct (in fpm) is found next. V = 4000
x (^Hv)ivg. Multiply V by the duct
area in sq ft to get the actual cfm de-
livered by fan. To compare with rated
cfm, first multiply Hv readings by K
to correct for density.
Fan horsepower is easily calculated
by reading volts and amps going into
the motor and applying electrical pow-
er equations. These are:
Fan hp = (watts x motor efficiency)
-•- 746 , and:
Watts = volts x amps; or for 3 phase
= V3 x volts x amps x power
factor
Horsepower is also corrected for
density by using the multiplier K.
Plot horsepower determination and
pitot tube static-pressure reading on
the fan's characteristic curve (sup-
plied by manufacturer when fan is in-
stalled) at fan's operating rpm, as de-
termined by a tachometer. These plot-
ted points should fall close to the line
representing calculated cfm.
POWER • MARCH 1968
S • 23
-------
New materials
mean more
uses for fans
One of the biggest challenges fan engineers face is develop-
ing ways to build units that can withstand high temperatures
and corrosive atmospheres. The problem has been alleviated
by the development of new materials, and the discovery of
new ways to apply existing materials.
Heat-resistant materials. Both structural and corrosive
considerations come into play when a fan is installed in a
high-temperature gas stream. Heat causes many chemical
reactions to take place, and the strength of most materials
drops with increasing temperatures. The ultimate strength
of steel, for instance—though it may improve slightly at
moderate temperatures—eventually decreases rapidly. Yield
strength decreases with temperature even when ultimate
strength is unaffected.
In normal atmospheres mild steel scales rapidly at temper-
atures over 900 F. There are ways to combat this, however.
A process combining metal spray and heat treatment, for ex-
ample, causes a protective steel-aluminum film to coat the
surface of steel.
Special metal alloy? are often used for extreme tempera-
ture applications. Stainless steel and monel, though expen-
sive, are particularly suited for high temperatures and, of
course, also have excellent corrosion-resistant characteristics.
Aluminum alloys are used frequently for corrosion resist-
ance, but their temperature limit is about 300 F
Corrosion resistance. When corrosion attacks, it does so
in one of two ways. Direct chemical attack is generally lim-
ited to high temperatures or to highly corrosive environ-
ments—or to a combination of the two conditions. These
reactions are prevented by controlling the temperature or
the concentration of the corrosive substance; making the fan
of a corrosion-resistant material, or applying an anti-corro-
sion coaling can also do the trick.
Corrosion may also be electrochemical in nature. Such a
reaction requires separate anodic and cathodic sections,
joined by solid material submerged in an electrolyte. These
conditions can be induced by the coupling of two materials,
slight variations in stress, dissimilar metal deposits, etc. Con-
densation can result in an electrolyte.
Coatings such as lead, rubber, plastic, etc provide protec-
tion against corrosion mainly because they are inert in the
corrosive media. Lead linings are bonded on, or attached
mechanically. Rubber is usually tacked to steel surfaces and
vulcanized in place; sometimes pans are rubber-coated by a
spraying or dipping process. Plastic coatings are usually
sprayed on.
Fiber glass is a leader in the war on corrosion. It's ex-
tremely resistant to the corrosive effects of most acids, gases,
organic materials, etc. In addition, it's generally as rugged
and tough as metals, has more design flexibility than metal,
is lightweight and economical. In applications where solid
fiber glass construction is not feasible, fiber-glass-coated
STAINLESS STEEL was used in this fan built for a glass man-
ufacturer. It will work in a corrosive atmosphere at 1450 F
metals can give the protection needed. Maximum tempera-
ture for fiber glass fans is in the neighborhood of 200 F
Example. One chemical manufacturer had to find a fan
that would handle corrosive gases (hydrochloric acid, water
vapor and organic acid) and operate satisfactorily in an out-
door, sub-freezing environment. Fan engineers licked his
problem with a nickel-molybdenum steel alloy; fan wheels
were solution-heat-treated for additional corrosion resist-
ance. Special coating material protects external surfaces
from extreme weather conditions. A specially designed syn-
thetic gasket provides a gastight fit, while an ethylene glycol
solution—acting as both coolant and anti-freeze—is circu-
lated through the shaft seal.
Crystal ball. And what else besides exotic materials can
we expect to see in tomorrow's world of fans? For one thing,
we'll be confronted by ever-larger units; fans with wheel di-
ameters of 150 inches will probably be common. Even now
there are fans whose diameters top 200 inches.
Fan testing is another area where techniques have im-
proved and are getting even better—and, in turn, fan per-
formance benefits. Manufacturers are definitely becoming
more quality-control conscious too; for instance, with many
companies it's now standard practice to nondestructively
test all welds.
Control techniques are also becoming more sophisticated.
Increased use of automatic controls is already evident; more
application of variable-speed devices seems inevitable.
Again, this means greater fan effectiveness.
Reprints available
For reprints of this special report, write to: POWER, Reprint
Dept., 1221 Avenue of the Americas, New York, N.Y. 10020.
24
FANS • A SPECIAL REPORT
-------
ITEM 4
Selecting Fans and Blowers
Robert Pollak
Chemical Engineering Journal
-------
^~*^m\
This discussion of available types of fans and blowers, and of the factors
that should be considered in their selection, maintenance and installation,
should help you choose the most adequate unit for your application.
ROBERT POLLAK, Bechtel. Inc.
Few pieces of equipment have as wide a range of
application in the chemical process industries as do fans
and blowers. Considering that they have such diverse
uses as exhausting or introducing air or other gases into
process reactors, dryers, cooling towers and kilns; assist-
ing combustion in furnaces; conveying pneumatically; or
simply ventilating for safety and comfort, these machines
can well be regarded as basic pieces of equipment.
In the last few years, fan-assisted, air-cooled heat ex-
-*
CENTRIFUGAL FAN: (a) entering air is turned 90 deg. as
it is discharged; (b) blade types—the airfoil kind is the most
efficient—Fig 1
changers also have made considerable inroads into the
CPI, as engineers have sought to solve thermal water-
pollution problems.
Because of an increasing demand for smaller, more-
reliable fans and blowers, and due to the new impetus
on occupational health and safety, these machines are
now receiving increasing attention. At the same time
that user requirements have forced manufacturers to
build fans for higher pressures—with resulting higher
speeds—environmental considerations have pressed for
lower noise levels and shorter noise-exposure times.
Because fan manufacturers are supplying machines at
higher compression ratios and at lower and higher flow-
rates than ever before, an in-depth engineering evalua-
tion of fans or blowers may be justified before selecting
one or the other. For this, a basic knowledge of what
the various types of fans and blowers can and cannot
do is essential.
Classification of Fans and Blowers
The word fan is ordinarily used to describe machines
with pressure rises up to about 2 psig. Between this
pressure and approximately 10 psig., the name applied
to the machine is blower. For higher discharge pressures,
the term used is compressor.
Fans are normally classified as axial (where air or gas
moves parallel to the axis of rotation), or centrifugal (air
or gas moves perpendicular to the axis). The National
86
JfiNMi D\
-------
'-•Vft
Assn. of Fan Manufacturers has established two general
categories of axial-flow (AF) fans: tube-axial and vane-
axial.
AF units are usually considered for low-resistance
applications because of their ability to move large quan-
tities of air at low pressure.
Centrifugal-flow (CF) fans are used for jobs requiring
a greater head, where moving air encounters high fric-
tional resistance. CF fans are classified by blade configu-
ration: radial, forward-curved, backward-curved or in-
clined, and airfoil (Fig. 1).
Blowers are generally single-stage, high-speed ma-
chines, or multi-stage units that operate at pressures close
to, or in the range of, compressors (Fig. 2). The term
blower is also applied to rotary (positive-displacement)
compressors that can handle relatively low flows at high
compression ratios.
Characteristics of Axial Fans
Classified into tube-axial and vane-axial types, the
characteristics of these machines are as follows:
Tube-Axial Fans—Designed for a wide range of vol-
umes at medium pressures, these consist primarily of a
propeller enclosed in a cylinder that collects and directs
air flow. A helical or screwlike motion is the typical
air-discharge pattern (Fig. 3).
Vane-Axial Fans—These are characterized by air-guide
AIR-TIGHT PRESSURE BLOWER can
handle air, natural gas, organic vapors,
helium, nitrogen, etc.—Fig. 2
CHEMICAL ENGINEERING/JANUARY 22, 1973
87
-------
s-:
Hfe
»5* C^.J*^^8S*ini5J^;:ieSBP<'*
-------
SELECTING FANS AND BLOWERS ...
FORWARD-CURVED FAN WHEEL has
large-volume capacity at low speed and
operates fairly quietly_Fig. 6
BACKWARD-INCLINED WHEEL devel-
ops much of its energy directly as pres-
sure—Fig. 7
AIRFOIL FANS have backward-inclined
blades with airfoil cross-section for less
air turbulence—Fig. 8.
Blades tend to be self-cleaning, and can be of high struc-
tural strength. Typical impeller types are shown in
Fig. 5. Normally, the machine is not used for ventilating
purposes.
Forward-Curved Type—This fan imparts a greater ve-
locity to the air leaving the blade than a backward-
inclined blade running at the same tip speed. Although
the machine discharges high-velocity air, it runs at slower
speeds than the other types, which makes it suitable for
process equipment requiring long shafts. The machine
operates fairly quietly and requires little space (Fig. 6).
Backward-Curved and Backward-Inclined Types—
These feature blades that are curved or tilted backward
to the optimum angle to develop much of the energy
directly as pressure (Fig. 1). This makes the units efficient
ventilators.
These fans operate at medium speed, have broad pres-
sure-volume capabilities, and develop less velocity head
than forward-curved units of the same size. Another
advantage of these backward-inclined fans is that small
variations in system volume generally result in small
variations of air pressure, which makes the units easy to
control.
Airfoil Centrifugal Fans—These are backward-curved-
blade units that have been given an airfoil cross-section
to increase their stability, efficiency and performance.
While operating, airfoil fans are also generally quieter,
and do not pulsate within their operating range, because
the air is able to flow through the wheels with less turbu-
lence (Fig. 8).
Tubular Centrifugal Fans—These are enclosed in a duct
so that air enters and leaves axially, and all changes in
direction of flow are within the fan (Fig. 9). Their design
produces a steeply rising pressure over a wide range of
capacity (Fig. 10). Being nonoverloading, these fans are
good for general building ventilation and air condition-
ing, as well as for fume removal, humidifying, drying,
motor cooling, and supplying combustion air.
Axial Versus Centrifugal Fans
In general, centrifugal fans are easier to control, more
robust in construction, and less noisy than axial units.
Their efficiency does not fall off as rapidly at off-design
conditions.
Inlet boxes* can sometimes be used without impairing
the pressure or efficiency of centrifugal fans, but they are
generally not recommended with axial-flow machines. If
possible, axial-flow fans should have about two diameters
of axial distance upstream and downstream without ob-
structions or changes in direction.
Centrifugal fans are less affected by miter elbows at
* Devices used to turn the air 90 cleg at the Ian inlet in a space close to one
diameter in the axial direction.
TUBULAR CENTRIFUGAL FAN is
.enclosed in a duct for air to enter
and leave axially—Fig. 9
CHEMICAL ENGINEERING/JANUARY 22, 1973
89
-------
SELECTING FANS AND BLOWERS ...
Nomenclature
A Barometric pressure corresponding lo site alti-
tude, psia.
B Factor, (A" - \)/KN
Ekfc Brake horsepower as read from standard per-
formance curve
E^ Brake horsepower required ai site
H Poiytropic head, (ft.-lb.)/lb.
K Ratio of specific heat at constant pressure to
specific heat at constant volume, cp/cr
M Molecular weight
N Poiytropic efficiency
/*, Inlet absolute pressure, psia.
P2 Discharge absolute pressure, psia.
pEA Equivalent air pressure to be used with standard
performance curves, for a compressor to provide
desired discharge pressure at site, psig.
p2 Discharge gage pressure at site, psig.
QM Volume of air entering compressor, cu.ft./min.
K Factor, 1,545/M
re Pressure ratio at standard inlet conditions
r, Ratio of absolute discharge pressure at site to
absolute inlet pressure at site, P2/P\
T Absolute inlet temperature, °R.
7"j Inlet temperature, °F.
VA Actual volume of air, cu.ft./min.
\'s Volume of air at standard conditions (68 F., 14.7
psia.), cu.ft/min.—actually a measure of mass
flow (air density of 0.075 Ib./cu.fl.)
W Mass flow, Ib./min.
xc Temperature factor to be used with standard
performance curve -when selecting a compressor
jr. Temperature factor for site conditions
Z Average compressibility factor
the inlet than vane-axia! fans, but losses in efficiency up
to 15% can be expected when abrupt changes in air-flow
direction occur at the fan inlets.
Inlet guide vanes usually provide smooth control down
to less than 30% of normal flow, but there have been
instances of vibration problems on large, induced-draft
. Percent rated static pressure
Percent
maximum
horsepower
STEEPLY RISING PRESSURE is produced by tubular cen-
trifugal fan over wide range of capacity—Fig. 10
and forced-draft fans when their inlet guide vanes have
been closed between 30 and 60%.
When high duct velocities are present with a fan
equipped with inlet guide vanes, extra consideration
should be given to having smooth air-flow patterns in
the inlet and outlet ducts, as well as making ducts as
strong as necessary to avoid vibration damage. Vibration
is aggravated by turbulence and improper inlet-guide-
vane setting.*
Axial fans have a narrower operating range at their
highest efficiencies (Fig. 11), which makes them less
attractive when flow variations are expected. The hump
on the axial-fan-performance curve (Fig. 12)—at about
75% of flow—corresponds to the stall point. Operation
* Ret. 5 provides a good general treatment of how tans work.
imiiiminiil imiiiiiuimiimmiii MI i iiniiin i imiiimmimiiin iiimiiium u liiimiiiiiiimmiiiiiniiiii mini mi m mimiimimii iiiiiiiimiini niiiiinmiiiiii iiiiiiinimiiiinmiii mm immiimimin miiiin
Typical Industrial Applications for the Various Types of Fans—Table I
Type of Fan
Application
Conveying systems
Supplying air for oil and gas
burners or combustion furnaces
Boosting gas pressures
Ventilating process plants
Boilers, forced-draft
Boilers, induced-draft
Kiln exhaust
Kiln supply
Cooling towers
Dust collectors and electrostatic
precipitators
Process drying
Reactor off-gases or stack
emissions
uiimMiiuiMfuuiiiiniiiiniiiimimimiHmnimnlHUiiHHUiimimiuniiniRnoiliilimimii
Tube-Axial Vane-Axial Radial Forward-Curved Backward-Inclined Airfoil
X
X
X
X
X X
X X
X
X
X X
mMmiiiiiiimimmihimimiMiiimmmiLinmiHiituiHmiiiiiiimiimiiiiiimiiiiiiiiniimiiiiiiiiiiiiiiiiiimmijtiii
X
X
X
X
X
X
X
X
90
JANUARY 22, 1973/CHEMICAL ENGINEERING
-------
Typical centrifugal
(100,000 cu.ft./min
EFFICIENCY CURVES for centrifugal and axial fans-Fig. 11
PERFORMANCE COMPARISON: total pressure and brake
horsepower of axial versus centrifugal fans—Fig. 12
•o
c
™ e
at t-
o>
I
«^>
o
c
I
i-i-5!!1"—
I t I
0)
3
e -1
Standard centrifugal ventilation fans ^
jj> 600,000 cu. ft. /min.1
20,000 40,000 60,000 80,000 100,000
Air inlet, actual cu. ft./min. (air at 70 F., 14.7 psia.)
120,000
To me this graph: (1 ) calculate the actual cu. ft./min. (ACFM) at inlet conditions (at fan flange), and
the total pressure rise— in inches of water— from the inlet to the discharge fan-flange; (2) locate the
ACFM value on the chart; if the region where the point falls can be served by more than one type of
fan (axial versus centrifugal, or different types of axial or centrifugal), decide on the type of fan by
making economic and engineering evaluations.
FAN SELECTION GUIDE, based on pressure rise versus air flow, according to catalog ratings— Fig. 13
CHEMICAL ENGINEERING/JANUARY 22, 1973
91
-------
SELECTING FANS AND BLOWERS . ..
of an axial-flow fan between this point and no-flow is
not desirable; performance is difficult to predict.
Fig. 11 also shows the efficiency curve for centrifugal
fans (CF). Bear in mind that both these curves are gen-
eral and are not intended to imply that axial fans are
less efficient than centrifugal ones.
Process applications, in general, are more apt to use
centrifugal fans, although there is a considerable amount
of overlap between centrifugal and axial units at the
lower end of the flow-pressure range. A performance
comparison of centrifugal versus axial machines is shown
in Fig. 12. Table 1 lists typical applications.
Fig. 13 shows the range of centrifugal and axial ma-
chines. This chart is based on catalog ratings. The stand-
ard, centrifugal ventilation fans operate up to approxi-
mately 22 in. of water. Beyond this, heavy-duty
centrifugal fans—with higher compression ratios at some
flows—may be made to specifications. The only area
where no fan is available is above 100 in. of water at
extremely low air flows.
When an application is outside the standard range for
fans, it is advisable to consult manufacturers to see if
a special heavy-duty unit can be built. At higher pressure,
it may be difficult to decide initially whether the process
requires a compressor or a fan. When this is the case,
it may become necessary to obtain estimated prices from
manufacturers of both types of equipment before making
a selection.
Sizing Procedure
To estimate the air-horsepower requirements of
fans—when density changes between inlet and outlet can
be neglected—the following formula can be used with
air:
Air hp. = (144 x 0.0361)2/1/33,000
(1)
where Q = inlet volume, cu.ft./min.; and h = static-
pressure rise, in. of water.
For estimating brake horsepower (BMP), an efficiency
value—obtained from Fig. 11—can be used with the
above formula (efficiency = output-air horsepower/input
horsepower). The actual efficiency will depend on the
type of fan. The driver horsepower is usually selected
so that a power margin of safety of at least 10% exists
at the expected operating point, and the required horse-
power at any flow is less than the driver horsepower. This
permits operation at other-than-design conditions.
Manufacturers' catalogs are usually arranged to show
a tabulation of standard cu.ft./min. versus pressure rise
across the machine. When air is not at standard condi-
tions, volume, pressure and horsepower corrections must
i i
Equivalent air pressure
for use with standard
performance curves
PRESSURE CORRECTION CURVES to be used for altitude and inlet temperature of air-Fig. 14
92 JANUARY 22, 1973/CHEMICAL ENGINEERING
-------
Value of x for Air Only*— Table II
3456
1.0
1.1
1.2
.3
4
.5
.6
.7
.8
1.8
2.0
0.0000
0.0273
0.0530
0.0771
0.0999
01216
01423
0.1620
0 1810
01992
0.2167
0.0028
0.03OO
0.0554
0.0794
0.1021
0.1237
01443
0.1640
0.1828
0.2010
0.2184
0.0056
0.0326
0.0579
0.0817
0.1043
0.1258
01463
0.1659
0.1847
0.2028
0.2202
0.0084
0.0352
0.0603
00841
0.1065
0.1279
01483
0.1678
0.1865
0.2045
0.2219
0.0112
0.0378.
0.0628
0.0864
0.1087
0.1300
01503
0.1697
0.1884
0.2063
0.2236
0.0139
0.0404
0.0652
0.0886
0.1109
0 1321
0.1523
0.1716
0.1902
0.2080
0.2253
0.0166
0.0429
0.0676
0.0909
0.1130
0.1341
01542
0.1735
0.1920
02098
O.E269
0.0193
0.0454
0.0700
0.0932
0.1152
0.1362
0.1562
0.1754
0.1938 -
0.2115
0.2286
0.0220
0.0480
0.0724
0.0954
0.1173
0.1382
0.1581
0.1773
0.1956
0.2133
0.2203
0.0247
0.0505
0.0747
0.0977
0.1195
0 1402
0.1601
01791
0.1974
0.2150
0.2320
•The table is used as in these examples if r - 1 00. x - 0.0000, if r - 1.01, * - 0.0028. if r •= 1.86. x - 0 1920
•UMnHiliimiiiimmjiiiiiiiiiHiiimiiiMuiimMiiiimiiimijiiiimiiimiiiHiummiimiiiiiiiHiiiiiiimiiiHijiiiimiiiimiM
be applied to be able to select a machine at an "equiva-
lent" volume and pressure.
Make the following corrections when inlet conditions
are not the standard 68 F. and 14.7 psia.:
Volume Correction
(2)
Pressure Correction
Method A: Use Fig. 14
Method B: r, = (A + P2)/A
x, = r,° 2R3 - 1 (see Table II for value of x)
rf = (xf + I)3-" (see Table II for value of re)
PEA = 14.7 (rc - 1) (3)
Horsepower Correction
«;TR \
PC W
When making calculations, bear in mind the following:
1. Make appropriate substitutions in Eq. (2) through
(4) if manufacturers' catalogs have been prepared for
conditions other than the standard 68 F. and 14.7 psia.
2. When an approximate value of the equivalent air
pressure (pEA) is needed, enter Fig. 14 on the left-hand
graph at the proper pressure, and read up to the corre-
sponding site elevation. From this point, draw a line to
the maximum inlet temperature expected (right-hand
graph) and proceed down from this intersection to the
equivalent air pressure on the X axis. For instance, to
obtain 6.0 psig. with inlet conditions of 4,000-ft. altitude
and 100 F., a blower must be selected that will develop
75 psig. under standard conditions.
3. For an accurate pEA value, use pressure-correction
Method B above—for Eq. (3)—with x factors shown in
Table II.
4. The brake horsepower needed for job-site conditions
is determined from standard performance curves.
5. For gases other than air, use Eq. (5) to calculate
head, and select a compressor that will develop this same
head on air. Brake horsepower may then be calculated
CHEMICAL ENGINEERING/JANUARY 22. 1973
by means of Eq. (6). For these applications, it is advis-
able to consult a manufacturer's representative.
Sample Calculations
Example 1—Calculate the brake horsepower required
for these conditions: suction flow = 10,000 standard
cu.ft./min.; P^ — 12.7 psia. at 4,000-ft. elevation; p2 = 4
psig.; TI = 120 F.
'46°+I20X 12,700 cu.f,./min.
-o-
528
r, = 16.7/12.7 = 1.318
x, = 1.3180-283 - 1 = 0.805
xf = 0.0805 (—) = 0.0884
\ 528 /
rr = (1 + 0.0884)3-53 = 1 35
PEA = 14.7 x 0.353 = 5.2 psi. (check value with Fig. 14)
Entering catalog rating tables with a 5.2-psi. pressure
rise, and 12,700 ACFM (actual cu.ft./min.):
Although, in general, fan-manufacturers' representa-
tives should be contacted when sizing for gases other
than air, the procedure that follows may be used to
estimate equivalent fan horsepower and flow. Fig. 15,
which is used in this procedure, was prepared by means
of the equation for polytropic head (similar to column
height in liquids), which applies to given speeds and inlet
flows, regardless of the type of gas:
(5)
Fig. 15 can be used with little error for efficiencies
between 0.60 and 0.80. Note that at lower compression
ratios, air (or gas) compressibility can be neglected.
To determine the horsepower required by a fan, Eq.
(6) can be used:
HW
33,000 N
(6)
Although A' in this equation is polytropic efficiency,
93
-------
SELECTING FANS AND BLOWERS ...
W =
POLYTROPIC HEAD versus compression ratio—Fig. 15
static efficiencies may be used as first approximations.
Example 2—Calculate the brake horsepower (Ehps)
required and the mass flow (W) attainable for a dry
carbon dioxide system, for a fan whose air-handling
characteristics are: P^ = 14.7 psia.; 7~, = 70 F.; actual
cu.ft./min. (VA) = 26,000; discharge pressure =18 in.
water gage; brake horsepower for air (Ehps) = 103 (as-
sume 98 hp. for air -f 5 hp. for bearing losses); speed =
960 rpm. Pertinent data for the CO2 system are: K = 1.3;
molecular weight (M) = 44; Ta = 100 F.
18 in. water gage = 18 x 0.03613 = 0.65 psi.
P2 = 14.7 + 0.65 = 15.35 psia.
PI/PI = 15J5/14.7 = 1.045
At R = 1.045, H= 1,260 (ft.-lb.)/lb. (from Fig. 15)
26,000 x 144 x 14.7 x 28.9
1,545 x 530
1,940 X 1,260
= 1,940 Ib./min.
33,000 x 98
= 0.756
With carbon dioxide at a head of 1,260 (ft.-lb.)/lb.,
the B factor is:
= 0.305
K >N) \ 1.3
And the pressure ratio:
o.756/
= 1.196
v ' " \KTf (1,545/44)(560)
P? B
— = VI.196 = 1.80
Therefore, at 26,000 actual cu.ft./min., and 100 F., the
mass flow ( W^) and the brake horsepower (Ehpl) for the
CO2 system are:
26,000 X 144 x 14.7 x 44
1,545 x 560
2,800 x 1,260
EI"" 33,000 x 0.756
= 2,800 ib. CO,/min.
hp. (plus 5 hp. for bearing losses)
Checking the discharge temperature:
T2 = 7j(—) = 560 x 1-196 = 670 deg. R. (210 F.)
\Pl /
Before the fan in this example is used on carbon
dioxide, the fan manufacturer must be consulted to de-
termine \vhether the equipment is suitable for the new
service. He could suggest changing the speed or restrict-
ing the flow to bring down the required power.
A lower speed would reduce the pressure ratio pro-
duced by the machine (fan laws may be used to estimate
the new performance). The flow would have to be re-
stricted within the stable flow for the fan, and a new
performance curve obtained from the manufacturer.
Ordinarily, fans are not switched from one service to
another, but the methods outlined above can be used for
estimating required power and—using general vendor
literature—selecting a fan size.
Specifications, Data Sheets
An essential part of correct sizing is an accurate defini-
tion of operating conditions and requirements. When a
fan is to be purchased, the usual procedure is to issue
a data sheet and specifications to fan manufacturers. The
data sheet should contain not only information to enable
the manufacturer to size the fan, but also a list of neces-
sary accessories; and sufficient space should be provided
to enter data supplied by the manufacturer. This aids in
evaluating the mechanical and aerodynamic charac-
teristics of the fan. A typical data sheet includes the items
listed in Table III.
Gas characteristics and operating conditions must be
defined as accurately as possible. Included should be the
widest expected range of gas components, pressures and
temperatures. For instance, a forced-draft fan for a boiler
in northern Canada may have to draw air at temperatures
from —50 to +90 F. It may therefore be necessary to
drive this fan with a motor that is nonoverloading at any
air-inlet temperature.
If fans are to be used in outside, unprotected areas,
the motor driver and other electric and control equip-
ment should be specified with enclosures suitable for the
environment (such as a totally enclosed, fan-cooled
motor). The fan itself can be protected with paint.
One should also keep in mind that the Air Moving
and Conditioning Assn. (AMCA)4 has standardized fan
and blower designations for spark-resistant construction,
wheel diameters, outlet areas, sizes, drive arrangements,
inlet-box positions, rotation and discharge, motor posi-
tions and operating limits.
All the foregoing points are covered in AMCA Stand-
ards 2401 through 2410. Reference to these standards
permits specifying fan characteristics in a precise manner.
Process plants generally use Class IV construction, which
covers fans for greater than 1225 in. water-gage total-
pressure-rise.
94
JANUARY 22, 1973/CHEMICAL ENGINEERING
-------
Investment Costs
Costs of centrifugal fans are difficult to estimate ac-
curately because of the many different fan types, classes
and arrangements available. Recent purchases indicate
about $30 to $40/brake horsepower (BHP) for fans of
about 50,000 cu.ft./min., 45-in. water-gage pressure
(500 BHP), and about S60/BHP for fans of 25,000
cu.ft./min., 40-in. water-gage pressure (250 BHP).
The costsjust mentioned include: fan; totally-enclosed,
fan-cooled (TEFC) motor drive; coupling; coupling
guard; baseplate mounting for the small units; and
standard materials of construction. The costs do not in-
clude starters, accessories or controls. A generalized guide
for fan costs—up to about 20 in. water-gage pressure and
1,000 BHP—can be found using a nomograph prepared
by J. R. F. Alonso.6
The price of small, high-pressure, single-stage fans (up
to 100 cu.ft./min.) is comparatively high. For example,
a 65-cu.ft./min. fan at 14-in. water gage recently cost over
$1,000/BHP. Blowers, in general, range from $50 to
$60/BHP. When stainless steel construction is required
by the process, prices may be two to three times as high
as those for standard materials.
Drives and Couplings
Whenever a fan is to be driven by a turbine or other
variable-speed device, it is important to ascertain that the
integrity of the rotating parts is assured up to the trip
speed of the driver. In the case of steam turbines, trip
speed is about 10 to 15% above normal running speed.
It is advisable to include in the specifications a rotating
assembly test at the trip speed. Advantages and disad-
iiwi(iiHHnHRiHtiiiiinmi(iHHiiiiiitiniiiiiiii(HiitiRiiiiHiniiitnHnnRnHHiiiiiiniiiiHnitiiiitiMiiiHnitiiiiHrniiiiiiitnmrirmm«niiiiiiiiiiiiii
Information That Should Be Provided on a
Data Sheet-Table III
CM Characteristic*
Composition
Molecular weight
Flow required
Operating Conditions and
Characteristics
Suction pressure and
temperature
Discharge pressure and
temperature
Required power
Speed of Ian
-Rotation of fan
Diameter ol impeller
Number of stages
Type of tan
Starting torque
Moment of inertia
Bearings and Lubrication
Type of bearings (radiaJ
and thrust)
Lubrication system and
recommended lubricant
Connection*
Size and rating
Location
Drain connections
Accessories Required
Driver (motor, steam turbine.
hydraulic turbine, other)
Coupling information (supplier,
type. size, etc.)
Gear required
Control (dampers, inlet guide-
vanes, variable-speed drive,
variable-pitch blades—axial,
actuators)
Safety devices (pressure,
temperature, vibration)
Filters (inlet), or screens
Cleanout holes
Noise-attenuation equipment
and lagging
Construction and Material*
Specifications
For case, impeller, shaft and
other parts
Type of seals
Shaft diameter
Testing and Miscellaneous
Required testing
Inspection
Witnessing tests
Test driver
Weights
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Pros and Cons of Variable-Speed
Drives for Fans—Table IV
Advantages
Disadvantages
Direct-Current Motor
Wide range of adjustable,
stepless. speed variation
High initial cost, requires
a.c -to-d.c. conversion
equipment; presents mainte-
nance and installation problems
A.C. Variable-Speed Motor
All the advantages of d.c. variable-
speed drives, many do not have'
commutators of brushes to
maintain.
High initial cost
Two-Speed A.C. Motor
Simple speed change
Limited choice of only two
speeds—in which category
are included single-winding.
• consequent-pole machines (with
a 2:1 speed ratio), and pole
amplitude-modulated motors,
which have a speed ratio of
3:2 to 3:1.
Hydraulic Drives
Low-cost; simple; they allow motor
to start against low torque;
generally trouble-tree.
Nllllimilllllllllllllllllllllllllllllllllllllllimimilllllimilllllllllllllllllllllllllimillimmilll
Inefficient at other than full speed:
some hydraulic clutches are
difficult to control near the full-
speed position; an auxiliary
lube-oil system is required.
iiiiniiiiiiiilliiiiiiiiiiimmiiliiiiiiiiiiililimiiilliiiuii
vantages of several variable-speed drives for fans are
listed in Table IV.
When a gear is used between the fan and the driver,
a torsional analysis of the entire train (including drive,
couplings, gear and fan) should be made. This analysis
can be done by the seller of the fan or driver, and should
be purchased with the unit.
The American Gear Mfrs. Assn. (AGMA)8 recom-
mends that a service factor be used with gears. The
following factors are usually applied to the power capa-
bility of the driver to obtain the rated horsepower of the
gear unit:
Type of Fan
Centrifugal units,
including blowers
and forced-
draft fans
Induced-draft fans
Industrial and
mine fans
Motor Turbine
1.4
1.7
1.7
1.6
2.0
2.0
Internal
Combustion
Engine
(Multi-Cylinder)
1.7
2.2
2.2
niiBiiiiniiiiiiiiiiiiiiiiisniiiiiiwniiiiiiniiiiiiiinDiiiiniiiiiinniiuiiiiiiiiiiiiiiiiiiiiiiniiiHnnmimiiiiiiiiui limniiuliiimJliiiH
Gear losses of about 2 or 5%—depending on the type
and quality of the gear unit—are added to the power
required from the driver.
Accessories for the gear, depending on its size, may
be bearing-temperature gages, vibration detectors, type
of thrust bearing (tapered-land,* tilting-pad, antifriction,
shoulder, etc.), and a type of lube-oil system. For fans
•Defined in Ret 1. pp 8-170. 8-171.
CHEMICAL ENGINEERING/JANUARY 22, 1973
95
-------
SELECTING FANS AND BLOWERS ...
Motor: 600 hp., 3,550 rpm.
Fan: WK2 •= 5,690 Ib.-ft.2
SPEED-VERSUS-TORQUE CURVE to estimate tan horse-
power ot air, or gas, flow—Fig. 16
in refinery applications, American Petroleum Institute
Standard 613 can be applied.9
In addition to the above considerations, a decision as
to whether to purchase the driver separately or with the
fan must be made. If the fan to be purchased is large,
it is advisable to buy the motor driver with the fan, to
avoid the coordination problems encountered in the se-
lection of the motor and coupling.
The fan manufacturer must determine an expected
speed-torque curve (Fig. 16), as well as the moment of
inertia of the fan. This will enable him to select a motor
to suit the electrical and area classification of the appli-
cation. The coupling, which must suit both the fan and
driver shaft, should be supplied by the fan vendor.
Other items subject to coordination are sole plates, or
a baseplate under fan and motor, and coupling guard.
Some of the complexities of coordination are illus-
trated by this actual example:
An industrial, forced-draft fan, for approximately
150,000 cu.ft./min. at 38-in. water gage, and requiring
1300 BHP on air, was to have a dual drive (motor and
turbine), with one-way clutches so that either motor or
turbine could be serviced with the fan running.
The fan and motor were to be purchased overseas from
different manufacturers; the turbine, gear, clutches and
coupling in the U.S., through the turbine manufacturer.
With so many vendors involved, none could be held
responsible for the unit, with the result that the fan could
not be run with either motor or turbine in the factory.
It is much simpler, and probably less expensive overall,
to purchase all the equipment through one vendor. This
is especially desirable when problems are encountered
in the field, and responsibility for repairs is difficult to
pinpoint.
A good treatment of fan motors, and how moment of
inertja (WR2), weight of the fan, and other factors affect
motor selection can be found in Ref. 7.
Fan Controls
The throughput of centrifugal or axial fans may be
changed by varying the speed of the fan, or by changing
pressure conditions at the inlet and/or outlet with damp-
ers or with inlet guide-vanes. Axial-flow fans may be
controlled also by varying the pitch of the blades.
Of these methods, the most efficient is changing the
speed. Since, however, this feature is not generally avail-
able—because fans are ordinarily driven by constant-
speed motors—other means of varying flow must be
resorted to. The next best way of accomplishing this is
by means of variable inlet guide-vanes, which must be
purchased with the fan.
The most common of the controls used with constant-
speed centrifugal fans is the inlet damper. As the damper
closes and reduces inlet pressure, the pressure ratio across
the machine increases, so that the operating point on the
fan curve moves in the direction of lower flow.
Sometimes, the extra pressure drop is taken by a dis-
charge damper, but the power wasted is greater than with
inlet-damper control. Partially closed dampers on axial
fans may increase power as they decrease flow, in accord-
ance with the general performance characteristics of the
axial fan.
Surge, which is a condition of unstable flow in dy-
namic-type compressors, can also occur in fans. This
happens at less-than-normal flowrates, when the fan (or
compressor) can no longer develop the required pressure
ratio. On fans or blowers of more than about 2 psi. (55
in. of water) and 150 BHP, surge can be damaging. Some
type of antisurge control should therefore be considered.
Occasionally, on high-head fans in services other than
air, it may be necessary to bypass some of the gas from
the discharge back to the suction side, to keep the flow
above the minimum needed to avoid surge. This gas must
be cooled, and it should be taken from a point in the
discharge line upstream of the discharge backflow pre-
venter (if used). On air service, flow can be maintained
above the surge point by dumping air to the atmosphere
or by bleeding some air into the suction side (on in-
duced-draft fans).
To avoid possible reverse rotation after shutdown,
some kind of backflow preventer should be considered
on fans exhausting gas from a closed system.
Vibration
Vibration limits depend on speed. A maximum peak-
to-peak amplitude (measured on the bearing caps), as
follows, would be classified as "good." Vibrations 2.5
times larger than the following values would be "slightly
rough," but still acceptable after some use.
Rpm. "Good" Vibration Amplitude, In.
400 0.003
800 0.002
1,200 0.0013
1,800 0.0008
3.600 0.0005
At the lower speeds—say, less than 800 rpm.—accept-
able vibration-amplitude values taken from charts may
96
JANUARY 22, 1973/CHEMICAL ENGINEERING
-------
not be a good criterion. It is then better to limit shaft
vibratory velocity to 0.1 in./sec.
Vibration monitoring should be considered for fans in
critical service, to provide an automatic warning when
the machine's vibration reaches a trouble-indicating
level. Fan vibration due to imbalance can be minimized
by asking the factory to balance the entire rotating as-
sembly (fan and shaft). With the larger fans, the impeller
may be shipped disassembled to the user. Fan manufac-
turers should be responsible for field balancing to a level
agreed upon with buyers.
Coping With Noise
Noise-attenuation equipment must be considered for
fans that exceed established noise limits. It is, however,
most difficult to specify a fan's maximum noise level.
The sound power generated by a fan depends on the
flow, fan-pressure level, and impeller type and configura-
tion. It is not possible to design a quiet fan at high-
pressure levels; for fans of 2 to 3 psi., a sound-power
level in the range of 110 to 130 db. is not uncommon.
Obviously, this type of fan must either be installed in
an unmanned area of a plant, or be suitably modified
with sound attenuators to bring the noise level within
acceptable limits. The Walsh-Healy Act10 and the Occu-
pational Safety and Health Act (OSHA)14 specify per-
missible sound levels in working areas.
To bring the noise level down, silencers, insulation
around ducts, and lagging—or a housing around the
unit—can be considered. Fan-pressure losses in cylindri-
cal attenuators are generally 2 in. water gage or less.
Silencing equipment can be placed in the inlet or
discharge ducts near the fan, or around the fan casing.
Fan suppliers can ordinarily furnish data on the sound
level generated by a particular fan. These data are usually
taken from tests at the factory on typical field installa-
tions of similar fans.
When rated under operating conditions, inlet and dis-
charge silencers usually provide the required noise atten-
uation. They are made for insertion in round or rectan-
gular ducts, in standard or special materials, and with
special acoustic fills for corrosive atmospheres.*
Sometimes, when an application is in the range of a
compressor manufacturer's equipment, the cost of a fan
plus associated noise-reducing accessories can be less
than that of a compressor that does not exceed the speci-
fied maximum sound level.
Flange Loading, Shaft Seats
Fan manufacturers generally require no load to be
transmitted to the fan casing due to attached ducts. This
is desirable, but when it is unavoidable to impose
loads—due to thermal expansion or weight—it may be
possible to strengthen the fan casing to avoid distortion
and misalignment.
Concerning leakage, a certain amount is usually toler-
able on fans and shaft seals because a prime consid-
eration in selecting fans is ease of seal replacement. Seals
can be made of felt, rubber, asbestos or other packing.
•To estimate noise levels, consult Ref. 11 and 12.
CHEMICAL ENGINEERING/JANUARY 77 1973
CONTACT-TYPE SEAL holds all seal faces in constant con-
tact with shaft to prevent leakage—Fig. 17
When leakage cannot be tolerated, a contact-type seal
may be considered. One of these (Fig. 17) has a centrally
located, annularly compensating feature, preloaded to
hold all seal faces in constant contact. This seal is claimed
to be suitable for liquids, gases, vapors and fine solids
in the chemical, petroleum, pharmaceutical and food
industries.
Systems Analysis
For a fan to function properly, one must understand
the effects of the fan system on the fan itself; otherwise,
neither the system nor the fan will work well. A fan
system consists of the whole air path—usually a combi-
nation of pipes or ducts, coils, filter, flanges and other
equipment.
In a fixed system, the volume flowrate in cu.ft./min.
will have an associated pressure loss, which is caused by
the system's resistance. Head loss for a fan system is
calculated similarly to tfae head loss for flow of fluids
in a process piping system. First, the complex system is
broken down into its component parts, with known pres-
sure-drop values. The summation of all these resistances
yields the total resistance of the system.
A system's total resistance would include the resistance
in the main duct to the fan inlet; the one in the main
duct from the fan discharge to the end of the duct; and
the ones in branch pipes (or ducts), filters, dust collectors,
grilles, or other pieces of equipment. A novice tackling
an extensive project would do well to consult a specialist
in the field.
In a typical fan-system curve (Fig. 18), the static pres-
sure (P,) of the system is a parabolic function. The point
of operation (PO) is located at the intersection of the
fan static pressure and the system's Pt.
Occasionally, fans operating at other than the design
PO are unstable and cause pulsation. This can damage
97
-------
SELECTING FANS AND BLOWERS . ..
FAN-SYSTEM CURVE locates operat-
ing point at intersection of fan's static
pressure and system's P,-Fig. 18
&$£&&&*^-^~-'*- Airflow,thousand? of^.^min^jjvife:; ^r"^bT 1
s3si3S?£^ii3kta--a'? ,v^fc;,\^;^/,'>;^"^«^s^iaii^?:^^asS^;L^^;-^i^.JW •- ,4
the fan, the system or both. To overcome the problem,
a fan should be selected so that its PO always falls in
the stable range—i.e., in the down-sloping portion of the
flow-versus-pressure-rise curve, and preferably at some
flow that corresponds to only one pressure-rise point. On
Fig. 18, for instance, this corresponds to flows exceeding
17,500 cu.ft./min.
Another important factor in a system's design is the
choice of fan blade. For example, according to Fig. 19,
there is less likelihood of paniculate buildup on the blade
if a forward-curved blade is used, but there is a tradeoff
in fan stability. The backward-inclined-blade fan is in-
herently more stable; forward-curved blades must be
carefully matched to the duct system.
Materials of Construction
The materials of construction and the types of seals
used in a fan depend on the composition of the gas
handled. Standard materials include cast iron and carbon
steel for casings; aluminum and carbon steel for im-
pellers; and carbon steel for shafts. In some cases, other
materials may be required. For instance, if the fan is
required to move a wet mixture of ammonia, carbon
dioxide and air, stainless-steel (304 or 316) construction
for all parts in contact with the gas may be necessary.
Plastics reinforced with fiber glass (FRP) are now also
accepted materials of construction for fans and blowers,
even though FRP units have pressure limitations. For
example, a 7J/2-in.-dia. fan at 120,000 standard cu.ft./
min. was built for only a 2-in. maximum static pressure.
With backward-inclined blades, FRP fans can handle
flows of 65,000 cu.ft./min. at 3-in. static pressure (8,200
ft./min. tip speed). With special supports and rein-
forcement, FRP fans with radial blades can handle pres-
sures up to 20 in. of water, at flowrates up to 45,000
cu.ft./min. (16,500 ft./min. tip speed).
Corrosion resistance can be enhanced with special
coating materials, which are often readily available from
fan manufacturers at lower costs than special materials.
However, the successful application of coatings depends
largely on experience. A coating's suitability is usually
proven by its previous use in a similar service.
Coatings are generally classified as air dry—such as
special paints, asphalt, epoxy, air-dry phenolic, vinyl,
silicone, or inorganic zinc—and baked—such as polyester
(with or without fiber-glass reinforcement), baked poly-
vinyl chloride, baked epoxy, and baked phenolic.
Whenever a coating is specified, its extent and thick-
ness must be clearly indicated. The surface preparation
and method of application must be in accordance with
recommendations of the coating manufacturer. Gener-
ally, sand-blasted or shot-blasted surfaces may be in-
cluded in the price quoted by the fan manufacturer, but
special preparations are not.
Coating the entire inside and outside surfaces may be
impossible with baked coatings. In some instances, just
coating the airstream surfaces may be satisfactory, and
much less costly. For example, in an application for a
blower to compress 100 actual cu.ft./min. of ammonia
and hydrogen sulfide from 18 to 21 psia., the blower
manufacturer recommended lining the internal parts only
with Heresite (Heresite and Chemical Co., Manitowoc,
Wis.) at about $1,000 per blower.
Allowable temperatures of coatings should be higher
than the expected operating temperatures by an apprecia-
ble margin. Rubber, also sometimes used, is limited to
about 180 F. The tip speed of wheels lined with rubber
is about 13,000 ft./min. (or lower for thick coats).
As a rule, the upper limit of tip speed for modern,
large, industrial fans is approximately 40,000 ft./min. At
such a high speed, pressure increases of 25% are attain-
able on air. Whenever the wheel is coated with any
material, it is necessary to use slower speeds. Thus, the
pressure ratio of the machines becomes limited.
Linings of epoxy resins, such as Corohne (Ceilcote Co.,
Berea, Ohio)^ or polyester resins, such as Flakeline (Ceil-
cote Co.), are also used successfully to protect fan sur-
faces from corrosive gases. These and similar coatings
can be used up to tip speeds of 20,000 to 28,000 ft./min.
98
JANUARY 22, 1973/CHEMICAL ENGINEERING
-------
BLOWER-PERFORMANCE for dif-
ferent impeller types—Fig. 19
However, if abrasive particles, dust or liquid droplets are
present in the gas stream, such coatings may fail. In such
instances, fans must be made of satisfactory metals.
To reduce costs, a manufacturer may recommend coat-
ing the shaft and some slower-moving portions of the
fan with a lining, and using bolted, riveted or sprayed-on
metal surfaces on the high-speed portions. Colmonoy 5
(Wall Colmonoy Corp., Detroit, Mich.), Stellite 3 and 5
(Stellite Div., Cabot Corp., Kokomo, Ind.), and
Inconel X (International Nickel Co., Huntington,
W. Va.) may be used satisfactorily in environments sub-
ject to hydrogen sulfide stress-corrosion. When the con-
struction of the fan allows, plates of Inconel X, Hastelloy
(Stellite Div., Cabot Corp.) or other suitable material
may be bolted or riveted on to minimize erosion.
Fan construction methods are as numerous as fan
manufacturers. Casings and impellers may be riveted,
welded, cast or bolted. Bearings may be sleeve-type or
antifriction and—depending on speed, load and temper-
ature—may be self-lubricated or require a lube system.
Minimum life and maximum allowable temperature and
antifriction bearings—according to American National
Standards Institute standards—should be specified. Gen-
erally, 30,000 hr. minimum is a satisfactory life, but for
heavy-duty fans 50,000 hr. is a conservative figure. Bear-
ing temperature, as measured internally, should not ex-
ceed 180 F.
Performance Tests
In the factory, fans are tested with open inlets and
smooth, long, straight discharge ducts. Since these condi-
tions are seldom duplicated in. the field, the result often
is reduced efficiency, impaired performance and—in
some extreme cases—failure of the fan or overloading
of the driver. Although fans are affected more by inlet
than by discharge conditions, care should be exercised
in both inlet and outlet ducts to provide proper flow
patterns.
Axial-flow fans are more prone to be affected by inlet
CHEMICAL ENGINEERING/JANUARY 22, 1973
conditions than centrifugal ones. Consider the case of an
actual 33-in.-dia., l,000-rpm., vane-axial fan, which had
a 70% efficiency and a total pressure of 1.0-in. water gage,
when the inlet was connected to a smooth 90-deg. elbow
(outside to inside radius ratio of 2).
At constant flow, when a miter elbow with turning
vanes was used instead of the smooth elbow, the effi-
ciency dropped to 54%, and the total pressure to 0.8-in.
water gage. With a miter without turning vanes, the
efficiency was 45%, and the total pressure 0.6-in. water
gage.13
Although shop testing a fan can disclose its mechanical
integrity or its aerodynamic performance, the ability to
perform tests may be governed by the size of the factory
test stand. For units in critical service, this may be an
important factor in fan selection.
A shop mechanical test—which should last for at least
2 hr. at maximum continuous speed—should certainly be
obtained if at all possible. This should be a witnessed
test to obtain readings of bearing temperatures, oil flows
and vibration amplitudes. For turbine-driven fans, an
over-speed test should also be included.
Performance tests are recommended when:
• The fans in question are large centrifugal or axial
units, or fans whose design has not been previously man-
ufactured, or is a scaled-up version of an existing fan.
• The quoted efficiency is at the upper end of the
scale for the type of fan, and utility costs are high.
• The fan is to be in critical service, and missing
the guaranteed operating point by any margin would
be costly.
Because testing codes (Ref. 2 and 4) outline only test
methods (no penalties are assessed for not meeting per-
formance promises), buyers should specify limits of ac-
ceptable performance in their purchase orders.
Field-performance tests are very difficult to make with
any degree of accuracy, but if the performance in the
field is not satisfactory, the purchaser should have the
option of requiring the seller to supervise such a test.
Even though factory conditions for measuring pressure,
-------
SELECTING FANS AND BLOWERS ...
temperature, system air humidity, and power consumed
cannot be duplicated in the field, major fan-performance
deficiencies can be demonstrated and corrective action
initiated. Sufficient instrumentation, or space for its loca-
tion, should be provided in the layout of fans that may
need field testing.
Installation Guidelines
Installing the fan on a heavy base is essential for a
good, long, trouble-free life. Fans mounted on concrete
slabs at ground level are ideally placed. If a fan must
be mounted on an elevated structure, such as on top of
a furnace, extra care in balancing must be taken to avoid
shaking the structure. For critical installations, vibration
analysis of the entire structure is necessary.
A rule of thumb for fans installed on concrete slabs
at grade level is to use a weight of concrete about six
times the mass of the rotating elements of the unit.
As the fan is installed on its foundation, shims and
sole plates should be used to aid in the alignment of
driver and fan (and gear, if used). Alignment is especially
critical on induced-draft fans running at elevated tem-
peratures. Here, allowances should be made for move-
ment as the casing, shaft and impeller reach operating
temperature. If possible, vibration should be continuously
monitored while the unit is heated to its normal running
temperature. A gradual increase in vibration is a good
indication of poor alignment due to temperature rise.
When ball or roller bearings are used on V-belt-driven
shafts, caution should be exercised to prevent excessive
preloading of the bearings (which could bend the shaft)
while the V belts are tightened. Units that have V-belt
drives should have the sheaves mounted with the im-
peller at the factory, at the time that balancing is done.
If the bearing temperature exceeds 180 F., special
lubricants may be used. If, however, the temperature is
lower than about —30 F., not only will special lubricants
have to be used but antifriction bearing metals may have
to be specially processed by the bearing manufacturer.
Safety devices available for fans are the same as those
used on centrifugal compressors. When a fan requires
a separate lubricating oil system, adequate pressure and
temperature protection must be provided to avoid run-
ning the fan unit without lubrication (even during coast-
down times due to power failure). Vibration switches are
recommended on high-speed fans that are in hot or dirty
service, as well as on most axial-type units.
A survey of plant startup problems for the last seven
years indicates that the incidence of failures attributed
to fans and their drives has been very small. The failures
on centrifugal fans were minor and were easily resolved.
The ones on axial fans were more severe but caused only
minor damage to other equipment. •
Acknowledgments
The following companies provided information and/or
illustrative material for this report: American Standard,
Inc., Industrial Products Div., Detroit, Mich. (Fig. 3, 4,
9, 10); Buffalo Forge Co., Buffalo, N.Y. (Fig. 8); Castle
Hills Corp., Piqua, Ohio; Clarkson Industries, Inc.,
Hoffman Air Systems Div., New York, N.Y. (Fig. 2, 14);
Ernest F. Donley's Sons, Inc., Cleveland, Ohio (Fig. 17);
Dresser Industries, Inc., Franklin Park, 111.; Fuller Co.,
Lehigh Fan & Blower Div., Catasauqua, Pa.; Garden
City Fan & Blower Co., Niles, Mich.; Joy Mfg. Co.,
Pittsburgh, Pa.; Lau, Inc., Lebanon, Jnd.; The New York
Blower Co., Chicago, 111. (Fig. 6, 7, 18); Niagara Blower
Co., Buffalo, N.Y.; Westinghouse Electric Corp., Westing-
house Sturtevant Div., Boston, Mass.; Zurn Industries,
Kalamazoo, Mich. (Fig. 5).
References
1 "Mark's Standard Handbook for Mechanical Engineers," 7th ed., T.
Baumeister, ed., McGraw-Hill, New York (1967)
2. "ASME Standard PTC-1I," Test Code for Fans, American Soc. of
Mechanical Engineers, New York.
3. "API Standard 617," American Petroleum Institute, New York.
4. "AMCA Standard 210-67," Test Code for Air Moving Devices, Air
Moving and Conditioning Assn.. Park Ridge, 111.
5. Fans, A Special Report, Paver. Mar. 1968.
6. Alonso, J. R. F., Estimating the Costs of Gas-Cleaning Plants, Chem.
£ng, Dec 13, 1971, p. 86.
7. Rajan, S., and Ho, T. T., Large Fan Drives in Cement Planis, IEEE
Transactions ICA. Vol IGA-7, No. 5, Sepl.-Oct. 1971.
8. American Gear Mfrs. Assn., Washington, D.C.
9. "API Standard 613," High-Speed, Special-Purpose Gear Units for
Refinery Service, 1st ed., American Petroleum Institute, New York
(1968).
10. Walsh-Hcaly Act, Federal Register, Vol. 34, No. 96, May 20, 1969;
revised. Jan. 24, 1970.
11. "ASHRAE Guide and Data Book," Chapter 31. American Soc. of
Heating, Refrigerating and Air Conditioning Engineers, New York
(1967).
12. Graham. J. B., How To Estimate Fan Noise, Sound and Vibration,
May 1972.
13. Christie, D., Fan Performance as Affected by Inlet Conditions,
ASHRAE Transactions, Vol. 77. Pan 1, 1971, pp 84-90, American
Soc. of Healing, Refrigerating and Air Conditioning Engineers, New
York.
14. Occupational Safety and Health Act. Federal Register, Vol. 36, No.
105, May 29, 1971.
15. "Fan Engineering," Buffalo Forge Co., 1970 ed., Buffalo, N.Y.
Meet the Author
Robert Pollak is engineering special-
ist with Bechtel, Inc.. Refinery and
Chemical Div , P O Box 3965. San
Francisco. CA 94119, where he is
engaged in the specrtication and
selection of centrifugal and recipro-
cating compressors and fans, as well
as of their motors, steam turtxnes
and other drivers A graduate of the
University of Illinois, he holds an M.S-
degree in mechanical engineering.
He is a member of the American Soc.
of Mechanical Engineers
Reprints
Reprints of this report on fans and blowers will be available shortly. To order, check No. 173 on the reprint order form in
the back of this or any subsequent issue. The price per copy is $2.
100
JANUARY 22, 1973/CHEMICAL ENGINEERING
-------
ITEMS
Fans And Blowers
Paul Cheremisinoff
Pollution Engineering Magazine
-------
Fans and Blowers
By PAUL N. CHEREMISINOFF, P.E., Field Editor, and RICHARD A. YOUNG, Editor, POLLUTION ENGINEERING
The most common method for moving gases under
moderate pressures is by using a fan. The fan is the
heart of any system that demands that air be supplied,
circulated and removed in a way that provides a safe
and comfortable environment. For industrial plants
the needs of heating, ventilating, air conditioning, and
pollution control are fulfilled by fans.
There are two general classes of fans: axial ajid
centrifugal. Axial fans employ propellers and are
classed into three sub-types: propeller, tube-axial,
and vane-axial. Centrifugal fan flow is principally
radial rather than axial. Centrifugal fans are also
divided into three groups: forward, backward, and
radial. A distinction is made in engineering practice
between fans for low pressure and centrifugal com-
pressors for high pressure. A boundary separating the
two classes of equipment is set at 7 percent increase
in density of air from the inlet to the outlet. Fan action
is below this density increase and the gas is assumed
to be uncompressed.
Choice of a fan depends on flow volume required,
static pressure, condition of air handled, available
space, noise, operating temperature, efficiency, and
cost. Consideration should also be given to the type
of drive system to be used—direct or belt driven.
Axial Fans
The axial fan is used in systems which have low
resistance to air flow. This fan moves the air or gas
parallel to the fan's axis of rotation. Axial flow fans
use the action of their propellers to move the air in
a straight-through path. This screw action of the
propeller causes a helical type flow pattern. Propeller
type axial flow fans move air at pressures from 0 to
1 in. of water. Variations of the axial flow fan can
move air at somewhat higher pressures.
Fig. 1 Typical Ian characteristic curves.
One variation of the axial flow fan is the tube-axial
fan, which is the basic axial flow fan encased in a
cylinder. The fan's propeller in the cylinder helps to
collect and direct the air flow. The tube-axial fan can
move air or gas at pressures between 1/4 and 21/2 in.
of water.
A second variation of axial fan is known as the vane-
axial fan. This is an adaptation of the tube-axial fan
using air guide vanes mounted in the cylinder either
on the entry or discharge side of the propeller. The
vanes improve the fan's efficiency and increase work-
ing pressures from Vz to 10 in. of water by straighten-
ing out the discharge flow.
Principal advantages of axial fans are their economy,
installation simplicity, and small space requirements.
Their principal disadvantage, aside from operating
pressure limitations, is noise. This noise is usually
apparent at maximum pressure levels. These fans are
seldom used in duct systems because of the relatively
low pressures developed. They are well adapted for
moving large quantities of air against low pressures
with free exhaust, as from a room to outside.
Centrifugal Fans
Centrifugal fans or blowers move air perpendicular to
the fan's axis of rotation. This is done by air being
sucked into the center of the revolving wheel which
is on a shaft bearing the fan's blades. The air enters
the spaces between the wheel's blades and is thrown
out peripherally at high velocity and static pressure.
As this occurs, additional air is sucked into the center
of the wheel. This type blower is used where the frac-
tional resistance of the system is relatively high.
There are various adaptations of the centrifugal fan
which differ principally in the type of blade used.
Fig. 2 Typical plot o1 dimensionless fan characteristics.
PRESSURE' IN. OF WATER
HORSEPOWER,hp
EFFICIENCY. PERCENT
5 10 100
90
4 8 80~,
4 6 B IO 12
VOLUME, cfm XIO3
16 18 20
hp IN PERCENT MAX. hp
PRESSURE IN PERCENT OF TOTAL MAX. PRESSURE
EFFICIENCY. PERCENT , /
100
90
80
70
60
RESSURE MORSE--
POWER
K> 20 SO 4O SO 6O 70 SO 90 KX>
PERCENT OF WIDE OPEN VOLUME
-------
Fig. 4 The series of Design II BCS
airfoil centrifugal fans have an effi-
cient airloil blade contour and rec-
ord of trouble-free service. Cover-
ing AMCA classes /, II, III, and IV,
19 sizes span a range of wheel
diameters from 12'/4 to 73 In.,
with capacities up to 253,690 cfm.
(ILG Industries, Carrier Corp.)
Fig. 5 The Port-O-Way mobile weld-
ing fume exhaust system. Other
fume exhaust systems for under-
f/oor and overhead systems. (Am-
merman Co., Inc.)
Fig. 7 Fiberglass tube-axial fan for
handling extremely corrosive fumes
often found in chemical, food pro-
cessing, metalworking, laboratories,
pharmaceutical, and other indus-
tries. All interior parts exposed to
the air stream have smooth contact
molded surfaces for minimum fric-
tion loss and maximum chemical
resistance. (Industrial Plastic Fabri-
cators, Inc.;
Fig. 6 Low-silhouette centrifugal
roof ventilator fan needs no damper
or rain cap to prevent rain or snow
from passing through the fan dur-
ing oft periods. Made of a tough
corrosion-resistant reinforced poly-
ester plastic, fan provides efficient
fume removal for severe-duty con-
ditions. Unit is capable of moving
15,000 cfm and developing pres-
sures up to l'/2 in. s.p., available
in ten sizes from 6 to 36 in. d/'a.
(Hell Process Equipment Corp.)
Blade types depend on space limitations, efficiency
demanded by the system for particular load conditions,
and allowable noise levels. There are three general
types of blades that are used in blowers: forward-
curved, backward-curved, and straight or radial type.
In the forward-curved type centrifugal fan, the blade is
inclined at the tip toward the direction of rotation. This
is the most widely used type centrifugal fan for general
ventilation purposes. It operates at relatively low
speeds and produces high volume air flow at low
static pressure. The fan is quiet, economical, space
efficient, and lightweight. Because of the inherent
design of its blade configuration and low operating
speeds, the forward-curved blade fan cannot develop
high static pressures. Fans with backward-curved
blades are more suitable for higher static pressure
operation. They operate at about twice the speed of
forward-curved centrifugal fans and have higher effi-
ciency and a non-overloading horsepower curve. The
higher operating speeds, however, require larger
shaft and bearing sizes and greater care In system
balancing. The radial type fan has a "blade curvature
tangent to the radius at it's outer tip': It Is designed to
handle low air volumes at relatively high static presr
sures. Because of its wheel design, it is also suitable
for handling heavily dust laden air,
When selecting a fan, one must consider which type
of fan will fit the purpose and be most economical to
operate. Cost considerations before purchasing
include those for operating, maintenance and equip-
ment. It is not necessary to design a new fan for each
new application.
Table 1 Relative Characteristics of Centrifugal Fans
Forward Backward
Curved Curved Radial
Blade Blade Blade
First cost
Efficiency
Operational stability
Tip speed
Abrasion resistance
Sticky material handling
Low
Low
Poor
Low
Poor
Poor
High
High
Medium
High
Medium
Medium
Medium
Medium
Medium
Medium
Good
Good
COUNTER-CLOCKWISE
TOP TOP TOP DOWN BOTTOM BOTTOM BOTTOM UP
HORI- ANGU- ANGU- BLAST HORI- ANGU- ANGU- BLAST
ZONTAL LA* LAR ZONTAL LAR LAR
DOWN UP- UP. DOWN
Fig. 3 Centrifugal fan rotation and discharge. Two directions
of discharge and 16 discharge positions are possible with
centrifugal fan. Rotation direction will be determined by the
tan function and is specified according to the view from
drive side.
POLLUTION ENGINEERING
25
-------
Fig. 8 The Barry Series 600 indus-
trial Ian is a complete, factory as-
sembled unit designed for use in
the exhaust systems, air circulation
and material conveying systems.
(Barry Blower Co.)
Fig. 9 Corrosion-resistant all-PVC
construction centrifugal tan. (Duall
Industries, Inc.)
Fig. 10 Regenerative blowers are
ideally suited lor low pressure or
low-to-medium volume airilow appli-
cations. The quiet, maintenance-free
units are capable of developing
flows up to 200 cfm and vacuums
or pressures to 4 psi. (Rotron Inc.)
Fig. 11 A 1/3-hp motor driven fan
lor small air flow requirement. Units
are available in cast aluminum, steel
in larger sizes, stainless or coated
for corrosion resistance. (Cincinnati
Fan & Ventilator Co.)
A fan's capacity is measured in ft3/min which is
equivalent to the number of Ib/min of air divided by
the density in Ib/ft3 at the system's inlet. In order to
meet the fan's capacity, the specified horsepower
motor must be used to drive the fan. Belt driven fans
are used for motor requirements generally between 1
and 200 hp. Direct drive motors are generally used for
fans requiring drive motors larger than 200 hp. Direct
drive fans are limited to the fan's motor speed. Prin-
cipal advantages of this type of drive are generally
lower maintenance cost and less power transmission
loss than for belt driven fans.
When ordering a fan, the following data are required:
Flow volume—the volume of air the fan will handle at
the prevailing temperature.
Composition of the gas handled—moisture, dust load,
corrosive gases present, etc.
Static pressure—the resistance the fan must overcome
to deliver the required volume of air from process
intake to exhaust stack exit.
Operating temperature—this parameter affects the vol-
ume of air handled and the materials of construction.
*
Efficiency—volume of air delivered per unit of elec-
trical energy. This parameter will determine the oper-
ating costs of the unit.
Noise—the best guide to the selection of a suitably
quiet fan is successful previous performance.
Space requirements and equipment layout—including
size and orientation of fan inlet and outlet.
Cost—initial cost of the equipment.
When ordering a fan, the pollution engineer must
include the purpose for which the fan is to be used
and information concerning the applicable size of
duct work. The supplier can then make sure the fan
will meet pressure requirements. The fan must be able
to accelerate the air from a given entrance velocity to
the velocity required at the exit. Fans exhausting
through stacks must maintain a minimum exit velocity
(usually 60 fps minimum) to ensure that the exhaust
gas will escape the turbulent wake of the stack. An
exit velocity on the order of 90 or 100 fps may be
necessary. The supplier should be advised of any
unusual conditions that would be encountered in
pollution control service.
When system requirements are known, the main points
to be considered in fan selection are: efficiency,
reliability of operation, size and weight, speed, noise
and cost. For help in choosing the most suitable fan,
consult manufacturers' tables or curves that show the
following factors for each fan size, operating against a
wide range of static pressures.
1. Air volume handled, in ft3/min at standard condi-
tions (68 F, 50 percent relative humidity, density
0.07496 Ib/ft3),
26
JULY 1974
-------
Fig. 13 Centrifugal fan of fiberglass
reinforced plastic. This style Ian
implements the non-overloading
backward!/ inclined tan wheel. (Jus-
tin Pacific Corp.)
\
Fig. 12 Fan capable of volumes to
165,000 cfm and static pressures
to 8 in. w.g. This series is available
in three impeller types: paddle
wheel (high strength, sell-cleaning),
forward curved radial (low tip speed,
high efficiency) and backward
curved (high efficiency, non-over-
loading characteristics). (Ceilcote
Co.)
i:
/&f*.
fr: 5^S^^r^;S5j
<£C^^*^
g&e&3&~g;£***~
Fig. 14 A Lehigh fan creates negative pressure for a Dracco
Mark /( dust co//ector system handling 120,000 scfm of
phosphate dust-laden air at a bulk loading facility in Florida.
Both the fan and filter-bag collector were furnished by
Fu//er Co.
Fig. 15 Centrifugal fan, belt drive
and motor mounted on the fan
pedestal. The fan wheel, in this case,
is designed to handle an air stream
containing granules, fumes, dust,
etc. (Carter-Day Co.)
Fig. 16 Blower series shown is
single stage, direct drive, 3600 rpm,
with precision permanent mold cast
aluminum housings and heavy
gauge fabricated aluminum im-
pellers mounted on strong durable
hubs. (North American Mfg. Co.)
2. Air velocity at the outlet,
3. Fan speed, rpm,
4. Brake horsepower,
5. Peripheral speed, or blade tip speed, fpm, and
6. Static pressure, in. of water.
Corrosion Resistance
Two major problems facing fan and blower users are
excessive temperatures and corrosive atmospheres.
Mild steel is suitable for fan construction in dry air up
to a temperature of 900 F. Temperatures exceeding
this cause scaling. For such service, steel may be
coated with a protective alloy. High gas stream tem-
peratures and/or corrosive atmospheres cause struc-
tural and corrosive problems. Methods used to solve
these problems include lowering gas temperatures and
controlling the concentration of corrosives in the
exhaust gas. With lower temperatures, the fan can be
coated with a layer of lead bonded to its surface, or a
layer of vulcanized rubber placed on the fan or a layer
of plastic sprayed on for corrosion protection. Fans
fabricated of higher resistance metals such as stainless
steel, monel and Hastelloy can be used with excellent
results if it is impractical to lower temperatures in a
corrosive atmosphere. These latter systems however
are substantially more expensive.
Fans fabricated from fiberglass-reinforced plastics are
also used under corrosive conditions. Fiberglass plas-
tics are strong, lightweight, economical and corrosion
resistant. Fans can also be coated with fiberglass-
reinforced plastics for protection. Unfortunately, the
maximum temperature at which fiberglass can be used
is 200 F. Aluminum and aluminum alloys also have
corrosion resistant properties, which can be taken
advantage of for specific applications with a maximum
operational temperature of 300 F.
Fan Noise
Fan noise is a complex mixture of sounds of various
frequencies and intensities. An absolute noise rating
cannot be measured; measurements are limited to
comparative intensities of noise produced at some
given point. Noise rating of a fan must specify the
measurement positions or points. The size of the room
and the form and material of the surfaces have an
effect on the noise intensity at a given point. It is
important, therefore, that measurements be compared
on a common basis such as the same room, at the
same location, with a satisfactory noise level meas-
uring instrument. These limitations should be recog-
nized and noise level values from manufacturers be
used as guides. The best guide to the selection of a
suitably quiet fan is successful previous performance
on a'job similar to the one under consideration. For
these reasons, there is no such quantity as an
absolute decibel rating of a fan.
POLLUTION ENGINEERING
27
-------
Fig. 17 Fan Separator un/t made of
rigid PVC are impervious to corro-
sive action of air stream contami-
nants. Design untilizes a centrifu-
gal fan and downstream filter to
separate liquid corrosive particles
from the air. Capacities range from
500 to 60,000 dm/single unit.
(Tri-Mer Corp.)
Fig. 19 Modern concept radial tip
fan featuring high efficiency and
sell-cleaning wheel that will tolerate
dirty atmosphere inside housing.
Heavy-duty construction for volumes
to 400,000 and 40 in. static pres-
sure. (Garden City Fan & Blower
Co.)
Fig. 18 Corrosion-resistant fiber-
glass vane-axial fan available in
capacities ranging from 2,150 to
106,000 dm. (Aerovent Fan Co.,
Inc.)
Fig. 21 Tube-axial fans are available
from 12 to 60 in. dia., capacities to
77,000 cfm, operation up to 3 in.
w.g. and 500 F. Cast aluminum ad-
justable pitch airfoil blades are
standard equipment with heavy
gauge welded steel housing and pil-
low block bearing construction.
Stainless steel, aluminum, or coated
fans are available for corrosive
duty. (American Fan Co.)
Fig. 20 Pressure air fan designed
to withstand the rugged service re-
quirements encountered in such in-
stallations as kiln exhaust, combus-
tion air, liquid agitation, pneumatic
conveying, product drying and cool-
ing. They are now available in 12
standard sizes for from 2500 to
44,000 cfm and up to 70-in. static
pressure, temperatures up to 800 F.
(Chicago Blower Corp.)
Fig. 22 This fan combines compact
size with low operating sound, high
efficiency and low total installed
cost. The aerodynamic configuration
of the fan is suited tor air condi-
tioning applications in the 1000 to
50,000 cfm range. (Trane Co.)
Noise may be caused by factors other than the fan
itself—excessive air velocity in the duct work,
improper construction of ducts and air passages, and
unstable housings, walls, floors and foundations. The
importance of selecting a fan to suit the characteristics
of the duct system cannot be over-emphasized. Where
noise responsibility can be laid to the fan itself, the
cause may be improper selection of fan type or exces-
sive speed for the size and blade configuration. A fan
operating considerably above its maximum efficiency
is usually noisy.
Fan Laws
When a given fan is used for a specific system, the
following fan laws apply:
1. Air capacity (cfm) varies directly as the fan speed,'
2. Pressure (static, velocity, or total) varies as the
square of the :fan speed,
3. Horsepower required varies as the cube of either
the fan speed or capacity,
4. At constant speed and capacity, the pressure and
horsepower vary directly as the density of the air,
5. At constant pressure, the speed, capacity and
horsepower vary inversely as the square root of the
density, and
6. At constant weight delivered, the capacity, speed
and pressure vary inversely as the density, and the
horsepower varies inversely as the square of the
density.
For conditions of constant static pressure at the fan
outlet or for fans of different size but same blade tip
speed, TrDR — constant
7. Capacity and horsepower vary as the square of the
wheel diameter,- r
8. Speed varies Inversely as the wheel diameter,
9. With constant static pressure, the speed, capacity,
and power vary inversely as the square root of the air
density,
10. At constant capacity and speed, the horsepower
and static pressure vary directly as the air density, and
11. At constant weight delivered, capacity, speed and
pressure are inversely proportional to the density.
Horsepower is Inversely proportional to the square of
the density.
These laws can be expressed mathematically singly or
In combination, as follows:"^. -«•-
Q = A RD3 H = B R2 D2d P = C R3 D3d
where: Q .= capacity, cfm -^
D = wheel diameter, ft ±~
H = static pressure head, ft fluid flowing
P = horsepower, hp
R = speed, rpm
d = density or specific weight of air or
gas, (Ib/ft3)
A,B,C = constants
28
JULY 1974
-------
Fig. 23 Blower for economical low
pressure air for: industrial combus-
tion systems, conveying, cooling,
drying, liquid agitation, smoke
abatement, vacuum cleaning, fume
and dust exhausting, and other ap-
plications where the air being
handled does not exceed 220 F.
Sizes available in pressure ranges
from 4 to 48 oz and capacities 6900
to 560,000 cfh. (Eclipse Combus-
tion Div., Eclipse Inc.)
Fig. 24 This 119-in. fan helps keep
the skies clear over a major steel
company. The fan handles 90,000
dm at 76-in. static pressure. It
operates at 1180 rpm and draws
contaminated air from scarfing
operations through a scrubber and
mist eliminator before exhausting
air out the stack. (Robinson Indus-
tries, Inc.)
Fig. 26 High-speed single-stage
centrifugal blower. The direct
coupled motor drives through a
single-step speed increasing gear-
box to speeds ranging from 26,000
to 41,500 rpm. By utilizing an open
radial blade impeller emitting into
a full emission discharge, high effi-
ciencies, compactness and servicing
simplicity are realized. (Sundstrand
Corp.)
Fig. 25 Single-stage compressors for
water pollution control plant. (Elliott
Div., Carrier Corp.)
Fig. 27 For service app/icatioi
which require maximum protectii
against erosion, wheels of this d
sign are furnished with renewal)
blade liners- covering the enti
blade area and including side pla
liners where necessary. The sci
(oped center plate of this design i
lows the use of a continuous blai
liner. (Sturtevant Div., Westinghoui
Electric Corp.)
If, when considering two fans, A = Af, then B = B(,
and C = C | , the fans are said to be operating at the
same equivalent orifice, ratio of opening, point of .
operation, corresponding points or point of rating.
This means the two fans are proportional and the
above three equations are applicable and the fans
have identical efficiencies./,
Example 1: A fan in a manufacturer's brochure Is
rated to deliver 20,500 cfm at a static pressure of 2 in.
water (wg) when running at 356 rpm and requiring
5.4 hp. If the fan speed is changed to 400 rpm, what is
the resulting cfm, static pressure and hp required at
standard air conditions?
Solution 1: By fan laws 1, 2 and 3
Capacity
= 20,500 / ^} = 23,042 cfm
y 356 J
Static pressure =
/40|\2 =
of water
Horsepower
= 5.4
/400\3 _ 7
\356j -7-
7.67 hp
Note: Standard airln fan tabulations is usually taken
as air at 68 F, 29.92 in. Hg, 50 percent relative
humidity, weighing 0.07496 Ib/ft3. (0.075 is most often.
used for approximate calculations).
Example 2: If, in the previous example, In addition to
speed change, the air handled was at 150 F instead
of standard 68 F, what capacity, static pressure and
horsepower would be required?,
Solution 2: Air density at 68 F and 29.92 in. Hg is
0.075 Ib/ft3.
Density at 150 F = 0.075f-460 + * )(*^2}= 0.065
\460-f150y \ 29.92 /
Density at 150 F and same barometric pressure is
obtained by multiplying initial density by absolute
temperature and pressure ratios.
By fan law 4:
From Example 1, Capacity =23,042 cfm at 150 F .
Static pressure
Horsepower
Ib/ftJ
= 767—U
\0.075j
Fundamental Formulas
Pressure exerted by a gas that is not moving is called
static pressure. The pressure resulting from velocity
impingement is called velocity pressure. The sum of
static pressure and velocity pressure is the total
pressure. Fan pressures are determined from duct
pressure readings by means of two impact tubes facing •
the air current. The total pressure, of a fan is the
POLLUTION ENGINEERING
29
-------
SELECTED MANUFACTURERS OF FANS AND BLOWERS
Data baud on available Information tupplied to editors by manufacturer*.
S8|
• £ | SELECTED MANUFACTURERS OF
-------
Data b«»ed on available Information aupplled to editor* by manufac1ur»r». "•.'
"S C E SELECTED MANUFACTURERS OF
E&Z FANS AND BLOWERS
443 Ingersoll-Rand
444 Jeffrey Manufacturing Co., Div. of Jeffrey Gallon Inc.
445 Jenn-AIr Corp. "_ '_ "_
446 Jones & Hunt, Inc.
447 Joy Manufacturing Co., Air Power Drv. ~ ' '" ~
448 Justin Pacific Corp.
"449 "' Lamson Blower Div., Diebold, lnc.~._J_~\ ~Li'!~ ...
450 M-D Pneumatics, Inc.
451 Maxon Corp. ". ' ' "' ^
452 McLean Engineering Laboratories
453 McQuay-Perfex, Inc. " ~ '
454 Wm. W. Meyer & Sons, InV "
455 " The Moore Co.\ -^ "'. J^' ~^_ CI7-l:~Vc-~ir -
456 NalgeCo.
457 New York Blower CdT >=~r^.- -_-.-,- ^-~Tr-~ ---'--
458 North American Manufacturing Co.
459 Peerless Electric Products, H. K. Porter Co., Inc.
460 Penn Ventilator Co., Inc.
461 "" Pesco Products Div., Borg-WarneTCorpT^ 'V^'~~~
462 Precipitair Pollution Control Inc.
463 Recold York Div., Borg-Wamer Corp^"7""'^" 7"
464 Robinson Industries, Inc.
465 Rotron Inc. "" ". _J ~ '"^'~72.-;~~~^~l^2'^
466 St. Louis Blow Pipe Div.^ "
467 Sanders Associates, Inc. :_T TL^;dm.^I T '
468 Senior! Process Corp.
469 " Sheldons Manufacturing CofpT^I^Jl~"r7"~r~
470 Spencer Turbine Co.
471 Standard Electric Manufacturing Co.,Tnc!'-;:.'jrri •',.-< ""
472"" Steelcraft Corp." ~ " " '
473 Sterling Blower Co. ^/ -- '^2SS-^-"'^i.*'* '•'" i—
474 Strobic Air Corp.
475" Sundstrand Co rp.'.?';*>7 '•'•'"' ^~'.<^7-'~'^ -~^~M T~
476 Sutton Manufacturing Corp.
477 Trane Co. " '--^ .''.- '-- ^j^-^JyT ~^.-^-y- .^'-^"''~ '-./"'•
478 Tri-Mer Corp.
479 Universal Fan' Co ?" " — Tr~rr;--~7-"^ ^~^-~^~ = - -
480 Walker Manufacturing & Sales Corp.
481 Wes-Co Blower & Pipe Co. ^ ^..^J^ m^I.I
482 Westinghouse Electric Corp., Sturtevant Div.
483 ^J Young & Bertke Co^. ."." '^'~^'C^2^^~^~^ ..
484 Zurn Air Systems Div Zum Industries
1
i
«
i
....
' •" •
A
.._,.-_,_
- - ,
-,_•,
•
— , „ „ ..
.r-".::.
e
m
—
~.
-,.
.—
-_
--
^**~.~-
-
. i—
Centrifugal Fan*
'.'.:
... •
:±
.„•:;_
j~^~~ -'.
-.-.-- --
•
• _._.
-..•..
.,:•
e
I
c
.
...
-1"*7..
—^. —
- •
^ •
•~i-
IL
X
• .
•
"•
-
.
...,.•
^
POLLUTION ENGINEERING
31
-------
un* ptatne pvrmra WM inotMMe o( m*ch*n*rr mupi
«•• e( mukt hoed to •oearwnoo*)* hMovoof* M lev
moow lor tug* M«
hood err pump W*rona or*, &K p
Fig. 28 The midget fan in this system is a member ol a family of nine
ventilation fans ranging up to 120 in. and volume capacities to 700,000
cfm. (Jeffrey Manufacturing Co.)
Fig. 30 Blowers available for han-
dling air; gases other than air, water-
sealed vacuum or dry vacuum.
(Roots Blower Oiv., Dresser Indus-
tries) x
Fig. 29 Four-bearing outboard turbo-
compressor (blower) used for sew-
age aeration. (Spencer Turbine Co.)
Fig. 31 Each of the four Lamson
blowers at the new $6 million water
pollution control plant at En field.
Conn., is an important element in
wastewater treatment. These units
are powered by 200-hp electric mo-
tors and can generate 4500 cfm.
(Lamson Blower Oiv., Diebold Inc.)
increase in total pressure through the fan as indicated
by a differential reading between the fan inlet and
outlet.
Static pressure (ps) is the total pressure rise p less the
velocity pressure in the fan inlet.
Velocity pressure (pv) is the velocity pressure in the
fan outlet, expressed in inches of water.
Velocity can be expressed in terms of velocity pressure
as follows:
V = 1
Air horsepower or power-output of the fan.
62.3pQ
Air hp =
= 0.0001575 pQ
12(33,000)
where: Q = volume of air, cfm
p = pressure rise in inches of water
Efficiency of a fan is the ratio between output horse-
power (air hp) and the input horsepower (bhp).
efficiency = air hp/bhp
Static efficiency of a fan is the ratio of static pressure
power and the input horsepower.
Standard air density is 0.075 Ib/ft3. Fan pressures and
horsepowers vary directly as air density.
Fan Characteristics
Fan performance can be best presented graphically.
It is common practice to plot volumes against pres-
sures, horsepower inputs and efficiencies. The forms
of the pressure and horsepower curves depend on
blade type. Figure 1 shows a typical plot of fan
performance, volume-cfm vs total pressure, static
pressure, horsepower and efficiencies; drawn for a
given size fan at a given speed. Plots of more general
application are also used since fans function closely
to dimensional theory. Dimensionless plotting of fan
curves is accepted practice. A dimensionless plot,
Fig. 2, shows percent of wide open volume vs percent
pressure, horsepower and efficiencies. These typical
performance curves show how efficiency, pressure and
power input vary with changing flow volume. Plots are
based on fans operating at constant speed and
standard air density.
Air Pollution Control with Fans and Blowers
Fans and blowers can be used alone as air pollution
control devices, or in conjunction with control equip-
ment such as wet scrubbers, baghouses, electrostatic
precipitators and combustion units. In any case, the
fan is the heart of the system.
Ventilation fans are used in heat control, removing
heat from rooms or closed areas. Fan size depends on
32
JULY 1974
-------
_j
Fig. 32 Twenty basic sizes ol small,
cast-iron blowers designed to handle
low-volume needs: low-pressure
units—4 psig, 2046 ctm; medium
pressure units—7 psig, 1169 cfm;
high-pressure units—12 psig, 548
dm. (Fuller Co., GATX)
Fig. 33 Complete design, fabrica-
tion and installation services are
available from the manufacturer.
Shown is a doffing-roll bin and re-
lated blower system of eight blow-
ers and fans for large paper mill.
(American Sheet Metal, Inc.)
Fig. 34 MD Series 1200 blowers are
designed to deliver maximum air at
low rotor speeds. Units are avail-
able in various sizes up to 48-in.
rotors.- (M-D Pneumatics, Inc.)
Fig. 35 Model SPN sidewall exhaust
fans incorporate a fan blade design
which moves more air with less
power and less noise. Available in 24
to 60 in. diameter sizes, with capa-
cities ranging up to 76,000 ctm.
(Greenheck Fan & Ventilator Corp.)
Fig, 36 All airstream parts of FRP
Radial Fume Exhauster's are fabricated
of top grade corrosion resistant FRP
making them practically impervious to
attack by most chemicals. They feature
a nonclogging radial bladed wheel and
are offered with 8, 10, 14 and 18 in.
wheel diameters to develop static pres-
sures to JO in. wg and allow capacities
to 5000 cfm. (The New York Blower
Co.)
the size of the room in which the work is being done
and equipment such as furnaces, milling machines,
etc. being used in the area. For industrial heat relief,
fans are used for spot cooling and in exhaust systems
employing hoods.
Roof ventilators help provide effective control of in-
plant environment. These compact units remove heat
and contamination from work areas efficiently at
modest cost. The equipment can also incorporate split
or combined heating control and room air circulation.
Mechanical ventilators have other advantages. Unit
efficiency can be maintained regardless of weather
conditions and equipment can often be located in
otherwise wasted space.
In-plant odor control involves the use of fans and
blowers to force or induce contaminated air through
various control devices. Industrial toxicants and
odiferous materials include substances such as
ammonia, solvent vapors, hydrogen sulfide, carbon
monoxide, pollutants from coal and petroleum
processing, irritants such as sulfur oxides, toxic dusts
from metal refining and working, or from asbestos
handling.
Activated carbon filters are used in conjunction with
fans to control odors and contaminants consisting of
organic substances. Fans are used to draw the
contaminated air through a bed of activated carbon
that absorbs the odors. All of the air may be passed
through the carbon bed, which is called a continuous
bed, or some may be diverted around the bed, making
it a discontinuous bed. Continuous carbon beds are
made of porous tubes that are filled with charcoal or
flat strips with charcoal granules glued to them. Most
applications use continuous beds made of pleated or
flat cells of charcoal or hollow cylinder canisters filled
with charcoal. Activated charcoal absorbs most odors
in a single pass at air velocities between 50 and 120
fpm, with maximum recommended velocities of
continuous bed absorbers recommended at 250 fpm.
Continuous bed absorbers are 95 percent efficient,
using from 5 to 50 Ib charcoal/1000 cfm capacity,
depending on application.
Dry filters consisting of a bed or mat of fiberglass or
fine synthetic fibers are also widely employed. This
type filter actually increases in efficiency as a dust
layer builds up to act as an additional filter surface.
Low air velocities, between 300 and 500 fpm, also
increase efficiency. When filters become dirty they
can be washed and reused, or thrown away and
replaced. S5
For m copy of this article circle 650
on Reader Service Card
POLLUTION ENGINEERING
33
-------
ITEMS
Fans And Fan Systems
John Thompson
Chemical Engineering Journal
-------
Pans and
fan systems
In the chemical process industries, the cost of installing
and operating fans may be substantial, so it is important to
know how to select a fan and apply it sensibly. This report
tells how fans work, and covers fan selection and fan-system
design. It also introduces fiberglass-reinforced-plastic
fans, which are used in corrosive environments.
John E. Thompson and C. Jack Trickier, The New York Blower Co.
Q Fans and their attendant components can represent
a substantial part of total plant cost, and their cost can
escalate dramatically if the established fundamentals of
selection, application, operation and maintenance are
not followed. Likewise, high energy costs demand that
fan efficiency get sufficient attention.
The engineer should know the major types of fans
and their recommended uses, and how to select fans for
dudes ranging from supplying fresh air to handling cor-
rosive, explosive and abrasive streams. Beyond this, the
fan user or specifier should be aware of fan-system de-
sign principles—i.e., how to ensure that an installed fan
works as expected. The engineer concerned about corro-
sive air or gas streams should know how fiberglass-rein-
forced-plastic (FRP) fans are different from steel or alloy
ones.
This report covers the basics of fans, fan selection,
system effects and FRP fans. * Still, it is important that
the engineer discuss specific duty requirements with
potential vendors to make sure that the best selection is
made and that all appropriate performance factors and
limitations have been considered.
•No one bu ever drawn • meaningful dnonraon btratm f»m mnd blowen.
Here, »11 nich ur moven will be called fun.
Fundamentals of fans
A fan's performance characteristics are determined
primarily by the shape and setting of the wheel blades.
On this basis, fans in general use today can be classified
in five groups. These are, roughly in order of decreasing
efficiency: backward-inclined, axial, forward-curved,
radial-dp and radial-blade. The axial fan wheel propels
air or gas straight through. The other types of wheels
are centrifugal.
Though the general performance characteristics of
these fan types are consistent among different manufac-
turers, specific capabilities, recommendations and limi-
tations will vary.
Backward-inclined fans
Fig. 1 shows the two backward-inclined wheel designs
in common use: one with single-thickness blades and
one with airfoil-shaped blades. The airfoil design is the
most efficient, being able to reach a peak mechanical
efficiency near 90%. It is generally die quietest type of
fan wheel
The single-diickness blades can handle fine airborne
paraculates or moisture that would damage airfoil
48
CHEM1CM. ENGINEERING MARCH 21 196?
-------
blades, but are slightly noisier and less efficient. Peak
mechanical efficiency is 84% or more.
An attractive feature of the backward-inclined types
is the non-overloading character of their input-power
curves. As Fig. 2 shows, brake horsepower (BHP) in-
creases to a maximum as flow increases, and then drops
off. This means that a fan motor selected to accommo-
date the peak BHP will not overload, despite variations
in the system's resistance or flow, as long as the fan
speed remains constant. Such flexibility is an asset when
resistance or flow can vary due to changing airstream
composition, or when it cannot be accurately de-
fined—e.g., in a pilot-plant situation.
The static-pressure curve in Fig. 2 is typical for most
backward-inclined fan designs in that there is a range of
instability to the left of the peak pressure—usually
where the curve has a pronounced dip. In t-hit range of
high pressure and low flow, airflow across the wheel can
change or break away from the blades so that perform-
ance is no longer stable. A fan having such a static-
pressure curve should be selected to operate well to the
right of the unstable range.
Fig. 2 also shows the performance curves for certain
airfoil designs that are stable throughout the entire
pressure range—from wide open to completely shut off.
The dip is much less pronounced, and the fan is stable
in this area. Such a feature is important in ventilation
and air-supply applications where volume and resist-
ance to flow can vary widely. Note that there are only
subtle differences between the static-pressure curves of
fans having stable and unstable characteristics; the fan
vendor should therefore be consulted.
Backward-inclined fan wheels may be installed in the
usual scroll-shaped centrifugal housing, where exhaust
is at a right angle to the inlet, or as inline centrifugal
fans, where net flow is straight-through. Both designs
exhibit the same non-overloading BHP curve, and their
static-pressure curves are generally the same, but the
inline design is slightly less efficient than its conven-
tional centrifugal counterpart. The advantage of inline
design is space saving; such a fan can be installed di-
rectly in a duct.
Axial fans
Axial fans are like inline fans in that air or gas flows
straight through. The most common type of axial fan is
the propeller fan, which can be found in window-,
wall-, or roof-ventilating applications. The same type of
propeller incorporated into a tubular housing is gener-
ally known as a duct fan. More-sophisticated wheels
such as the one in Fig. 3 have airfoil rather th=»" propel-
ler blades.
Single-thickness bl«d«s
Airfoil bl*d«
Axial fans installed in tubular housing are typically
called tube-axial if the housing ha* no guide vanes,
vane-axial if it does. Tube-axial fans having airfoil
blades are found in low-pressure ventilating applica-
tions; vane-axial ones are used for dean-air handling at
pressure ranges to 8-10 in. water gage. Vane-axial de-
signs are generally more efficient; some offer peak effi-
ciencies above 85%. There are more-sophisticated
vane-axial designs that can operate at much higher
pressures, and some that ran tolerate airborne particu-
lates, but these are specialized for uses such as induced-
draft boiler exhaust in power plants.
Fig. 4 shows typical brake-horsepower and static-
CHEMICAL ENGINEERING MARCH 2J. 1983
49
-------
FANS AND fAN SYSTEMS
Typical wheel has an unstable range
Q.
to
0.
X
m
Flow
Some airfoil wheels are stable throughout
a.
X
CD
Flow
pressure curves. Axial fans have a significant stall re-
gion, so they should always be operated to the right of
the intermediate pressure-peak on the static-pressure
curve. Further, axial fans are unlike any others dis-
cussed here in that horsepower increases with decreas-
ing flow, and peaks at shutoff (no flow).
The most common axial fans contain the motor, or
the bearings and drive components, within the air-
stream (this is also the case for inline centrifugal fans).
Even when drive components are protected by a tube
assembly, as in Fig. 3, airborne particulates and explo-
sive or corrosive fumes could come in contact with these
moving parts. If the air is hot, the drive components
may be heated beyond their recommended tempera-
tures. Therefore, most axial fans are limited to clean-air
applications at relatively low temperatures. There are,
however, special designs for air that is contaminated or
at high temperature.
Axial fans are slightly noisier than inline centrifugal
fans, but their noise is mostly in higher-frequency bands
and dius easier to attenuate. In other words, high-fre-
quency sound waves peak in a shorter distance than do
low-frequency ones, so sound-deadening devices can be
smaller and less expensive.
Forward-curved fans
The forward-curved design, also called the squirrel
cage, is used to handle low to medium volumes at low
pressure. The many cup-shaped blades tend to retain
airborne contaminants, so this design is limited to the
cleanest airstreams,
Fig. 5 shows typical performance curves for a for-
ward-curved fan. There is an area of instability to the
left of the pressure peak, so the fan must be operated to
the right of that point. Brake horsepower increases with
increasing Sow throughout the entire range, in contrast
to the other fans discussed so far.
The forward-curved wheel turns slower than any of
50
CHEMICAL ENGINEERING MARCH 21, 1983
-------
Flow
the other wheel types for the same level of performance,
which makes it preferable for high-temperature appli-
cations. This is especially true when high temperature
;mposes limits on speed because of reduced material
strength—e.g., in a heater box. Lower speed is also a
plus in applications that require long spans of fan shaft
between bearings—e.g., air redrculation in a dryer.
Although airborne noise is directly related to me-
chanical efficiency, the forward-curved fan is typically
quieter than other types having similar efficiency. This
is because its lower speed produces Jess vibrational
noise—i.e., noise caused by vibrations transmitted
through the structure.
Radial-tip fans
The radial-tip design fills the gap between the clean-
air fans discussed so far and the more rugged radial-
blade fans used for materials handling. The radial-tip
fan wheel shown in Fig. 6 has a relatively low angle of
attack on the air, which allows air to follow the blades
with minimal turbulence. At the blade tips, the air is
accelerated for pressure generation as the blades change
toward a straight radial shape. Hence the designation
radial-tip.
This type of fan wheel is ideal for contaminated
airstreams that the backward-inclined, axial and for-
ward-curved types cannot handle. However, it is not
intended for the bulk-materials-handling and air-con-
veying applications that the radial-blade fan wheel is
used for.
The radial-tip design combines the static-pressure
•diaracteristics of the backward-inclined fan with the
BHP characteristics of the radial-blade fan. This is
shown in Fig. 7. Peak mechanical efficiencies can be
75 % and above. Many housings for such fans are avail-
Flow
able, but the most common ones are similar to those
used in backward-inclined fans. These handle medium
to high volumes of air or gas in a unit physically smaller
than a typical radial-blade fan.
Radial-blade fan*
The radial-blade fan is the workhorse of industry—
the type most commonly used for handling low-tc-
medium volumes at high pressures and for handling
airstreams containing high levels of particulates. Appli-
cations range from moving dean air to conveying dust,
woodchips and even metal scrap.
The radial-blade design is well suited for materials
Rotation
CHEMICAL ENGINEERING MARCH 21. 1983
51
-------
FANS AND FAN SYSTEMS
Flow
handling because the flat blades limit material buildup,
and the design can be adapted to abrasion-resistant-
alloy construction. Also, radial-blade fan wheels turn at
a lower speed than all but forward-curved wheels, so
abrasive panicles move along the surfaces at relatively
low velocities.
Generally, radial-blade fans are stable from wide
open to dosed off, as shown by the static-pressure curve
of Fig. 8. This is important in handling contaminated
5 10 15
Flow, 1,000 ftVmin
airstreams whose density may vary, since the fan may
have to accommodate a broad range of airflows. Here
again, increasing flow will result in increased brake
horsepower.
Efficiency is not usually the key criterion in selecting
a radial-blade fan; the most common designs sacrifice
some efficiency in favor of materials-handling ability.
However, some radial-blade fans designed for dust han-
dling can achieve mechanical efficiencies up to 75%.
Selecting a fan
Engineers often admit that fan equipment in a chem-
ical-process plant is sometimes taken for granted: Fans
tend to cause fewer problems than do other machines
and system components. True, fans are relatively simple
machines, but reliability depends on proper selection
and application.
Fan selection depends first on the flow-and-pressure
performance required for the application. Other con-
cerns, which may eliminate certain fans or fan types,
include: particles and chemicals in the airstream; size
and space constraints; airstream temperature; and
noise. Finally, capital and operating-cost considerations
identify one of the fans as being most economical
Fan-system performance
Performance is described as volume of airflow
(frVmin) and static pressure (in. water gage) required
to overcome resistance to flow. Finding a fan that meets
or exceeds the required performance seems a straight-
forward tMlc but there are several pitfalls that should
be considered.
First, how accurate and dependable is the system-
resistance calculation? A fan having a relatively steep
static-pressure curve would deliver the specified volume
of air despite minor errors or changes, while a fan hav-
ing a flat curve would see a large change in airflow.
Also, a backward-inclined fan would not overload de-
spite changes in system resistance, so the motor for such
a fan could be sized with greater confidence.
Another factor is that fans are not all rated at the
same conditions. Propeller-type fans and roof ventila-
tors are typically rated by themselves, with no duct-
work, while most other fans depend on inlet or outlet
ducts (or both) to perform as rated. Fortunately, details
of rating can usually be found next to rating tables in
catalogs or product brochures, and fans tend to be rated
in configurations similar to those most commonly
found.
The dean-air fans used to supply air to buildings or
process systems often have no inlet ductwork. Back-
ward-inclined, forward-curved and inline centrifugal
fans in such applications normally have a smooth ven-
turi-shaped inlet cone that serves to minimize losses.
(Radial-tip fans are usually fitted with such cones.)
Fig. 9 shows such a setup.
Fans having inlet cones may or may not have inlet
52
CHEMICAL. ENGINEERING MARCH 21. 1983
-------
ducts when they are rated, but it is standard
practice that they have outlet ducts. * Fans not having
inlet cones should have either inlet ducts or external-
•enturi inlets.
Axial fans are usually installed within a duct, and
they are typically rated for operation with inlet and
outlet ductwork. However, some manufacturers rate
their fans with diverging outlet transitions that convert
velocity pressure (kinetic energy) to static pressure. This
can be confusing, especially when one manufacturer's
fan is compared with another's. Large, high-horsepower
centrifugal fans may also be rated with different outlet
conditions. A transition known as an cease increases the
outlet area, thereby gaining static pressure.
Conversion of kinetic energy to static pressure is rou-
tinely taken into account in fan-system design. When
airflow enters an enlargement in the -duct, static pres-
sure will increase, because velocity, and thus kinetic en-
ergy, is reduced. Total pressure remains constant, except
for a slight efficiency loss due to the abruptness of the
duct enlargement.
Fig. 10 shows an example. There is a 1-in.-water-gage
velocity-pressure difference across die enlargement in
the duct—between points B and C. An enlargement of
this shape (1.4:1 area ratio, 7-deg angle) could be as
much as 94% efficient in converting this velocity pres-
sure to static pressure. In other words, the overall A-
B-C-D resistance is only about 14 in. water gage, rather
than the 15 in. water gage it would be without the
static-pressure regain.
This same principle applies in the case of evase or re-
•Rating procedures an developed jointly by the Air Movement and Control
Ann. (AMCA), an industry group, and the American See. of Heating, Refriger-
ation and Air^Conditioning Engineer! (ASHRAE), a profeaional tociecy.
AMCA publication 201 ("Fani and Systems") deuUs the outlet-duct length!
and condition! used to measure fan performance consistently, and includes
correction factors for absent or different connections.
Velocity pressure - 2 in
Velocity prenure • 1 in.
Distance
gain-cone oudets on fans. The efficiency is not as great,
however, due to turbulence.
Since connections are implied in fan ratings, the engi-
neer should make sure that die raring applies to die
application. If not, it can be corrected for the actual
connections to be used. It is also important to allow
space for any connections required, and to realize diat
pressure-regain connections may reduce velocity below
die minimum needed to keep airborne panicles aloft.
Fan class
Fan class is another aspect of describing performance.
The Air Movement and Control Assn. (AMCA) pre-
scribes minimum static-pressure and air-velocity per-
formance for Class I, n and HI fans.11 There are separate
standards for backward-inclined single- and double-
widdi fans, forward-curved single- and double-widtii
fans, and backward-inclined inline fans. As an
example, the class standard for backward-inclined sin-
gle-widdi fans is:
Class I: 5-in.-water-gage static pressure at
2,300 ft/min to 23-in. static pressure at 3,200 ft/min.
Class El: 8.5-in. static pressure at 3,000 ft/min to
4.25-in. static pressure at 4,175 ft/min.
Class ETI: 13.5-in. static pressure at 3,780 ft/min to
6.75-in. static pressure at 5,260 ft/min.
Class IV: Above Class HI minima
There is a common belief tiiat fan class also dictates
construction requirements such as gage of metal to be
used. This is not so. A fan sold for Class I duty may
meet Class II requirements (this is convenient for some
manufacturers), but it is not necessarily superior in con-
struction. To assure quality construction, a minimum
metal gage, not fan class, should be specified.
'AMCA Standard 24O8-69.
CHEMICAL ENGINEERING MARCH 21, 1983
53
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TANS AND FAN SYSTEMS
3
ff
10
Airstream composition
After performance, the composition of the airstream
is the most important concern in fan selection. Mois-
ture, corrosive chemicals, flammable or explosive fumes
or gases, and airborne particulates each impose limits
on the choice of fans. In many casm, the airstream com-
position requires materials of construction that are in-
compatible with certain fan designs. And particle-laden
airstreams narrow the choices to only the most rugged
radial or radial-tip fans.
Paniculate loading can be denned by the maximum
gr/scf (grains per standard cubic foot of air) paniculate
content and the maximum (not average) particle size.
Most dean-air fans can handle loadings to 0.02 gr/scf
and particle size to 0.05 micron without buildup. Be-
yond these levels, there is a chance that moisture or
particles will build up on backward-inclined, forward-
curved or axial fan blades, causing imbalance, erosion
and even performance deficiency.
Corrosion can be dealt with in many ways. Most fans
can be protected with a variety of paints or protective
coatings, and most centrifugal fans are available in alu-
minum or stainless-steel construction. In recent years,
fiberglass-reinforced-plastic (FRP) fans have been devel-
oped as reasonably economical alternatives for corro-
sive service. These will be discussed in detail later.
Flammable or explosive fumes require careful consid-
eration of all system components. There are standards
for explosionproof electric motors, but no such stand-
ards for fans. Manufacturers do offer various forms of
spark-resistant construction, where some parts are built
of nonferrous alloys to minimize the potential for
spark generation from two fan components rubbing
together or striking each other. However, such construc-
tion does not eliminate the potential for spark genera-
tion by foreign influences such as airborne particles, nor
does it provide any guarantee of safety.
Abrasion is a major problem and significant cost in
materials handling. But there are no reliable methods
for predicting the abrasive characteristics of a particu-
lar material or airstream, and thus no way to accurately
predict the life of a fan exposed to abrasives. There are
construction modifications and special devices that can
extend the life of a fan in abrasive service, but fan type
and features are best determined case by case.
For any contaminated airstream, certain basic fea-
tures should be provided. A shaft seal or closure will
contain the contaminants and thus protect the external
bearings and surroundings. Flanged inlet and outlet
connections aid in sealing or gasketing to prevent leak-
age, though FRP fans often have slip connections that
can be physically bonded to FRP ductwork for positive
sealing. A drain at the lowest point in the fan housing
prevents moisture buildup and allows periodic wash-
down to remove corrosives or contaminants that
might adhere. Fans often have an access or deanout
door; this allows inspection of the fan interior.
Size and space constraints
Limits on physical space available for an installation
may impose limits on fan selection. There are ways to
accommodate such space constraints, often by sacrific-
ing some other feature. Whenever possible, and espe-
cially in outdoor installations, such constraints should
be removed. That allows more latitude in meeting other
specifications that may be more important.
Space saving is one of the key reasons for choosing an
axial or inline centrifugal fan. When installed in the
ductwork or in ceiling or rooftop areas, such fans elimi-
nate the need for separate equipment rooms and save
-------
valuable floorspace. And of course these fans may also
be the most economical choices for low-pressure, me-
dium-to-nigh-volume applications.
When the application demands a centrifugal fan, the
choice of drive and bearing arrangement affects space
requirements. Fig. 11 shows the most common arrange-
ments denned by AMCA.
Belt-driven fans are generally available in arrange.
ments 1, 3, 9 and 10. In 1, both bearings are on a pedes-
tal, and the motor may be mounted on the floor or on a
unitary base. Arrangement 3 takes less floorspace than 1
because it has one bearing on each side of the fan, but it
is limited because one bearing is in front of the fan inlet.
Arrangement 9 is like 1 except that the motor is
sidemounted to conserve fioorspace, and 10 saves space
by having the motor within the bearing pedestal; but
both 9 and 10 have Emits on motor size.
Direct-driven fans are generally available in arrange-
ments 4, 7 and 8. The fan wheel is mounted directly on
the motor shaft in 4, so application is limited by the
motor's temperature limits. Arrangement 7 is like 3, but
with a motor pedestal. Arrangement 8 is like 1 with a
motor pedestal; it is well suited for elevated tempera-
tures or contaminated airstreams since die motor is far
from the fan.
Arrangements 3 and 7 are available in double-width,
double-inlet (DWDI) designs as well as in the common
single-width, single-inlet (SWSI) ones. The SWSI ar-
rangement 3 and 7 is not recommended for wheel sizes
less than 30 in., since the bearings obstruct the inlet;
DWDI types are used for all sizes. Generally, a DWDI fan
is about 75% as tall as an SWSI one, but it takes more
floorspace. Fig. 12 illustrates this difference.
Temperature
Minimum and maximum temperature limits depend
on the type of fan and on the drive arrangement. Fans
that have motor, drive and bearings within the air-
stream impose limits on airstream temperature.
Fan arrangements 1, 8, 9 and 10 do not have these
components in the airstream, but such fans may require
a shaft cooler or "heat slinger" device located between
the fan housing and the inboard bearing to block the
flow of hot air from the shaft opening to the bearing.
Fig. 13 shows such a cooler, which is basically a set df
fan blades, fitted with a safety guard.
Airstream temperature also affects the safe operating
speed of a fan; this depends on the materials of con-
struction. Generally, steels lose strength as temperature
increases, and become brittle as temperature falls well
below 0°F, so speed must be adjusted downward in ei-
ther case. Most fan applications fall in a range from
— 25 to 1,000'F or more, and any case other than 70°F
may require correction of standard operating-speed
limits.
Noise
In general, the most efficient fans produce the least
airborne noise, but vibrational noise due to the struc-
tural surroundings, and mechanical noise due to the
drive and motor, -may be more important in some situa-
tions. Also, an improperly sized fan may not be operat-
ing in its peak-efficiency range. While it is generally
Single-wheel, lingte-inlet
Double-wheel, double-inlet
true that an airfoil fan at peak efficiency will be quieter
than a radial fan in the same service, a radial fan at its
peak efficiency may be quieter than an airfoil fan oper-
ating outside its peak-efficiency range.
Thus, concerns about noise must be considered case
by case, as part of the overall fan-selection problem,
rather than in a general manner. In comparing relative
noise levels, it is also important to use the uniform
measure of fan-rated sound power (watts or dB) rather
than a nonuniform measure such as sound pressure
level at some reference point.
Efficiency and economics
As in any equipment selection, the decision hinges on
economics once the choice of fan types and manufactur-
ers has been narrowed. Of course, the analysis must in-
clude the operating, maintenance and service costs as
well as the initial cost.
Due to todays expensive energy, more-efficient fan
CHEMICAL ENGINEERING MARCH 21. 1983
55
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FANS AND FAN SYSTEMS
types may be the better choice despite a higher price.
For example: Two types of fan are available to handle
3,000 frVmin at 12-in.-water-gage static pressure. The
first fan needs 9.2 brake horsepower to do the job; the
second needs 8.2, but costs S80 more. If the user values
energy at a conservative $250 per horsepower-year, the
second fan will pay for itself in five months. Motor effi-
ciency and power factor may alter this payback figure,
but the potential saving is there nonetheless.
The best way to compare energy costs for competitive
fans is to look at the brake-horsepower ratings for the
required performance. AD such ratings should of course
have the same basis: volume, pressure, density and dis-
charge velocity. A way to specify power-consumption
criteria is to stipulate a mini™"™ mechanical efficiency
(ME) or static efficiency (SE). These are calculated as
follows:
SE =
(Flow) (TP)
(BHP) (6,356)
(Flow) (SP)
(BHP) (6,356)
X 100%
X 100%
where TP is total pressure (static and velocity pressure),
in. water gage; SP is static pressure, in. water gage; BHP
is brake horsepower; and flow is ftVmin.
Fan-system effects
The fan installation shown in Fig. 14 is typical, at
least in appearance, of many exhaust-fan setups. What
is unusual is that the fan and its inlet and outlet ducts
have been designed and installed so that the fan system
performs exactly as expected. As already mentioned,
fans may be rated independent of any system. All too
often, a fan does not perform as expected because the
effects of the system were not considered.
In system design, calculated volume and pressure
requirements are used to select and size the fan. But
there is rarely a chance to build and test a pilot system
to assure that the calculations prove out before the ac-
tual equipment is installed. If system effects are not
fully considered, there may be unexpected pressure or
velocity losses that would require fan-speed and motor-
horsepower increases to compensate.
For example: The resistance of a given elbow to a
given flow can be determined accurately, unless the
elbow is located too dose to the fan inlet or outlet.
Then, there will be an added resistance that cannot be
measured or even detected by field instruments. In ef-
fect, the proximity of the elbow reduces the fan's per-
formance, and the problem caused by the elbow loca-
tion may wrongly be attributed to the fan.
AMCA test codes define the inlet and outlet duct
connections required for performance testing.* (AMCA
also certifies fans if they develop their rated flow and
pressure within a 2.5% speed tolerance and a 5% horse-
power tolerance.) If the installed system includes the
same connections, and system flow and resistance have
been calculated accurately, the fan should perform as
expected.
The fan's actual operating point is the intersection of
its static-pressure curve and the system's flow-vs.-resist-
ance curve. Fig. 15 shows this relationship; note that
system resistance varies as the square of Sow. If the re-
sistance is different than expected, the operating point
will be elsewhere on the fan's static-pressure curve. And
the static-pressure and horsepower curves themselves
will be altered if system effects prevent the 'fan from
achieving its rated performance.
The four most common causes of system-induced per-
formance defects are: eccentric flow into the fan; spin-
ning flow into the fan; improper outlet ductwork; ob-
structions at the inlet or outlet. We will now cover these
in greater detail
•AMCA Bulletin 210, "Labormmiy Method! of Taring Fmm far Rating
56
CHEMICAL ENGINEERING MARCH 21, 1X3
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Eccentric flow
A fan can perform correctly only if air flows straight
into the inlet with a uniform velocity profile. As shown
in Fig. 16b, this distributes the air load evenly over the
fan wheel. In Fig. 16a, there is an elbow at the inlet.
This produces both turbulence and maldistribution of
air over the fan wheel, causing a drop in performance.
The severity of the effect depends on the shape of the
elbow. A mitered elbow is worse than a smooth one; a
larger radius is better than a smaller one. Even more
important is the length of straight-run duct between the
elbow and the inlet—this is usually expressed in terms
5 10 15
Flow, 1,000 ft'/min
20
of duct or fan-inlet diameters. The greater the straight-
run length, the more chance the air has to straighten
out and fill the duct, and the lower the static-pressure
loss. The loss becomes negligible if the straight-run
length is more than 5-7 duct diameters; this varies with
air velocity.
Table I shows system-induced pressure losses for
round and square elbows having a given ratio (R/D) of
turn-radius to diameter or width. These losses must of
course be added to the calculated system resistance to
determine the correct pressure for fan selection.
For example: A system's resistance is 3 in. water gage
-Fan
Inflow
-Fan
Inflow
a. Elbow at inlet cau«« eccentric flow.
b. Straight inlet dritnbutra flow evonly.
CHEMICAL ENGINEERING MARCH 21, 1983
57
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FANS AND FAN SYSTEMS
Inlet box-x
Fsn
Inflow
'Inlet box
Egg-crate
flow divider
Verrturi'
at 4,000-ft/min velocity. If a multipiece mitered elbow
having a two-diameter turn radius is to be located at
the fan inlet, it will cause a system-effect loss of 1 in.
water gage. Therefore, the fan should be selected and
sized for 4-in.-water-gage static pressure. If the same
elbow were to be located five duct diameters away, the
system-effect loss would be only 0.3 in. water gage.
Nonuniform inlet flow can also be- caused by a poorly
designed inlet box, like the one shown in Fig. 17. Air is
not weightless, and forcing it past the fan inlet as in
Fig. 17 leads to turbulence.
There are many possible inlet-box configurations:
The box may be shallow and wide, to conform to a
narrow space; it may be a square elbow diat turns the
air 90 deg at the fan inlet; and it may be equipped with
vanes diat straighten Sow. Tabulating such boxes' sys-
tem-effect losses here would be impractical, but manu-
facturers can usually predict losses for their standard
inlet-box designs.
Spinning flow
If the entering air is spinning in die same direction as
die fan wheel is rotating, the fan produces less "lift"
than it would if die air were not spinning. This is analo-
gous to launching an airplane with the wind rather
than against it—again, less lift and poorer performance.
If die air is spinning counter to the fan-wheel rota-
tion, horsepower and noise increase. There is some in-
crease in static-pressure performance, but far less than
die increased power consumption would suggest
Prespinning flow is more difficult to evaluate than
eccentric flow because of die variety of potential causes.
Prespin may occur in conjunction with eccentric
flow—this would happen in the inlet, box shown in
Fig. 18. Or it may be caused by an air-cleaning device
that spins air to remove airborne contaminants. The
cyclone in Fig. 19 is an example; here die system in-
cludes an "egg-crate" flow straightener diat removes
most of die spin.
In general, die most efficient fans, such as backward-
inclined ones, are die most sensitive to prespin, but pre-
spin can cause a significant performance reduction in
any type of fan. The only way to get predictable per-
formance when prespin may exist is to test die installed
fan system, or a pilot model, and determine die neces-
sary fan-speed and horsepower corrections.
Correcting for poor inlet conditions
The ideal fan inlet creates neidier spinning nor eccen-
tric flow. If mere is no inlet duct, die system should
have a smooth venturi-shaped inlet or add-on venturi
to negate entrance losses. When an inlet duct is re-
quired, die best kind is a long run straight into die fan.
When space constraints m?^ such a duct impracti-
cal, tiiere are two further options: install corrective de-
vices, such as egg-crate flow dividers (Fig. 19) or turn-
ing-vanes (Fig. 20), in die duct; or increase fan speed
and power to compensate for expected losses. The latter
is usually easier to accomplish, and may be required in
addition to corrective devices in extreme cases where
die devices add significant resistance.
If fan speed is increased, static pressure will increase
by die square, and brake horsepower by die cube, of die
58
CHEMICAL ENGINEERING MARCH 21. 1983
-------
increase. Such a waste of power indicates that system-
related deficiencies should be avoided in the first place
whenever possible.
When there is a problem, and performance must be
corrected in the field, it may be possible to change the
speed without getting a new fan and motor. For exam-
ple: Suppose that the inlet box in Fig. 20 produces an
unexpected 10% system loss. If the fan is belt-driven,
there may be enough reserve to accommodate the re-
quired 10% speed increase and 33% power increase. If
the fan is connected directly to a fixed-speed motor, on
the other hand, the solutions are more limited and
nearly always more expensive.
Unexpected system effects can also move a fan's per-
formance into an unstable zone. When fan and system
are properly matched, the operating point should fall
within the fan's stable range. (This was illustrated in
Fig. 2).-However, a system loss can move the operating
point to the left into the unstable range. If this occurs,
the system must be altered to produce greater flow
through the fan without increasing resistance—e.g., by
installing larger ducts—so that the operating point
moves back into the stable range. The alternative is to
replace the fan with one that is inherently stable or is
smaller.
It is important to remember that a system-effect loss
cannot be observed in system tests; the loss occurs
within the fan. But it must be considered all the same in
selection and sizing.
Discharge ductwork
Air discharged from a fan has a nonuniform velocity
profile, like that shown in Fig. 21, rather than a uniform
one. This is because the centrifugal acceleration in the
fan forces air to the outside of the scroll. Since velocity
pressure (kinetic energy) is proportional to the square of
the velocity, it is greater at the fan outlet than down-
stream—where velocity has evened out. Since total
pressure is about constant, static pressure is not fully
developed until some point downstream.
A duct length of 2.5-6 duct diameters is usually re-
quired at the outlet for the fan to develop its full rated
pressure. If there is no outlet duct, a static-pressure loss
equal to half the outlet velocity pressure will result. This
must be considered part of the system resistance, in
specifying fan performance.
The outlet velocity determines the length of duct
needed to render the static-pressure loss negligible. For
velocities of 2,500 ft/min and less, 2 J> duel diameters
are sufficient. Beyond 2,500 ft/min, each additional
1,000 ft/min requires an additional duct diameter.
Elbows should be avoided at the outlet as at the inlet.
If an elbow or other turn is necessary because of space
constraints, the turn should be in the same direction as
the wheel rotation. A turn in the counter direction, as
shown in Fig. 22, creates a static-pressure loss. The se-
verity of this loss depends on the distance between out-
let and turn.
Inlet and outlet obstructions
Obstructions that add to system losses may be as ob-
vious as a cone-shaped stack cap, which can produce a
loss equal to the velocity pressure. Or they may be
Turning vanes
Fan.
"Inlet box
housing
Square elbow
CHEMICAL ENGINEERING MARCH 21, 1983
59
-------
FANS AND FAN SYSTEMS
Outlet duct
Plenum
Inflow
subtle—e.g., a belt drive mounted directly in front of
the fan inlet (as in the double-width, double-inlet fan
shown in Fig. 12).
When a fan is located in a plenum, or there is an
obstruction nearby, the effects on inflow have to be con-
sidered. Fig. 23 shows how the plenum can cause non-
uniform flow, which creates the system-effect loss.
Table II lists typical system-effect losses caused by
inlet obstructions. Losses increase with velocity and de-
crease with distance between the fan and the obstruc-
tion. Like the other system-effect losses, these are added
to system resistance when one is specifying or sizing the
fan.
Fiberglass-reinforced-plastic fans
Fiberglass-reinforced plastic (FRP)S made from
chemical-grade polyester or vinyl-ester resin resists cor-
rosion as well as, or better than, higher-priced materials
such as titanium or high-nidcel alloys. In general, FRF is
widely used in handling the fumes of acids and of many
inorganic and organic chemicals, but not organic sol-
vents. Applications are limited to about 250 "F and
below.
When FRP is the selected material for an air-handling
system, it is logical that the fan also be made of FRP.
For example: The acids used in stainless-steel pickling
are necessarily those that attack stainless steel In such a
pickling system, the acid-holding tanks, fume-control
hoods, ducts, scrubbers and fans are often made of FRP
because it resists acid corrosion and costs less than metal
alloys having comparable resistance.
Potential applications for FRP fans include any proc-
ess in which corrosive fumes must be captured, moved,
cleaned or vented. Currently, FRP fans are most often
used in fume-scrubber systems: The scrubber itself may
be FRP or an exotic alloy; FRP seems to be the preferred
fan material Galvanizing, etching and pickling proc-
esses often have FRP exhaust hoods and ducts, as al-
ready mentioned, and more and more of the fans used
to convey fumes in such systems are being built of FRP.
Wastewater-treatment plants and laboratory exhaust
systems are other potential applications.
In general, FRP fans may be an economical alterna-
tive to stainless-steel or other metal-alloy ones when cor-
rosion is a concern and temperature is below 250°F. An
FRP fan may even provide better performance than spe-
*FU> a aljo referred to u gUn-reinforced plastic (C»P), reinforced plinir,
and reinforced thennotetnng rain. We will all it Ttf throughout.
cial alloys in handling airstreams that are particularly
corrosive to metals.
The makeup of FRP
The ts/m FRP describes a broad spectrum of fiber-
reinforced plastic materials—for example, cabinets for
office machines might be. made of non-corrosion-resist-
ant plastics reinforced with mica and loosely called
FRP. However, the FRP used in making process vessels
and equipment (such as fans) is composed of: about
30% by weight of glass fibers, or sometimes other fibers,
that have been given a coating (sizing) to enhance their
bonding with the resin; and about 70% by weight of
corrosion-resistant polyester or vinyl-ester resin.
The fibers provide physical strength, and the resin
provides the corrosion resistance and rigidity that
make FRP a workable solid. Sometimes, non-glass fiber
materials are used in FRP to impart special properties:
e.g., graphite fibers add tensile strength, and aramid
fibers (e.g., Kevlar) add toughness. But FRP for process
equipment usually has glass fibers because they are
more economical and easier to work with; graphite fi-
bers, for example, are more difficult to handle and do
not bond as well as glass.
The corrosion resistance of FRP depends on the resin.
Resins used in FRP for process equipment are formu-
lated for maximum corrosion resistance, and are conse-
quently two or three times as costly as those used in
everyday products—e.g., polyester boat hulls.
An FRP fan typically uses different resins for the fan
wheel and housing. Vinyl esters are more ductile, and
form stronger joints. Thus fan wheels, which must with-
stand dynamic stresses, are typically made of vinyl-
MAACH ::. 1983
-------
ester-based FRP. Fan housings, on the other hand, are
typically made of polyester-based FRP.
How FRP fans are built
FRP fabrication is similar to metal-casting. A pattern,
called a plug, is used to make a mold for the FRP part.
In a fan, the airstream surfaces of the housing should be
smooth, to minimize resistance and prevent buildup of
airborne contaminants. Thus, male molds (rather than
female ones as in casting) are required: The smooth
outside surface of the mold shapes the inside surface of
the housing.
Parts made with male molds must be removable, so
FRP-fan housings are usually made in two halves, with
matching flanges. In larger fans, these two halves are
bonded together permanently, as shown in Fig. 24. This
is by means of FRP filler between the flanges; a lamina-
tion laid over the joint on the inside of the housing
provides a smooth surface. The joined flanges form a
ridge that adds strength to the housing. The inlet subas-
scmbly is typically bolted into place to allow access.
Smaller FRP-fan housings are also molded in halves,
but they are typically bolted together as shown in
Fig. 25. Removing the inlet side of the housing allows
installation or removal of the fan wheel
Fan-wheel construction is also different for large and
small FRP fans. Small wheels are typically made by
casting or press-forming in fully enclosed molds; Fig. 26
shows an example. Larger wheels, such as the one in
Fig. 27, are made by assembling and bonding molded
parts: wheel blades, frontplates, backplates, and hubs.
Solid-FRP balancing rings are often built into the
outside diameters of the frontplates and backplates.
These allow the fabricator to balance the fan wheel—
statically and dynamically—by grinding the rings.
Glass fiber itself has limited chemical resistance. The
resin provides the corrosion resistance in FRP, and a
pure-resin surface provides the greatest resistance. Un-
fortunately, pure resin is weak and brittle; if applied in
too thick a layer, it may crack.
In an FRP fan, surfaces that require the greatest cor-
rosion resistance are coated with a thin layer of pure
CHEMICAL ENGINEERING MARCH 21, 1983
61
-------
FANS AND FAN SYSTEMS
resin, which may incorporate a thin fiber mat (called a
veil) for reinforcement. The veil may be of glass fiber,
but for fumes that attack glass agressively a polyester
veil is preferred.
FRP-fan standards
If the fan in a fume-control system fails, the entire
process may come to a halt. The importance of reliabil-
ity has led to development of a standard for FRP fans—
American Soc. for Testing and Materials (ASTM)
D4167.-
This standard defines minimum specifications for
construction of major fan elements. Here are some of
the details:
Fan-housing construction must conform to the
ASTM C 582 specification, which applies to all FRP
process equipment The same resin must be used
throughout the housing unless the manufacturer and
user agree to use different resins in different layers of the
laminate. The structural rigidity of the housing (or a
prototype) is tested by running the fan with the inlet
dosed and die outlet open. Inward flexing may be no
greater than 03% of the fan-wheel diameter.
Fasteners, hubs and shafts must be either corrosion-
resistant or encapsulated in a material that is. This
means that bolts may be embedded in the housing and
completely covered, and that shafts may be protected
by an FRP or alloy sleeve that extends out through the
housing.
Safe wheel-operating speed is determined either by past
experience or by destructive testing—Le., running the
fan wheel at increasing speeds until it fails, and reduc-
ing the failure speed by a safety factor. Safe operating
speed depends on the square root of the flexural modu-
•"Standard Specification for Fiber-Ron/orecd Plastic Fan and Blown."
lus of die wheel material, which in turn depends on
temperature.
For example: If an FRP fan wheel has a safe operat-
ing speed of 1,000 rpm at 70 °F, and its flexural modulus
at 200 "F is only 88% as great, then its safe operating
speed at 200 *F will be 94% of die 70°F speed. That is,
940 rpm. The flexural modulus of FRP falls off rapidly
beyond 250"F, so FRP fans are seldom used above that
temperature.-
Spark resistance. FRP is spark-resistant in the sense that
contact of FRP pans does not generally produce sparks.
However, FRP fans handling dry air can develop elec-
trostatic charges on wheel and housing surfaces because
FRP is a poor conductor. Still, an FRP fan can be made
spark-resistant by incorporating graphite fibers in the
wheel and housing surfaces to make diem conductive,
and grounding the surface layers of die housing as
shown in Fig. 28. The standard defines acceptable resis-
tivity as no greater than 100 megohms between all
points on the airstream surfaces and ground.
Dynamic balance is achieved either by balancing die
wheel/shaft assembly as a separate unit or by balancing
the wheel once it is installed in die fan. Some manufac-
turers do both. Imbalance is corrected eitiier by grind-
ing down balance rings built into die wheel for tiiat
purpose or by adding metal weights (corrosion-resistant
or encapsulated) where needed.
Fan specifications
The selection and specification criteria already dis-
cussed apply to FRP fans and fan systems as well as to
metal ones. But FRP fans do have some additional
complexities.
One is resin selection. Whenever possible, select from
die manufacturer's standard resins. This keeps costs
62
CHEMICAL ENGINEERING MARCH 21, 1983
-------
down, and avoids delivery delays, since the manufac-
turer can use standard parts. If a standard fan-wheel
resin is not acceptable for a particular duty, encapsulat-
ing a standard-resin wheel in a special resin is more
sensible than building a custom wheel—not all resins
are suitable for wheel construction, and data on safe
operating speed may have to be developed from
scratch.
A fire-retardant fan housing may be desirable for
fans that handle combustible fumes or that may be ex-
posed to fire. Antimony trioxide added to the fan-hous-
ing resin imparts fire-retardance. The ASTM standard
does not allow any such additives in fan-wheel resins
because they cut the translucency of FRP and thus ob-
struct visual inspection of fan-wheel defects.
An FRP fan handling combustible fumes should also
be nonsparking. We have already discussed how a fan
may be built to conduct static electricity to the ground.
The user completes the installation by grounding the
fan base.
Fans in corrosive service require drive arrangements
that keep bearings and motors outside and away from
the gas stream. Fig. 11 on p. 54 shows seven common
arrangements. Of these, 1, 8, 9 and 10 are considered
acceptable for corrosive service. Fig. 29 shows an FRP
fan in arrangement 1. Arrangement 4, which has the
motor shaft in the gas stream, and 3 and 7, which have
the bearings in the gas-stream, are not acceptable. In
any arrangement, the bearings should be visible and
accessible.
Shaft-hole leakage is generally inward when a fan is
running. Where fan housings are pressurized, or there is
another possibility of fumes leaking out, it is advisable
to provide a shaft seal. Lubricated-lip seals or parking
glands are usually available as accessories.
Duct connections do not often receive the attention
they deserve. Too often, small FRP fans are fastened to
ducts by means of large, costly bolts because the speci-
fier is accustomed to high-pressure-pipe connections.
FRP duct does not need 150-psi bolting flanges; there is
a more practical specification.*
Ducts may be fastened to FRP fans by a flange, a slip
joint (flexible sleeve), or a butt joint. Outlet connections
are usually flanged; fan manufacturers typically offer
transition pieces so that the installer can match a round
outlet to a rectangular duct. Inlet connections are either
flanged or slip-jointed. Slip connections prevent trans-
mission of vibrations from the fan to the duct, but they
require caution since inlet suction may pull the sleeve
material into the inlet, causing an obstruction. Butt
joints should be selected with care, since they are per-
manent and cut off access to the fan's interior.
Mark Lifiaunex, Editor
•National Bureau of Standard! PS 15-69 applia to nr duct.
The authors
John £. Thompson is Senior Product
Manager for The New York Blower Co.,
7660 Quincy St., Willowbrook, ft.
60521. He began hit career with the
company at the manufacturing facility,
where he worked in engineering, research
and field service before transferring to a
marketing function. Hii recent
reiponiibiliaa have included managing
the marketing lupport-service group,
writing and ^i"ing much of the
company1! trrhninal literature, and
aiding product designer! in applying
marketing-research data.
C. Jack Trickier b Vice-President of
Corporate Development for The New
York Blower Co., 171 Factory St.,
LaPone, IN 46310. Prior to his current
position, he was the company1! Chief
Engineer and then General Manager of
the manufacturing operation. Mr.
Trickier has a B-S. in mechanical
engineering from the Illinois Institute of
Technology, and is registered in Illinois
and Indiana. He chaired the AMCA
committee that developed the industry!
sound-ten m^. and he recently chaired
the ASTM committee that wrote the
standard for rw fans.
Reprints
This 16-page report on fans and fan systems will soon be available as a reprint. To order, check No. 095 on the Reprint
Order Form in the back of this or any subsequent issue. For a free catalog of reprints, check No. 305 on the Reader
Service Card.
Equipment sources
For information on suppliers of fan* consult Section 1 of the annual Chemical Engineermg Equipment Bvyen' Guide.
CHEMICAL ENGINEERING MARCH 21. 1983
63
-------
ITEM 7
Basic Courses In Fan Selection, Fan Density Correction, And
Fan Arrangements And Classes
Chicago Blower Corporation
-------
'resented by
CHICAGO BLOWER CORPORATION
-------
COURS
'-^-^•'-J^^^^^-^^t^^Sef^sfs^y^^^
.
I
Course 100 is part of a series developed by Chicago Blower Corp. to assist those interested in
fan selection and engineering. While we do desire your business, we have tried to keep these
courses "generic" and applicable to the many good fan manufacturers in our industry. Please
write us if you have questions or critiques on any material presented.
Purposely we have kept Basic Course 100 simplified because many of you are not interested in
further detail. If you want to go beyond this level, write us. We will be happy to send courses (at
no charge) as follows:
Course 100:
Course 200:
Course 300:
Course 400:
A BASIC COURSE IN FAN SELECTION.
How to select fan size and type from manufacturers' catalogues.
(This course — if you would like another copy.)
A BASIC COURSE IN FAN DENSITY CORRECTION.
How to compensate for temperatures other than 70°, elevations
above or below sea level and suction pressure.
A COURSE IN FAN ARRANGEMENTS AND CLASSES.
Illustrated definitions and guides to which class and arrangement
to use. Also includes rotations and discharges.
ADVANCED COURSE IN FAN DENSITY CORRECTIONS.
How to correct density for gas other than dry air.
We also are developing courses in sound, fan curves and other topics. Availability of these
additional courses will be announced.
Terms such as "Fan Engineering", "Engineered Selections", etc. abound in the air moving
industry. They tend to infer a certain mystique to a sometimes simple fan selection process.
Some selections are tricky and critical, but most are easy and straight-forward. Our purpose
here is to show the easy way to fan selection. Warnings are provided where needed so you
won't get in overyour head in unusual situations.
"FAN ENGINEERING" is the development and manufacturing of a product suitable for air
moving applications. It's a difficult process that is already done for you by the fan
manufacturer. Your interest is to pick the size and type fan for your use.
'TCoovnqht 1981. Chicago Blower Corooration
-------
.
| f*
There are many fan "types". Generally you can group any type into one of two broad
categories, Centrifugal Fans and Axial Fans. Centrifugal Fans have an airflow which enters the
rotor and is turned 90° in all directions. Usually the air is then captured in a scroll-shaped
"housing" and pushed through the fan outlet. Most residential forced airfurnaces and
automobile heater fans are "centrifugals". Axial Fans have propeller type rotors and the
airflow is straight-through. A common window fan is a type of "axial"
-'ix
Within the centrifugal and axial categories, there are many sub-groups. Centrifugal
I. | sub-groups are typically named by their "wheel" type ("wheel" is rotating part that moves air,
• * * often called a squirrel cage or propeller). Axial fans have sub-groups named by theircasing
pa styles.
r "^s^^^s^Ka^^^f.--"-^y^gg^^
| | o Further descriptions of the fan sub-groups (types) and typical applications are included in the
chart on the next page.
To make your selection, you need to know What is in the Alrstream? Clean air? Dust? Moisture?
This leads you to the correct fan type.
If you can put your face in the airflow and breathe, the airflow should be clean enough to pick a
,2 1 good efficient Airfoil Centrifugal Fan (a specialized variation of the Backwardly Inclined Fan)
or Axial.
If the airflow is dusty, or could be dusty if something in the system breaks down, pick a
1 22 Backwardly Inclined Fan. They require more energy than an Airfoil but will run longer in an
erosive atmosphere.
OI
Wnen tne airf|ow 9ets really dirty, or you are handling material, you should consider a Radial
Tip or Radial Bladed fan.
jffiaEMR»^e&«,.irfMg:
EXCEPTIONS; Custom Airfoil fans are available in construction which allows fordirty air.
These are practical in power plant and other large installations. Radial blades are sometimes
used in clean air because they are aerodynamically more suitable for systems requiring low air
volumes at very high pressures.
I B
There are other fan type details required (Arrangement, Class, Rotation, etc.) but since most
manufacturers limit tneir standard types to certain "sizes", size selection is the appropriate
next step.
-------
CENTRIFUGAL FAN TYPES
Bali,)
v^
u
U-
or
DESCRIPTION
FORWARD CURVED:
The wheel's blades are small and curved
forward in the direction of the wheel's rotation.
This fan runs at a relatively low speed to move
a given amount of air. This wheel type is most
often called a squirrel cage wheel.
I APPLICATIONS
Primarily for low pressure heating, ventilating
and airconditioning such as domestic
furnaces, central station units and packaged
airconditioning equipment.
RADIAL BLADE:
a This wheel is like a paddle wheel.. .wither
| without side rims.The bladesare
perpendicularto the direction of the wheel's
rotation and the fan runs at a relatively medium
speed to move a given amount of air.
BACKWARD INCLINED:
The wheel's blades are flat and lean away from
the direction of the wheel's rotation. This fan
runs at a relatively high speed to move a given
amount of air. It is more efficient than the
above listed types.
The radial blade type is designed for material
handling applications, features rugged
construction and simple field repair. Also used
for high pressure industrial requirements.
General heating, ventilating and air
conditioning systems. Used in many industrial
applications where the airfoil blade might be
subjected to erosion from light dust.
AIRFOIL BLADE:
Although not a "Basic Type", this is an
important refinement of the Backward Inclined
wheel design. It has the highest efficiency and
runs at a slightly higher speed than the
standard flat blade to move a given amount of
air.
RADIAL TIP:
The wheel's blades are somewhat cupped in
the direction of the wheel's rotation but the
blade leans back so that its outside tip
approaches a radial position. This fan runs at
approximately the same speed as a backward
inclined wheel to move a given amount of air.
Most efficient of all centrifugals. Usually used
in both larger HVAC systems and clean air
industrial applications where the energy
savings are significant. Can be made with
special construction for dusty air.
«s*"?^
This type is also designed for material handling
or dirty or erosive applications and is more
efficient than the radial blade.
TVPPQ
, Jl il riut3
PROPELLER:
Wheels usually have two or more single
thickness blades in a simple ring enclosure.
Efficiencies are generally low and use is
limited to low pressure.
High volume air moving applications such as
air circulation within a space or ventilation
through a wall without attached duct work.
mum
TUBEAXIAL:
The wheel is similar to the propeller type
except it usually has more blades of a heavier
design. Wheel enclosed in a drum or tube to
increase efficiency and pressure capability.
VANE AXIAL:
Most efficient axial type fan. Uses straight-
ening vanes to improve efficiency and pressure
capability. Blades often have airfoil shapes and
may be available with adjustable pitch.
Pressure capabilities are medium to high.
^B^^^B^«»^HHBH8BBOBB5MBIi^H^^^BKBBE
— i^BBHBiBiircri'ifcvfTfiinTmm iT.ffi^&"lrTiM
INLINE CENTRIFUGAL:
This type is actually a centrifugal fan, with
airfoil or backward inclined wheel in a
Vaneaxial Casing. Good efficiency but lower
than a similar centrifugal type.
4
Ducted HVAC applications where air
distribution on the downstream side is not
critical. Industrial applications include drying
ovens, paint spray booths and fume exhaust
systems.
General HVAC systems especially where
straight thru flow and compactness is required.
Good downstream air distribution. Used in
many industrial applications.
Used primarily for low pressure return air
systems in heating, ventilating and air
conditioning applications. Has straight thru
flow.
-------
I TO SELECT THE FAN SIZE YOU NEED TO KNOW —
-- CFM: Cubic Feet per Minute or Q. This is the volume of the air flowing. If the "pounds" of airareaFso
2.10 important, such as when used to support combustion, refer to Course 200. CFM is usually used "as is" in
ventilation, exhausting, conveying.
;*s*M«^^
2,20
SP: Static pressure or Ps. This is most common term used to identify the "push" needed to overcome the
system's resistance to airflow. The unit of measure is "INCHES" (inches water guage) such as 3" SP
Pressure is often corrected before a selection is made. If fan will be in that rare environment where the air is
near70°F., near sea level, and is "dry", proceed without correction to SP. In the real world, a 10% addition to
SP will usually get you close enough for most applications. To be safe, you should refer to SP corrections in
the Chicago Blower Engineering Guide EG-1B. Use table and instructions, page 1, lower right.
The most commonly used reference for fan selection is the "Multi-Rating Table" Usually the table lists the
CFM vertically on the left side and SP horizontally across the top.
2.30
Example
(simplified)
SIZE 24% Airfoil
CFM
2400
2550
2700
2850
3000
1"SP
RPM
1254
1307
1360
1415
1470
BHP
.66
74
83
.93
1 04
2"SP
RPM
1471
1515
1562
1612
1661
BHP
1.09
1 19
1.30
1 42
1.55
3"SP
RPM
1683
1713
1748
1787
1829
BHP
1.59
1 70
1.82
1.96
2.10
4"SP
RPM
1934
1963
1996
2032
2115
BHP
241
2.55
2.70
2.86
3 24
2.40
c
8
Size selection is somewhat "trial and error" because you pick a size and observe what it will do with the
required CFM and SP. Again, the basic wheel type was selected by what is in the airstream. This gets you into
the right catalogue with its multi-rating tables.
You observe what a size will do by finding the crossing point... of the row of figures next to the CFM and the
column of figures under the SP. Important: You have to use SP corrected to "near 70°F, near sea level and dry"
(see paragraph 2.20).
•*-"——-
Using the simplified multi-rating table above (2.30), let's try 2700 CFM at 3" SP. Following the 2700 CFM row
across to the 3" column, you read 1748 RPM and 1.82 BHP. You now know that our hypothetical 24% Airfoil will
work in your system if it is run at 1748 RPM, and at that condition, will require 1.82 BHP (Brake Horsepower). A
ra 2-HP motor, for example, will work fine. In any case where a SP correction is made for conditions other than
2,52 dry airat 70°F, sea level, the BHP from the multi-rating table must be reduced by the same factorus.ed to
increase the SP. Example: If the above 3" resulted from increasing the SP by 10% (ie. SP x 1.10), the BHP can
be reduced by dividing it by the same 1.10 factor. 1.82 BHP is really1.65 BHP if the SP was corrected. MORE
ABOUTTHIS IN COURSE 200.
Another example. . .Try 2850 CFM at 1" SP. Congratulations if you read 1415 RPM and .93 BHP! You have
learned the common fan selection method.
Not every system will have nice even values such as 2700 at 3". In-between values can be determined by
interpolation. Since the tables were developed from relatively smooth performance curves, you can
interpolate between the SP columns and CFM row. Straight-line interpolation calculations are usually
suggested as the way to find intermediate values however that tedious chore is usually more for personal
AQ satisfaction than necessity.
It is easier to try "eyeball interpolation" ... which is another way of saying "look at the printed numbers and
eyeball the intermediate ones". For 2605 CFM at 1", will it matter if an "eyeballed" 1320 RPM and .76 BHP
miss a mathematically interpolated 1326 RPM and .77 BHP? With motor speed, drive selection and system
variations from design, further calculations are usually unnecessary.
Traditionally, the fan industry spaces out the multi-rating tables in even increments
of Outlet Velocities (0V) rather than CFM. CFM & 0V are directly proportional so
— instead of 2400, 2550, 2600, etc. you will find the column at the right more typical
2,70 in a multi-rating table.
The non-even increment CFM values as published show the need for eyeball
interpolation.
Note: OV = CFM/OA (Outlet Area). In this example the fan has an Outlet Area of 1.59 sq. ft.
CFM
2385
2544
2703
2862
3021
OVFPM
1500
1600
1700
1800
1900
2.80
Paragraph 2.40 mentions "trial and error", which to some means "drudgery". Take the drudgery out of it by
following a simple rule. If a rating (intersection of desired CFM row and SP column) is at the top, or beyond the
top of a table, there are smaller sizes to consider. Conversely, if the rating is at the bottom, or below the
bottom of a table, there are larger sizes to consider. If the rating appears to be beyond the left or right side of
the table, you should try another fan type- or catalogue-or call your fan vendor for help.
-------
PRACTICE
H£Sf&si&£!£i!i/}^:iy*zX>'-'3i»?:2
The size 15 (in table below) is one selection for 1600 CFM at 11/." (1295 RPM, .46 BHP). A smaller
fan should be considered for 1000 CFM at I1/.". A larger fan should be picked for 4500 CFM at
I1/." and the 15 may not be able to do 2%'at any CFM. Note — there are ways to select
off-table ratings but the technique is beyond the scope of this first basic course.
3.10
CFM
660
792
924
1056
1188
1320
1452
1584
1716
1848
1980
2112
2244
2376
2508
2640
2904
3168
3432
3696
0V
FPM
500
600
700
800
900
1000
1100
1200
1300
1400
1500
1600
1700
1800
1900
2000
2200
2400
2600
2800
1/4" SP
RPM
559
607
662
719
778
839
902
966
1031
1098
1165
1234
1302
1370
1439
1508
1647
1787
1927
2071
BHP
.03
.04
.06
.07
.09
.11
.14
.17
.21
.25
.30
.35
.41
.47
.54
.62
.81
1.02
1.28
.1.58
1/2" SP
RPM
752
787
834
887
942
998
1057
1117
1178
1240
1304
1369
1434
1500
1566
1702
1839
1975
2115
BHP
.08
.10
.12
.14
.17
.20
.24
.28
.33
.38
.44
.50
.57
.65
.73
.93
1.15
1.42
1.73
3/4" SP
RPM
942
983
1031
1085
1139
1196
1253
1312
1373
1433
1495
1559
1623
1753
1855
2020
2156
BHP
.17
.20
.23
.27
.31
.35
.40
.46
.52
.59
.67
.75
.84
1.05
1.28
1.56
1.88
T'SP- 1-1/4" SP
RPM
1080
1118
1164
1214
1268
1323
1379
1437
14%
1557
1617
1678
1803
1931
2063
2195
BHP
.26
.29
.33
.38
.43
.48
.54
.61
.69
.77
.86
.95
1.17
1 .41
1.70
2.03
RPM
f-
^
1205
1242
|f2B7
>S3fc_
1389
1444
1499
1556
1613
1673
1733
1853
1978
2106
2236
BHP
..
'
.36
.40
—W
.56
.63
.70
.78
.87
.96
1.06
1.29
1.55
1.84
2.18
1-1/2" SP
RPM
1000
1290
1321
1'Mo
1402
1451
1504
1558
1613
1669
1725
1783
1904
2024
2148
2275
BHP
:FM
.44
.48
"FlW
.58
.64
.72
.79
.88
.97
1.07
1.17
1.41
1.68
1.99
2.33
4500 CFM
ai£&£
|5| Most manufacturers will offer more than one size that is capable of performing to your
requirements. Theoretically there is only one perfect size for a performance point but for
3,20 economical or space considerations, it is more often the smaller fan that is picked. The tables
should not allow a selection much larger than the perfect size because a too-large fan can be
unstable (pulsate) or noisy.
The multi-rating tables reproduced on the facing page show four consecutive sizes of Airfoil
fans. SISW (means Single Inlet, Single Width as opposed to DIDW which is Double Inlet-2
inlets - Double width). Try following a practice rating like 2200 at 1" through the four tables.
3.21
#15
#16%
#18'/4
#20
1478 RPM
1195 RPM
968 RPM
822 RPM
.66 BHP
.57 BHP
.51 BHP
.48 BHP
So, trial and error is reduced by opening the manufacturer's catalog to any table and
visualizing "where" your rating is with respect to up (too large?) or down (too small?). Then
move in the proper direction and zero in on the best size. All four sizes will "fit" the
requirement. Which is best? That can be determined only after considering the following.
3.22
Physical size
Purchase cost including motor
Operating cost (Note #15 costs 38%
more to run than #20)
Future needs
Availability of the product
Noise • Generally most efficient is
quietest
Erosion • Higher speed sizes erode
faster if air is dirty.
Remember, assistance is always available from qualified fan engineers.
That's all there is to basic fan selection. You must observe several cautions when finalizing
your pick. Check the selection's RPM against the manufacturer's RPM limit published in the
catalogue. Usually a deration table to those limits is published for elevated airstream
4>00 temperatures.
Remember, if you are interested in other courses, write us.
-------
161/
SISW
181/
SISW
SISW
H
.1
•fi
at
-'V?
BHP
.44
.48
.53
.58
.64
.72
79
.88
97
1 07
, .17
1 .68
,.99
2.33
RPM 1 BHP
,400
,432
,469
,513
16,3
1668
1722
177B
834
960
2070
2190
2315
.56
.61
67
RPU
1504
1536
,575
80 ,6,9
89 1667
.98 ,718
,.07 i ,773
,.,7 : ,828
, JS
,.53
1.82
2 13
2 49
,382
,996
2114
2233
2355
SF
BHF
70
76
82
90
98
1.07
, .17
1.28
1.40
1.65
1.95
2J8
2.64
RPM
1668
,700
1735
1775
1820
1869
192,
1976
2084
2197
2313
2432
7-SP
BHP
.95
1.02
, 10
1.18
1 28
1.39
,.50
63
1.90
2.21
2.56
2.96
J
KPU
SP
•HP
1-1 /
RPU
185 , 3,
188S 1 40 ,994
,923 1.50 2025
,966 | ,6, 2063
2012 1 73 2,04
2062
2169
2278
2390
2506
i 86
2.15
2 48
2.85
3.26
^w±i^izi-^j^^jfi*£4H£tj^LMMjscFjrsFSf&&niP8sjm
2150
2248
2356
2466
2578
2" SP
BHP
1 .64
1 .74 1
1 85
, .97
2 11
2.4
2.76
3.14
3.57
CFM
795
954
1113
1272
1431
1590
1749
1906
2067
2226
2385
2544
2703
2862
3021
3160
3498
3816
4134
4452
OV
FPM
500
600
700
800
900
1000
1100
1200
1300
1400
1500
1600
1700
1800
1900
2000
2200
2400
2600
2800
1/4" SF
RPM
506
552
601
653
708
762
820
873
938
998
•HP
1/2" SP
RPM
.04
05 ' 684
.07 716
.09 ! 753
, , . 806
.17
.21
.25
.30
856
907
961
,016
1071
,069 .36 1127
1122
1183
1246
1308
1370
1497
1624
,751
1883
42
.49
57
.76
.99
1 J4
135
1.91
1186
1244
,304
1364
1424
1547
1672
1795
1923
BHP
.10
.12
.14
17
J5
J9
.34
40
.46
.53
.61
.69
.79
.89
1.12
1.40
1.72
2.00
1/4"SP
RPM
857
894
938
986
1036
,087
,139
,193
,248
,303
1359
1417
1475
1593
1713
1836
1960
BHP
.21
.24
.28
J2
.37
.43
.49
.56
.64
.72
.61
.91
1.02
1 J7
136
1.89
2J8
1" SF
RPM
981
1016
1058
1104
1153
,203
1254
1307
1360
1415
1470
1525
1639
1756
1875
1996
BHP
.3,
.35
.40
.46
.52
.58
.56
.74
.83
.93
1.04
1.15
1.41
1.71
2.06
2.46
1-1/4" SF
RPM
096
129
170
2,5
263
3,3
363
414
466
520
575
685
796
915
2033
BHF
49
.54
.6,
68
76
85
95
1.05
.17
1 J9
1.56
, .88
2.23
2.54
1-1/2" ST
RPM
1173
,20,
,235
,275
,3,9
,367
14,6
1466
1517
1568
1621
1731
1840
,953
2068
•HP
1-3/4" SF
RPM
I
.53
.58 ,272
64 ,301
71 i 1336
78 1376
BHF
2"SF
RPM
BHF
68
74 ,367 : 85
81 ,396 i 92
.69 1431 i 1 00
2-1/2" SP
RPM
BHP
,515 , ,5
,545 ' ,.24
.87 ' ,4,9 I 96 1471 i 09 ' 1578 1.33
.96
.06
.17
J9
.42
.71
2.03
2.41
2.83
1466
1516
1566
1616
1667
1773
,882
1991
2104
,.,8
,.30
1 42
136
1.35
2JO
2.58
3 02
1562
let:
1661
171,
,8,4
,221
2030
2140
1.30
1.42
1.55
1.69
2.00
2J6
2.76
3.20
1 655
1 699
1746
1796
895
1997
2102
221,
, 55
,.68
1.82
1.97
2JO
2.68
3.10
3.59
3"SP
RPM
BHP
3-1/2" SP
RPU
Bh*
683 ' 1.59
748
787
829
875
972
2071
2173
2277
1.70 1812 ' 1 93
1.82 84, 210
1.95 ,875 2.24
2,,0 ,912 2.39
2J5
2.6,
3.01
3.45
3 95
1954
2044
2142
2241
2343
2.55
2.92
3.34
3.81
4.33
,-*
'*%>
I
i
a
j
CFM
r/s
1170
1560
1755
I960
2145
2340
2536.
2730
2925
3120
3316
3S10
3705
3900
4290
4680
5070
5460
OV
FPM
500
600
700
800
900
1000
1100
1200
1300
1400
ISOO
1600
1700
1800
1900
2000
2200
2400
2600
2800
1/4" SP
RPM
459
499
544
59,
640
689
74,
794
848
902
957
1014
1070
1126
1181
1239
1354
1469
1583
1702
•HP
.05
.07
.09
.11
.14
.17
J,
J6
J,
J7
.44
32
.60
.70
.81
.92
1 JO
132
1.89
2J4
1/7" SP
RPM
618
647
685
729
774
820
869
918
966
1019
1072
1126
1179
1233
1287
1399
1512
1623
,738
•HP
.13
.15
.18
Jl
JS
JO
J6
.42
49
.56
.65
.74
.85
.96
1.09
IJ7
1.71
2.11
236
3/4" SP
RPM
775
808
848
89,
936
983
1030
1079
1128
1178
1229
1281
1334
1440
1549
1660
1772
•HP
J6
J9
J4
40
.46
32
.50
.69
.78
.88
.99
1.12
US
135
1.90
2J1
2.78
1"SP
RPM
887
919
956
998
1042
1087
1134
1181
1230
1279
1329
1379
1482
1587
1695
1804
•HP
J9
.43
.49
.56
.63
.72
.81
.91
.02
.14
.27
.41
.73
2.10
232
3.01
1-1/4" $P
RPM
991
021
056
098
142
187
232
279
372
375
424
523
626
731
838
•HP
34
.60
.67
.75
.84
.94
.04
.16
J9
.43
38
.91
JO
2.73
3J3
1-1/7" SP
RPM
1060
1086
1117
1152
,193
1236
1281
1326
1372
1418
1465
,565
,663
1765
1870
•HP
.65
.71
.78
.87
96
,.06
1.18
1 JO
1 .44
,38
1.74
2.09
2.49
2.95
3.46
1-3/4" SP
RPM | BHF
1150
,,77
,207
,244
1283
1326
1371
1416
1461
1506
1603
1701
1800
1902
.83
.91
.99
1.09
1.19
1 Jl
1.45
139
1 .74
1 .90
2J7
2.69
3.16
339
2-SP
RPM
1236
'263
1294
1330
1370
1412
1457
1502
1547
1640
1737
1835
1935
BHF
1 .04
1.12
1 22
1 J3
1 46
1 39
1.74
1.90
2.07
2.45
2.89
3J8
3.92
2-1/2" SP
RPM
BHP
,371 , 4,
,397 i ,3,
1426
1459
14%
1536
1579
1624
1713
806
1901
1999
1-.63
1 .75
1.90
2.05
2J3
2.4,
2.82
3J8
3.79
4.39
1"
RPM
1522
1549
1581
1616
1654
1695
1783
1872
1965
2059
SF 3-1/T
BHP
1.95
2.06
2J3
2J9
2.57
2.76
3.19
3.68
4J2
4.83
RPM
1638
1665
169S
1729
1767
1848
1936
2026
2118
• SP
BHF
2.43
237
2.74
2.92
3.13
3.57
4.09
4.66
5JO
••••••••••(•••••••••••••••B
I
1
i
CFM
1170
1404
1638
1872
2105
2340
2574
2806
3042
3376
3510
3744
37T8
4212
4446
4680
5148
561«
9084
6562
OV
FPM
500
600
700
BOO
900
1000
1100
1200
1300
1400
ISOO
1600
1700
1800
1900
2000
7200
2400
2600
2800
1/4" SP
RPM
419
4SS
496
539
584
629
676
725
774
823
873
925
976
1028
1079
1131
1235
1340
1445
1553
•HP
.06
.08
.10
.13
.17
J6
Jl
J7
.45
33
.62
.73
.84
.97
1.11
1 .44
1.83
2J8
2-tl
1/7" SP
RPU
564
590
625
665
706
748
793
838
883
930
978
1027
1076
1125
1175
127«
1379
1481
1586
•HP
.15
.18
Jl
J6
Jl
J8
.43
30
38
.58
.78
.89
1.02
1.16
I Jl
1.65
2.06
233
3.08
J/4~ SP
RPM
707
737
774
813
354
897
940
984
030
075
121
169
217
314
414
515
817
•HP
Jl
JS
.41
.48
.53
.72
.83
.94
1.06
1.19
1 J4
130
1 M
2J9
2.78
3-35
1-SP
RPU
8,0
839
873
911
951
992
1035
1078
1122
1157
1213
1258
1352
1449
1547
1646
IMP
.46
32
39
.67
.76
.86
.97
1.09
1 J3
1 J7
133
139
2.08
232
3.03
331
1-1/4" SP
RPM
904
932
965
002
042
083
125
167
210
254
300
390
483
580
677
BHP
.65
.72
.30
.90
1.01
1.12
1 JS
1 J9
135
1.72
130
2J9
2.76
3J8
3.88
1-1/7" SP
RPM
968
991
1019
1052
,068
1128
1169
1210
1252
1294
1337
1428
1518
1611
1706
•HP
.78
.85
.94
.04
.15
J8
.42
36
.73
.90
2.09
232
2.99
334
4.16
1-3/4" SP
RPU
1050
1 074
1103
1135
1171
1210
I2S1
1292
1333
1376
1463
1553
1643
1736
BHP
1.00
1.09
1.19
Ul
1 .44
138
1.74
1.91
2.09
2J9
2.73
3J3
3.79
4.43
•^^M
r-sp
RPM
1128
1152
1181
1214
1250
1289
1330
1371
1412
1497
I58S
1675
1766
•HP
1 JS
1 J5
1.47
1.60
1.75
1 .91
2.09
2J9
2.49
2.94
3.47
4.08
4.70
2-1/7" SP
RPU
1251
1275
1302
1331
,365
1402
1441
,482
1563
1648
I73S
1824
•HP
1.69
1.82
1.96
2.11
2J8
2.47
2.67
230
3J8
3.94
436
5J7
••••
1"SP
RPU
1389
1414
1442
1474
1509
IS47
1527
1708
1793
1879
BHP
2J4
230
2.68
2.87
3.06
3J1
334
4.42
5.07
5.80
3-1/7" SP
RPM
495
519
547
578
612
586
767
849
933
•HP
2.91
3.09
3 JO
331
3.75
4J9
4.32
5.60
fije
-------
,
-
resented by
CHICAGO BLOWER CORPORATION
-------
FAN DENSITY
Course 200 is part of a series developed by Chicago Blower Corporation to assist those
interested in fan selection and engineering. While we do desire your business, we have tried to
keep these courses "generic" and applicable to the many good fan manufacturers in our
industry. Please write us if you have questions or critiques on any material presented.
This course is an extension of Course 100, A Basic Course in Fan Selection, for conditions
other than dry air at 70°F. and sea level. These conditions change the density of air and may
require corrections to obtain rated fan performance. There is no mystery to density corrections
in fan (and system) calculations, however it is often ignored. Ignoring density is one of the
major causes of insufficient air. This course covers the common corrections required for
temperatures other than 70°, elevations above sea level and suction pressure. If you want to go
beyond this course, write us. We will be happy to send courses (at no charge) as follows:
Course 100:
Course 200:
Course 300:
Course 400:
A BASIC COURSE IN FAN SELECTION.
How to select fan size and type from manufacturers' catalogues.
A BASIC COURSE IN FAN DENSITY CORRECTION.
How to compensate for temperatures other than 70°, elevations
above or below sea level and suction pressure. (This course - if you
would like another copy.)
A COURSE IN FAN ARRANGEMENTS AND CLASSES.
Illustrated definitions and guides to which Class and Arrangement
to use. Also includes Rotations and Discharge.
ADVANCED COURSE IN FAN DENSITY CORRECTIONS.
How to correct density for gas other than dry air.
We also are developing courses in sound, fan curves and other topics. Availability of these
additional courses will be announced.
We walk or run freely through air. We don't normally feel it or see it, and tend to forget that it is
there. Even more likely, we don't think of air as a mechanical mixture, of various gas
molecules, which has weight. If there is more weight of air (or any gas) in a given space, we say
that the density or air (or any gas) is higher. When air gets heavier, more work is required to
move it. Conversely, lighter weight (or thin) air is easier to move.
Measured fan performance will change as the airstream density changes. To avoid chaos in
published ratings, the fan industry has adopted a standard density of .075 Ibs. per cu. ft. All
ratings are made at, or adjusted to thfs standard. When a fan is applied to a system with a non-
standard density gas, corrections must be made for accurate results.
"Non-standard" densities are caused by... Temperatures other than 70°F. (21 °C)...
Elevations above or below 0' or sea level.. . Barometer readings higher or lower than 29.92*
Hg ... Partial vacuum in airflow caused by suction of the fan ... Gas other than "air"
Relative humidity over 0%.
CBC 200, covers Temperature, Elevation and Suction. The other causes of non-standard
densities are covered in CBC 400.
©Copyright 1981, Chicago Blower Corporation
-------
COURSE OUTLINE
This course will introduce you to one of the often important, and often ignored, details of fan engineering
... knowing the density of the gas at the fan's inlet.
1 • 1U The word "knowing" is intentionally used above. Density is an important detail only when it is not close to
*" .075 Ibs. per cubic foot. If you follow the 10% addition guide in CBC-100 paragraph 2.20, you are
conservatively covered for densities down to .067.
1.20
2.10
?*
2.11
This course CBC-200 approaches the density subject in two steps:
A. How density affects fan performance
B. How to find density
Step B is further broken down into two levels. This course covers dry air at various temperatures and
elevations, and systems with high suction on the fan's inlet. Course CBC-400 expands the subject into
gasses and gas mixtures other than air, and moist or saturated air.
Let's explore a hypothetical but typical fan application to demonstrate the effect of density on fan
performance.
Location: Wichita, Kansas (avg. elevation 1372')
CFM: 8570 CFM '
SP: - 15" at fan inlet, 0" on outlet
Temp: 225° F.
Application: Exhaust clean air
Using only the selection process learned in CBC 100, you may try to select a 221A C/3.
SISW
S&fJL
CFM
4930
5220
5510
saoo
6380
6960
7540
8120
8700
9280
9860
10440
11020
11600
12180
12760
13340
13920
0V;
FPM
1700
1800
1900
2000
2200
2400
2600
2800
3000
3200
3400
3600
3800
4000
4200
4400
4600
4800
^ *
RJ
1
1i
1?
17;
17
11 1
U 1
19-
200
207
214
.
• OUTLET
21-5/8x19-3/8 in. ;
inside i
B CLASS II RPM 2483 (
f --
JHP
3.94
1499
16.13
7.36
'.69
J.12
V*
11"
A
90"
I 16
A 3-
1 *>L
\ I1
----- 12- SP
RPM
2347
2389
2435
2483
'2532;
2584
2639
2697.;
2759>
2823 ~
2889
IP
BHP
1646
1765
18.93
2031
zteer
23.40
25.11
2633
28.85
30.88
33.03,
'-f^r-'i
^
-- 13- SP
RPM
2424
2464
2508
2555'
2603
2653-;
2706
2762,,
-2821 •-
2883
S?^
•"SiJT-
BHP
1797
19.20
20.54
21.97
23.51
25.17
26.94
28.82
30.82
32.92
.yVcsR"
>.90 sq.ft.
nside area
:LASS III RPM 2943
-- 14- SP
RPM
2499
2537..
2579
2625
2672
272U
2772
2825 ,
2882-
2943'
.-• •-'>
-t,^. - •
ffe
BHP
19 SO
20.80
22.18
23.66
25.26
26.97.
28.80
30.74
32.81
34.99
it
~&\* \-T
15' SP ' •
RPM
'2608-
2648
269?
2739 -
2787^
2836
2888.
2943J
-^j^:*^.
1®s*
'.fi-JVj;
BHP
22.42
23.86
25.39
27.04
28.80
3068
32.69
34.82:
•- ££>'.
jg&fe
-rsW™
-f, „«• ''rffl
The selection "looks" good. Fan
speed is 2682 RPM, BHP is 25.05.
Efficiency is good and the fan is
in stock.
The selection "looks" good but the published ratings are based on .075 density air. We will show later in
this course that the actual condition described in para. 2.10 is .053 Ibs/cu. ft. density. We will also show you
that this lower density will cause your above selection to be short on volume, high on pressure and high on
horsepower. The selected fan will not do your job.
The density of a gas affects two elements of your fan selection criteria. 1) It affects the resistance to flow
2.20 and therefore the pressure requirement of the fan. 2) It affects the power or Brake Horsepower needed to
move the air.
I
-------
PY FACTOR
To use any manufacturer's fan ratings, you must convert your requirements to the density used in the
ratings. That density is .075 Ibs. per cubic foot. The conversion is very simple once you know the density of
OO the 9as entering the fan's inlet. All you need is a "factor" to adjust yourapplication condition to .075.
Factor = where d = density at fan inlet.
i
3.11
3.12
In our hypothetical application (in Wichita), we said that the inlet density is .053, therefore:
Factor = -075
.053
= 1.415 (1.42 is OK)
Examples:
Dry air, 200°F, 2000'elev. has density of .0559. Factor is 1.34
Dry air, 325°F, 0'elev. has density of .0506.
Factor is 1.48
Dry air, 600°F, 1500'elev. has density of .0355. Factor is 2.11
-™--iJ*-lj—•—f-1-^1-JJ.ii.Jl,—,-'>-J,..._-ti-,;-?TV ,-. -. _ Juyra
•.0---,*;",^-.-. f-^^jj'.in ••TiW^i" = = r.-_^-_r^^. _, ,\. -• •-. ^ -. - j
To select a fan from a performance table, first determine whether or not the table conditions are based on
the .075 standard. This is often referred to as "COLD" conditions (vs. "HOT" which are the "actual"
conditions at some other density). This common terminology can cause confusion below 70°F. Usually job
conditions are stated "Hot" In either case, correct as follows:
rx
Ci
3.2
T7
r.
&
r;
s-a
1
ki
i
y
JO CFM:
SP:
BMP:
RPM:
eraa J*^5aa5aa^
If ACFM, which is "ACTUAL" CFM, do not convert.
If SCFM, which is "STANDARD" CFM, you must convert to ACTUAL by
multiplying it by the factor. A more detailed discussion follows in this course.
Multiply by the factor.
Select fan, determine BHP for selection and divide by factor.
Do not convert.
3.30
Applying the above to the Wichita example of 8570 CFM, 15" SP, 1.42 factor:
8570 CFM (assuming ACFM)
15" x 1.42 = 21.3" SP
Selecting a 22V< SQA (using off-table selection methods which will be covered in a later course), we can
make 8570 @ 21.3" running at 3077 RPM and 36.40 BHP. The 36.40 BHP represents power requirements at
.075 so we divide by the factor and find we will require 25.63 BHP at .053 density. Unfortunately the C/3 fan
won't run this fast which further dramatizes the problem created by overlooking density.
ACFM means ACTUAL CFM. "Actual" represents the conditions of the job - not corrected to any density.
O 4Q SCFM means STANDARD CFM. This means someone has converted the values to those which would exist
if the job conditions were at some fixed "standard" density . . . such as .075.
3.50
Correct interpretation of the basis used in determining the CFM and SP requirements of a fan are every bit
as important as the type and size. Most tables that are used to calculate system losses are at .075 standard
conditions. The system designer is more interested in "actual" values so the conversions will usually
result in actual CFM (ACFM) and actual SP (ASP). This is why it is typical in ventilation work to convert only
the SP since, to repeat, we are looking for ACFM and standard SP (SSP). We will use the "unofficial" terms
ASP and SSP, for actual and standard static pressure, in this course as a convenient tie-in to ACFM and
SCFM.
-------
Fan selector's procedure
A. Select with specified ACFM. Multiply ASP by factor and select with resulting SSP Divide
selection's BMP by factor.
3.5 1 B- Multiply SCFM by factor and select with resulting ACFM. Multiply ASP by factor and select with
resulting SSP. Divide selection's BHP by factor.
9 Note: To select we had to know basis of designer's specifications. In both cases SP is converted. With
CFM, the key memory jogger is "A" for Always . . . Always use ACFM.
Caution should be used on the BHP conversion. Dividing by the factor is correct, howeveryou should
select the motor for the maximum job condition's BHP. For example, system density may change as hot
3,60 9as is drawn into the fan. Before the gas warms up, the fan may be handling gas at close to .075 density and
f* draw the BHP shown in the performance table for the selection. In this case, you would size motor for a
si cold start.
In the following sample specifications, what values of CFM and SP would you use fora proper fan
selection?
Examples:
A. 3705 ACFM, 1"SP at .070 density.
4.10
B. 2703 ACFM, 1 %" SP at .060 density.
C. 4000 CFM, %"SP both at .075
density.
D. 4000 CFM,'/." SP both corrected by
designer to .075 standard but actual
conditions are at .035 density.
Answers:
A. Always find density factor first.
Factor = .075/.070 = 1.07.
Since CFM is ACFM, do not convert.
SP is given as being "at" .070 which must be
converted to .075. 1.07x1" = 1.07"SP.
Therefore, select for 3705 CFM and 1.1" SP.
B. Factor = .075/.060 = 1.25
No CFM conversion. 1.25x1%" = 1.56" SP
Select for 2703 CFM and 1.6" SP.
C. Telling you that both are at .075 says that CFM
is SCFM and must be converted. However, the
factor .075/.075 is 1.0 so no conversion is
necessary. Everything is at .075 just like the
performance tables, so select for 4000 CFM
and'/."SP.
D. Like C, the CFM must be converted and this
time the factor is not 1.0. It is .075/.035 = 2.14.
2.14 x 4000 = 8560 ACFM. The SP is supposed
to be at .075 so no conversion is needed. Your
selection should be for 8560 CFM and 'A" SP.
You are now an expert at applying the density factor. Now lets move on to finding the factor.
FINDING DENSITY FACTOR
The formula for the factor is 075/d where fordrygas" d — 075 x
„ BP „
- 29.92 -
AIP
AP
- x SG
5.00
where: d = actual density at fan inlet.
T = Temperature, °F. (dry bulb) at inlet.
BP = Barometric Pressure, "Hg."
AIP = Absolute Pressure at fan Inlet, (any unit).
AP = Absolute Pressure, (same unit as AIP).
SG = Specific Gravity of the gas.
This course covers T, BP, AP and AIP. For SG other than 1.0 for non-air gasses and gas mixtures including
those with moisture, refer to course CBC-400.
-------
1
^TEMPERATURE AND ELEVATION FACTORS
The mos
The maU
t
S
.00
>_-*
:3
~]
"il
-j
5
•j<
j
.-«
-j
;s
^J Actual den
_J and 2500 '
1 ,«u.»
t com
i for.
TEMP.
°F
-40
0
40
70
80
100
120
140
160
180
200
250
300
350
400
450
500
550
600
650
700
750
800
850
900
950
1000
mon influences on density are the effects of: T = Temperature other than 70" F
BP = Barometric Pressures other than 29.92
caused by elevations above Sea Level.
075/d for temperature and pressure is done for you in the following table.
^*S?K.3:?f^?ALTITUDE (FEET) WITH BAROMETRIC PRESSURE IN "Hg^S^tf -fe"95S&t
- .0 '.;•
',2332
.79
.87
.94
1.00
1.02
1.06
1.09
1.13
1 17
1.21
1.25
1 34
1.43
1.53
1.62
1.72
1.81
1.91
2.00
2.09
2.19
2.28
2.38
2.47
2.57
2.66
2.76
500' .
29.38
.81
.88
.96
1.02
1.04
1.08
1 11
1.15
1.19
1.23
1.27
1.36
1.46
1.56
1.65
1 75
1.84
1.94
2.04
2.13
2.23
232
2.42
2.52
2.6.1
2.71
2.81
1000'
f 28.86 -
.82
.90
.98
1.04
.06
10
.13
.17
.21
.25
.29
1.39
1.49
1.58
1.68
1.78
1.88
1.98
2.07
2.17
2.27
2.37
2.46
2.56
2.66
2.76
2.86
1500'
.28.33
.84
.92
1.00
1.06
1.08
1 12
1 16
1 20
1.24
1.28
1.32
1 41
1 51
1 61
1.71
1.81
1.91
2.01
2.11
2.21
2.31
241
2.51
2.61
2.71
2.81
2.91
2000'.
27 .82;
.85
.93
1.01
1.08
1 10
1 14
1.18
1.22
1.26
1.30
1 34
1 44
1.54
1.64
1.75
1.85
1.95
205
2.15
2.25
2.35
2.46
2.56
2.66
2.76
2.86
2.96
2500'
27,3t:
.87
.95
1.03
1.10
1.12
1.16
1.20
1.24
1.28
1.32
1.36
1 47
1.57
1.67
1.78
1.88
1.98
2.09
2 19
2.29
2.40
2.50
2.60
2.71
2.81
2.91
3.02
3000'-
26.82j
.88
.97
1.05
1.12
1.14
1 18
1.22
1.26
1.31
1 35
1.39
1 49
1.60
1.70
1.81
1.92
2.02
2.13
2.23
2.34
2.44
2.55
2.65
2.76
2.86
2.97
3.07
3500'
26.32
.90
.99
1.07
1.14
1 16
1 20
1.24
1.29
1.33
1.37
1 42
1.52
1 63
1.74
1.84
1.95
206
2.17
2.27
2.38
2.49
2.60
2.70
2.81
2.92
3.02
3.13
4000'
25.34
.92
1.00
1.09
1.16
1.18
1.22
1.27
1.31
1.35
1.40
1 44
1.55
1.66
1.77
1.88
1.99
2.10
2.21
2.32
2.43
2.53
2.64
2.75
2.86
2.97
3.08
3.19
4500'
25.36
.93
1 02
1.11
1.13
1.20
1.25
1.29
1.34
1.38
1 42
1 47
1 58
1.69
1.80
1.91
2.03
2 14
225
2.36
2.47
2.58
2.69
2.80
2.92
3.03
3.14
3.25
5000'
24.90,
.95
1.04
1.13
1.20
1.22
1.27
1.31
1.36
1 41
1 45
1 50
1.61
1.72
1.84
1.95
2.06
2.18
2.29
2.40
2.52
2.63
2.74
2.86
2.97
3.08
3.20
3.31
5500'
24.43
.97
1.06
1.16
1.22
1.25
1 29
1.34
1.39
1.43
1 48
1.53
1 64
1.76
1 87
1.99
2 10
2.22
2.33
245
2.56
2.68
2.80
2.91
3.03
3 14
3.26
3.37
6000%
23.98|
.99
1.08
1.18
1.25
1.27
1.32
1.37
1.41
1.46
1.51
1.55
1.67
1.79
1.91
2.02
2.14
2.26
2.38
2.50
2.61
2.73
2.85
2.97
3.08
3.20
3.32
344
sity of dry air at above conditions can be easily calculated, d = .075/factor. Example: density at 350* F.
s. 075/1. 67 = .045.
""' • ••—•»— *•»•*
": To use the temperature/elevation table, find the factor number at the crossing point of your temperature (go
il
across) and altitude (go down).
6.10
Tryafew... 120° F. and 3000' = 1.22
180° F. and 0' = 1.21
190° F. and 0' = 1.23
(OK to eyeball interpolate)
600° F. and 5000' = 2.40
Now you try
the next two:
800° F. and 2000' =
100° F. and 700' =
If you found factors of 2.56 and 1.09, in 6.10's last two conditions, you are ready to try more practical
examples...
Chicago (614' elev.), 200° F.. .Need 5000 SCFM at 6" SP.
Solution: Factor from table is 1.27. Using the "rules" from 3.20, we must select our fan for
5000 x 1.27 = 6350 CFM and 7.62" SP. Our BHP read from the multi-rating table will be
derated by dividing it by 1.27.
New York City (10' elev.), 650° F.. .Need 8500 CFM at 10" SP.
6. 1 1 Solution: Here we have a most common problem. Is the CFM actual or standard? It's not a
good rule to follow, however ACFM is often assumed in HVAC and industrial ventilation and
conveying applications. In combustion applications, SCFM is usually specified in the form
of lbs./hour. These guide lines are dangerous because a double correction on CFM (in the
case of NYC-6500 F with a 2.09 factor) would mean a selection that will produce 50% excess
air, depending on the fan's characteristics. Neglecting to correct is equally dangerous on
the short side.
"Assuming" ACFM, we would select for 8500 CFM at 20.9" SP. The BHP read from the multi-
rating table would then be a divided by 2.09.
-------
SUCTION CORRECTION FACTORS
7.
Another common influence on density, especially on exhaust systems, is suction. When resistance is
placed on a fan's inlet, the suction of the fan creates a partial vacuum at the inlet. This partial vacuum
lowers the density of the gas at the inlet. Since fans are rated with .075 density at the inlet, corrections may
be required. Since low pressure corrections are small, suction less than 10" is usually ignored.
7.
!
\
fc=
t
k
i
r
L
M
7.:
n
M
cC
ff
f^/iTfT
*
^:
16 1.
2J*—
S^ffff
^
" ' I '
17 1
^pf*^
18 1
:±r: r^:
$&
19
This graph must be used with temperature and altitude corrections unless conditions are equivalent to 70° SL.
The graph is used with appropriate factors from the altitude/temperature chart 6.00. For example, the
suction factor for a - 30" SP at the fan inlet, with an ambient condition equivalent to 1000' elevation is
1.083. Combine this factor with any temperature in the 1000' elevation column of 6.00 for the total
correction.
A. 1000 ',600° Fair, - 30" SP at fan inlet:
Factor = 2.07(1000' @ 600 ° F from 6.00) x 1.083(1000' with - 30" SP from 7.20) = 2.24.
B. 1372', 225° Fair, - 15" SP at fan inlet:
7.21
7.
(Density = .075/1.414 = .053. This is the Wichita, Kansas example in para. 2.10. Interpolation of
chart 6.00 was used for 1.36 factor.)
Now, your try the next two:
C. 0',350° Fair,-13" SP at fan inlet:
Factor = x =
D. 5200•', 400° F air, - 40" SP at fan inlet:
Factor = x =
i
Give yourself a pat on the back if you got factors of 1.57 and 2.24 for C and D. Note eyeball interpolation is
OK on 5200'.
-------
TEST YOURSELF
8.
Here is a little final exam for you to demonstrate how easy these common density corrections are to use.
The( ) refers to paragraph numbers in this course .. for explanation.
FIND:
Factor =
Inlet Density =
5332' elev.
70°F
#/k;=Factorfor5332'_and700 F is 1.21
1 =^.062 (6.00
EJECTOR FOR CONVEYING
SP-
8.10
FIND:
Factor =
Inlet Density =
f#2T Factor for 0' and 225° F is 1.30 (6.00K ^•'-^'-•.
plir-Factor for 0_' and 7" suction is 1.017 (7.20) or may be ignored.
jgOjotal factor is"l.30x1.017 =""l.32(7.21).~-f-77:;- V"-
teV-Density = .075/1.32 =..057 (6.00 footnote).' r^- ->."_.'
SPEC/F/ED:
I.D. Fan = 200,000 SCFM @ 8"
Gas Temp = 400°F
SP
QUESTION:
If I.D. Fan is handling air, what CFM & SP
would I.D. Fan be selected for? If BHP from rating
curve or table is 995, what is actual BHP?
:#3.;lf.air,495', 400° FFactor> 1.65(6.00). We don't know what portion of 8" is inlet suction but it would '-
^|^besomething~|ess than 8*." We can ignore it or assume a suction factor of 1.01 so that we can't be off
-^-'•' than;l%^200,000 SCFM x"1.65 x.1.01 = .333,300 ACFM (3.30). 8' x 1.65 x 1.01 _= 13.3" SP, ,,
^^NOTErWith gases'of combustion'ln a coal fired boller^the I.D.'fan would not handle "air«.^Typically-^s
of .078.This will be coveredjn course CBC-400 but for now~
"J078 inWdur^haiK'^f.cor/ectjpn factorsj Going back to ''Factojr.;=i.p75/d'j'/.S:r'.075/.078 =="^
^•-^^f*--l^^""^-^->^*----|»----'j'bemVde\fo^ ^
575 ^S^J^^^^^^^^M^'^-^^^M^^
If you have had difficulties, please write us or contact your local fan sales engineer. If you are interested in
_pther courses, write us.
i!ed in U S
CORPORATION
1675 GLEN ELLYN ROAD • GLENDALE HEIGHTS, ILL. 60137 • AREA CODE 312 858-2600 TELEX 72-1552
CHICAGO BLOWER (CANADA) • 901 REGENT AVENUE WEST, WINNIPEG, CANADA, R2C 2Z8
-------
WITH ROTATIONS AND
DISCHARGES DEFINED
Presented by
CHICAGO BLOWER CORPORATION
-------
A COURSE IN
WITH ROTATIONS AND
DISCHARGES DEFINED
^^^l^^^V&b^—**i> ^trfi^rv
^&@iP^iig£i£&
.l%i£if£is££«£ifei£&££lei
Course 300 is part of a series developed by Chicago Blower Corporation to assist
those interested in fan selection and engineering. While we do desire your business,
we have tried to keep these courses "generic" and applicable to the many good fan
manufacturers in our industry. Please write us if you have questions or critiques on
any material presented.
Course 300 covers the mechanical specifications of the selection. Because of the
subject material, this course will contain definitions and guidelines rather than
calculations. It is a "language course" which will help you communicate with your fan
vendor and will assist you in checking the best selection by wheel type and size. If you
want to go beyond this course, write us. We will be happy to send courses (at no
charge) as follows:
Course 100:
Course 200:
Course 300:
Course 400:
A BASIC COURSE IN FAN SELECTION.
How to select fan size and type from manufacturers'
catalogues.
A BASIC COURSE IN FAN DENSITY CORRECTION.
How to compensate for temperature other than 70°,
elevations above or below sea level and suction pressure.
A COURSE IN FAN ARRANGEMENTS AND CLASSES.
Illustrated definitions and guides to which Class and
Arrangement to use. (This course - if you would like
another copy).
ADVANCED COURSE IN FAN DENSITY CORRECTIONS.
How to correct density for gas other than dry air.
We also are developing courses in sound, fan curves and other topics. Availability of
these additional courses will be announced.
In addition to the fan size and types in Course 100, fans should be further described to
define the physical configuration of the equipment.
Copyright 1981, Chicago Blower Corporation
-------
SINGLE INLET-SINGLE WIDTH
DOUBLE INLET-DOUBLl
#fam"mmm "••"•• "-' —
II ft Centrifugal fans are either SISW or DIDW. This stands for Single Inlet-Single Width or
• A " Double Inlet-Double Width. Axial fans do not use width designations.
The Single Inlet-Single Width fan has air entry from one side. This is convenient for
attaching ductwork to the fan's inlet. SISW is more adaptable to construction
arrangements not having bearings in the airflow so they can be more suitable for
handling elevated temperatures or dirty air. On some mechanical draft installations,
inlet boxes are provided by the fan manufacturer. These rectangular or contoured
"boxes" are plenums welded to the inlet side of the fan. Boxes are used for the
attachment of rectangular ductwork and inlet damper control.
..^HSaM'IHlMMjn^^M.'CTgMEaiMMaKMMiaraa
S. J
-T-
,'""\
t
^^
1
" ~~
— — — -i
k J
-T-
,-'-,_
DIDW
In illustrations,
this indicates a
centrifugal fan's
discharge...
facing
you. (Wheels
omitted).
Double Inlet-Double Width fans can be thought of as having two single width wheels
mounted back-to- back on a common shaft in a single housing. Air enters both sides
of the fan. The DIDW fan is more suitable for many higher volume selections. DIDW is
less common in smaller sizes because the fan bearings are in the airstream and are
relatively large in proportion to the fan's inlet which can reduce performance. It is
more difficult to attach ductwork to DIDW fans but they are very suitable for "open
inlet" systems. In mechanical draft installations, inlet boxes are used with bearings
mounted outside of the boxes. This eliminates several disadvantages when ducting
into a DIDW fan. Unless boxes are used, the bearings in the airstream restricts the
DIDW fan to clean cool air.
ARRANGEMENTS
Fan arrangements are assigned numbers abbreviated A/# or Arrgt. #. The
_ _ arrangements describe the running gear placement in the fan. Over the years, some
1 O arrangements (such as 5 and 6) became unpopular and are now rarely used.
Commonly specified arrangements are shown below.
OR
A/1 Axial
A/1 Centrifugal
12 Centrifugal
Arrangement 1 • This is perhaps the most common arrangement for industrial and
other applications. It is available in SISW and usually belt driven. Two bearings are
mounted on a pedestal and the wheel is overhung to one side. The bearing pedestals
are internal on Axial fans.
Arrangement 2 • This arrangement is similar to Arrangement 1 except the bearing
2_ _ pedestal is supported by the fan housing. Two separate bearings may be used
• A A although at one time it was more common to see this arrangement with the 2 bearing
races in a common bearing housing. A/2 is usually belt driven.
A/3 SISW
Centrifugal
OR
A/3 Axial
A/3 DIDW
Centrifugal
OR
A/4 Centrifugal
A/4 Axial
Arrangement 3 - This arrangement is available in both SISW and DIDW. A bearing is
bracket mounted on each side of the housing ... or on axial fans, on each side of the
2,22 wheel. This results in a compact unit. Since one or both of the bearings are in the
airstream, A/3 is usually not used on any application where dirt and/or heat will run
through the fan. The bearing bracket supports can make it difficult to add ductwork to
the inlet or inlets of the fan. A/3 is usually belt driven.
Arrangement 4 • In this arrangement, the wheel is directly mounted on the motor's
shaft (& bearings). The fan itself does not have a shaft or bearings. This arrangement is
more common in axial or smaller centrifugal fans where proportions allow the motor
_ shaft to reach the wheel hub. However, some manufacturers offer A/4 centrifugals up
2.23 to ^OO HP as standard. Due to the close coupling of the motor, Arrgt. 4 fans are
normally restricted to a maximum temperature limit of 200°-F or less and it is common
to add some type of volume control to the fan since variable speed motors are often
not economically available.
-------
ARRANGEMENTS (continued)
A/7 Axial A/7 SISW Centrilugal
A/9 Axial A/9 Centrifugal
A/8 Centrifugal Belt Driven
CW
Arrangement 7 • An Arrangement 7 fan is an Arrangement 3 fan with a motor base
attached to the drive side. It is designed to be direct driven through a flexible coupling,
with the motor (or turbine) mounted on the attached base. The same cautions that
2 24 apply to Arran9ement 3 fans aPP'Vto Arrangement 7. A very practical use is in large
mechanical draft fans which use inlet boxes. In that case, it eliminates the need for
separate independent bearing pedestals which can simplify installation. Arrangment 7
is available in SISW and DIDW.
Arrangement 8 • This arrangement is similar to an Arrangement 1 fan. A smaller
"outrigger" motor or turbine base is customer provided, or shipped attached to the
bearing pedestal, for direct connection through a coupling. It is most often used
where V-belt drives are not appropriate, as with very high horsepowers.
Arrangement 9 • This is an Arrangement 1 belt driven fan with the motor mounted on
the fan rather than on the "floor1. It allows factory assembly of the motor and drives.
On higher horsepower fans, the industry sometimes refers to Arrangement "9H". In
this modification of Arrangement 9 the motor is mounted on the structural steel base
that is furnished by the fan manufacturer. It also allows factory assembly of the motor
and drives. See A/9H sketch in paragraph 6.20.
II^inTiTiiifWil"^/iV^"T'^T'^^M^TMr^
Arrangement 10 • This SISW arrangement is similar to an Arrangement 9 fan except
the motor is mounted inside of the bearing pedestal. This offers some degree of
weather protection to the motor however it restricts the motor size. It is most
important to provide adequate ventilation to the motor in A/10.
Technically, the definitions given for arrangements do not include all of the variations
allowed by standards published by the Air Movement and Control Association (see
"AMCA", para. 5.20). AMCA standards include the option of either belt drive or direct
drive for most arrangements. In this publication, rare and often impractical drive
options have been omitted.
^r _
2.3O
ROTATION
sSslPPSi^s^.^-;- —.-.^
It should be easy to leam the fan rotations ... because there are only two. A fan turns
either clockwise (CW) or counterclockwise (CCW). Sounds simple except errors are
3. 10 made because many of us tend to view the fan from wrong side. We tend to view them
from the inlet. The fan industry is contrary. It views everything from the "drive side",
which in itself, is a term that can be confusing.
3.20
CCW
The drive side is intended to mean the side that is driven by the motor, turbine, etc. On
SISW fans, the drive side is always the side that is opposite the fans inlet. On DIDW
fans the drive side is the side that has the drive. What do you do if there is the driver on
both sides of the fan? It is not uncommon on larger fans to have a turbine driving on
one side and a motor on the other. With dual drive, the drive side is the one that has
the highest horsepower driving unit. What do you do if both units have the same
horsepower? Draw a picture! On axial fans, users normally do not specify rotation but
rotation must be observed when wiring the motors . Watch out for the small single
phase motors or large TEFC motors which are built to run in only one rotation
direction.
3.30
Someday we may have to define clockwise and counterclockwise in a world of digital
watches!
-------
DISCHARGE
Fan "discharge" is always important because the fan must be aimed in the proper direction for connection to the ductwork.
Like rotation, fan discharge is always viewed from the drive side of the fan. (Referto paragraph 3.20, drive side). Drawings of
the various discharges are somewhat self-explanatory.
CW FANS . . CCW FANS
4.10
Top Horizontal Top Angular Down Down Blast Bottom Angular Down
Bottom Horizontal Botto
Up Blast Top Angular Up
Top Horizontal Top Angular Up Up Blast Bottom Angular Up
Bottom Horizontal Bottom Angular Down Down Blast Top Angular Down
Note: CCW Bottom Horizontal illustration modified to
show discharge of the popular "square" or"cube" fans.
Shape of scroll is concealed by sides. Not available in
angular discharges.
2 The discharge positions are typically abbreviated and the abbreviations may end in "D" for "discharge". For example, top
411 horizontal discharge can be shown as TH orTHD. If you mount a fan on a wall or ceiling, the discharge does not change. It is
always viewed as if the fan was sitting on the floor.
4.
With "Angular" discharges, the tilt or angle does not have to be 45° as shown in the sketches. If available from the
manu'acturer' other angles can be specified. The alternate angles are commonly measured either above or below the
horizontal centerline (angle 6 in sketches). New standards are being considered that will change the reference. If adopted,
angles will be measured from the upper vertical centerline, in the direction of rotation (angle ff in sketches).
Under the current system, an angular fan without specified degrees is assumed to be at 45°. Other angles must be called out,
4.2 1 such as 30°TAU'27% ° BAD etc-To avoid errors.or conflict if standards are changed, it is always best to call out the angles
and their reference line, such as "30° above horizontal centerline TAU."
Axial fans do not have the complications of discharge direction. Incoming or Outgoing ducts are "angled" to suit the straight
axial fan tube.
For economic reasons, many fan manufacturers are offering standardized discharge positions only. In these cases, it is
30 usua"v less cost'y f°r the user *° reroute ductwork than to buy a special-built fan. A broader selection of discharges is
available in large or heavy duty fans where ductwork modifications are less practical.
Almost all manufacturers offer fans in different classes. "Class" by itself, means different things to different
5« 10 manufacturers. The most useful meaning of class is that it defines the RPM limits of the fan. "AMCA Class" does
have definite meaning which also relates to RPM limits.
Classes are designated by number. .. 1, 2, 3 etc. (Class is usually shown in roman numerals eg. Class I, Class II... or
5.11 C/l' C/" etc''- Eacn higher number represents higher RPM (and thus air performance) capabilities of the fan. Higher
_ RPM capability may mean heavier or stronger construction.
-------
CLASS (continued)
5.20
AMCA Class: Most fan manufacturers comply, in some way, with the industry standards published by AMCA, the Air
Movement and Control Association Inc. AMCA has standards for "class" Those standards are set for the ventilation
markets and cover only Forward Curved, Backward Inclined, Airfoil Blade and Inline Centrifugal fans. Manufacturers'
m class designations for other fan types such as Radial Blad
'i uniform standard.
5
I For certain wheel types, AMCA publishes
"'.. curves (Standard 2408-69 available from AMCA,
r 30 W. University Drive, Arlington Heights, IL
j 60004). Each curve represents a segment of a
\ typical fan performance curve.
• Pressure-Velocity ranges were chosen
: somewhat from fan industry standards
j (pre-1969) and plotted to closely follow RPM
Jl changes of the typical fan.
u 1
n
^ Within a percent or two, AMCA says that the
:1 theoretical class 2 (C/2 or C/ll) fan would run
'<:• 25% faster than class 1 and a class 3 fan would
« run 30% faster than class 2. They do not show
•"- classes beyond class 3.
f:
14
12
FIG PRESSURE
<• t
<
tn 6
4
2
^^*
es, Vaneaxials, etc. do not necessarily conform to any
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1000 2000
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3000 4000
OUTLET VELOCITY
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OPERATING LIMITS FOR SISW CENTRIFUGAL FANS
\
AJ
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^•
__.
6000
EXAMPLE
5.:
When a fan is manufactured with standard steel gauges and readily available shaft and bearing sizes, its maximum
speed capability will rarely coincide with the AMCA standard limit. For this reason, most manufacturers now publish
their own limits (which exceed the AMCA standards).
SBSta
i^-^^
Fan manufacturers have extended the class concept to higher classes and to fan types not covered by AMCA. For
5.30 example, Chicago Blower markets a "C/IV" Radial Blade (D/16A) fan which run faster than its C/lll. . . their C/lll runs
faster than their C/ll. . . etc. AMCA doesn't have a standard for C/IV or Radial Blade fans.
Which class to use? Vendor "X" may have a C/lll fan with larger shaft and bearings than vendor "Y" but "Y" may have
Q 1 a heavier duty bearing and stronger alloy wheel. It's hard to tell which is better for you. Unless you are thoroughly
familiarwith all of the construction differences between manufacturers or within one manufacturer, class selection
is best left up to the vendor.
Perhaps the most valid reason to specify "class" is to get the longest trouble free life from your selection. A guide to
C 4O f°"ow 's the vendor's published RPM and HP limit. The further below the maximum conditions you are, the safer (but
possibly more expensive) your selection.
5.50
C
M
1
At the pressures for which you normally self-select, housing gauges are not important. The housing must be built to
withstand its own weight and handling during shipment. These forces are higher than most static pressures. Wheel
gauges are important but without access to the manufacturers' "finite element" analysis data (or maybe
understanding of it), comparing wheel gauges is not really helpful. You are right if this sounds like the old "trust me"
line and a better idea is to ask for references. Are other users satisfied with the vendor? If so, the vendor is most
likely reliable in all product designs IF the product is part of his standard line.
Manufacturers grow by extending their lines. A vendor of ventilation equipment will stretch ventilation concepts into
heavy duty equipment. This results in an inexpensive (to buy) fan that is somewhat heavier than a ventilating fan. The
heavy duty manufacturers will stretch heavy duty concepts into ventilating equipment and sell an expensive fan that
is a little lighter than a heavy duty fan. The idea is to look at your needs and compare them with the standard product
line of your vendor. Rely on his choice of class. 5
-------
"^'^^ftDDITICWAllDEtAILlS
10
Motor positions for belt driven centrifugal fan
Positions W &Z are the most compact IF motor will fit close
to fan. With the current trend to fans on unitary bases (Arrgt.
9H), this is more common than ever. Positions X & Y will
typically take more space.
1
1
r
X
i
The first sketch shows the motor close to the fan shaft but it
may be more practical to move it out so that a standard
6 20 adJus*able motor slide base will give a greater adjustment
" fg range for belt tension. Longer belt centers (C.D.) also result
in higher horsepower per V-belt ratings.
C.D. = Center Distance
Arrangement 9 motor base positions:
A/9R
Arrgt. 9. ..
Motor on right-
A/9T
Arrgt. 9. ..
Motor on top
A/9L
Arrgt. 9...
Motor on left-
on bearing pedestal
A/9SL
Arrgt. 9 . . .'
Motor on side
(left)
A/9H
(typical)
A/9SR
Arrgt. 9...
Motor on side
(right)
Note again that
RIGHT & LEFT
are viewed from
the drive side.
""-"•on fan housing
Axial and Inline Centrifugal
A/1
A/1L
Arrgt. 1 ...
7.20 Motor on 'eft
A/1R
Arrgt. 1 ...
Motor on right
Alpha designations are AMCA Std. 2410-66
for Inline Centrifugal Fans. Alternate
designations are o'clock positions.
D = 4:30
E = 6:00
F = 7:30
G = 9:00
H = 10:30
Drive or outlet end
That covers most-of the "configuration" definitions in fan engineering. Purposely we have left
out detail on inlet boxes and other specifics that relate to custom heavy duty fans.
Q | Q You should enlist the help of professional fan engineers in heavy duty applications.
Remember, if you are interested in other courses, write us.
-------
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