Flow Fans
Flow Fans
Flow Fans
Page
FOREWORD ........................................................................................................................ 3
FIGURES ............................................................................................................................. 3
NOMENCLATURE ............................................................................................................. 4
REFERENCES ..................................................................................................................... 34
DEDICATION ...................................................................................................................... 34
Hudson Products Corp. Page 2 of 35 The Basics of Axial Flow Fans
FOREWORD
This manual is designed to familiarize users It should be noted that final fan selection should
with applications for axial flow fans, be made by using Hudson’s Tuf-Lite®‚ Fan
velocity recovery stacks, seal discs, and Selection Program or by contacting Hudson
variable flow fans. Calculations are provided Products Corporation at 713-914-5700 or
for estimating fan power consumption and 1-800-634-9160.
noise.
FIGURES
Page
1. Typical Air-Cooled Heat Exchanger Fan .................................................................. 6
2. Typical Fan Curve .................................................................................................... 12
3. Approximate HP Savings with Velocity Recovery Stack. ........................................ 14
4. Stall Points on Fan Curve ......................................................................................... 14
5. Tip Vortex (Leakage) ................................................................................................ 15
6. Tip Clearance ............................................................................................................ 15
7. Airflow with No Inlet Bell ....................................................................................... 16
8. Airflow with Inlet Bell ............................................................................................. 16
9. Typical Large Cooling Tower Fan with VR Stack.................................................... 17
10. Approximate Cost of Power to Operate 8400 hrs/yr. ............................................... 17
11. ASP vs. CFM for Air-Cooled Heat Exchanger ......................................................... 18
12. Velocity Profiles - Tapered vs. Constant Chord Blade ............................................. 22
13. Typical Seal Disc Prior to Installation ...................................................................... 23
14. Typical Auto-Variable® Fan ...................................................................................... 23
15. Inverted Actuator-Variable Pitch Fan ....................................................................... 23
16. AV Force Diagram .................................................................................................... 24
17. Airflow vs. Signal for Variable-Pitch Hub ............................................................... 24
18. Typical Auto-Variable® Hub ..................................................................................... 25
19. Throttled Flow Through Fan .................................................................................... 27
20. Air Delivery vs. Temperature ................................................................................... 28
21. Power Study Results ................................................................................................. 28
22. Weighted Sound Levels ............................................................................................ 30
ρ = density
Subscripts
1 = inlet, condition 1
2 = outlet, condition 2
in = inlet
out = outlet
SL = sea level
des = design
eff = effective
An axial flow fan moves air or gas parallel ACFM - Actual cubic feet per minute of air
to the axis of rotation. By comparison, a moved by the fan.
centrifugal or radial flow fan moves air
perpendicular to the axis of rotation. Axial Actual Conditions - Resistances related to
flow fans are better suited for low-resistance, actual inlet or outlet temperature and fan
high-flow applications, whereas centrifugal elevation above mean sea level compared to
flow fans apply to high-pressure resistance, standard conditions.
low-flow conditions. Typically, the types of
fans discussed in this manual can handle Air Density - Air density at the plane of the
“resistances” up to approximately 1 in. of fan based on standard or actual conditions.
water. Axial fans can have widely varied
operating characteristics depending on blade Beam Passing Frequency - Number of times
width and shape, number of blades and per revolution that one fan blade passes over
tip speed. a beam or strut thought of as “how the
structure interacts with the fan blade” ex-
The most common type of fan for air-cooled pressed in cycles/sec (Hz).
heat exchanger (ACHE) is less than 14 ft.
diameter and has four blades. The most Blade Natural Frequency - Frequency at
common type for wet cooling towers is 28 ft which a blade freely vibrates when it is struck
diameter and has eight blades. A typical in cycles/sec (Hz).
ACHE fan is shown in Fig. 1.
Blade Passing Frequency - Number of times
per revolution that a fan tip passes a point on
the fan ring expressed in cycles/sec (Hz)
thought of as “how the fan interacts with the
structure”.
I-P: ft2
• PRESSURE metric: m2
SI: m2
I-P: inches of water (in. - H2O)
metric: mm of water (mm - H2O)
SI: Pascals (N/m2)
• POWER
I-P: HP (horsepower)
metric: kW (kilowatt)
SI: kW (kilowatt)
9.806 * mm-H2O = Pa
35.314 * m3 = ft3
0.02832 * ft3 = m3
• VIBRATION FREQUENCY
(oC * 1.8) + 32 = oF
2.2 Useful Conversion Factors
(oF - 32) / 1.8 = oC
=
• AIRFLOW
Many parameters are important in selecting (a) Confirm that the operating point is not
a fan to meet the specified operating condi- close to a “stall” condition. (Explained
tions such as: in detail later).
Fan diameter - to suit bay or cell size, Typically, a 2° pitch angle safety
margin is desired before reaching a stall
Hudson blade type: “H”, “HW”, “B”, point on the curve or overloading the
“C”, “D”, “K”, motor. The American Petroleum
Institute has specific limitations that
ACFM - Airflow required for design heat must be met for ACHE applications.
transfer duty, (b) Confirm that the fan brake horsepower
(BHP) is low enough that when
ASP - Actual static pressure from the sum environmental, drive and motor losses
of all resistances to airflow, are added, the installed motor HP is
not exceeded.
Air temperature at plane of fan,
(c) Be sure that the “horsepower per blade”
Fan elevation above mean sea level or mechanical limit (for Hudson fans) is not
air density at fan, and exceeded or select a fan with additional
blades.
Fan speed in either tip (FPM) or rotational
speed (RPM). (d) Calculate the blade passing frequency,
and compare it to the first mode
Also, a velocity recovery stack installed? Is resonance frequency for the selected
there a noise limitation? How much allow- blade (for Hudson fans).
able horsepower can the fan use? Will the
fan avoid resonant frequency problems? (e) Repeat step (d) for the beam passing
frequency. (Most structures have four
In the most simple case, design airflow, static main beams.)
pressure, and density are calculated for the (f) Repeat step (d) for 1x harmonic of
fan “curve” conditions at 12,000 FPM tip fan RPM, and compare to first mode
speed and standard density with no noise resonant frequency.
limitations. The operating point is then
plotted on a fan curve of the appropriate (g) Check resulting noise level against any
diameter. noise limit specifications. Noise level
is most often specified as sound
The fan curve shows the correct pitch angle pressure level (SPL) at a given distance
and horsepower required by the fan at the from the tower or ACHE. Convert the
design point. SPL to a sound power level (PWL) to
evaluate the selected fan. Some adjust-
ments to tip speed and/or the number of
blades may be required.
2
SP = ASP * 1 / DR • Velocity Pressure = V for std air
4005
Note: Hudson fan curves are based on 2
ρ2
AVP =
V
standard conditions and 12,000 FPM tip
4005
speed. If operating conditions are given in ρ1
actual conditions, they must be converted
to standard conditions before referring to the
where ρ = density at the fan
fan curve to determine performance. 2
ρ = std air density (0.075 lb/ft3)
1
required ACFM.
Horsepower
10˚
6˚
• Basic fan laws: 2˚
CFM = fn(RPM)1
Airflow varies in direct proportion to 10˚
Total Pressure
RPM 6˚
in.-H2O
2˚
Pitch
Angle
SP or TP = fn(RPM)2 Velocity
Pressure
up to 14 ft dia. 4 blades
16 - 20 ft 6 blades
24 - 30 ft 8 blades
Stall Points on Fan Curve
Fig. 4
36 - 40 ft 8 blades
3. Horsepower ranges:
Allow a 2° pitch angle margin of safety
before stall for a conservative fan 6-10 ft dia 7.5 - 15 HP
operating point. 12 ft 20 - 30 HP
14 ft 30 - 60 HP
16 - 20 ft 50 - 100 HP
24 - 30 ft 100 - 250 HP
36 - 40 ft 150 - 300 HP
Hudson Products Corp. Page 14 of 35 The Basics of Axial Flow Fans
4. Temperatures at the fan (Farenheit):
5. Fan stacks:
6.1 Effect of Tip Clearance Fig. 6 shows the API recommended tip
clearances for smaller (16ft and less)
Tip clearance between the blade and diameter fans. Also included in this
inside of the fan cylinder or fan ring is figure are the nominal tip clearances for
critical to proper axial flow fan perfor- larger diameter fans.
mance. Hudson fan curves are derived
from tests conducted at Texas A&M
University’s Engineering Laboratory, Fan Diameter Minimum Maximum
based on very close tip clearances with
3ft through 9ft 1/4 in. 1/2 in.
optimum inlet conditions. If close tip
>9ft through 11ft 1/4 in. 5/8 in.
clearances are not maintained, a loss of >11ft through 16ft 1/4 in. 3/4 in.
performance will be noted, in pressure 18ft through 40ft 1/2 in. 1 in.
capability and airflow. The primary
reason for close tip clearance is to mini- Tip Clearance
mize air loss or leakage around the tip, Fig. 6
known as a tip vortex. (See Fig. 5).
Hudson recommends the use of safety
factors to correct for the environment in
which a fan is placed.
kW HP
149 200
131 175
112 150 $
Typical Large Cooling Tower Fan .04 .06 .08 .10 kW-HR
with VR Stack 93 125
Fig. 9 75 100
56 75
37 50
Consider the following fan:
19 25
28 ft diameter, 8 blades, Type H 0
14 ft stack, 9 ft venturi height 0 20 40 60 80 100 120 140
1,180,000 ACFM Cost (Thousands of $)
0.38 in. - H20 ASP Approximate Cost of Power
136 RPM to Operate 8400 hrs/yr
105° F, at sea level Fig. 10
Use r = 0.10D inlet condition
30
A 25% increase in airflow yields
250,000 ACFM. The new ASP 20
increases to 1.49x0.33 or 0.49 in. - H20.
10
Results from the Tuf-Lite program
0
show a 14H - 4 fan is still a choice, but
0 5 10 15 20 25 30
33.9 HP is now required to produce the % Increase ACFM
250,000 ACFM. A 40 HP motor
would be required instead of the ASP vs. CFM for ACHE
original 20 HP. Fig. 11
• Airfoil twist
6.9 Effect of Blade Shape
• Tangential velocity squared
Airflow across the plane of the fan is
not uniform varying from positive at At mid radius (0.25D), tangential
the tip to negative at the center of the velocity is only 25% of the velocity at
fan. Blade shape and twist of the air- the tip. To compensate for this
foil along the blade affects the shape decrease in velocity, the chord width
of the velocity profile. and twist must be increased. This is
the reason for the increased efficiency
Velocity profile of a well-designed and more uniform airflow from a
tapered blade with a generous twist tapered blade. Note that for a constant
compared to a constant chord blade chord width blade, exit velocity
with minimal twist is shown in Fig. 12. decreases rapidly inboard of the tip and
typically becomes negative outboard
of the seal disc. A typical aluminum
blade’s chord width and twist do not
vary along the blade.
Note that the airflow direction becomes Variable-pitch or Auto-Variable fans auto-
negative near the center of the fan. This matically adjust the pitch angle to provide
is the result of the torque applied to the the precise amount of airflow for controlling
fan which creates a “swirl” process temperature while saving substantial
effect on the air vectors. Air at the tips amounts of energy. This section discusses
is axial or parallel to the fan axis. basic operating characteristics, process
Toward the fan’s center, air velocity temperature control methods, energy savings
decreases by the square of the radius and economic comparisons with other axial
and the air vectors lean further toward fan airflow control systems.
the horizontal. At or before the center
of the fan, the air is actually moving Variable-pitch applications are typically
opposite to the airflow at the blade tips. 10 -14 ft (3.04 - 4.3m) in diameter consum-
A well-designed fan will have a center ing up to 40 horsepower (30 kW). Fig. 14
seal disc of about 25% of the fan’s shows the typical type actuator system found
diameter. Referring to Fig. 12, note that in ACHEs. Fig. 15 shows the inverted
the air vectors at the center of the fan actuator system popular in Europe for the
actually turn to a negative direction. ease of access to the actuator.
Tests have shown that the seal disc pre-
vents this negative airflow and
improves fan efficiency by about
4 - 5%. The seal disc can be as large as
14 ft in diameter for 40 ft fans. Fig. 13
shows a typical seal disc.
60
Force
Spring
40
Rotation
20
AV Force Diagram
0
Fig. 16
Signal-(psi)
-20
Negative
As the blade moves the air, the aerodynamic
moment acts to turn the blade to a lower pitch. -40
Louvers are the first step to modulated The variable pitch fan can also provide from
airflow; however, fan horsepower is wasted 0 - 100% positive or from 0 - 60% negative
as airflow is throttled by the louver. At airflow at the same horsepower. Negative
complete shut off, the fan stalls and horse- airflow is useful, along with louvers, in
power actually increases. winterized ACHEs to seal off freezing out-
side air and recirculate warm air inside the
Variable-speed fans for fully modulated plenum chambers. Internal recirculation
airflow are available in two types: hydrau- systems utilize a positive and negative
lic and electric drive. Either type conserves airflow pair of fans to recirculate warm air.
energy and offers good airflow control. The
latest development in electrical variable-
speed control for fans is the variable- Power Consumption and Cost Comparison:
frequency drive (VFD). There are three ba- When evaluating power consumed by any
sic types: VVI (Variable Voltage Inverter), axial fan operating at a fixed pitch, consider
PWM (Pulse Width Modulation) and CS not only the fan performance curve but the
(Current Source). These drives utilize a system resistance characteristics as well.
standard induction motor and automatic
control is obtained by a process control Fan performance curves are obtained in a
device to interface the 4 - 20 milliampere wind tunnel with a means of varying the
temperature controller output with the VFD. resistance or static pressure head against
which the fan must work, and measuring
The older less common hydraulic drive resulting airflow and horsepower.
system consists of a motor/variable volume
pump/reservoir unit connected to a slow- System losses are more difficult to determine,
speed, high-torque, direct-drive motor. and judgement is necessary in many cases
Advantages are variable fan speed and to evaluate the losses due to poor inlet
eliminating a reduction belt or gear drive. conditions, excessive tip clearance and
The main disadvantage is inferior system unusual structural conditions. If a cooling
drive efficiency. Efficiencies are approxi- service is critical, it may be wise to model
mately: motor 0.97, pump 0.92 and the unit and establish the system losses in a
hydraulic motor 0.92. Therefore, the wind tunnel. What must be determined is
optimum drive system efficiency is the sum of the static pressure resistances
(0.97)*(0.92)*(0.92) = 0.82, not counting versus system airflow. Since the fan’s air de-
hydraulic line loses. This must be compared livery characteristics have been accurately es-
to a typical motor/belt drive efficiency of tablished, the operating point for any
0.95 or 0.97 with a gear or timing belts. airflow requirement can be predicted. This
operating point is where the fan output in
Advantages of VFDs are reduced noise and terms of pressure and flow, exactly meet
vibration during slow-speed operation. system requirements and is referred to as the
Disadvantages are high cost per horsepower equilibrium point of operation.
for control of a small number of fans in the
Hudson Products Corp. Page 26 of 35 The Basics of Axial Flow Fans
Let’s examine a typical case of an ACHE with a Energy Comparison:
fixed-pitch fan and a louver for throttling the
airflow at the design point and several other A typical case study to evaluate cost
reduced airflow operating points. differentials between fixed-pitch (single-or
dual-speed) or variable-pitch fans for
For example, consider a 14 ft diameter fan op- air-cooled heat exchangers is conducted
erating at a 14° design pitch at point 1 (Fig. as follows:
19). The system resistance line, shown as the
dashed line, is the locus of points Step 1. Thermal studies are needed
obtained by summing the static and velocity to determine total airflow required
pressure losses vs. flow through the bundle. The as a function of ambient tempera-
pitch angle line for 14° represents the fan’s ture for the particular application.
delivery characteristics. Assume density Plot flow versus temperature.
remains constant and the decreased airflow is a
result of a reduction in required airflow. Step 2. Obtain climatological data for the
area where the air cooler will be
If the louver throttles airflow to control the located, and derive a table of
outlet temperature, the points shown as 1, 21, “degree-hours” for incremental
31 and 41 in Fig. 19 are total pressure output temperature ranges. This tabulates
and horsepower consumed by the fixed-pitch the number of hours per year that
fan. If we were controlling airflow with a each temperature range occurs.
variable-pitch fan instead of a throttling
device, total pressure output would exactly Step 3. For a fixed-pitch fan, flow is a
match the system resistance line as shown at direct function of speed. Create a
points 2, 3 and 4, and the horsepower require- plot of fan output versus ambient
ments are significantly reduced. In this case, temperature for each case to be
the variable-pitch fan used 26, 51, and 73% studied. This yields HP-hours per
of the horsepower required by the fixed-pitch temperature range.
fan at points, 21, 31 and 41 respectively.
Once the airflow is defined for each range,
fan horsepower for any point can be approxi-
Fixed-Pitch HP
mated by the following relation:
21 1
Horsepower
60 41 31
2
40
20 3 cfm1 2.8 ρ1
Where: HP1 = HPdes
4
Variable-Pitch HP
cfmdes ρdes
Total Pressure, IN.-H2O
1.2
HP1 = Horsepower at cfm1
ρ1
1
4
1.0 31
21 = Density at point 1 lb/ ft 3
.8 1
.6 14˚
cfm1 = ft 3/min flow at point 1
2
.4 3
4 2˚ 6˚
10˚ System
Resistance
HPdes = Design horsepower
.2
325 Line
ρdes = Design density
150 175 200 225 250 275 300
1. Two fixed-pitch fans. Fig. 21 gives the power study results which
show that the variable-pitch fan would pay
2. Same, but with one 1800/900 RPM motor. for itself within two years. In this particular
application, the two-speed motor in Case 2
3. One fixed-, one variable-pitch fan (single- made only a small difference in energy cost.
speed).
-5
C B and C
-10
Octave Bands: Noise is categorized by -15
dividing it into separate frequency bands -20
B
Frequency Responses for
of octaves or 1/3 octaves. Generally, we -25 Weighting Characteristics
use 63, 125, 250, 500, 1K, 2K, 4K and -30
8K center frequencies to define these -35 A
bands in Hertz (cycles/sec). -40
-45
-50
Sound Power Level: Acoustical power 20 50
200
100
000
0
0
000
0
500
500
200
100
Monroe, Robert, CTI Paper TP74-03, “Cooling Tower Fans - Today and Tomorrow”,
Jan. 1974
Monroe, Robert, CTI Paper TP97-12, “Maximizing Fan Performance”,
Feb. 1997
API Standard 661, Fourth Edition, “Air-Cooled Heat Exchangers for General Refinery
Services”, Nov. 1997
API Standard 631M, First Edition, “Measurement of Noise from Air-Cooled Heat
Exchangers”, June 1981
DEDICATION
The Basics of Axial Flow Fans is a compilation of work performed by many Hudson
Products’ engineers and employees. However, one individual deserves special recognition.
Mr. Robert C. Monroe, who worked tirelessly for Hudson from January 28, 1967 to his
retirement on August 1, 1997, was not only a major contributor to the contents of this book but
also to the success of Hudson Products Corporation in manufacturing quality fiberglass-
reinforced, plastic fan blades. Bob’s expertise, vision and hard work were
® ®
instrumental in Hudson’s development of the Tuf-Lite and Tuf-Lite II families of superior
fan blades.
The Basics of Axial Flow Fans is dedicated to the memory of Mr. Monroe who passed away on
May 31, 1998.
Hudson, Auto-Variable, Combin-Aire, Exact-A-Pitch, Fin-Fan, Heatflo, Hy-Fin, Solo Aire, Split-Flo, Stac-Flo, Steamflo, Thermflo, Tuf-Edge, Tuf-Lite
and Tuf-Lite II are registered trademarks of Hudson Products Corporation.