CEMENT COURSE

FAN MANUAL FOR THE CEMENT

INDUSTRY

W. Zeller / P. Hunziker

PT 97/14220/E
November 1997

"HOLDERBANK“
Management and Consulting Ltd.
Process Technology

C:\TEMP\Wz14220e.doc The copyrights for this document and all appendices are
reserved by HMC. The right to reproduce them entirely or in
part in any form and to make them available to third parties is
subject to the authorization of HMC.
 Contents

1. FAN APPLICATIONS IN THE CEMENT INDUSTRY..........................................................1

2. SELECTION OF FAN IMPELLER ........................................................................................2

2.1 Fan impeller types .............................................................................................................2

2.2 Selection criteria................................................................................................................3

2.2.1 Fan efficiency............................................................................................................3
2.2.2 Continuous operation................................................................................................5
2.2.3 Mechanical design ....................................................................................................5

3. WEAR PROTECTION............................................................................................................7

3.1 Erosion types.....................................................................................................................7

3.2 Protection of parts subjected to abrasion .........................................................................7

3.2.1 Direct protection .......................................................................................................7
3.2.2 Protection by wear plates .........................................................................................7

3.3 Deflection of abrasive particles .........................................................................................8

3.3.1 Deflector plates on impeller (Fig. 3.1) ......................................................................8

4. FAN IMPELLER ARRANGEMENT ....................................................................................10

4.1 Overhung assembly (Fig. 4.1).........................................................................................10

4.2 Center hung assembly (Fig. 4.2).....................................................................................10

5. FOUNDATIONS ...................................................................................................................12

6. FAN PERFORMANCE .........................................................................................................13

6.1 Fan Capacity Adjustment ................................................................................................14

6.1.1 Fan capacity too low ...............................................................................................14
6.1.2 Fan capacity too high..............................................................................................15

6.2 Adjusting fan performance curves ..................................................................................16

7. FLOW CONTROL................................................................................................................19

7.1 Damper control (Fig. 7.1 & 7.2).......................................................................................19

7.2 Radial inlet vane (Fig. 7.3)...............................................................................................19

7.3 Speed control (Fig. 7.4)...................................................................................................19

 7.3.1 Hydraulic transmission with fixed speed motor......................................................19
7.3.2 Speed-controlled electric motors ...........................................................................20

8. START-UP OF FANS...........................................................................................................26

9. POSSIBLE PROBLEMS WITH FANS................................................................................27

9.1 Vibration...........................................................................................................................27
9.1.1 Material build-up......................................................................................................30
9.1.2 Variable speed operation ........................................................................................33
9.1.3 Thermal effects.......................................................................................................33
9.1.4 Hot shutdowns ........................................................................................................33

9.2 Erosion ............................................................................................................................34

9.2.1 Improper duct connection.......................................................................................34
9.2.2 Effects of impeller speed and wheel inlet velocity..................................................34
9.2.3 Liner materials ........................................................................................................35

9.3 Bearings ..........................................................................................................................36

10. TROUBLESHOOTING......................................................................................................37

Note: This Report replaces the following previous Reports No.: VA 89/5657/E
VA 89/5657/S
VA 91/5844/E
VA 93/4030/E
VA 94/4242/E
VA 95/4359/E
 -1-

1. FAN APPLICATIONS IN THE CEMENT INDUSTRY

Dust Blades Max. max. Flow Rotor Stator Effi-

load mainly Temp speed regulation protection protection ciency
Location
[g/Nm3] used [°C] [rpm]
[%]
Coal Mill
Filter exhaust < 0.15 F/C/A 150 1800 VC/ILD (WP) - 70 - 85
Separator / cyclone ex- < 100 R/F/C 150 1200 ILD HSWP WP( HSWP) 55 - 75
haust
Raw Mill
Filter exhaust < 0.15 F/C/A* 300 1800 VC/ILD/VS - - 70 - 85
Separator / cyclone ex- < 100 F/C 300 1200 ILD/VS (HW/HWSP) WP 65 - 75
haust
Cement Mill
Filter exhaust < 0.15 F/C/A 200 1800 VC/ILD - - 75 - 85
Separator / cyclone ex- < 100 F/C 200 1200 ILD (WP/HSWP) WP 65 - 75
haust
Dry Process
Preheater exhaust < 100 R/F/C 450 1200 ILD/VS (WP) - 55 - 75
Kiln line filter exhaust < 0.15 F/C/A* 350 1200 VC/ILD/VS - - 70 - 85
Semi-dry process
Recirculation fan < 20 R/F 450 750 ILD/VS HSWP WP 60 - 70
Kiln line filter exhaust <0.15 F/C/A* 450 1800 VC/ILD/VS - - 70 - 85
Cooler
Filter exhaust <0.15 F/C 450 1200 VC/ILD/VS (WP) - 70 - 88
Filter exhaust 0.5 F/C 450 1200 ILD/VS WP (WP) 70 - 80
Cyclone exhaust <5 R/F 450 1200 ILD/VS HSWP WP 65 - 75
Recirculating < 15 R 450 750 ILD HSWP WP(HSWP) 60 - 70
Cooler F.D. fans Traces F/C/A 50 2200 VC/VS - - 75 - 85

* in case of bag filter

Abbreviations:
a. Blade Type
R: radial blade
F: backward inclined flat plate
C: backward curved plate
A: backward airfoil blade

b. Flow regulation
VC: Vane control set at the fan inlet
ILD: Inlet louver damper set on the inlet box
VS: Variable speed

c. Rotor and stator protection against abrasion

WP: bolted or welded wear plates
- mild steel
- quenched steel
- wear-resistant steel
HSWP: bolted mild steel wear plates with hard surfacing
- chromium carbide
- tungsten carbide

(WP/HSWP): optional
 -2-

2. SELECTION OF FAN IMPELLER

2.1 Fan impeller types

radial blade
Ø good reduced tip width capabilities
Ø static efficiency 60 - 75 %
Ø high tip speed capabilities
Ø reasonable running clearance

Ø best for high dust loads > 100 g/m 3

Ø most resistant to build-up
Ø minimal build-up problems
Ø high wear resistance

backward inclined
Ø limited reduced tip width capabilities
Ø static efficiency to 80 %
Ø low to medium tip speed capabilities
Ø running clearance important to % -ES

Ø dust load up to 100 g/m 3

backward curved
Ø medium to high tip speed capabilities
Ø static efficiency to 82 %

Ø dust loading capability < 100 g/m 3

airfoil
Ø limited reduced tip width capabilities
Ø static efficiency to > 84 %
Ø medium to high tip speed capabilities
Ø running clearance inlet important to % - ES

Ø for clean air applications

 -3-

2.2 Selection criteria

It is of great importance that critical process equipment, such as fans, be selected on the
basis of proven ability to provide dependable service in the system rather than on an
efficiency rating.
There are four basic blade forms used in industrial induced draft service. They are
Radial, Backward Inclined, Backward Curved, and Airfoil.
In situations where more than one blade form will meet a performance requirement, it
then becomes necessary to select the one form that will be most overall cost-effective.
For the selection process the supplier should provide the operating and service manual
for the equipment type being considered.
The operating and service report should include all occurrences that require fan
maintenance.
To assist in fan type selection, there are at least four important points to be considered.
Ø fan efficiency
Ø continuous operation
Ø mechanical design
Ø equipment costs

2.2.1 Fan efficiency

Because many of the higher efficiency fans can only achieve their stated efficiency within
a narrow operating range, a true energy evaluation must consider the actual operating
point and alternate operating points on a time basis.
Many systems include a built-in safety factor, which results in reduced efficiency when
operated at constant speed with damper regulation.
With higher efficiency equipment, clearance between wheel and inlet collar and shape
are significant factors. To maintain the high efficiency, these clearances must be kept
very precise, which becomes difficult when thermal expansion is involved. If additional
clearance is allowed for thermal expansion, the efficiency is adversely affected.
In the airfoil design, the leading edge contour is extremely important in maintaining
optimum efficiency, to the point that material build-up or the addition of blade liners can
have a significant adverse effect on efficiency. It should also be noted that, as a blade is
tilted back from radial, the possibility of build-up on the blade usually increases.
Efficiency is also affected by the number of blades on a wheel. For a given design, there
is an optimum number of blades based on efficiency, but for structural purposes (high
pressure, high tip speed applications), additional blades are often required. This, in turn,
often reduces fan efficiency.
For direct connected fans, only one wheel diameter can be selected for a given pressure
requirement and blade form. As a result, to meet the corresponding air flow
requirements, it may be necessary to reduce the standard tip width of the wheel.
Reduced tip width significantly affects fan efficiency in some blade forms, but is more
tolerable in others. The Radial Blade fan has proven to be the most tolerable one.

The following table shows the approximate allowable percentage of tip width reduction
without significantly changing efficiency characteristics.
 -4-

Radial Blade 50 %
Backward inclined 25 %
Airfoil 10 %

Because of reduced tip width capabilities, a manufacturer often cannot substitute a single
inlet fan by a double inlet fan at the same speed, even though it might be preferred. For
example, if a single-inlet full-width airfoil fan has a wheel inlet velocity of 60 m/s (erosion
factor of 9 - see Figure 2.2), a double inlet fan should be considered, because its lower
erosion factor of 2.2 is an improvement of 400 %. However, this may not be a good
recommendation because reduced blade tip width results in reduced fan efficiency. In
this case, it is preferable to retain the single inlet and choose a larger, slower-speed fan
that would provide a lower inlet velocity and erosion factor.
 -5-

2.2.2 Continuous operation

Any type of equipment will require maintenance. In blade form selection, blade build-up
and erosion have the greatest effect on fan operation. Build-up on the wheel results in
reduced performance. There is an increased tendency for material to build up on blades
as the blade angle is tilted back from radial. This build-up can accumulate to the degree
that it restricts and alters air passages, reducing both efficiency and performance.
As a result of build-up and uneven dislodging of material, the wheel becomes unbalanced
and begins to vibrate. If not corrected, vibration will reduce bearing life and possibly lead
to serious structural problems in other fan components.
In the specific case of the Airfoil blade fan, any infiltration of material into the fan's hollow
sections is critical because the problem cannot be corrected by cleaning the external
surfaces.
Blade erosion is the second risk factor affecting continuous fan operation. It is commonly
accepted that erosion varies as the second to third power of the speed of a particle that
hits the blade. For a given rotor dimension, this particle speed is proportional to the rotor
speed, therefore erosion varies as the second to third power of fan rpm.
Impeller erosion is also affected by the wheel inlet velocity. The wheel inlet velocity is
calculated by dividing the fan flow by the smallest section of the stationary inlet piece.
The kinetic energy of a dust particle entering the fan wheel can be expressed as W =
2
mv /2.
Graph 8 shows the relative erosion factor as a function of wheel inlet velocity. This graph
also reveals why it is often recommended that a double-inlet fan be utilized rather than
single inlet. If the solid line represented a single inlet fan, for example, then the dotted line
would represent the corresponding erosion factor of a double-inlet fan.

Erosion factor

16
Volume flow
w =wheel inlet velocity=
area of inlet throat 14

12

10 single inlet

8
w At
6
Area of inlet throat
4
double inlet
2

0
10 20 30 40 50 60 70 80

wheel inlet velocity [m/s]

Fig. 2.1 Fig. 2.2

2.2.3 Mechanical design

All fan rotors are subject to centrifugal force. Depending on blade form and angle,
different types of stress occur in a blade. The radial Blade is in tension, while bending
 -6-

and tensile stresses act on the Backward inclined/Airfoil design. Bending stresses are
more subject to fatigue stresses.
The effect of the number of blades on fan efficiency was mentioned earlier. Consideration
must be given to the stresses resulting from the span of the shroud between blades.
Large diameter, high tip-speed wheels with long shroud spans between blades can have
inherent resonant frequency responses that could result in vibration and/or structural
problems unless adequate thickness or blade numbers are used.
 -7-

3. WEAR PROTECTION

3.1 Erosion types

It has been established that maximum erosion occurs when the angle of incidence
between particle and surface is between 20 and 40 degrees. This erosion, referred to as
"ductile", is thought to be the consequence of microscopic melting, which occurs when
sharp-edged, hard particles scratch the surface.
Particles striking at an angle of about 90 degrees to the surface erode according to
another mechanism. The kinetic energy of the particle creates a stress at the contact
surface that can exceed the elastic limit of the material, thus forming a surface crack.
This type of erosion is called "fragile".

3.2 Protection of parts subjected to abrasion

3.2.1 Direct protection

Ø Increased thickness of wear parts
This should only be done when abrasion is very light (ID behind electrostatic
precipitators)
Ø Direct hard surfacing on impeller by deposits by electrode or fusion projection
This process is very efficient if the deposits are well chosen and properly applied.
However, the thickness of the deposit is limited. The base structure of the impeller
can be affected (dilution of hard surfacing and significant addition of energy) and
the mechanical characteristics of the steel can deteriorate, especially after several
maintenance operations.
Ø Glued-on ceramics
Their hardness is very high and they are very resistant to ductile abrasion
Temperature must be limited and the ceramics tend to come off at the blade end
(high centrifugal forces). Application on static parts is easier and more reliable.

3.2.2 Protection by wear plates

Ø Corners or mild steel plates (or steel of the same grade as the base structure)
added on to the blade and on the center plate. This protection is applicable only
where abrasion is low. If the protection area is not wide enough, the base plate can
also be attacked.
Ø Wear-plates in abrasion resistant steel. This process is applied particularly on flat-
bladed impellers. The wear-plates are fitted on with countersunk screws.
Replacement is easy, but the solution can be insufficient if abrasion is very high.
 -8-

Ø Mild steel wear plates with hard surfacing by electrode or hardsurface spray
coating. This solution has the double advantage of easy replacement and high
resistance to Abrasion when the type of deposit is well chosen. Moreover, hard
surfacing on site is easily carried out with no risk for the base structure.

3.3 Deflection of abrasive particles

3.3.1 Deflector plates on impeller (Fig. 3.1)

Fig. 3.1a shows the fan inlet and the rotating impeller. A, B and C are flow lines for the
gas and lines 1,2 and 3 represent the trajectories for particles of varying size. Line 1
refers to a very small particle that closely follows the gas flow line. With increasing
particle size the trajectories 2 and 3 deviate from the gas flow lines. The particles hit the
back-plate of the impeller and erode it.
Fig. 3.1b depicts the passage of two different particle sizes between two blades. Line 1
is the trajectory of a small particle, line 2 of a coarser one. Most particles will hit the blade
during their passage through the impeller.
With suitably located deflector plates (Fig. 3.1c), the particle trajectories can be altered
so that no particles will hit the blades. The location and direction of the deflectors
depends on the particle size distribution, density of gas and particle, and fan speed and
size.
 -9-

Fig. 3.1 Wear protection by particle deflection

 - 10 -

4. FAN IMPELLER ARRANGEMENT

4.1 Overhung assembly (Fig. 4.1)

This is possible if the fan operates in an environment with little risk of clogging and
associated imbalance and the impeller diameter is not too large (less than approx.
2.7 m).
Advantages:
Ø lower purchase and installation costs than center hung assembly
Ø connection of upstream duct directly to fan inlet is possible (no inlet box)

4.2 Center hung assembly (Fig. 4.2)

This assembly is more robust and absorbs the dynamic effects of rotor imbalance better
than the above solution.
Its use is recommended if
Ø the dust load is high with a risk of clogging or wear of the impeller, causing rotor
imbalance
Ø when the size and weight of the impeller makes overhung assembly delicate

The following figure indicates the application range for the two arrangements.
wheel dia. [m]

2.7

2.5
center hung
2.3

2.1 overhung

1.9

1.7

1.5

900 1000 1100 1200 1300 1400 1500 1600 1700 1800

fan speed [rpm]

Fig. 4.3
 - 11 -

Fan arrangement
 - 12 -

5. FOUNDATIONS

1. separate motor baseframe 2. common baseframe

3. common baseframe on anti-vibration mounts 4. concrete slab

1. Directly onto concrete with separate motor-bearing base-plate (this is the most
economical solution and most widely used).
2. Onto concrete with common base plate with centerline axis support maintaining the
rotor-bearing-motor alignment with the stator (more expensive solution, but easy to
install onto the foundations).
Used for hot gas fans where thermal expansion must be taken into account.
3. Common base-plate with centerline axis support resting on anti-vibration mounts (this
chassis must be perfectly rigid and is thus very expensive).
Used only in very special cases (fans mounted on steel structures).
4. With spring-supported concrete block resting on anti-vibration mounts and supporting
the fan (a spring supported block is less expensive than a common base-plate with
centerline support and allows for remarkable vibration absorption. Its weight, between
10 - 60 tons must be included in design calculations).
Used only in very special cases.
 - 13 -

6. FAN PERFORMANCE
The fan performance curve is derived from laboratory data when the flow conditions to
and from the fan is ideal. Since these conditions seldom exist at cement plant fan
locations, the fan curve data cannot be considered reliable when applied to field
conditions. Because of this uncertainty of prevailing conditions, we find that engineers
specify oversized fans during the plant design stage. For example, typical factors that
can effect the performance of an under-grate cooler fan include: a dirty inlet screen; a
structural column too close to the inlet; a silencer at fan inlet or the influence of an
adjacent fan. Likewise, similar factors influencing an induced-draft fan include
asymmetrical inlet duct which effects streamline distribution to the fan or heavier than
normal blade wear pads that restrict blade passage geometry.
The typical fan curve shows the quantity of air on the horizontal axis and the fan static
pressure and fan power plotted on the vertical axis. The conditions of density and flow
are at the inlet of the fan. The actual operating point of the fan will be the intersection of
the fan curve with the system resistance curve.
Definitions :
Ø Fan static pressure is the difference between the static pressure (measured
perpendicular to the direction of air flow) at the fan outlet and the total pressure at
the fan inlet.
For most fans in cement plants, it is satisfactory to assume that the fan's static
pressure is the difference between the static pressure at the inlet and outlet of the
fans.
Ø Fan power is the power at the fan shaft. For most plant applications where the fan
has constant speed and is direct coupled, this can be assumed to be 96 % of the
power consumed by the drive motor when the motor is fully loaded. Since power
factor and motor efficiency vary with motor load, it is difficult to measure fan shaft
power input without a kilowatt meter. However, near full motor load conditions, the
amp reading is a good indicator of shaft power. For fans with variable speed drives,
especially hydraulic or magnetic clutch drives, the efficiency of the drive can be
very poor so the power at the fan shaft cannot be assumed to be a linear function
of the power consumed by the motor.
Ø The System Resistance Curve is the relation between pressure and volume flow
for the given system. For a system in which only air is moved and the geometry of
the ductwork is constant, the pressure drop is proportional to the square of the flow
rate.
In most of the applications around the cement plant, this curve is constantly
changing because of the changing system conditions. For instance, the cooler
under-grate fan system resistance curve depends not only on the geometry of the
ductwork and cooler, but also on size distribution and level of clinker on the grate
above the compartment.
The system resistance curve of a roller mill system is dependent on the geometry
of the ducts, the amount of material being transported by the gas, the composition
of the gases and the speed of the classifier. Because of these variables, it is
essential to understand that a system resistance curve can vary significantly from
the slope of the curve illustrated in e.g. Fig. 6.1.
 - 14 -

6.1 Fan Capacity Adjustment

Fan capacity adjustments in the field are practically limited to the speed adjustment and
this within a limited range: ~ 0 - 15 % for flow increase and ~ 0 - 30 % for flow reduction.
Other capacity adjustments require mainly dimensional modifications of the fan wheel
which have to be carried out in the workshop. Those are more costly and critical to
execute.

6.1.1 Fan capacity too low

Important variables that are hampered by insufficient gas flow are e.g.
Ø kiln production
Ø separator efficiency
Ø mill performance (throughput, drying capacity, mill venting)

Possible remedies :
Ø Design changes in the system (reduce false air, reduce pressure drop)
no additional energy is consumed by the fan!
© reduce/rectify false air inleaks
© reduce unnecessary pressure drops caused by e.g.
© not fully open dampers
© improper design of inlet/outlet connections
© no turning vanes in bends (refer to Annex II 2/4 & 3/4)
© dust deposits in ducts and bends
© too narrow ducts (appropriate air speed must however be maintained to
avoid settling of dust)

Ø Changes in fan design, speed

© Speed increase
© volume flow is directly proportional to fan speed
© fan absorbed power rises with the third power of fan speed
© fan works with lower than design efficiency
© normally only possible with V-belt drives
© sound emission rises with fifth power of speed
© higher wear, if fan handles dust-laden gas
© increased sensitivity to rotor imbalance caused by dust deposits on
blades
© speed increase is limited by mechanical strength of rotor
© speed increase is limited by fan critical and resonant speed
 - 15 -

© Fan wheel diameter increase

© practically limited to wheels where the blades do not extend to the full
diameter of the rotor sideplates (plates can then be welded-in to
increase the effective fan wheel diameter)
© volume flow rises with the third power of wheel diameter
© fan absorbed power rises with the fifth power of wheel diameter

© Replacement of rotor with inherent low efficiency by one designed for high
efficiency
© e.g. replacement of radial-blade wheel by backward-inclined or
backward-curved blade wheel.
© limitations with regard to diameter, width (and speed, if wheel is directly
coupled to motor via flexible coupling) of the new wheel exist if fan
housing and motor are to be kept.

Ø New fan
Can be designed for optimum performance at the new operating point

6.1.2 Fan capacity too high

May be due to e.g. oversized fan, decline in production, process changes
Possible remedies:
Ø Flow reduction by damper
Widely used solution, but inefficient and expensive (fan energy consumption)
Ø Flow reduction by inlet vane damper
More energy-efficient than damper, but can be recommended only for flow
regulation purposes, not for permanent use at lower capacity (expensive, may be
difficult to fit into existing equipment)
Ø Changes in fan design, speed
© Lower speed
© volume flow decreases proportional to fan speed
© absorbed power decreases with third power of fan speed
© lower sound emission
© lower wear, if handling dust-laden gas
© lower sensitivity to rotor imbalance caused by e.g. dust deposits on
blades
© Rotor width reduction
© applicable if fan must deliver the same pressure as before, but at lower
volume flow
© Rotor diameter / width reduction
© volume flow decreases with the third power of rotor diameter
 - 16 -

© volume flow decreases proportional to rotor width

© fan delivery pressure decreases with the square of rotor diameter
© absorbed power decreases with the fifth power of rotor diameter
© absorbed power decreases proportional to width
© lower sound emission
To maintain a good efficiency, it is not sufficient to decrease the diameter only,
but also to adapt the rotor width.
For a permanent capacity reduction this is a good solution
If the old transmission is kept, speed reduction is normally reversible, i.e. fan capacity
can be increased again if necessary, whereas a reduction in rotor diameter / width are
not.

6.2 Adjusting fan performance curves

It was mentioned above that the fan performance curve is given at specific conditions of
density and fan speed. Most often, it is necessary to correct the fan curve for density and
speed other than the predicted conditions. Fig. 6.1 shows the influence of density
changes on the fan performance.
A calculation example is added in Annex 1.
Gas density correction
Ø temperature changes
Ø changes in chemical composition of gas
Ø changes in altitude (height above sea level)

Index 1 : reference conditions

2 : actual conditions

ρ 
∆p2 = ∆p1 •  2 
 ρ1 
Volume flow and efficiency of the fan are unaffected by gas density changes.

Calculation of actual density:

M p 273
ρ= • •
22.4 1013 T + 273

ρ density [kg/m 3]

M molecular weight of gas [kg/kmol]

p actual pressure [mbar]
T actual temperature [°C]
ambient pressure for higher altitudes can be calculated by:
p = 1013 •exp (− 0.0001255 • H )
 - 17 -

H altitude above sea level [m]

Gas Density Molecular weight

[kg/Nm3] [kg/kmol]
O2 1.429 32
CO2 1.964 44
N2 1.250 28
Air 1.292 29
H2O 0.804 18

Fan rpm correction

Index 1 : original fan curve characteristic
2 : actual

n 
V& 2 = V& 1 •  2 
 n1 
2
n 
∆p 2 = ∆p1 •  2 
 n1 
3
n 
N2 = N1 •  2 
 n1 
Fan wheel dimensions correction
Index 1 : original
2 : actual
d... impeller diameter
b... impeller width
3
d  b 
V& 2 = V& 1 •  2  •  2 
 d1   b1 
2
d 
∆p 2 = ∆p1 •  2 
 d1 
5
 d  b 
N2 = N1 •  2  •  2 
 d1   b1 
 - 18 -

Influence of density changes on system resistance and fan

performance curves (Fig. 6.1)
 - 19 -

7. FLOW CONTROL
Damper control generally results in higher power consumption and so more if the
damper is installed at the fan outlet. Therefore the damper should always be installed at
fan inlet. Inlet vane control may be satisfactory from 100 % down to about 70 % of
maximum flow, but the power demand becomes high when the flow is reduced further.
Speed control is virtually ideal.
Figure 7.5 shows how fan power consumption is affected by the different flow regulation
methods.

7.1 Damper control (Fig. 7.1 & 7.2)

A parallel blade inlet damper is preferred over either outlet damper or an opposed-blade
inlet damper. The parallel-blade inlet damper pre-spins the incoming air in the direction of
wheel rotation, resulting in a lower energy consumption in the regulation range of 100 - 80
% of maximum flow.

7.2 Radial inlet vane (Fig. 7.3)

The radial inlet vane mounted direct at the fan inlet pre-spins the incoming air still better
in the direction of wheel rotation, resulting in a wider range of stable regulation (100 -
70%) and less energy consumption.
Their use is mainly recommended in connection with over hung arrangement fans with
low rates of dust, thus limited to applications after filters or in clean air, e.g. for cooler
under-grate fans.
This arrangement is normally more costly.

7.3 Speed control (Fig. 7.4)

7.3.1 Hydraulic transmission with fixed speed motor

Hydraulic transmission in connection with a fixed speed motor can be an option for speed
ranges from 100 % down to 85 % of maximum speed, but the energy efficiency becomes
low when the flow is reduced further.
 - 20 -

7.3.2 Speed-controlled electric motors

Flow control by variation of the fan speed is most efficient with regard to energy savings
and permit also the reduction of wear on the fan wheel.
Ø DC motors have limitations in high speed and power (roughly 1000 rpm for 1500
kW, 600 to 800 rpm for 2000 kW motor) and require a lot of maintenance work
(motor ventilation circuits, carbon brushes etc.), so their use is not recommended
any longer.
Ø AC motors with slip recovery. This motors have a limited speed range down from
100 - 30 %. Its cost increases with the width of the range.
Ø Synchronous motors with AC variable frequency control
Both types of AC motors are well suited for high power (500 to 5000 kW)
 - 21 -

Flow regulation by outlet damper (Fig. 7.1)

 - 22 -

Flow regulation by inlet box damper (Fig. 7.2)

System resistance
80
74 mbar
curve

70
s tatic p re s s u re [m b a r]

60

50
41 mbar open damper
40

30

20
partially closed damper

10

100 200 300 400 500

900 open damper

800 kW

700

565 kW
partially closed damper
500
p o w e r [k W ]

300

100
241'500 m3/h 325'000 m3/h

100 200 300 400 500

Volume flow [m3/h . 1000]

 - 23 -

Flow regulation by inlet vane damper (Fig. 7.3)

 - 24 -

Flow regulation by speed control (Fig. 7.4)

 - 25 -

Comparison: Types of Flow Regulation (Fig. 7.5)

100
outlet damper
90

80 inlet damper
absorbed power [%]

70

60

50
inlet vane
40

30

20
variable speed
10

0
0 20 40 60 80 100
flow [%]
 - 26 -

8. START-UP OF FANS
Before starting fan the first time, complete the following list

1. Uncouple motor from fan and check motor (fan) for proper rotation.
2. Shut off power by disconnecting motor main breaker.
3. Check and tighten hold-down bolts.
4. Check and tighten rotor set-screws.
5. Check couplings and bearing for proper alignment.
6. Move rotor to see if it is rotating freely and maintains proper inlet piece/rotor
clearance.
7. Check that fan wheel is balanced.
8. Check fan and ducts for any foreign material or dirt build-up.
9. Check that physical position of damper corresponds to indication at actuator and
control panel.
10. Secure all access doors.
11. Check lubrication of bearings, couplings, drive unit etc.
12. Couple motor again to fan and secure and check safety guards for clearance.
13. Close dampers for adequate system resistance to prevent drive unit from
overloading.
14. Supply water to water cooled bearings and start lubrication pump.
15. Make sure that all persons are away from fan and out of any other equipment of the
system to which the fan is connected.
16. Connect electric motor by closing main circuit breaker of the motor. Start
equipment according to recommendations of drive unit and starting equipment
supplier.
17. Allow fan to reach full speed, then shut down. Make immediate corrections if any
vibrations or unusual sounds have been detected.
18. During a run-in period make observations of bearings at least once an hour. Higher
bearings temperatures may occur if bearings are over-lubricated.
19. Refer to trouble-shooting guide for any unusual occurrences encountered during
the run-in period. Only after any vibrations, misalignments, etc. have been
corrected, may the fan be restarted.
 - 27 -

9. POSSIBLE PROBLEMS WITH FANS

9.1 Vibration
General Machinery vibration severity as per ANSI S2.41 (Fig. 9.1 a)
for use as a guide in judging vibration as a warning of impending trouble)
1. Rigid support
The fundamental natural frequency of the machine/support system is higher than the
operating speed

excellent 0. to 2.54 mm/s vibration velocity (Peak)

good 2.55 to 6.35 mm/s
alarm 6.36 to 12.7 mm/s
shutdown > 12.7 mm/s
2. Flexible support
The fundamental natural frequency of the machine/support system is lower than the
operating speed

excellent 0. to 3.81 mm/s vibration velocity (Peak)

good 3.81 to 10.16 mm/s
alarm 10.17 to 19.1 mm/s
shutdown > 19.1 mm/s

Vibration severity criteria (10 Hz to 1 kHz) per ISO 2372 (Fig. 9.1 b)
1. Large machines with rigid foundations whose natural frequency exceeds machine
speed

good 0.0 to 2.54 mm/s vibration velocity (Peak)

allowable 2.55 to 6.35 mm/s
just tolerable 6.36 to 15.84 mm/s
not permissible > 15.84 mm/s
2. Large machines operating at speeds above foundation natural frequency

good 0 to 4 mm/s vibration velocity (Peak)

allowable 4 to 10 mm/s
just tolerable 10 to 25.4 mm/s
not permissible > 25.4 mm/s
If the tolerable vibration levels are exceeded, the fan must be shut down, which usually
results in costly production losses. To extend the periods between shutdowns due to
vibration, the use of automatic balancing devices, mounted on the fan shaft may be
considered. Depending on their size, they are capable of automatic compensation of a
certain rotor unbalance. For manufacturers of such devices, refer to the information
source on page 40.
 - 28 -

Fig. 9.1 a Vibration severity chart (ANSI S2.41)

 - 29 -

Fig. 9.1 b Vibration severity chart (ISO 2372)

 - 30 -

9.1.1 Material build-up

After balancing, the rotor of a fan will still have a certain residual imbalance, the value of
which depends on the balance quality grade. The unbalance force can be calculated with
the formula

F = m • ω 2 • eper

m rotor mass [kg]

ω angular velocity 2π.n/60 [s -1]

n rotor speed [rpm]
eper permissible residual specific unbalance [m]

Industrial fans often handle dust-laden gases and dust deposition on the impeller may
occur, increasing the initial unbalance. Asymmetrical wear on the impeller has the same
effect. These additional unbalances can be considerable compared to the permissible
unbalance according to the balance quality grade. If the resulting vibration exceeds the
tolerable limits, the fan has to be shut down to clean the wheel. See Fig. 9.1 c for
permissible unbalance.

Kiln exhaust fan build-up

One idea about the causes of build-up is that some particles are "sticky" at temperatures
above 300 °C and begin to build up on the rotor surfaces. The impact energy of the
particles striking the rotor surface (especially at an angle of 90 °) is also converted to
heat and results in additional softening. Other particles with a higher softening point are
caught in the sticky material and increase the coating layer thickness. The originally soft
build-up gets harder under the influence of heat and pressure (from centrifugal force and
the impact of other particles).

Recommendations against build-up :

Ø The fan rotor should be designed for the smoothest possible flow lines to reduce
the impact energy of dust particles. Backward curved and airfoil are the best blade
forms. Airfoil blades must be designed carefully to prevent material from getting
inside and regular inspection of the fan blades is mandatory.
Backward curved blades must be inclined enough to prevent the "hard" build-up on
the front surface and radial enough to prevent the "soft" build-up on the back-
surface.
 - 31 -

Ø Fans should be designed for low gas and particle velocity at the fan inlet. This
reduces the impact energy of particles against the rotor and can be achieved by
1. Double inlet instead of single inlet fans
2. Larger diameter / lower speed fans
3. If possible, the peripheral speed at the rotor inlet opening should be limited to
about 76 m/s, and wheel inlet velocity should not exceed 38 m/s
4. The fan shaft should be oversized to reduce unbalance effects.
The design critical speed (considering a bearing oil film thickness and a build-up
thickness of 25 mm on all leading surfaces of the rotor blades) should be at least 1.25
times the operating speed of the fan.
Suppliers generally do not recommend to spray water directly onto a hot fan wheel, but in
practice it has been done over years with success. The rapid cooling effect removes
build-up effectively, however, the strength properties of the wheel material may be
affected adversely if the injection rate is too high.
 - 32 -

Fig. 9.1 c Balance quality grades

 - 33 -

9.1.2 Variable speed operation

All electrical variable speed drive systems can generate harmful harmonics that result in
torque pulsation. Such harmonics can be predicted and filtered, but often at high cost.
For fans with variable speed control it must be verified by the motor supplier that the
pulsating torque do not cause excessive vibrations.

9.1.3 Thermal effects

Some typical problem areas are :
Ø due to expansion joint problems, forces due to thermal expansion of ducts are
transmitted to the fan housing, resulting in damage to the housing or interference
between wheel and housing. It can also cause excessive force on foundation bolts
which sometimes can result in foundation cracks.
Ø rapid temperature changes in a system require proper design of the wheel-to-shaft
fit. This will assure that looseness and resulting vibration sensitivity will be avoided.

9.1.4 Hot shutdowns

Thermal shaft set is a usual concern on center-hung fans above 120 °C when shut down
in the hot condition. It is generally agreed that this thermal bowing of the shaft occurs due
to uneven thermal gradients across the wheel and shaft assembly at hot shutdown. This
results in small asymmetrical distortions, often sufficient to cause excessive unbalance
forces during startup.
In some cases the resulting vibrations are within acceptable limits. If so, it is usually
found that the unbalance forces disappear after 12 to 36 hours of operation.
The preferred, but expensive solution is to have an auxiliary drive to slowly rotate the
wheel and shaft assembly, which should be engaged immediately after hot shut-down to
avoid the undesirable thermal distortion. Auxiliary drives are typically designed to maintain
a minimum speed (40 to 60 rpm) as the fan slows down. They are not intended for use in
starting the fan rotor from a dead stop.
Most hot gas fans work well without auxiliary drive, since hot shutdowns are infrequent
and of short duration and the natural draft of the chimney keeps the rotor in slow motion
for some time.
The tendency is to install hot gas fans without auxiliary drives. Often it is also the case
that variable speed drives have a turndown ratio of 10:1 which corresponds normally to
less than 70 r.p.m.
 - 34 -

9.2 Erosion

9.2.1 Improper duct connection

Figure 9.2a shows a duct arrangement encountered on a raw mill system. Due to the
uneven material distribution one side of the double-inlet impeller wore out much faster.
Figure 9.2b shows the recommended modification.

Fig. 9.2 a Fig. 9.2 b

9.2.2 Effects of impeller speed and wheel inlet velocity

As mentioned earlier, erosion is proportional to the square of wheel inlet velocity and to
the second to third power of the relative gas velocity w1 at the rotor inlet. With a given
rotor size this velocity is proportional to rotor speed [rpm].

w1

c1

u1
 - 35 -

9.2.3 Liner materials

Due to the need for ductility in the structural members of the wheel, the wheel itself is
usually not capable of high resistance to erosion. It is therefore necessary to install liners
with a higher hardness classification to provide sufficient protection against wear.
Material hardness is an indication of its resistance to erosion. A very good liner material is
chromium carbide with an average hardness of 600 Brinell.

Rate of Erosion for different Materials

Erosion Rate

1 2 3 4 5 6 7 8 9

Fig. 9.2 c Erosion Test on Fan (Solyvent-Ventec)

The figure above shows erosion test results of nine different materials.
1. Alloy of tungsten and nickel-chromium. Flame spray coating.
2. Alloy of nickel, chromium and cobalt. Flame spray coating followed by fusion.
3. Alloy of tungsten, cobalt carbides and nickel-chromium. Flame spray coating followed
by fusion.
4. A special chromium cast iron. Special electric arc welding.
5. Chromium cast iron with chromium carbides. Electric arc welding.
6. Chromium cast iron. Semi-automatic electric arc welding.
7. Ceramic powder containing basically aluminum oxide. Flame spray coating.
8. Same as 7, but of different hardness.
9. Ceramic tiles containing basically aluminum oxide. Glued to the blades.
 - 36 -

Carbon steel sheets with a protective layer of chromium carbide are normally available in
standard sizes and various thicknesses. They can be cut to the necessary form and size
by cutting discs or plasma.
The liner fixation can be done by bolting or by welding the carbon steel base plate to the
existing structure of the wheel or casing.

9.3 Bearings
Bearing problems are usually caused by operating conditions and not by deficiencies in
the bearings. By following regular operating and maintenance procedures, many bearing
problems will be avoided.
Most problems are due to hot bearings. For hints see section 10.
 - 37 -

10. TROUBLESHOOTING

Problem Check for

Noise 1. Squealing V-belts, due to misalignment or improper tensioning
2. Defective bearings, or bearing seal rubbing
3. Misalignment of bearing seal
4. Misaligned housing-shaft seal
5. Foreign matter in fan housing
6. Rubbing of shaft seal, wheel to inlet piece, or wheel to housing
7. Heat flinger is contacting guard
8. Coupling failure
9. Untreated expansion joints
Poor per- 1. Incorrect fan rotation
formance 2. Wheel is off-center, poor inlet piece fit-up allows recirculation of
air
3. Fan speed too low/high
4. Poor duct design, installation of elbow or turning vanes could
remedy problem
5. Inlet damper installed backwards (counter-rotation)
6. System resistance is excessive compared to design require-
ments (partially closed damper may be the cause)
7. Density may be different from design density
High bearing 1. Defective bearings caused by electrical arc due to improper
temperature grounding of nearby welding
2. Over-lubrication
3. Improper lubrication or contaminated lubricant
4. Lack of lubrication, cooling fluid, or circulation
5. High ambient temperatures or direct exposure to sunlight
6. Misalignment
7. Excessive thrust loading
8. High vibration
9. Inadvertently exchanged bearing caps (mismatched)
10. Bearing race turning inside housing
11 Moisture in bearing
12. V-belts too tight
13. Improper location ; not enough room for free axial movement of
floating bearing in its housing at elevated temperatures)
14. Heat flinger missing
Excessive 1. Motor improperly sized for fan wheel WR2
starting time 2. Inlet dampers not closed during start-up
3. Properly selected time-delay starter/fusing required (many indus-
trial fans take up to 20 - 25 seconds to reach operating speed)
4. Temperature at inlet is excessively low (high density)
5. Low voltage at motor terminals
6. Inadequate system resistance

Vibrations 1. Loose bolts in bearings and pedestals, or improper mounting

 - 38 -

Problem Check for

2. Defective bearings
3. Improper alignment of bearings or couplings
4. Out-of-balance fan wheel
5. Loose set-screws holding wheel to shaft
6. Weld cracking
7. Improper fan wheel clearance to inlet piece(s)
8. Material build-up and/or wear on wheel
9. Ensure expansion joints in ductwork are not fully compressed
10. Misalignment or loose V-belts
11. Improper wheel rotation
12. Operation near system critical speed
13. Shaft bent or distorted during high-temperature shutdown
14. Defective motor
15. Resonant frequencies of structural steel mounting
16. Beat frequency with other fans on common base
17. Loose hub-to-shaft fit

Duct 1. Control volume with a radial inlet damper

pulsation 2. Install speed variation
3. Change to a special "surgeless" blower design

High motor 1. Improper ventilation of cooling air to motor (may be blocked by

temperature dirt)
2. Input power problems (especially low voltage)
3. High amperage
4. High ambient temperature
 - 39 -

Component Problem Probable cause

Bearings Noise 1. Imperfection in bearing elements
2. Improper clearance
3. Internal wear of bearing parts
Freezing water jacket 1. When stopping water flow in freezing
weather, blow out lower portion of bearing
housing water cavity
Fan Wheel Erosion 1. Reduce dust loading
2. Reduce rpm
3. Redesign inlet ductwork
4. Damper setting
5. Damper design
6. Better liner material
7. Alternate blade design
Buildup 1. Reduce dust loading
2. May be affected by system temperature
change
3. Blade form
4. Alternate wheel material, apply "slippery"
material
Vibration 1. Rectify build-up
2. Rectify erosion
3. Tighten foundation bolts
4. Correct misalignments
5. Improve supporting structure
6. Check effects of ductwork thermal expan-
sion
7. Shaft bow due to "thermal set"
8. Special considerations, refer to factory
Shaft Cracks at section - Get new shaft that is machined to eliminate
change stress raisers
Natural frequency too - Redesign shaft
close to running value
Out of round at bearing - Replace shaft
Bowing and torsion - Refer to factory
problems
Shaft dropped or dam- - Get new shaft
aged during transit or
installation
Hubs Loose fit on shaft - Tighten interference fit
Insufficient stiffness - Redesign
Cracks in casting - Change to weldment
Erosion - Build up or replace
 - 40 -

Information Sources:
Ø Robinson Industries, Inc., Zelienople, PA, USA
Ø TLT-Babcock, Inc., Akron, Ohio, USA
Ø Venti Oelde, Oelde, Germany
Ø Solyvent-Ventec, Chalon-Sur-Saône, Cedex, France
Ø Balance Dynamics Corporation, Ann Arbor, Michigan, USA, Fax # 313 994 3690
 Annex I 1/4

Fan location

Fan type [-]

Supplier [-]
Location [-]
Wheel diameter [m]

Ambient conditions

Temperature [°C]
Ambient pressure [mbar]

Gas characteristics (fan inlet)

Pressure (absolute) [mbar]

Temperature [°C]
Density (see questionnaire 'Gas characteristics')
ρ =N [kg/Nm3]
ρ= [kg/m 3]

Fan operating data

Pressure fan inlet p1 [mbar]

Pressure fan outlet p2 [mbar]
p [mbar]

Temperature fan inlet T1 [°C]

Temperature fan outlet T2 [°C]
T [°C]
3
Volume flow (per fan) [m /h]
Nm3/h (from gas volume determination)

Fan drive
Motor type [-]
Motor Supplier [-]
Power at shaft [kW]
Motor speed [RPM]
Fan speed [RPM]

Plant: Fan evaluation

Detail:
Date:
 Annex I 2/4

Gas velocity

cross section fan inlet [m2]

inlet v1 = [m/s]

cross section fan outlet [m2]

outlet v2 = [m/s]

*v2*v 2
2 1
*v1*v1
p Vent. = p2 - p1 + ----------- - -----------
2 2

p Vent
= [mbar]

Fan efficiency (measured) : [%]

Fan characteristics
- Conversion of measured data to conform with basis
of fan characteristics

real (measured) data : data according to fan characteristic

- V1 [m3/s] - V2 [m3/s]
- p1 [mbar] - p2 [mbar]
- N1 [kW] - N2 [kW]
- n1 [RPM] - n2 [RPM]
- ρ1 [kg/m3] - ρ2 [kg/m3]

1. Correction of fan RPM

V2 = V1 * n 2/n1 (1)

p2 = p1 *(n22)/(n12) (2)

N2 = N1 *(n23)/(n13) (3)

Plant: Fan evaluation

Detail:
Date:
 Annex I 3/4

2. Correction of gas density

∆p2 = ∆p1 * ρ2 / ρ1 (4)

N2 = N1 * ρ2 / ρ1 (5)

(V is not influenced by ρ )!

Example

What are gas volume flow and fan motor power consumption
according to the fan characteristic ?

example our fan

Measured data: p1 = 50 [mbar]
T1 = 250 [°C]
ρ1 = 0.7 [kg/m3]
n1 = 800 [RPM]
N1 = [kW]
V1 = [m3/h]

Basis of fan characteristic: T2 = 20 [°C]

ρ2 = 1.2 [kg/m3]
n2 = 1000 [RPM]
1. Correction of density

∆p2' = ∆p1 * ρ2/ρ1 = 85.7 [mbar]

N2' = N1 *(n23)/(n13) = [kW]

2. Correction of RPM

∆p2 = ∆p2' * (n22)/(n12) = 133.9 [mbar]

N2 = N2' *(n23)/(n 13) = [kW]
V2 = V1 * ( n2 / n1 ) = [m3/h]

Plant: Fan evaluation

Detail:
Date:
 Annex I 4/4

With the pressure of 133.9 mbar we get 20 m3/s volume flow

from the fan characteristic :

V2 = 20.0 [m3/s]
N2 (optional) = 335 [kW]

The real volume V1 is calculated by transforming equation (1) :

V1 = V2 * n1/n2 = 16 [m3/s]

And power consumption with equations (3) and (5) :

N1 = N2 *(n13)/(n23)*ρ1/ρ2 100 [kW]

Hence fan efficiency is : 80 [%]

Plant: Fan evaluation

Detail:
Date:
Annex II 1/4
Annex II 2/4
Annex II 3/4
Annex II 4/4