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Process Engineering
Training Program
MODULE 5
Fan Measurement and Testing
Section Content
1 The Chemical Engineering of Fans
2 Fans and Airflows
3 Fans and Fan Systems, Chemical Engineering March 1983
4 Fans, FL Smidth
5 Harleyville Preheater Exhaust Fan performance Verification
6 Fan Applications
PRESENTATIONS
Bowmanville Calciner ID Fan Problem
Fan Technology
Centrifugal Fans and Systems
Blue Circle Cement
PROCESS ENGINEERING TRAINING
PROGRAM
Module 5
Section 1
The Chemical Engineering of Fans
C0NTENTS
Page No.
1. Definition of Fan Pressure 1
2. Fan Curves 3
3. The Fan laws 8
4. Control of Fan Output 11
5. Unsatisfactory Fan Performance 14
6. Series Fans 16
7. Parallel Fans 18
8. Blade Types 26
9. Fan Noise 31
10. Other Gas Pumping Equipment 32
11. Power Required for Compression 41
12. Further Reading 42
Appendix 1- Effect of Speed and Temperature
on Fan Operation 43
Appendix II- Prediction of Fan Noise 46
Appendix III - Effect of Water Injection on Fan
Performance 48
1.0 DEFINITION OF FAN PRESSURE
Fan total pressure PT, is the difference between the total pressures at the fan inlet and outlet.
Fan static Pressure, PS,is the fan total Pressure minus the fan velocity pressure.
Fan, velocity pressure, PV, is the velocity pressure 







g
2
u2
corresponding to the average velocity at the fan outlet.
i.e. the fan static -pressure is NOT equal to the difference between the static pressures at the inlet Ar!-.1 outlet.
It should be measured as the difference between the reading of a facing tube at the inlet and a side tube at the fan
outlet (see Figure 1).
FIG. 1 MEASUREMENT OF FAN STATIC PRESSURE
2.0 FAN CURVES
These show the effect on fan pressure, Power and efficiency of a change in fan throughput (Fig.2). The
horse-power on a fan curve is that which must be delivered to the shaft and does not take into account losses in
couplings and motors.
A calculation of the work actually done on the air will therefore give the efficiency of the fan itself,
Work done on air = pressure x flowrate x HP
33000
2
.
5






= 





6360
PV
HP (inches wg and cfm)
or 9.81 PV watt (mm wg and m3/sec).
P may be either the fan static or total pressure and the efficiencies thus calculated are the static and total
efficiencies respectively.
Because work is being done on the air, its internal energy will rise, and this, together with losses due to the
inefficiency of the fan will cause the air temperature to rise.
Temperature rise = ( )( )
heat
specific
V
density
Power
fan








For air at ambient conditions = efficiency
fan
total
24
.
0
V
075
.
0
779
2
.
5
PV
=
η














×
×








η
×
×
= 0.37 







η
P o
F
or about ½
o
F per inch wg.
The resistance of the system served by a fan depends upon the hardware in it and for any system there will be a
relationship between flowrate and pressure drop. If the system contains only dampers, ducting, expansions etc.,
of fixed resistance, then the pressure drop may be taken as being proportional to V2 (approximately). This would
not be true in the case of a resistance (as opposed to pressure drop), that varied with flowrate, such as a
Lepol grate. The intersection of the system resistance line with the fan pressure/flowrate curve gives the
point of operation of the fan. The effect of, for instance, closing a damper is to steepen the system resistance
line, giving a new operating point (see Fig.3), System resistance lines may be calculated from fluid flow theory
or one point may be measured end the remainder drawn in assuming, for instance, a squared relationship.
FIG. 2 TYPICAL FAN CUIRVEES
FIG. 3 EFFECT OF A CHANGE IN SYSTEM RESISTANCE
2.1 The Shape of Theoretical and Practical Fan Curves Fan curves may be calculated theoretically and the
shape of these curves are shown in Fig. 4. However, in practice losses modify the curves to give what we
normally recognize as a typical fan curve.
(a) Friction Losses:
Friction at the fan inlet, casing, and blade passagescauses a loss which is proportional to the square
of the velocity (or volume) and is zero at zero flow.
FIG. 4 FAN CURVES CALCULATED FROM THEORY
(b) Eddy Losses:
These occur at the blade leading edge and in the fan casing. Depending upon the relative velocity
between the air and the impeller leading edge, the angle of attack will change. The smoothest flow
will occur when the above relative velocity is in the same direction as the blades themselves. Here
eddy loss is zero and increases as the angle of attack moves away from zero.
A similar situation occurs in the fan casing, there being only one flow where eddies will not occur. It is usual to
design casings such that the flowrate corresponding to eddyless flow is the same for both impeller and casing.
(c) Leakage:
Backflow between casing inlet and impeller inlet causes a small loss.
(d) Disc friction:
The resistance offered by the air to the impeller rotation does not affect the head developed but is seen as
increased power absorbed.
(e) The theoretical no-lose characteristics shown above assume that the air follows the blade passages
absolutely. This only be true for an infinite number of blades.
The effect of the above on fan characteristics is shown in Figures 5A and. 5B for a backward curved impeller.
3.0 THE FAN LAWS
The performance of a fan may be changed by several methods, the main ones being a, change in
a) fluid density, p
b) impeller speed, n (rpm)
c) fan size * Most text books only quote the fan laws for "fans of geometric similarity", so that
diameter, width, outlet duct size, etc., are all increased at the same rate. Any convenient dimension
may then be taken as being representative of the fan, e.g. Impeller Diameter, Dim,
The effect of these on fan pressure,volume and horse-power is summarised in Table I.
TABLE I - The Fan Laws
ALTERATION IN
CHANGE IN FAN
PRESSURE
VOLUME POWER
Density, p α p No change α p
Speed, n α n 2 α n α n3
Fan Size*, Dim α Dim
2 α Dim
3 α Dim
5
Fan pressure may be either total, static or velocity. In order to help remember the above table, it is worth noting
that fan pressure can be shown to be theoretically proportional to the square of the tip speed whereas volume is
proportional to the swept volume of the impeller (fan size/volume is a cubic relationship since width increases
we well as diameter). Power is proportional to the product of volume and pressure.
A change in impeller diameter/width ratio will change the characteristic of the fan and direct scale-up is not
possible, although if a fan is running at somewhere near its peak efficiency (i.e. just to the right of the peak on
the pressure - Volume graph) then for a small increase in diameter, the volume will increase as the cube and the
pressure as the square of the tip speed, i.e. the fan laws in Table I would appear to apply. However, it must be
remembered that a new fan type has been produced by the change in diameter/width ratio and throttling of the
fan will not follow any easily calculable curve.
Presented with curves for a fan running at any set of conditions, it should be possible to predict the curves at any
other set of conditions.
Before setting out on the calculations, it is worth checking that the fan is actually operating close to the
manufacturer's fan curve. Their fan tests are performed with straight inlet and outlet ducts and, for instance, a
bend close to the fan inlet on site could lead to adverse whirl and a reduction in the fan efficiency. The method
of calculation is best shown by example. Appendix I shows the effect of a change in density and speed on a fan,
these being the variables normally encountered in practice.
Care must be taken in applying the fan laws, they are only applicable over a limited range because they are only
an approximation. A better estimate is given by
V = kg Dim
3
n (Re)x
(4)
P = kp Dim
2
n2
(Re)z
(5)
where, kp, kg, x, z are constants.
Re = The Reynolds Number (Dimensionless)
i.e.: 







viscosity
fluid
D
x
speed
tip
density x
fluid im
The indicies x and z are so small, however, that the effect of the Reynolds Number is usually negligible in
practical cases.
Even British Standard 848 states that Reynolds Numbers may be ignored in scale up provided they do not vary
by more than 40%.
4.0 CONTROL OF FAN OUTPUT
For a centrifugal fan there are three basic methods of controlling output
1. Damper Control
2. Variable Radial Inlet Vanes - these give a beneficial .swirl to the air at the fan inlet and therefore reduce the
work the fan has to do.
3. Speed Control.
Initially, dampers are the cheapest to install and speed control the most expensive, but there is a reversal in a
running cost comparison. - A cost comparison must be made in each individual case based on the capital cost,
running cost, and the amount of turndown that will be needed. However, speed control usually only becomes
economic on the larger machines.
Whereas variation in a damper setting will change the system resistance and not the fan curve, the other two
methods produce new fan curves, that for speed control retaining the same fan efficiency. Since the use of speed
control introduces the inefficiency of the device itself, it is worth noting that for small changes from the
maximum throughput, speed control has, in fact, the highest running cost.
Referring to Figure 6, the effect of turndown to 3 V/4 and V/2 by the three methods is shown.
Graph A - Damper Control - Fan curve retained but system resistance line altered.
Graph B - Variable Inlet Vanes - New fan curves produced for each vane setting.
Graph D - Comparison of three methods on basis of power taken
Graph C - Speed Control - New fan curves produced for each speed but efficiency retained.
(Note: loss due to coupling is exaggerated).
This graph assumes no pressure drop across a fully open damper. This may be a reasonable assumption for some
dampers operating in clean air, but would certainly not be true for some I.D. fan dampers where build-up will
occur. The expected pressure drop across an open damper, based on experienced, should be taken into account
when deciding on the method of control.
Radial inlet vanes are not cheap to install and can increase the cost of a fan by 20%. However, they are an
efficient means of control over a limited range and should be considered as an alternative to speed control. For
larger variations in volume a two or three speed motor with final adjustment by inlet vanes would be cheaper
than infinitely variable speed control.
If considerable operation at less than maximum capacity is expected, speed control had the added advantages of
an increase in life and a reduction in noise level.
5.0 UNSATISFACTORY FAN PERFORMANCE
Fan performance may be affected by worn impellers, inefficient drives, etc., but the most common
causes of fan performance being substandard are bends close to the inlet or outlet of the fan.
If the bend is at the inlet, the air is pushed over to one side of the eye, whereas a manufacturer's fan curve is for
uniform inlet conditions. Uniformity of flow at inlet is easily checked by a pitot traverse.
The best solution to this is an inlet box which can be sized so that the air enters the impeller uniformly, failing
that a square bend with cascade (Fig.7) will give good performance. Retention of the original bend and fitting of
long turning vanes will improve matters greatly (Fig.8).
Useful pressure in a fan originally comes from the acceleration of the air, - As the air is again slowed down by
expansion, velocity pressure is converted into static pressure. For this to be done efficiently the expansion
should be slow and gentle and clearly a bend immediately on the fan outlet will not help matters. This problem
is not so easy to remedy but again a cascade will help.
6.0 SERIES FANS
For fans in series, the second fan will handle -the same mass of gas as the first fan. However, due to
compression effects and temperature effects (Equation 3). The second fan will not handle the same volume. To
be absolutely accurate, the fan curve should therefore be modified to correct for the density difference. The two
effects, however, tend to cancel each other out, and in practice it is usually acceptable to use the original curve
for the second fan. The combined pressure at any volume is found by adding the total pressures of the individual
fans and subtracting the losses for the ducting between the fans (usually negligible) - see Fig. 9. Here to obtain
the complete fan curve it has been necessary to know the characteristic of one fan at volumes greater than it
would handle with free inlet and discharge (i.e. being "force-fed"). This is not normally known, and in any case
there would be little point in operating under these conditions since the fan would be hindering the flow of air
rather than helping it.
Unless the fans are identical it is likely that one of them will have to operate at a point that does not give its
maximum efficiency. For this reason, if a series set up is necessary, it is customary to install identical units. This
may be necessary if the space available renders one large fan impractical or if it is necessary to keep the air
flowing while one of the fans is being maintained (each fan would then have a by-pass).
A similar exercise is then performed on Fan 2 and its associated system. The two curves (P-F)1 and (P-F)2 are
then added together to give the characteristic of the parallel unit. Any system resistance line now drawn in a
refers only to the resistance of ducting, etc., that is common to both fans. It should be noted that alteration of any
damper in the non-common ducting will change the parallel unit characteristic rather than the external system
resistance line.
7.2 Unsatisfactory Operation of Parallel Fans
Not all fans have such simple characteristics as those in Fig. 10. For instance, parallel operation of two identical
forward curved blade fans is shown in Fig.12.
At the pressure Pv, there are three possible operating volumes for the single fan, namely Vx, Vy, and Vz. For the
single fan the system resistance line would determine the point of operation. However, these three points rive
rise to six possible points of operation for the parallel system. Three of them are simply 2Vx, 2Vy and 2Vz and
they form the main part of the characteristic for the parallel arrangement However, the three other possible
additions (Vx + Vy), (Vx + Vz ) and (Vy + Vz ) from off-shoots to the main curve.
It can be seen that, depending on the position of the system resistance line there will be one, two or three
possible operating points. For the system resistance line drawn in Fig.12 there will in fact be three possible
points of operation. The fans may settle down at any of these positions or oscillate between all three. C would be
the desired operating point, when both motors will develop the same load, whereas there will be progressively
more imbalance as the system moves to-point B and then point A. In practice both motors may cut out
immediately on start-up, the second fan because it is overloaded and the second fan because it then tries to run
at the point D.
7.3 Prediction of Unsatisfactory Parallel Operation
Hagen's method is a rather neat way of predicting the above unsatisfactory performance of identical fans. The
system resistance and total pressure single fan curve are drawn and, for any total pressure, the magnitude of the
term [ ]
pressure)
at that
flow
fan volume
possible
(maximum
-
flow)
volume
(system is superimposed. If the line
(Eagen's line) thus formed crosses the fan curve at more than one point then operation will be unsatisfactory
(Fig. 13).
7.4 Correction of Unsatisfactory Parallel Operation
This can be effected by fitting balancing dampers to each fan as shown in Fig.14. The system resistance line, D,
crosses the original parallel fan curve in three places leading to unsatisfactory operation. The balancing damper
and original fan curves are combined to give the curve (P-R) for each fan (the balancing damper bring
considered part of the fan) and when the curve is drawn for the two fans (P-R) in parallel it can be seen that the
system resistance line cuts it at only one point. Hence with careful adjustment of the balancing dampers it
should be possible to run the two motors with balanced loads, with only a small loss in volume flow.
8.0 BLADE TYPES
Basically, three types of impeller are used in centrifugal fans :
1. Forward Curved Blades (Blade angle > 90o)
2. Radial Tipped Blades (Blade tip = 90o)
3. Backward Sloping Blades (Blade anele < 90o)
Fig. 15 shows typical fan curves for each type. In general for a given impeller size, the pressure developed
increases with blade angle and higher volume flows can be obtained by use of a double inlet.
8.1 Forward Curved Blades
These fans give the highest pressure for a given size and speed. They are also used for high volume applications
such as ventilation, where their compact size and low speed (hence low noise) is an advantage. The low speed
also allows a lighter, less expensive construction to be used.
The blades are of short radial depth to reduce the inlet throttle and consequently more of them (30-60 compared
with 6-16 for other fans) are required to have an effect on the air. It can be seen (Fig.15) that a decrease in
restriction and hence an increase in airflow will cause an increase in power. The fan is said to be of the
overloading type and this effect may cause the motor to trip out.
These fans are unsuitable for dirty air, the narrowly spaced tips tending to bridge. Efficiencies are not greater
than 75%
8.2 RADIAL BLADED FANS
These are of two types
8.2.1 Radial Bladed Paddle Fans, e.g. firing fans.
These have their main application in handling high dust burdens. They are of simple construction and
therefore the wearing parts can be of substantial material or even replaceable. The impeller is
unshrouded but may have a back plate in the case of fibrous dust to prevent build-up on the shaft.
For corrosive applications, it is a simple matter to rubber-cover the impeller. Disadvantages of the
paddle fan are its low efficiency (about 605), high speed of rotation (noisy operation) and
overloading power characteristics.
8.2.2 Radial Tipped Forward Curves Fans
These fans combine some of the ruggedness of the paddle fan with the higher efficiency and higher
pressure development of the forward curve blade. The impeller is self-cleaning to a large extent and
is used extensively for dusty flue gases. Blades themselves are not easily made removeable but
replaceable wear plates and nose pieces can be fitted if dust burdens are high. Efficiencies higher than
75% are obtainable with this fan and it is therefore commonly used where high dust burdens prevent
the use of the more efficient aerofoil or backward inclined bladed fans.
8.3 Backward Bladed Fans
In the main, these fans are used for clean air applications, although the use of a reinforced nose, or
leading edge, will allow low concentration of dust to be handled. Efficiencies up to 90% are
obtainable with fans having aerofoil section blades (Fig.16) and therefore the fan is preferred for
the larger machines.
FIG.16 SECTICN THROUGH AEROFOIL
For a given duty the fan must run at a higher tip speed and the higher efficiency may be partially offset by the
capital cost involved in producing accurate aerofoil blading of sufficient robustness to withstand the increased
rotational stress.
The high efficiencies are a result of laminar flow over the blade profile and the reduction of eddies leads to
quieter operation. 'Backward bladed fans are nonoverloading (see power curve on Fig.15).
Summing up, the progressive use of fan types would be
Forward Curved
Backward Aerofoil
Backward. Inclined
Forward Curved Radial Tip
Paddle
as the dust burden increases.
9.0 FAN NOISE
The main source of fan noise is caused by the formation of shock eddies as the air passes through the blades of
the impeller and therefore quietness depends upon a smooth passage of the air. Fans generally operate most quietly
at their highest efficiency and dampering back may cause an increase in sound level. The outlet throat velocity does
not in itself control the noise level - the clearance between the impeller and the tongue being more important. The
rate of expansion of the air caused by the casing shape is also of importance. An expansion that is too rapid will
give a low outlet velocity but a higher noise level. The noise generated by the fan is transmitted to the casing both
upstream and downstream causing resonance in ducting. Fan noise can be reduced by speed reduction (if the fan is
already heavily dampered), duct and casing stiffening, or the fitting of inlet sound attenuators of the "Quietflow"
type. Appendix II gives a method of predicting octave band sound pressure levels from which a dB(A) rating can
be calculated.
10.0 OTHER PUMPING EQUIPMENT FOR GASES
Table 2 shows the applications of the various types of pumps available, though the figures should be taken as a
general guide only :
TABLE 2 - Gas Pumping Equipment
PUMP TYPE MAX. DELIVERY
PRESSURE
MAX.
THROUGHPUT
ft3/min (STP)
Reciproacting
compressors 4000 ats 5000 ats
Rotary Compressors:
High Compression
Ratio 75 psi 4000
Low Compression
Ratio 15 psi 5000
Centrifugal Types:
Fans 40" WG 150,000
Single Stage
Blowers 5 psi 12,000
Multistage Water
Cooled Compressors 150 psi 50,000
(10 stage)
10.2 Other Fans
Apart from the centrifugal fan, which is by far the most common in our industry, use is also made of Crossflow
and Axial flow fans.
10.2.1 Cross flow fans
Here the air enters and leaves radially (see Fig.17). This fan is used when a very low fan pressure is required
Such as in the domestic fan heater and some hair driers.
FIGURE 17 - CROSSFLOW FAN
10.2.2 Axial Flow fans
Here, as the name implies the air enters and leaves the fan axially. They are a development of the propeller fan
(e.g. Xpelair) which is only suitable for low pressures (i.e. free inlet and outlet). The modern axial fan, however,
is much more versatile and has the advantage that the motor may also be mounted on the axis of the duct giving
a very compact unit. The motor may be sited outside the gas stream by use of a bifurcated duct (Fig.18). The
axial flow fan is generally not used in dusty conditions since the blade tip clearance tends to bridge
FIGURE 18 - AXIAL FAN - EXTERNAL MOUNTING OF MOTOR
The circular motion of the fan gives the air leaving it a rotational component which tends to increase the
downstream pressure drop. For this reason guide vanes are installed either Upstream (in which case it is the fan
itself that straightens out the flow) or downstream. Two fans in series (often built as a single unit) are often
made contra-rotating so that guide vanes are not required. The air then leaves the second fan axially. The output
of axial fans can be changed by altering the pitch of the blades, usually when the fan is stopped, though in
specialized applications such in aircraft engines, the pitch may be altered while running.
Typical axial fan curves are shown in Fig.15.
10.3 Reciprocating Piston Compressors
These are the only machines capable of very high delivery pressures (100 pai). Their high efficiency must be
balanced against the lower capital cost of a rotary compressor.
10.4 Rotary Blowers and Compressors
High compression types include the sliding vane blower (Fig.19) and the Nash Hytor (Fig.20). The Hytor has an
elliptical casing, partially filled with fluid which is thrown to the outside of the casing by the impeller. The
spaces in vanes act as liquid pistons so that the air is sucked in, compressed and then discharged by the
advancing interface. The liquid may be chosen to be inert to the gas being pumped.
Neither of the above two machines are affected by tip wear. The sliding vane compressor has loose blades which
simply move radially outwards as they wear whilst it can be seen that the Hytor will function provided the blade
tips are always covered by liquid. For this reason these machines will maintain their efficiency over long
periods. Typical of the lower compression ratio type is the Rootes Cycloidal Blower (Fig.21 ), where
Increased wear and therefore back-leakage will reduce efficiency.
10.5 Centrifugal Compressors (Turbo-Blowers) -(Fig. 22)
These machines are used at anything from ½ psi to 100 psi and operate at high speeds (3500 to 30,000 rpm).
Water cooling may be employed in-between the stages and in-between casings which normally contain 6 or 7
stages.
The axial flow compressor is used for special applications (e.g. jet engines) where its lightness is an advantage.
It is a high efficiency machine but has a limited operating range, greater vulnerability to erosion and 00=03ion
and susceptibility to deposits.
Fig. 23 shows an axial flow compressor. The rotating element c6nsists of a single drum to which are attached
several rows off aerofoil blades, of decreasing height. Between each row is a stationary row which straightens
the airflow before it meets the next moving set of blades.
10.6 Vacuum Pumps
Here, a high compression ratio is necessary and therefore the machines used include the reciprocating piston,
rotary vane (Fig.24) and Nash Hytor, Piston and rotary pumps will give pressures down to 10-2
mm. Hg. but the
Hytor is limited by the vapor pressure of the sealing liquid (usually water).
11.0 POWER REQUIRED FOR COMPRESSION
This depends on the method of compression, which can be isothermal, adiabatic (no heat removed from
system)or something in-between the two. Isothermal compression requires the least work and is therefore the
most efficient so that the isothermal line on a P-V diagram should be approached as closely as possible, which
means that the maximum cooling between stages should be used.
Work dome on the gas in isothermal compression =








=
1
2
e
1
1
i
P
P
log
V
P
W
Work done on gas for non-isothermal compression =










−














−
=
=





 −
1
P
P
1
n
n
V
P
W
n
1
n
1
2
1
1
n
where, W = work done per unit mass of gas
P1 = inlet pressure
P2 = outlet pressure
V1 = initial volume of unit mass of gas
n = value of exponent in equation of compression
n
PV = constant
For adiabatic compression n = γ = ratio of specific heats. Compressor efficiencies (based on the
isothermal power required vary between 45 (low compression rotary compressors)
and 60% reciprocating piston compressor)).
12.0 FURTHER READING
Most fan manufacturers have published their own technical books, the most comprehensive being
1) "Fan Engineering" Buffalo Forge Company, Buffalo, New York. (1961).
A cheaper book, and perhaps a little more readable is
2) "Fans" W.C. Osborne Pergamon 1966
Of use for definitions is
3) British Standard 848 : Pt I ; 1963 (Fans)
APPEINDIX I
Effect of Speed and Temperature Change
Given the fan curves shown at 50o
C and 200 rpm what are the fan curves for running conditions of 0 o
C and
220 rpm.
Step 1
Calculate effect of speed and temperature on individual points of curve:-
Speed change factor = 10
.
1
200
220
=






Density change factor = 18
.
1
273
323
=






Pressure factor = ( ) 18
.
1
10
.
1
p
p
n
n 2
1
2
2
1
2
×
=
















= 1.43
Volume factor = 10
.
1
n
n
1
2
=








HP factor = 57
.
1
p
p
n
n
1
2
3
1
2
=
















Step 2
Pick out reasonably spaced points on the original fan curves as follows:-
Volume
(cfm x 1000) 0 10 20 30 40 50 60 70 80 90 100
Fan Total
Pressure
(ins Wg) 9.7 9.8 9.9 10.0 9.7 9.2 8.2 6.8 5.1 3.4 1.5
HP 45 50 62 76 87 96 99 99 97 93 80
Step 3
Multiply by the appropriate factors to obtain the new fan curves:-
Volume 0 11 22 33 44 55 66 77 88 99 110
PT 13.9 14.0 14.2 14.3 13.9 13.1 11.7 9.7 7.4 5.0 2.1
HP 71 78 97 119 137 149 155 155 152 146 126
APPENDIX II
Prediction of Fan Noise Levels : Graham and Christie Method (1966)
PWLB = PWLs + 70 Log (D) + 50 Log (S) - 259
PWLB = Octave band power level (re 10-12
watt)
PWLs = Specific power level in each octave band (re 10-12 watt)
D = Impeller diameter (ins)
S = Fan Speed (rpm)
PWLs is shown in Table 3.
5 dB should be added to the octave band containing the blade passage frequency to account for rotational noise.
This frequency is given by f = N x S/60 cycles/see
where N = number of blades
APPENDIX III
Effect of Water Injection on Fan Performance
Water injection into an air stream before a fan will cool the gas and the effect will be an increase in density, the
cooling effect being much larger than density reductions due to the lower molecular weight of steam. The fan
curve will be changed in accordance with the fan laws, the denser gas resulting in the fan doing more work. In
fact this effect is so great that water injection will result in an increased mass flow of gas measured at a point
upstream of the injection point.
As an example, the original Hope Preheater fan curve has been taken. The system has been simplified to that in
Fig. 26.
Example
Fan curve as given (Fig.27)
Initial fan gas flow = 4900 m3/min at 350
o
C
Gas density = 0.64 kg/m3
Gas molecular weight = 32.7
Water injection rate = 1820 gph = 137 kg/min
Latest heat of steam = 540 cal/g
Specific heat of steam = 0.5 cal/g
o
C
Initial water temp. 15
o
C
Problem
Given the above information what will the increase in kiln gas flowrate be on water injection ?
Step 1
Find temperature after injection
Mass of gas = 4900 x 0.64 = 3140 kg/min
Let new temp = T
o
C
Heat balance:-
3140 x 0.25 (350 – T) = 137 (540 + 85(l) + (T - 100) 0-5)
T = 230
o
C
Step 2 : Calculate new gas density
Total gas flow = 3140 kg/min at MW 32.7
Plus 137 kg/min at MW 18
= 3277 kg/min at MW 31.6
New density = 0.64 x =






×






503
623
7
.
32
6
.
31
0.766 kg/m
3
Step 3 : Modify fan curve
Original fan curve :-
Volume 0 1 2 3 4 5 6 7
Pressure 612.5 665 725 745 725 675 600 510
Power 290 410 525 630 720 790 840
Density ratio = 195
.
1
64
.
0
766
.
0
=






Therefore, volume ratio = 1.0
Pressure ratio = 1.195
Power ratio = 1.195
New fan curve:-
Volume 0 1 2 3 4 5 6 7
Pressure 732 795 868 891 868 808 718 610
Power 348 490 630 755 860 945 1010
Step 4
Here, we correct the new fan curve so that the volume it applies to are at the conditions upstream of the water
injection point. This is done because the system resistance line we draw in will refer to volumes at these
conditions.
The calculation is carried out by considering successive volumes of gas at 350
o
C reaching the injection point.
e.g. Volume flow = 3000 m
3
/min at 350
o
C
Therefore, mass flow = 3000 x 0.64
= 1920 kg/min
Water required to cool.this to 230
o
C
= 137 x 





3140
1920
= 83.8 kg/min
Therefore, total mass of gas reaching fan
= 83.8 + 1920 = 2003.8 kg/min
Volume of gas at fan = =






766
.
0
8
.
2003
2616 m
3
/min
Fan static pressure at this fan volume (from curve l)
= 887 mm wg
The following table is completed in the same way:-
Volume approaching Volume.Reaching Fan Static
Injection point (m
3
/min) Fan (m
3
/min) Pressure (mm wg)
0 0 732
1000 872 787
2000 1744 846
3000 2616 887
4000 3488 885
5000 4360 850
6000 5232 735
7000 6104 705
A plot of the first column against the last column is then the requiread curve (curve 1K)
Step 5 : Determine System Resistance Line
Assume P = kV2
(P = 680, V= 4900) is the original, operating point
Hence system resistance line is:-
Volume 4 5 6 7
Pressure 453 707 1019 1385
Then the new operating point will be at the intersection of the S.R.L. and the new fan curve.
i.e. 5400 m3
/min
(intersection with curve 1K)
Step 6 : Determine gas flows
Kiln gas mass flow = 5400 x 0.64
= 3460 kg/min
Therefore, increase in gas flow = 3460 - 3140
= 320 kg/min
However, we now reach a contradiction in the calculation. if the kiln gas flow increased to 3460 kg/min,
then the addition of 137 kg/m, in of water would not quite cool it down to 230o
C. Unfortunately, 230o
C
is the temperature upon which we have based our fan curve calculations. To cool the extra kiln gas down
to 230o
C we would need a further
137 x =






3140
320
14 kg/min of water
giving a total injection rate of 151 kg/min
In other words we have assumed an injection rate of 137 kg/min but found that we in fact need 151
kg/min to cool the gas down to the temperature at which we have made our calculations.
Step 7
What we must do now is to make a better guess' at a water rate that we may assume to carry out our fan
curve calculation, so that the total water rate, calculated at the end of Step 6, 13 equal to the true water
rate.
This may be done by proportion
New guess = 137x =






151
137
124 kg/min
We then use this value, 124, in place of 137 kg/min and repeat Steps 1-6.
Step 5 then gives us an operating point of 5350 m3
/min (intersection with curve 2K) = 3420 kg/min.
Extra gas = 3420 - 3140
= 280 kg/min
Water required to cool extra gas down to new temperatures of 240o
C
=124 x =






3140
280
11 kg/min
Therefore total water flow = 124 + 11 = 135 kg/min
This is sufficiently close to the true water rate (137) and further iteration is not necessary.
Power Drawn
When the operating point is at 5350 m3
/min (curve 2K) the fan will handle 4700 m3
/min (curve 2). At this
volume the fan shaft power will be 820 kW (curve 2P), whereas prior to water injection the fan power was 715
kW. Hence we may draw up the following table:-
Before Af ter Increase
Water Water
Injection Injection
Kiln gas flow (kg/min) 3140 3420 8.9%
'Water rate (kg/min) - 135
Fan Temp (00 350 240 -
Fe.n Power (kW) 715 820 14.7%
Water injection, therefore, may in some cases be used as a means of uprating the useful Volume handled by a
fan.
If the extra gas that is drawn were not required one would, of course, damper back along the curve 2K, or reduce
the fan speed until the original kiln gas flow was attained.
Blue Circle Cement
PROCESS ENGINEERING TRAINING
PROGRAM
Module 5
Section 2
Fans and Airflows
FANS AND AIRFLOWS
1. Principles of Fan Performance
2. Terms Used In Fan Engineering
3. Fan Blade Types
4. Fan Curves
5. Control of Fan Output
6. Unsatisfactory Fan Performance
Appendix I The Fan Laws
Appendix II Fan Checks
FANS AND AIRFLOW
1. PRINCIPLES OF FAN PERFORMANCE
As a first step to an understanding of fan performance it is necessary to get a clear conception as to the
manner in which a fan actually operates.
The purpose of a fan is to establish and maintain between the fan inlet and fan discharge a difference in
pressure of the air or gas it is handling, and it is in consequence of this maintained difference in pressure
that, circumstances permitting, the volume of air or gas flows through the external circuit to which the
fan is connected.
The fan has to set up a difference in pressure sufficient to overcome the losses in the system, including its
own internal losses.
2. TERMS USED IN FAN ENGINEERING
2.1 Energy
When considering fan problems it is seldom that the energies imparted to the air are referred to in terms
of potential and kinetic energy. Instead the term water-gauge is used (usually expressed in inches).
a) Potential Energy
The air within a duct system connected to a fan will be at a different pressure to the atmosphere
outside the duct and will have a capacity for doing work in virtue of that difference in pressure,
i.e.,'it will have potential energy.
b) Kinetic Energy
The air will be in motion and will have energy by virtue of that motion, i.e., it will have kinetic
energy.
c) Total Energy
The total energy is the algebraic sum of the potential and kinetic energies.
In all fan work the pressure of the atmosphere surrounding the fan at sea level is taken as the datum, and
pressures are considered positive or negative, according as they are above or below atmospheric pressure.
Therefore, since atmospheric pressure is taken as a datum and since the fan produces a difference in
pressure it follows that the air only possesses potential and kinetic energy by reason of being connected to
the fan, and therefore the energy must be supplied by the fan.
Within limits of the total energy in the air, the ratio of the amount of energy present as potential to the
amount of energy present as kinetic to the amount present as kinetic is governed entirely by the external
system to which the fan is connected.
This can be summarized as a fundamental principle:
“The resistance of the external system controls the volume which the fan can handle at a fixed speed."
2.2 Pressure
Fig. I shows diagrammatic sketches of fan pressure terms.
d) Static Pressure
Static pressure is the bursting or collapsing pressure according to whether the difference in pressure
between the inside and outside of the duct is positive or negative, that is, above or below atmosphere.
Thus the static pressure is a measure of the capacity of the air within the duct for doing work, and is
therefore a measure of the potential energy.
e) Velocity Pressure
Velocity pressure is a measure of the kinetic energy which the air possesses by virtue of its motion can be
measured by the Pitot Tube.
f) Total Pressure
Total pressure is defined as the algebraic sum of the static pressure and velocity pressure.
Since the velocity pressure can only be measured by impact it must always be a positive quantity, whether
the flow be on the discharge side of the fan, where the static pressure is positive, or on the suction side
where the static pressure is negative.
It follows, therefore, that upon the discharge side of the fan the total pressure will be numerically greater
than the static pressure, and upon the suction side of the fan the total pressure will be numerically less
than the static pressure.
g) Water-Gauge
The pressures to be measured are usually quite small and it is usual to refer the pressures to the height of a
column of water which could be sustained, hence the term "water-gauge".
In fan work, since the pressures generated are so small, the variations in air volume, due to compression or
expansion, are also quite small, and such variations are usually neglected. Variations in temperature and
barometric pressure, however, produce considerable -changes in volume and air density.
(A cubic foot of water weighs 62.4 lbs at 60 o
F so that a cube of water with 12" sides would exert a
pressure of 62.4 lbs over the area of its base which is one square foot, so that a water-gauge of 12"
corresponds to a pressure of 62.4 lbs per square foot, or 1" water-gauge represents 5.2 lbs per square foot.
Hence to transform a pressure expressed in terms of inches water-gauge to a corresponding pressure in lbs
per square foot the water-gauge is multiplied by 5.2).
2.3 Manometers
The apparatus which actually indicates the water-gauge is known as a manometer (Fig. 2). For use on site
it generally consists of a plain glass U tube, one leg of which is connected to the duct where the pressure
is to be measured, the other leg being open to atmosphere. The pressure is then obtained by the difference
in the height of the water in the two legs of the tube.
The correctness of the readings of a manometer will be dependent upon the accuracy with which the
difference in level of the height of the columns of water in the two legs of the manometer is measured. It
is essential that the U tube be held in a vertical position. The difference in the levels of the water can then
be measured carefully by means of a scale attached to the gauge.
The water in the tube will tend to cling to the surface of the tube, and hence the surface of separation
between the air in the tube and the water will be cup-shaped, the smaller the bore of the tube the more
curved will be the surface of the water. If is usual to measure to the bottom of the cup-shaped surface (or
meniscus as it is called). For accurate work the cup-shaped surfaces must be the same, i.e. the tubes must
be of the same bore. It is usual to specify that the bore of such a U tube shall not be less than, say, 3/8”
diameter.
2.4 Measuring Fan Static Pressure
Fan total pressure PT, is the difference between the total pressures at the fan inlet and outlet.
Fan static pressure, PS, is the fan total pressure minus the fan velocity pressure.
Fan velocity pressure, PV, is the velocity pressure corresponding to the average velocity at the fan outlet.
i.e. the fan static pressure is NOT equal to the difference between the static pressures at the inlet and
outlet. It should be measured as the difference between the reading of a facing tube at the inlet and a side
tube at the far outlet (see Fig. 3).
3. FAN BLADE TYPES
Basically, three types of impeller are used in centrifugal fans:
a) Forward-Curved Blades (Blade angle 90o).
b) Radial Tipped Blades (Blade tip 90o).
c) Backward Sloping Blades (Blade angle 90o).
Fig. 4 shows typical fan curves for each type. In general for a given impeller size, the pressure developed
increases with blade angle.
3.1 Forward Curved Blades
These fans give the highest pressure for a given size and speed. They are used for high volume
applications such as ventilation, where their compact size and low speed (hence low noise) is an
advantage. The low speed also allows a lighter, less expensive construction to be used.
The blades are of short radial depth and consequently more of them (30 - 60 compared with 6 - 16 for
other fans) are required to have an effect on the air. It can be seen (Fig. 4) that a decrease in restriction
and hence an increase in airflow will cause an increase in power. The fan is said to be of the overloading
type and this effect may cause the motor to trip out.
These fans are unsuitable for dirty air, the narrowly spaced tips tending to bridge.
Efficiencies are not greater than 75%.
3.2 Radial Bladed Fans
These are of two types:
3.2.1 Radial Bladed Paddle Fans e.g. firing fans
These have their main application in handling high dust burdens. They are of simple construction and
therefore the wearing parts can be of substantial material or even replaceable. The impeller is
unshrouded, but may have a back plate in the case of fibrous dust to prevent build-up on the shaft.
Disadvantages of the paddle fan are its low efficiency (about 60%), high speed of rotation (noisy
operation) and overloading power characteristics.
3.2.2 Radial Tipped Forward Curved Fans
These fans combine some of the ruggedness of the paddle fan with the higher efficiency and higher
pressure development of the forward curve blade. The impeller is self-cleaning to a large extent and is
used extensively for dusty flue gases. Blades themselves are not easily made removable but replaceable
wear plates and nose pieces can be fitted if dust burdens are high.
Efficiencies higher than 75% are obtainable with this fan and it is therefore commonly used where high
dust burdens prevent the use of the more efficient aerofoil or backward inclined bladed fans.
3.3 Backward Bladed Fans
These fans are used mainly for clean air applications, although the use of a reinforced nose, or leading
edge, will allow low concentration of dust to be handled.
Efficiencies up to 90% are obtainable with fans having aerofoil section blades, and therefore the fan is
preferred for the larger machines. The high efficiencies are a result of laminar flow over the blade profile.
Backward bladed fans are non-overloading (see power curve on Fig. 4).
3.4 Blade Section
Summing up, the progressive use of fan types as the dust burden increases would be:-
Forward Curved
Backward Aerofoil
Backward Inclined
Forward Curved Radial Tip
Paddle
4. FAN CURVES
These show the effect on fan pressure, power and efficiency of a change in fan throughput (Fig. 5). The
horse-power on a fan curve is that which must be delivered to the shaft and does not take into account
losses in couplings and motors.
P may be either the fan static or total pressure and the efficiencies thus calculated are the static and total
efficiencies respectively.
4.1 System Resistance
The resistance of the system served by a fan depends upon the hardware in it and for any system there
will be a relationship between flowrate and pressure drop. If the system contains only dampers, ducting,
expansions etc., of fixed resistance, then the pressure drop may be taken as being approximately
proportional to the square of the flowrate (V2). This would not be true in the case of a resistance (as
opposed to pressure drop), that varied with flowrate, such as a cooler grate. The intersection of the system
resistance line with the fan pressure/ f lowrate curve gives the point of operation of the fan. The effect of,
for instance, closing a damper is to steepen the system resistance line, giving a new operating point (see
Fig. 6).
5. CONTROL OF FAN OUTPUT
For a centrifugal fan there are three basic methods of controlling output-
a) Damper Control.
b) Variable Radial Inlet Vanes - these give a beneficial swirl to the air at the fan inlet and therefore
reduce the work the fan has to do.
c) Speed Control.
Initially, dampers are the cheapest to install and speed control the most expensive, but there is a reversal
in a running cost comparison. However, speed control usually only becomes economic on the larger
machines.
Whereas variation in a damper setting will change the system resistance and not the fan curve, the other
two methods produce new fan curves, that for speed control retaining the same fan efficiency.
Referring to Fig. 7, the effect of turndown to ¼ V and ½ V by the three methods is shown.
Graph A - Damper Control - Fan curve retained but system resistance line altered.
Graph B - Variable Inlet Vanes - New fan curves produced for each vane setting.
Graph D - Comparison of three methods on basis of power taken. (Note: loss due to coupling is
exaggerated.)
Graph C - Speed Control - New fan curves produced for each speed but efficiency retained.
This graph assumes no pressure drop across a fully open damper. This is a reasonable assumption for
some dampers operating in clean air, but would not be true for some ID fan dampers where build-up will
occur. The expected pressure drop across an open damper, based on experienced, should be taken into
account when deciding on the method of control.
6. UNSATISFACTORY FAN PERFORMANCE
Fan performance may be affected by worn impellers, inefficient drives, etc., but the most common causes
of fan performance being substandard are bends close to the inlet or outlet of the fan.
If the bend is at the inlet, the air is pushed over to one side of the eye, whereas a manufacturer's fan curve
is for uniform inlet conditions. Uniformity of flow at inlet is easily checked by a pitot traverse.
The best solution to this is an inlet box which can be sized so that the air enters the impeller uniformly,
failing that a square bend with cascade (Fig. 8) will give good performance. Retention of the original
bend and fitting of long turning vanes will improve matters greatly (Fig. 9).
Useful pressure in a fan originally comes from the acceleration of the air. As the air is again slowed down
by expansion, velocity pressure is converted into static pressure. For this to be done efficiently the
expansion should be slow and gentle and clearly a bend immediately on the fan outlet will not help
matters. This problem is not so easy to remedy but again a cascade will help.
APPENDIX I THE FAN LAWS
The performance of a fan may be changed by several methods, the main ones being a change in:-
a) Fluid density, p.
b) Impeller speed, n (rpm).
c) Fan size*. Most text books only quote the fan laws for “fans of geometric similarity", so that diameter,
width, outlet duct size, etc., are all increased at the same rate. Any convenient dimension may then be
taken as being representative of the fan, e.g. Impeller Diameter, Dim.
The effect of these on fan pressure, volume and horse-power is summarized in Table I.
TABLE I - The Fan Laws
Alteration In
Change In
Fan Pressure Volume Power
Density, p α p No change α p
Speed, n α n2 α n α n3
Fan Size*, Dim α Dim
2 α Dim
3 α Dim
5
Fan pressure may be either total, static or velocity.
In order to help remember the above table, it is worth noting that fan pressure can be shown to be theoretically
proportional to the square of the tip speed whereas volume is proportional to the swept volume of the impeller
(fan size/volume is a cubic relationship since width increases as well as diameter). Power is proportional to the
product of volume and pressure.
APPENDIX II FAN CHECKS
The following is given as a guide to minimize losses in fan & associated plant efficiencies due to
mechanical/process deteriorations of fans.
1) Check to ensure that fan inlet & discharge ductwork is not restricted with material drop out ( or plates, etc.
left inside following major repairs.)
2) Check to ensure that control dampers are functioning correctly, indicating correct position externally ( eg.
with louvre type dampers that all louvres can move freely, and together
3) Check to ensure fan speed is correct ( eg. on belt drives check belts are not slipping. Check pulley sizes are
correct).
4) Check fan power absorbed at motor.
5) Check fan inlets to ensure.
a) At entrances no material deposits exist.
b) At fan eye that there exists a close concentric fit of duct to impeller, to eliminate recirculation, within the
fan casing.
c) Within casing, no wear exists at nose bridge causing turbulence.
6) Check to eliminate all inleaks at duct Joints, fan inlet joints, doors, etc., on fan suction side.
7) Check to verify fan impeller dimensions ( dia. etc. ) are as designed.
8) Check to ensure there is no wear on fan blades ( and repair as necessary).
9) General inspections inside fan casing are necessary to ensure no obstructions exit, or wear of body, that
turbulence could be created.
The above checks are necessary at shutdowns, particularly at the beginning to assess work/repairs necessary
and also before start up to ensure correct repairs have been carried out, & no extraneous materials have been
left inside casings & ductwork.
In addition to the above physical condition checks, Technical Departments should carry out process fan
efficiency checks during normal stable fan operation. This is essentially to verify that the fans are operating on
their fan curves ( ie. as designed ) and therefore corrective measures can be planned/ programmed for
shutdowns if necessary.
To check fan curve operation the following measurements are necessary.
1) Gas volume handled by fan, by anemometer or pitot tube measurements as applicable.
( Care should be taken to maximize accuracy of such measurements ).
2) Fan total pressure & static pressure.
( Care Should be taken to maximize accuracy and also use of this data according to fan static pressure as defined
previously).
3) Fan impeller R.P.M.
4) Fan motor power. ( Note curve power is shaft power only)
The above data can then be compared with curve specifications using the fan laws.
Note coupling/ralley /belt efficencies compared to motor power mean that motor power is greater than fan shaft
( curve ) power.
If the operating point of the fan is found to be significantly off the design curve, it is suggested that the test is
repeated, for consistency, & than corrective actions/inspections can be programmed as necessary.
1.0 DEFINITION OF FAN PRESSURE
Fan total pressure, PT, is the difference between the total pressures at the fan inlet and outlet
Fan static pressure, PS, is the total pressure minus the fan velocity pressure.
Fan velocity pressure, PV, is the velocity pressure 







g
u
2
2
corresponding to the average velocity at the fan outlet.
i.e. the fan static pressure is NOT equal to the difference between the static pressures at the inlet and outlet. It
should be measured as the difference between the reading of a facing tube at the inlet and a side tube at the fan
outlet (see figure 1).
These show the effect on fan pressure, power and efficiency of a change in fan, throughput (Fig.2). The
horse-power on a fan curve is that which must be delivered to the shaft and does not take into account losses in
couplings and motors.
A calculation of the work actually done on the air will therefore give the efficiency of the fan itself.
Work done on air = pressure x flowrate x 





000
,
33
2
.
5
)
fm
and
wg
(inches
HP
6360
c
PV






=
or 9.81 PV watt (mm wg and m3/sec).
P may be either the fan static or total pressure and the efficiencies thus calculated are the static and total
efficiencies respectively.
Because work is being done on the air, its internal energy will rise, and this, together with losses due to the
inefficiency of the fan will cause the air temperature to rise.








×
×
=
heat
specific
V
density
power
fan
rise
e
Temperatur
For air at ambient conditions
( )
0.24
V
075
.
0
779
5.2
PV
×
×






×
×
=
η
efficiency
fan
total
=
η
P
0.37 







=
η
o
F
or about ½ o
F per inch wg.
3.0 THE FAN LAWS
The performance of a fan may be changed by several methods, the main ones being a change in:-
a) fluid density, p
b) impeller speed, n (rpm)
c) fan size* most text books only quote the fan laws for “fans of geometric similarity”, so that diameter,
width, outlet duct size, etc., are all increased at the same rate. Any convenient dimension may then be
taken as being representative of the fan, e.g. Impeller Diameter, Dim.
The effect of these on fan pressure, volume and horse-power is summarized in Table I.
TABLE I - The Fan Laws
ALTERATION IN
CHANGE IN FAN
PRESSURE
VOLUME POWER
Density, p α p No change α p
Speed, n α n 2 α n α n3
Fan Size*, Dim α Dim
2 α Dim
3 α Dim
5
Fan pressure may be either total, static or velocity. In order to help remember the above table, it is worth noting
that fan pressure can be shown to be theoretically proportiona1 to the square of the tip speed whereas volume is
proportional.
TULSA FAN Number 2 COMPARTMENT
1750 RPM 1284 RPM 1687 RPM
0.7337 0.5383 0.3950 0.964 0.9293 0.8958
Vol. I
cfm x
1000
Press I
“ w.g.
HP 1
B. H. P.
Vol. 2
cfm x
1000
Press 2
“w.g.
HP 2 Vol. 3
cfm x
1000
Press 3
“w.g.
HP 3
0 17.6 12 0 9.47 4.74 0 16.36 10.75
5 18.8 32 3.67 10.12 12.64 4.82 17.47 28.67
10 19.5 47 7.34 10.50 18.56 9.64 18.12 42-10
15 19.0 63 11.01 10.23 24.88 14.46 17.66 56.44
18 18.0 71 13.21 9.69 28.04 17.35 16.73 63.60
20 17.2 76 14.67 9.26 30.02 19.28 15.98 68.08
25 13.8 92 18.34 7.43 36-34 24.10 12.82 82.42
30 9.3 104 22.01 S.01 41.08 28.92 8.64 93.17
35 3.2 114 25.68 1.72 45.03 33.74 2.97 102.12
Blue Circle Cement
PROCESS ENGINEERING TRAINING
PROGRAM
Module 5
Section 3
Fans and Fan Systems, Chemical Engineering
March 1983
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Blue Circle Cement
PROCESS ENGINEERING TRAINING
PROGRAM
Module 5
Section 4
Fans
FL Smith
FANS
Abstract
In this lecture the most important fan types used in cement plants are presented and their advantages and
disadvantages are outlined.
The definition of fan efficiency and simple fan calculation methods are discussed and it is demonstrated
how to save energy by optimum fan regulation and modification of existing fans with high efficient
impellers.
Table of contents:
1. Introduction
2. Fans functioning principles
3. Centrifugal fan types used in the cement industry
3.1 Single suction type with open overhanging impeller
3.2 Single suction type with closed overhanging impeller
3.3 Single suction type with closed impeller suspended between the bearings
3.4 Double suction type with closed impeller
4. Pan calculation and dimensioning methods
4.1 Basic definitions
4.2 Theoretical fan power and efficiency
4.3 Model fan curves
4.4 Calculations assuming imcompressible flow
4.5 Fan dimensioning assuming imcompressible flow
4.6 Calculations considering gas compressibility
4.7 Correction for dust laden gas
5. Fan drives
5.1 Direct/indirect drive
5.2 Motor type and speed regulation
5.3 Drive efficiencies
6. Capacity regulation
6.1 Damper regulation
6.2 Inlet guide vanes
6.3 Fan speed regulation
7. Fan modifications
7.1 Impeller replacement
7.2 Energy savings
Fans
1. Introduction
Fans are used in all parts of a cement plant and the operational success of nearly all main machinery in the
cement plant is completely dependent on the reliability and capacity of the connected fans.
The specific electrical energy consumed of the total number of fan motors included in a modern cement
production unit normally amounts to 30-40 kWh/t clinker which makes the fan motors the biggest power
consumers at the cement plant after the mill motors.
Thus, it is very important that the fans included in a cement plant are very carefully selected and
dimensioned, and the plant operators must know how to operate the fans most efficiently.
2. Fans functioning principles
Very broadly speaking the function of a fan is to create a flow of gas or air in the system to which the fan is
connected. In the fan the necessary energy to create the desired flow through the system is transferred to the
air or gas by a rotating impeller driven by the fan drive (f.inst. an electrical motor).
Fans are normally characterized as either axial or centrifugal fans. In an axial fan the flow is predominantly
axial, that is in the direction of the impeller axis whereas the flow in the centrifugal fan is predominantly
radial (that is perpendicular to the impeller axis) in the region where the energy transfer takes place.
In the cement industry centrifugal fans are the most commonly used and in the following we will therefore
only deal with centrifugal fans though much of the information, formulae etc. may also apply to axial fans.
3. Centrifugal fan types used in the cement industry
A centrifugal fan can be categorized according to its various design details such as: inlet design, impeller
type, blade type, impeller position etc., see table 3.1.
Table 3.1: Design details which characterize a centrifugal fan.
Impeller type: Open, paddle type
Open with back-plate
Closed (shrouded)
Inlet type: Single axial inlet
Single, with inlet box
Double, with inlet boxes
Blade form: Radial
Forward curved
Straight backwards inclined
Curved backwards inclined
Air foil backwards inclined
Impeller position: overhanging
Between bearings
Bearing types: Roller bearings, grease lubricated
Roller bearings, oil lubricated
Journal bearings
Other features: Inlet vanes
Wear protection
Shaft cooling
Further, fans are often characterized according to their function or position in the system in which they
create a flow of gas or air, namely as:
1. Blowers, i.e. fans which supply (fresh) air or gas to a system
2. Exhausters or induced draught (i.D.)fans, i.e. fans which remove air or exhaust gas
from a system
3. Circulation or recirculation fans are fans which maintain a closed loop air or gas flow
in a system.
In the cement industry the fans are often termed by the machine system to which they are connected, f.inst.:
1. Kiln I.D. fan
2. Raw mill fan
3. Cement mill fan
4. Kiln by-pass fan
5. Precipitator or filter fan
6. Dedusting fan, etc.
See also fig. 3.1.
Sometimes fans are also categorized as either low pressure, medium pressure or high pressure fans,
according to the differential pressure rating of the fan. However, the pressure range for the different
categories has not been standardized.
In the following, four fan types commonly used as kiln I.D., and/or mill circulation fans in the cement
industry will be presented and the advantages and disadvantages of the individual types will be
explained.
3.1 Single suction type with open overhanging impeller
Fig. 3.2 shows a centrifugal fan with open impeller. The air or gas is introduced in the fan in axial
direction, proceeds radially through the impeller and leaves the fan through a tangential outlet in the fan
casing (for the terminology used here,,-please refer to the legend of the figure).
The impeller has bolted on plane, radial blades which can be changed when worn or major changes in
the fan performance are required.
The impeller is overhanging, that is both shaft bearings are placed on the same side of the fan casing.
This provides for the use of a simple axial inlet to the fan, but the load on the bearing closest to the
impeller is very big and the bearing has to have a diameter corresponding to maximum shaft diameter.
Normally this fan type is designed for a fairly low specific speed, that is for a given rating the fan will
be designed with a relatively large impeller running with a low rotational speed.
The advantage of this type of fan is that with the radial blades the fear of dust accumulation on the
backside of the impeller blades is eliminated and the low rotational speed makes the impeller less
exposed to wear from dust contained in the gas being conducted through the fan. Further, should the
fan blades get worn, they are easily replaceable.
on the other hand this fan type exhibits a very low efficiency (65-70%) and the front side of the radial
blades are often exposed to hard coating formation when operating with kiln exhaust gas.
Due to its low efficiency this type of fan will probably only be used in special cases in the future, f-inst.
when the gas to be transported contains high amounts-of abrasive dust.
Fig. 3.2: Single suction fan with open overhanging paddletype impeller
1. Fan outlet 5. Blade 9. Thermal shield 12. Bearing (locating bearing)
2. Fan casing 6. Access door 10. Shaft guard 13. Shaft
3. Fan inlet 7. Cooling disc 11. Base frame 14. Foundation
4. Impeller 8. Beating
3.2 single suction type with closed overhanging impeller
Fig. 3.3 shows a centrifugal fan with closed overhanging impeller. The impeller blades extend between the two
impeller shroud plates also termed the cover plate and the back plate.
Thus, the gas or air flows through the impeller in "closed" channels.
This design allows for a wide variation of the blade form and inclination (angle with the impeller tangent).
Normally backwards inclined blades are used as this ensures a higher fan efficiency than radial and forward
curved blades.
The fan in fig. 3.3 has relatively short, moderately backwards inclined S-shaped blades. This blade form
increases the impeller stability, and at the same time reduces the impeller surface area exposed to dust
accumulation. The efficiency of this fan type is approx. 78%.
Fig. 3.3: Single suction fan with shrouded overhanging impeller with S-formed blades.
1. Fan Outlet 5. Blade 8. Bearing 11. Base Frame
2. Fan Casing 6. Access Door 9. Shaft guard 12. Shaft
3. Fan Inlet 7. Cooling Disc 10. Bearing (Locating) 13. Foundation
4. Impeller
3.3 Single suction type with closed impeller suspended between the bearings.
For a given tip velocity an impeller with backwards inclined blades exhibits at lower pressure rating than
impellers with radial or forward curved blades. Thus, a high pressure rating affords a very high impeller tip
speed for which it is difficult to design the fan shaft and bearings in case the impeller is overhanging.
Thus, for high pressure fans with backwards inclined blades the impeller is normally placed between the shaft
bearings.
Fig. 3.4 shows a high efficient fan designed according to this principle.
Fig. 3.4: Single suction fan with shrouded impeller with backwards curved blades suspended between the
bearings.
1. Fan outlet 5. Blade 9. Shaft
2. Fan casing 6. Cooling disc 10. Bearing bracket
3. Fan inlet 7. Beating (locating) 11. Access door
4. Impeller) 8. Thermal shield 12. Foundation
As the bearing on the inlet side of the fan makes a simple axial gas inlet impossible the fan is provided with
an inlet box integrated with the fan casing into which the gas flows radially.
The fan shown in fig. 3.4 has backwards curved blades with a tip angle of 50
o
. The maximum efficiency of
this fan type exceeds 80% despite the inevitable pressure loss in the inlet box. It is designed for an impeller
tip speed up to 180 m/s.
The major draw back of this fan type is that when working with dust laden gas it is often exposed to soft
coating formations on the back side of the impeller blades and/or wear of the impeller due to its high
rotational speed. However, various solutions to remedy these draw backs are offered, such as pressurized air
cleaning systems, hard facing of the impeller blades etc.
3.4 Double suction type with closed impeller
For very large flow capacities double suction fans are used, see figure 3.5.
A double suction impeller can be considered as two (laterally reversed) single suction impellers placed on the
same shaft and having a common back plate. The impeller is placed in a fan casing having an inlet box on
each side. Thus, the double suction fan has two inlet ports, which makes the inlet ducting to such a fan a little
more complicated, but on the other hand the capacity of the double suction fan will correspond to two similar
single suction fans with the same impeller diameter, and thus the double suction fan will often be the most
economical solution for large capacities.
Some fan suppliers provide single suction as well as double suction fans with backwards inclined airfoil
blades. See figure 3.6.
With this blade form a very rigid rotor is obtained even with a small plate thickness in the blades. This blade
form also ensures a very high fan efficiency (up to 85%). on the other hand the thin blades will naturally be
very sensitive to wear.
Fig. 3-5: Double suction fan.(Deutsche Babcock)
Fig. 3.6: Cross section of impeller with airfoil blades (Solyvent)
4. Fan calculation and dimensioning methods
In this section some basic definitions and "fan laws" are presented and methods for "manual" calculations of
fans are outlined.
Normally, fans are dimensioned using a number of computer programs available at the fan suppliers, and the
purpose with the discussion here is only to give an idea of the principles used for dimensioning of large fans.
Further the importance of distinguishing between static and total pressure rating and different fan efficiency
definitions is emphasized.
4.1 Basic definitions
The operation of a fan (working with a clean gas) can be characterized by:
'
1
Q = the actual flow of gas to the inlet of the fan.
'
1
ϑ = the specific density of the gas of the inlet to the fan and
'
t
p
∆ = the fan pressure rating expressed as the difference between the total pressure of the gas in the fan
outlet and the fan inlet, respectively:
1
t
2
t
t p
p
p −
=
∆
Refer also to fig. 4.1.
Fig. 4.1 Nomenclature
When using small p's for pressure we refer to the gauge pressure that is the pressure stated relatively to the
ambient pressure, B (barometric pressure), at the elevation at which the fan is placed.
The absolute gas pressure, corresponding to the relative pressure p, is
P = p + B
(p being stated as a negative pressure when below the ambient pressure)
Pressure is measured in a great number of units. In the following mainly the SI unit Pascal is used (1 Pascal =
1 N/m2 = 0.01 mbar = 0.102 mm WG).
The total pressure at a certain point in a flow of gas is the sum of static pressure, ps, and dynamic pressure, pd,
of the gas at the apoint considered
pt = ps + pd
pt is the pressure measured through the center hole in a pitot tube directed against the gas flow direction and p
is the pressure measured through the side holes of the pitot tube (perpendicular to the flow direction).
The dynamic pressure is calculated as pd = pt - ps = ( ) 2
u
2
1
ϑ






in which
ϑ is the gas density and
u is the gas velocity.
Often it is the static pressure rise ∆ps = ps2 – ps1
across the fan which has the interest.
This can be calculated from the total fan pressure rating ∆pt as follows:
( )( ) ( )( )2
1
1
2
2
2
t
s u
2
1
u
2
1
p
p ϑ






+
ϑ






−
∆
=
∆
( ) ( )
2
1
1
1
2
2
2
2
t
A
Q
2
1
A
Q
2
1
p 







ϑ






+








ϑ






−
∆
Herein Q1 and Q2 designate the flow through the inlet/ outlet of the fan, respectively, and A1 and A2 the cross
sectional area of the inlet and outlet of the fan.
The actual gas density ϑ can be calculated, if the normal density N
ϑ (that is the density at 0
o
C and 760 mm
Hg) of the gas is known, together with the temperature t and relative (static) pressure p:





 +






+
ϑ
=
ϑ
760
p
B
t
273
273
N
in which B (the barometric pressure) and p (with sign should be inserted in mm Hg (1 mmHg = 133.3 Pa).
Likewise an inlet flow expressed in normal cubicmeters QN can be transformed to actual flow volume Q by








+





 +
=






ϑ
ϑ
=
p
B
760
273
t
273
Q
Q
Q N
N
N
4.2 Theoretical fan power and efficiency
Assuming that the compression of the gas flow in the fan is an isentropic (or reversible adiabatic) process and
that the gas velocity is the same in the fan inlet and outlet we can calculate the fan power input (using a normal
thermodynamic nomenclature):
( ) ( ) ( ) ∫
∫
∫ ∫ =
+
=
+
+
=
=
−
=
2
1
2
1
2
1
2
1
1
2
is dP
v
m
dP
v
ds
T
m
dP
v
dv
P
du
m
dh
m
h
h
m
P
as ds = 0 (isentropic process)
The latter expression can for an ideal gas be integrated to (see ref.(l)):
( )( ) ( ) ( ) ( )










−







 ∆
+






−
=










−














−
=
−
−
1
P
P
1
P
1
k
k
Q
1
P
P
P
1
k
k
v
m
P
k
1
k
1
t
1
1
k
1
k
1
2
1
1
is
Here k is the adiabatic constant for the ideal gas k = v
p c
/
c
Developing the last parenthesis and rearranging we find
( )( ) ( )( )( )
( )
α
∆
=















 ∆







 +
+
∆






−
∆
= t
1
2
1
2
2
1
t
t
1
is P
Q
......
P
P
k
6
1
k
P
P
k
2
1
1
P
Q
P
The last factor is called the isotropic compressibility coefficient. The relation between α and 







1
t
P
P
for atm.
air (k = 1.4) is shown in figure below (from ref. (1)):
Fig. 4.2: Isotropic compressibility coefficient as a function of relative pressure rise.
The isentropic efficiency of a fan is thus defined as
( )( )( )





 α
∆
=






=
η
P
P
Q
P
P t
1
is
is
in which P is the actual power delivered to the fan shaft.
As α is close to 1 when the relative pressure rise across the fan 






 ∆
1
t
P
P
is small it has been common practice
to use the efficiency defined as
( )( )





 ∆
=
η
P
P
Q t
1
1
for fans generally. This efficiency is also termed the inlet efficiency.
It must be noticed that 1
η is not a real efficiency as 1
η would be greater than 1.0 for an ideal (high pressure)
fan.
One should always be aware that the fan efficiencies stated by different sources might have a different
definition.
4.3 Model fan curves
The dimensioning of large industrial fans is normally based on the experimental results from a performance
test with a smaller fan model similar to the large fan. The results from the model performance test might be
expressed as performance curves, as illustrated in figure 4.3.
The "left" (full line) curve establishes the relation between the total pressure increase tm
P
∆ obtained across the
model fan as a function of the volume flow lm
Q measured at the inlet to the fan.
Correspondingly the "right" (full line) curve establishes the dependence between the volume flow lm
Q and the
fan inlet efficiency lm
η defined as
( )( )







 ∆
=
η
m
tm
lm
lm
P
P
Q
in which m
p is the measured fan power consumption (at the fan shaft) at the flow lm
Q .
4.4 Calculations assuming imcompressible flow
If the compressibility of gas flow through the fan can be neglected (that is assuming the volume of the gas
does not change when passing the fan) the performance of a fan similar to the model fan can be calculated by
means of the three basic relations:
( ) ( )
1
Q
D
D
n
n
Q m
3
m
m
















=
( ) ( )
2
P
D
D
n
n
P tm
2
m
2
m
m
t ∆
























ϑ
ϑ
=
∆
( )
3
lm
m η
=
η
=
η
m
Q is the flow through the model fan (= ml
Q for incompressible flow)
Q is the flow through the actual fan
tm
P
∆ is the total pressure rating of the model fan
t
P
∆ is the total pressure rating of the actual fan
D is the actual fan diameter
m
D is the model fan diameter
n is the rotational speed of the actual fan
m
n is the rotational speed of the model fan
ϑ is the specific weight of the gas drawn through the actual fan
m
ϑ is the specific weight of the gas drawn through the model fan
η is the inlet efficiency of the actual fan and
m
η is the efficiency of the model fan
Thus, the flow/pressure curve at any rotational speed and gas density for any fan size in the same series as
the model fan can be calculated from the model fan curves using the "fan laws" (1), (2) and (3).
Likewise a flow/fan power characteristic curve can be calculated using the expression
( )
( )
4
Q
P
P t








η
∆
=
based on the definition of the fan efficiency η
Figure 4.4 shows actual fan performance curves for a kiln I.D.fan, which has been calculated from the
model fan curves in figure 4.3. (Note the curve has been drawn by a computer which also corrects for the
gas compressibility, thus the curves derivates slightly from the curves found using the above mentioned
calculation methods).
4.5 Fan dimensioning assuming incompressible flow
The optimal fan (that is the fan operating with maximum efficiency) for given values of t
P
∆ , Q and ϑ is
characterized by a diameter Dideal and a rotational speed n ideal
Dideal and n ideal can be calculated from (1) and (2) by insertion:
( ) ( )
5
Q
D
D
n
n
Q mideal
3
m
ideal
m
ideal
























=
( ) ( )
6
P
D
D
n
n
P tmideal
2
m
ideal
2
m
ideal
m
t








∆
























ϑ
ϑ
=
∆
where mideal
Q is the flow at which the model fan obtains maximum efficiency max
m
m η
=
η and tmideal
P
∆
is the corresponding model fan pressure.
Combining (3) and (4) we obtain the following expressions for direct calculation of Dideal and n ideal
( ) ( )
7
D
Q
Q
P
P
D m
5
.
0
mideal
25
.
0
t
tmideal
m
ideal
























∆
∆
×
ϑ
ϑ
=
( ) ( )
8
n
D
D
Q
Q
n m
3
ideal
m
mideal
ideal
























=
once the ideal fan diameter has been calculated a standard fan diameter close to the ideal diameter can be
chosen. The corresponding rotational speed might be calculated by iteration using the model fan curves
and the expressions (1) and (2), as illustrated by the following example.
Example 4.1
A raw mill fan (working with dust free air) of the HAF type should be dimensioned with max. efficiency
at the following operational data:
Flow at inlet s
/
m
100
Q 3
1 =
Temperature at inlet C
100
t o
1 =
Pressure at inlet Pa
9810
-
WG
mm
1000
p1 =
−
=
Specific gravity of gas at inlet 3
1 kg/m
83
.
0
=
ϑ
Total pressure increase across the fan Pa
9810
WG
mm
1000
pt =
=
∆
Barometric pressure B = 720 mmHg
Neglecting the flow compressibility in the model fan as well as in the actual fan the calculation goes as
follows:
From figure 4.3 we find that the model fan obtains max. efficiency at mideal
Q = 6.5 s
/
m3
corresponding to a
pressure rating of tmideal
P
∆ = 4140 Pa.
The ideal fan diameter can next be found by insertion in (7).
( ) m
17
.
3
m
1
.
1
5
.
6
100
9810
4140
2
.
1
83
.
0
D
5
.
0
25
.
0
ideal =




















×
=
and the ideal rotational speed (according to (8)) to
( ) rpm
932
rpm
1450
17
.
3
1
.
1
5
.
6
100
n
3
ideal =




















=
Choosing the standard fan impeller HAF 315/315 and assuming a rotational speed of 940 rpm we find from (1)
( ) ( ) s
/
m
57
.
6
s
/
m
100
15
.
3
1
.
1
940
1450
Q
D
D
n
n
Q 3
3
3
3
m
m
m =




















=




















=
( ) ( ) Pa
9820
4120
1
.
1
15
.
3
1450
940
2
.
1
83
.
0
P
D
D
n
n
P
2
2
tm
2
m
2
m
m
t =


























=








∆
























ϑ
ϑ
=
∆
which is very close to the desired t
P
∆ . Thus further iteration is not necessary and a HAF 315/315 operating at
940 rpm would be the right choice.
The inlet efficiency for the actual fan will be equal to the efficiency for the model fan at 6.57 s
/
m3
which is
read to 81.2% thus the shaft power for the actual fan is calculated to
( )( ) ( )( ) kW
1208
812
.
0
N/m
9810
s
/
m
100
P
Q
P
2
3
1
t
1
=






=






η
∆
=
With a safety factor for the fan motor of 15% the fan motor must have a minimum shaft power of 1390 kW.
4.6 Calculations considering gas compressibility
In practice a fan always works with compressible flow and a treatment using the above mentioned calculation
methods, which assume an incompressible medium, always introduces some error in the calculations.
As long as the fan works with a pressure difference t
P
∆ which is approximately the same as the pressure
difference tm
P
∆ read in the model fan curve used for the calculations, the error will be small.
However, for accurate dimensioning of high pressure fans it will be necessary to take the flow compressibility
into consideration.
Such calculations are quite complicated and they are normally always performed by means of computers.
Thus, we shall refrain from dealing with these here.
It can be mentioned, that recalculating example 4.1 considering the-gas compressibility results in the same
selection of fan diameter but the fan rotational speed is calculated at 926 rpm, and the fan power at 1170 kW
rather than the 940 rpm and 1208 kW, respectively, calculated on neglecting the gas compression.
4.7 Correction for dust laden gas
Up till now we have assumed that the fans to be calculated operate with clean gas.
Many fans used in a cement factory operate with dust laden gases f. inst. all kiln I.D. fans and most mill air
circulation fans.
The dust concentration in the gas passing these fans are typically 30-60 3
m
/
g .
Very little can be found in the literature concerning calculation of fans for gas with such a dust load.
If the dust is so fine that the dust particles could be assumed to pass the fan evenly distributed in the gas, the
fan could be calculated as a fan operating with (clean) gas with an equivalent (or apparent) gas density of
( )
9
s
1
leq +
ϑ
=
ϑ
s is the dust concentration in the gas (in kg/m3)
In this case the dust loading of the gas would not affect the fan power input, but the calculations would
suggest the use of a slightly bigger fan operating at lower speed than in the case of the fan operating with (the
same volume of) clean gas.
However, the dust is known to concentrate along the front side of the impeller blades on its way through the
rotor and the dust particles - leaving the rotor with a velocity close to the tip speed of the rotor, will at least to
some extent hit the fan casing and thereby loose their kinetic energy (obtained in the rotor).
According to ref. (2) it will be a good approximation to account for the dust loading simply by multiplying the
fan power found for the fan operating with clean gas with the ratio between the apparent density of the dust
laden gas and the clean gas density thus
( ) ( )
10
P
1
P
1
leq
dust 







−
ϑ
ϑ
=
The following example will illustrate the order of magnitude of the "dust correction".
Example 4.2
If the fan from example 4. 1 was to operate with gas with a dust loading of s = 40 3
m
/
g the approximate fan
power increase according to (10) is estimated to
( )
kW
1208
1
83
.
0
04
.
0
83
.
0
Pdust 





−
+
= = 58 kW
Thus according to this calculation the fan power increases with almost 5% 5rom 1208 kW to 1266 kW when
the gas is loaded with 40 3
m
/
g dust.
5. Fan Drives
5.1 Direct/indirect drive
The fans used in the cement and similar industries are nearly always driven by an electrical motor.
The motor might either be connected directly to the fan shaft via a flexible coupling or the motor power might
be transmitted to the fan through a V-belt drive (up to 400 kW) or a gear unit.
If direct drive by A.C. motors shall be used the fan must be dimensioned for a synchroneous or asynchroneous
motor speed cf. table 5.1.
Table 5.1
Rotational speeds for A.C. motors ( > 55 kW)
Rotational Speed in RPM Frequency No. of poles
Hz 2 4 6 8 10 12
Synchroneous motors 50 3000 1500 1000 750 600 500
60 3600 1800 1200 900 720 600
Asynchroneous motors ( > 55 kW ) 50 2970 1480 985 735 590 490
60 3565 1775 1180 880 705 585
In many cases a direct drive will not allow for dimensioning the fan with optimum efficiency at the actual
operational conditions and in this case an indirect drive must be chosen.
For large fans it will also some times be cheaper to choose an indirect drive via a reducing gear coupled to a
motor with high rotational speed (few poles) than to choose a direct drive with a slow running motor (many
poles).
5.2 Motor types and speed regulation
To-day the following combinations of motor types, starters and speed regulation systems are used for (larger)
fans:
1. Fixed speed drives:
a) Squirrel cage motors with direct start
b) Squirrel cage motors with Y/.t, starter.
c) Synchroneous slip ring motors with rotor rheostat starter.
d) Synchroneous AC motors.
2. Variable speed drive:
a) Squirrel cage motors with frequency converter.
b) Asynchroneous slip ring motor with rotor rheostat regulator.
c) Asynchroneous slip ring motor with subsynchroneous converter cascade.
d ) DC motors with thyristor converter.
For smaller shaft outputs ( <100 kW) fixed speed squirrel cage motors with direct start will always be preferred
due to the low price and easy maintenance of the squirrel cage motor.
For larger motors the large starting current for the squirrel cage motor with direct start (4-8 times the full load
current) can be a problem.
The starting current can be reduced by using an Y/∆ starter, but also the starting torque of the motor is
reduced, which results in a larger starting time which might not be permissible. In this case a slip ring motor
with rotor rheostat starter has to be used.
Synchroneous AC motors are relatively seldom used and only for large outputs ( >-500 kW). One advantage
of this motor type is that it can act as (capacitive) phase compensator (regulating the magnitisation of the
rotor).
For variable speed drives slip ring motors with rotor rheostat regulator and DC motors with thyristor
converter are the drive types most often used.
The advantage of the slip ring motor with rotor regulator is the relative simplicity of the system, but
connected to a fan the regulation range is relatively narrow (down to 60-75% of full speed) and the
efficiency of the drive goes drastically down when the motor speed is reduced as the "slip power" is wasted
(as heat) in the rotor rheostat (see below).
The speed of the DC motor with thyristor converter (often termed thyristor drive) is controlled by regulating
the armature voltage delivered by the thyristor converter. The drive system is somewhat more complicated
than the slip ring motor with rotor rheostat regulation but the control range is normally very wide (down to
10% of full speed) and the drive system maintains a good efficiency even at low motor speed.
The drastical reduction of the efficiency of a slip ring motor drive at reduced speed can be avoided if the slip
power instead of being wasted in a simple rheostat is treated in a "slip recovery system" in which it is
rectified, subsequently converted back to AC with grid frequency and then fed back to the (stator) terminals
of the motor. Such a system is termed a subsynchroneous converter cascade (or just cascade regulation).
However, this system is only price competitive to the DC motor drive system for large motor output (above
1000 kW).
For motor outputs below 1000 kW a squirrel cage motor connected to a static frequency converter might be
chosen. This drive system combines the easy maintenance of the squirrel cage motor with the wide
regulation range and low regulation power loss of the DC drive system.
5.3 Drive efficiencies
The electrical power used to drive a fan can be calculated if the necessary fan shaft power P is known by
means of
( )( )







η
η
=
el
trans
el
p
p 1
trans
η is the efficiency of the power transmission system
trans
η 1.0 for direct drive and
trans
η 0.98-0.99 for indirect drive (gear or V-belt)
el
η is the efficiency of the motor inclusive a possible speed regulation system.
For large fixed speed AC motors the efficiency el normally lies between 92% and 96%. For various variable
speed drives the efficiency as a function of the relative speed reduction ratio appears from figure 5.1.
Drive efficiency, %
Rotational speed, % of max.
Fig. 5.1: Drive efficiencies of various variable speed fan drives.
a) 400 kW squirrel cage motor with frequency converter
b) 1000 kW slip ring motor with rotor rheostat regulator
c) 1000 kW cascade drive
d).1000 kW DC motor with thyristor converter
6. Capacity regulation
The total pressure drop t
P
∆ across most flow systems in-, creases approximately with the second power of
the flow through the system, that is
( )( )
( ) ( )
1
Q
K
P
2
t =
∆
At full speed the fan which transports the gas through the system is also characterized by an unique
relation between total pressure rating and gas flow
t
P
∆ = f(Q) (2)
Thus, at full speed the fan will generate a certain flow, Q., through the system which can be found from
the equa,tion
( )( )
( ) ( ) ( )
3
Q
f
Q
K l
2
1 =
or from the fan curve, cf. figure 6.1.
If it is desired to reduce the flow through the system this can be accomplished in a number of ways which
are more or less economical as seen from an operational point of view.
Fig. 6.1: Simple damper flow regulation
6.1 Damper regulation
By aamper regulation the resistance coefficient in the system, K, is changed by opening or closing a damper
normally placed near the inlet to the fan.
Closing the damper somewhat will increase K and reduce the flow in the system according to (3) (assuming
that f(Q) is a decreasing function in the relevant flow interval)), cf. also figure 6.1.
In the example in figure 6.1 the flow through the fan can be reduced from 1
Q to 2
Q by closing a damper in
the system so as to increase the total pressure drop in the system from 1
t
p
∆ to 2
t
p
∆
By this regulation method the pressure rating of the fan increases, when the flow through the fan is reduced.
Therefore the power consumption of the fan only decreases slightly when the flow is reduced.
Accordingly damper regulation is a fairly uneconomical flow regulation method.
However, as the damper is a very simple and cheap flow control device this regulation method is normally
preferred for smaller fans with limited power consumption.
6.2 Inlet guide vanes
By means of inlet guide vanes placed just at the inlet to the fan the gas can be given a prerotation which - if
in the rotor rotation direction - reduces the pressure rating and the power consumption of the fan for a given
gas flow. Thus by changing the adjustment of the inlet guide vanes the fan curve (or the function f (Q)) can
be changed and thereby the flow through the fan can be reduced.
As the guide vanes not just introduce a pressure drop in the system, but also reduce the fan pressure rating
and power consumption for a given flow, the flow regulation by means of guide vanes is more economical
than simple damper regulation.
Fig. 6.2 illustrates this. The fan with inlet guide vanes operates in a system with the system curve K.
With fully open guide vanes the fan characteristic curve designated 0
o
applies. The resulting flow through
the fan is 1
Q , and the power consumption 1
P .
If the guide vanes are turned 45
o
the flow through the system is reduced to 2
Q and the fan power
consumption to 2
P .
If the flow had been reduced by a simple damper in the system, the considerably higher fan power
consumption 2
P would have resulted.
However, guide vanes are more expensive than simple dampers and they are rather sensitive to wear when
operating with dust laden gas.
6.3 Fan speed regulation
Reducing the speed of the fan will reduce the pressure rating of the fan at a given flow through the fan, cf.
fig. 6.3.
If the fan is connected to a second order flow system in which the pressure drop varies with the flow
according to (1) the fan will maintain the same efficiency independent of the fan speed. That is, if the fan is
designed to operate with maximum efficiency when operating with full speed the fan will still operate with
maximum efficiency in the case the flow is reduced by changing the speed of the fan shaft.
Thus, this form of regulation gives a loss-free regulation of the fan itself, and the operating economy depends
on the losses in the speed regulation system.
Various methods can be applied for speed regulation of the fan shaft.
Smaller fans can be coupled to a fixed speed motor through a mechanical variator or a hydraulic or eddy
current coupling.
For larger fans motor speed regulation is nearly always applied to reduce the fan shaft speed.
If the motor is a slip ring motor with rheostat regulation the regulation loss is considerable whereas a near to
loss-free regulation is obtained with an AC motor with frequency convertors, slip ring motor with cascade
regulation and DC motors with thyristor rectifiers as explained in section 5.
In fig. 6.4 the fan power consumption as a function of the flow for the various flow regulation methods is
compared.
Fig. 6.3: Flow regulation by changing fan speed.
Fig. 6.4: Fan motor power consumption curves
Example 6.1
A raw mill fan is dimensioned for:
Q = 100 s
/
m3
t
p
∆ = 9810 Pa
l
ϑ = 0.83 3
m
/
kg
It has been calculated that HAF 315/315 fan will provide the desired pressure and flow at a rotational speed of
n = 926 rpm. However, the fan is supplied with a directly coupled 1600 kW, 985 rpm slip ring motor.
The actual fan curves for the fan at 985 rpm and 926 rpm, respectively, are shown in fig. 6.5.
A (second order) system resistance curve corresponding to the above mentioned flow and pressure for which
the fan was dimensioned is also indicated on the diagram.
From the diagram it can be read that the fan operating at full speed, 985 rpm, and with fully open 3 damper
would produce a gas flow in the system of 106.7 s
/
m3
and the total pressure difference across the fan would
correspond to approx. 11200 Pa.
The shaft power can be read at 1425 kW.
The desired flow 100 s
/
m3
can be obtained closing the fan damper somewhat to obtain the system resistance
curve 2
K shown in the diagram fig. 6.5.
From the diagram it can be read that the fan shaft power consumption hereby will decrease only slightly to
1395 kW. The pressure difference across the fan will increase to 11600 Pa of which (11600 - 9810) Pa = 1790
Pa is lost in the damper.
With a motor efficiency of 0.96 the electric power consumption of the fan regulated by damper is calculated at
1450
0.96
kW
1395
Pel =






= kW
If the motor speed instead is reduced to 926 rpm the desired flow would be reached with fully open fan
damper. From the diagram we read in this case a fan shaft power consumption of 1175 kW.
If the slip ring motor is rheostat regulated the actual motor efficiency (including losses in the slip ring
rheostat) can be calculated to
( ) ( ) 90
.
0
96
.
0
985
926
n
n
motor
nom
el =












=






η








=
η
and the electrical power consumption of the fan would be






=
90
.
0
1175
Pel kW = 1305 kW
If the motor was provided with cascade regulation (slip recovery system) the efficiency of the motor and
speed regulator would be 0.96 - 0.015 = 0.945. In this case the electrical power consumption of the fan would
be






=
945
.
0
1175
Pel kW = 1245 kW
This example illustrates the importance of using speed regulation of the fan motor rather than damper
regulation even when a relatively small reduction of the gas flow in a system is desired.
Thus, in this case where the flow is reduced with only 6% we save 150 - 200 kW (depending on the variable
speed drive type) if the flow is reduced by reducing the motor speed rather than closing the fan damper
somewhat.
Fig. 6.5: Performance curve for Fan type: HA.F 315/315
(used for example 6.1)
7.0 Fan modifications
Rising electricity prices have created an interest in changing low efficient fans with high efficient fans.
However, to change a large existing fan with a complete new fan not only involves buying the new fan but due
to another design of the new fan the existing fan foundation cannot be used and it will normally also be
necessary to modify the ducting around the fan which increased the installation costs and the down time.
7.1 Impeller replacement
Another solution is to install a new impeller in the existing fan in order to increase its efficiency without
changing the fan casing, fan drives etc.
As an example we will describe the equipment which is now offered for modification of a fan with open radial
impeller like the one described in section 3.1.
Fig. 7.1: HP impeller
By the modification the open impeller is replaced by a high efficient closed impeller with backwards curved
blades cf. fig. 7.1.
Fig. 7.2 shows a cross-section in the modified fan.
Together with the new impeller an inlet cone and inlet nozzle are installed in the existing fan casing.
Further, it is necessary to modify the rear plate of the fan casing somewhat and install a new shaft sealing and
cooling disc (if the fan is to operate with hot gasses).
In many cases the existing shaft can be reused after a machining of the end on which the new impeller is to
be mounted.
As the new impeller has backwards curved blades, it will normally have to operate with slightly higher tip
velocity than the old impeller to obtain the required pressure rating.
Normally, this problem is solved by increasing the impeller diameter in relation to the existing impeller.
However, in some cases the casing does not allow for a sufficient increase in the diameter and in that case the
shaft speed has to be increased introducing a new gear box or by changing the gearing ratio of an existing
gear box.
7.2 Energy savings
As-the operational conditions for the fan which is to be modified are normally known very well, the new
impeller can be designed to operate with optimum efficiency under these conditions, and an efficiency of the
modified fan exceeding 82% can normally be guaranteed.
Comparing this with a typical efficiency of 68% of the existing fan with open impeller the relative power
saving by installing the new impeller can be calculated at
S = ( )
%
100
82
.
0
68
.
0
1 





− = 17%
assuming the same drive efficiency before and after the modification.
If the existing fans always have been operated with reduced motor speed and/or with partially closed dampers
the new impeller can be designed to allow for operation with almost full motor speed and fully open damper.
In this case the relative power saving on installing the new impeller will be even greater.
Fig. : 7.2: Fan originally provided with open impeller after modification with new high-efficient impeller
etc. New parts indicated.
If on the other hand the existing fan is operating at full capacity and a further increase in the production
capacity of the system to which the fan is connected is desired, it will often be possible to design the
replacement impeller for a pressure rating which allows for the desired production capacity increase.
Finally, it should be mentioned, that if the existing open impeller is exposed to excessive wear, it cannot
be recommended to replace it with a high efficient impeller as this will probably be exposed to even
greater wear and does not allow for an easy replacement of worn impeller blades.
Example 7.1
Data for an existing raw mill fan with open ¿ 4500 mm impeller:
Yearly operating time 7500 h
Motor power consumption 1300 kW
Electricity price 0.06 USD/kWh
Estimated fan efficiency 67%
Budgetary quotation for new closed fan impeller etc. for modification of the fan:
Impeller diameter: ¿ 4950 mm
Expected efficiency: 84%
Rotational speed: Same as existing impeller
Price f.o.b.: 170,000 USD
Fan motor power consumption after modification:
1300 kW =






84
.
0
67
.
0
1040 kW
Annual electricity saving:
(1300 - 1040) 7500 kWh = 1.95 x 106 kWh
(1.95 x 106 ) (0.06 USD ) = 117,000 USD
Estimated installation costs: 30,000 USD
Total investment: 200,000 USD
simple pay-back period: 200/117 years = 1.7 years.
References: (1) Bruno Eck: "Ventilatoren". 5th Edition
Springer Verlag, 1972
(2) "Fan engineering" - 8th Edition
Buffalo Forge Company, 1982.
Blue Circle Cement
PROCESS ENGINEERING TRAINING
PROGRAM
Module 5
Section 5
Harleyville Preheater Exhaust Fan Performance
Verification
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391861703-Mod-5-Fan-Measurement-and-Testing.pdf

  • 1. Process Engineering Training Program MODULE 5 Fan Measurement and Testing Section Content 1 The Chemical Engineering of Fans 2 Fans and Airflows 3 Fans and Fan Systems, Chemical Engineering March 1983 4 Fans, FL Smidth 5 Harleyville Preheater Exhaust Fan performance Verification 6 Fan Applications PRESENTATIONS Bowmanville Calciner ID Fan Problem Fan Technology Centrifugal Fans and Systems
  • 2. Blue Circle Cement PROCESS ENGINEERING TRAINING PROGRAM Module 5 Section 1 The Chemical Engineering of Fans
  • 3. C0NTENTS Page No. 1. Definition of Fan Pressure 1 2. Fan Curves 3 3. The Fan laws 8 4. Control of Fan Output 11 5. Unsatisfactory Fan Performance 14 6. Series Fans 16 7. Parallel Fans 18 8. Blade Types 26 9. Fan Noise 31 10. Other Gas Pumping Equipment 32 11. Power Required for Compression 41 12. Further Reading 42 Appendix 1- Effect of Speed and Temperature on Fan Operation 43 Appendix II- Prediction of Fan Noise 46 Appendix III - Effect of Water Injection on Fan Performance 48
  • 4. 1.0 DEFINITION OF FAN PRESSURE Fan total pressure PT, is the difference between the total pressures at the fan inlet and outlet. Fan static Pressure, PS,is the fan total Pressure minus the fan velocity pressure. Fan, velocity pressure, PV, is the velocity pressure         g 2 u2 corresponding to the average velocity at the fan outlet. i.e. the fan static -pressure is NOT equal to the difference between the static pressures at the inlet Ar!-.1 outlet. It should be measured as the difference between the reading of a facing tube at the inlet and a side tube at the fan outlet (see Figure 1). FIG. 1 MEASUREMENT OF FAN STATIC PRESSURE
  • 5. 2.0 FAN CURVES These show the effect on fan pressure, Power and efficiency of a change in fan throughput (Fig.2). The horse-power on a fan curve is that which must be delivered to the shaft and does not take into account losses in couplings and motors. A calculation of the work actually done on the air will therefore give the efficiency of the fan itself, Work done on air = pressure x flowrate x HP 33000 2 . 5       =       6360 PV HP (inches wg and cfm) or 9.81 PV watt (mm wg and m3/sec). P may be either the fan static or total pressure and the efficiencies thus calculated are the static and total efficiencies respectively. Because work is being done on the air, its internal energy will rise, and this, together with losses due to the inefficiency of the fan will cause the air temperature to rise. Temperature rise = ( )( ) heat specific V density Power fan         For air at ambient conditions = efficiency fan total 24 . 0 V 075 . 0 779 2 . 5 PV = η               × ×         η × × = 0.37         η P o F or about ½ o F per inch wg.
  • 6. The resistance of the system served by a fan depends upon the hardware in it and for any system there will be a relationship between flowrate and pressure drop. If the system contains only dampers, ducting, expansions etc., of fixed resistance, then the pressure drop may be taken as being proportional to V2 (approximately). This would not be true in the case of a resistance (as opposed to pressure drop), that varied with flowrate, such as a Lepol grate. The intersection of the system resistance line with the fan pressure/flowrate curve gives the point of operation of the fan. The effect of, for instance, closing a damper is to steepen the system resistance line, giving a new operating point (see Fig.3), System resistance lines may be calculated from fluid flow theory or one point may be measured end the remainder drawn in assuming, for instance, a squared relationship. FIG. 2 TYPICAL FAN CUIRVEES
  • 7. FIG. 3 EFFECT OF A CHANGE IN SYSTEM RESISTANCE 2.1 The Shape of Theoretical and Practical Fan Curves Fan curves may be calculated theoretically and the shape of these curves are shown in Fig. 4. However, in practice losses modify the curves to give what we normally recognize as a typical fan curve. (a) Friction Losses: Friction at the fan inlet, casing, and blade passagescauses a loss which is proportional to the square of the velocity (or volume) and is zero at zero flow.
  • 8. FIG. 4 FAN CURVES CALCULATED FROM THEORY (b) Eddy Losses: These occur at the blade leading edge and in the fan casing. Depending upon the relative velocity between the air and the impeller leading edge, the angle of attack will change. The smoothest flow will occur when the above relative velocity is in the same direction as the blades themselves. Here eddy loss is zero and increases as the angle of attack moves away from zero.
  • 9.
  • 10. A similar situation occurs in the fan casing, there being only one flow where eddies will not occur. It is usual to design casings such that the flowrate corresponding to eddyless flow is the same for both impeller and casing. (c) Leakage: Backflow between casing inlet and impeller inlet causes a small loss. (d) Disc friction: The resistance offered by the air to the impeller rotation does not affect the head developed but is seen as increased power absorbed. (e) The theoretical no-lose characteristics shown above assume that the air follows the blade passages absolutely. This only be true for an infinite number of blades. The effect of the above on fan characteristics is shown in Figures 5A and. 5B for a backward curved impeller. 3.0 THE FAN LAWS The performance of a fan may be changed by several methods, the main ones being a, change in a) fluid density, p b) impeller speed, n (rpm) c) fan size * Most text books only quote the fan laws for "fans of geometric similarity", so that diameter, width, outlet duct size, etc., are all increased at the same rate. Any convenient dimension may then be taken as being representative of the fan, e.g. Impeller Diameter, Dim, The effect of these on fan pressure,volume and horse-power is summarised in Table I. TABLE I - The Fan Laws ALTERATION IN CHANGE IN FAN PRESSURE VOLUME POWER Density, p α p No change α p Speed, n α n 2 α n α n3 Fan Size*, Dim α Dim 2 α Dim 3 α Dim 5 Fan pressure may be either total, static or velocity. In order to help remember the above table, it is worth noting that fan pressure can be shown to be theoretically proportional to the square of the tip speed whereas volume is proportional to the swept volume of the impeller (fan size/volume is a cubic relationship since width increases we well as diameter). Power is proportional to the product of volume and pressure. A change in impeller diameter/width ratio will change the characteristic of the fan and direct scale-up is not possible, although if a fan is running at somewhere near its peak efficiency (i.e. just to the right of the peak on the pressure - Volume graph) then for a small increase in diameter, the volume will increase as the cube and the pressure as the square of the tip speed, i.e. the fan laws in Table I would appear to apply. However, it must be
  • 11. remembered that a new fan type has been produced by the change in diameter/width ratio and throttling of the fan will not follow any easily calculable curve. Presented with curves for a fan running at any set of conditions, it should be possible to predict the curves at any other set of conditions. Before setting out on the calculations, it is worth checking that the fan is actually operating close to the manufacturer's fan curve. Their fan tests are performed with straight inlet and outlet ducts and, for instance, a bend close to the fan inlet on site could lead to adverse whirl and a reduction in the fan efficiency. The method of calculation is best shown by example. Appendix I shows the effect of a change in density and speed on a fan, these being the variables normally encountered in practice. Care must be taken in applying the fan laws, they are only applicable over a limited range because they are only an approximation. A better estimate is given by V = kg Dim 3 n (Re)x (4) P = kp Dim 2 n2 (Re)z (5) where, kp, kg, x, z are constants. Re = The Reynolds Number (Dimensionless) i.e.:         viscosity fluid D x speed tip density x fluid im The indicies x and z are so small, however, that the effect of the Reynolds Number is usually negligible in practical cases. Even British Standard 848 states that Reynolds Numbers may be ignored in scale up provided they do not vary by more than 40%. 4.0 CONTROL OF FAN OUTPUT For a centrifugal fan there are three basic methods of controlling output 1. Damper Control 2. Variable Radial Inlet Vanes - these give a beneficial .swirl to the air at the fan inlet and therefore reduce the work the fan has to do. 3. Speed Control. Initially, dampers are the cheapest to install and speed control the most expensive, but there is a reversal in a running cost comparison. - A cost comparison must be made in each individual case based on the capital cost,
  • 12. running cost, and the amount of turndown that will be needed. However, speed control usually only becomes economic on the larger machines. Whereas variation in a damper setting will change the system resistance and not the fan curve, the other two methods produce new fan curves, that for speed control retaining the same fan efficiency. Since the use of speed control introduces the inefficiency of the device itself, it is worth noting that for small changes from the maximum throughput, speed control has, in fact, the highest running cost. Referring to Figure 6, the effect of turndown to 3 V/4 and V/2 by the three methods is shown. Graph A - Damper Control - Fan curve retained but system resistance line altered. Graph B - Variable Inlet Vanes - New fan curves produced for each vane setting. Graph D - Comparison of three methods on basis of power taken Graph C - Speed Control - New fan curves produced for each speed but efficiency retained. (Note: loss due to coupling is exaggerated). This graph assumes no pressure drop across a fully open damper. This may be a reasonable assumption for some dampers operating in clean air, but would certainly not be true for some I.D. fan dampers where build-up will occur. The expected pressure drop across an open damper, based on experienced, should be taken into account when deciding on the method of control. Radial inlet vanes are not cheap to install and can increase the cost of a fan by 20%. However, they are an efficient means of control over a limited range and should be considered as an alternative to speed control. For larger variations in volume a two or three speed motor with final adjustment by inlet vanes would be cheaper than infinitely variable speed control. If considerable operation at less than maximum capacity is expected, speed control had the added advantages of an increase in life and a reduction in noise level. 5.0 UNSATISFACTORY FAN PERFORMANCE Fan performance may be affected by worn impellers, inefficient drives, etc., but the most common causes of fan performance being substandard are bends close to the inlet or outlet of the fan. If the bend is at the inlet, the air is pushed over to one side of the eye, whereas a manufacturer's fan curve is for uniform inlet conditions. Uniformity of flow at inlet is easily checked by a pitot traverse. The best solution to this is an inlet box which can be sized so that the air enters the impeller uniformly, failing that a square bend with cascade (Fig.7) will give good performance. Retention of the original bend and fitting of long turning vanes will improve matters greatly (Fig.8). Useful pressure in a fan originally comes from the acceleration of the air, - As the air is again slowed down by expansion, velocity pressure is converted into static pressure. For this to be done efficiently the expansion
  • 13. should be slow and gentle and clearly a bend immediately on the fan outlet will not help matters. This problem is not so easy to remedy but again a cascade will help.
  • 14.
  • 15. 6.0 SERIES FANS For fans in series, the second fan will handle -the same mass of gas as the first fan. However, due to compression effects and temperature effects (Equation 3). The second fan will not handle the same volume. To be absolutely accurate, the fan curve should therefore be modified to correct for the density difference. The two effects, however, tend to cancel each other out, and in practice it is usually acceptable to use the original curve for the second fan. The combined pressure at any volume is found by adding the total pressures of the individual fans and subtracting the losses for the ducting between the fans (usually negligible) - see Fig. 9. Here to obtain the complete fan curve it has been necessary to know the characteristic of one fan at volumes greater than it would handle with free inlet and discharge (i.e. being "force-fed"). This is not normally known, and in any case there would be little point in operating under these conditions since the fan would be hindering the flow of air rather than helping it. Unless the fans are identical it is likely that one of them will have to operate at a point that does not give its maximum efficiency. For this reason, if a series set up is necessary, it is customary to install identical units. This may be necessary if the space available renders one large fan impractical or if it is necessary to keep the air flowing while one of the fans is being maintained (each fan would then have a by-pass).
  • 16.
  • 17.
  • 18.
  • 19. A similar exercise is then performed on Fan 2 and its associated system. The two curves (P-F)1 and (P-F)2 are then added together to give the characteristic of the parallel unit. Any system resistance line now drawn in a refers only to the resistance of ducting, etc., that is common to both fans. It should be noted that alteration of any damper in the non-common ducting will change the parallel unit characteristic rather than the external system resistance line. 7.2 Unsatisfactory Operation of Parallel Fans Not all fans have such simple characteristics as those in Fig. 10. For instance, parallel operation of two identical forward curved blade fans is shown in Fig.12. At the pressure Pv, there are three possible operating volumes for the single fan, namely Vx, Vy, and Vz. For the single fan the system resistance line would determine the point of operation. However, these three points rive rise to six possible points of operation for the parallel system. Three of them are simply 2Vx, 2Vy and 2Vz and they form the main part of the characteristic for the parallel arrangement However, the three other possible additions (Vx + Vy), (Vx + Vz ) and (Vy + Vz ) from off-shoots to the main curve. It can be seen that, depending on the position of the system resistance line there will be one, two or three possible operating points. For the system resistance line drawn in Fig.12 there will in fact be three possible points of operation. The fans may settle down at any of these positions or oscillate between all three. C would be the desired operating point, when both motors will develop the same load, whereas there will be progressively more imbalance as the system moves to-point B and then point A. In practice both motors may cut out immediately on start-up, the second fan because it is overloaded and the second fan because it then tries to run at the point D. 7.3 Prediction of Unsatisfactory Parallel Operation Hagen's method is a rather neat way of predicting the above unsatisfactory performance of identical fans. The system resistance and total pressure single fan curve are drawn and, for any total pressure, the magnitude of the term [ ] pressure) at that flow fan volume possible (maximum - flow) volume (system is superimposed. If the line (Eagen's line) thus formed crosses the fan curve at more than one point then operation will be unsatisfactory (Fig. 13). 7.4 Correction of Unsatisfactory Parallel Operation This can be effected by fitting balancing dampers to each fan as shown in Fig.14. The system resistance line, D, crosses the original parallel fan curve in three places leading to unsatisfactory operation. The balancing damper and original fan curves are combined to give the curve (P-R) for each fan (the balancing damper bring considered part of the fan) and when the curve is drawn for the two fans (P-R) in parallel it can be seen that the system resistance line cuts it at only one point. Hence with careful adjustment of the balancing dampers it should be possible to run the two motors with balanced loads, with only a small loss in volume flow.
  • 20.
  • 21.
  • 22.
  • 23. 8.0 BLADE TYPES Basically, three types of impeller are used in centrifugal fans : 1. Forward Curved Blades (Blade angle > 90o) 2. Radial Tipped Blades (Blade tip = 90o) 3. Backward Sloping Blades (Blade anele < 90o) Fig. 15 shows typical fan curves for each type. In general for a given impeller size, the pressure developed increases with blade angle and higher volume flows can be obtained by use of a double inlet. 8.1 Forward Curved Blades These fans give the highest pressure for a given size and speed. They are also used for high volume applications such as ventilation, where their compact size and low speed (hence low noise) is an advantage. The low speed also allows a lighter, less expensive construction to be used. The blades are of short radial depth to reduce the inlet throttle and consequently more of them (30-60 compared with 6-16 for other fans) are required to have an effect on the air. It can be seen (Fig.15) that a decrease in restriction and hence an increase in airflow will cause an increase in power. The fan is said to be of the overloading type and this effect may cause the motor to trip out. These fans are unsuitable for dirty air, the narrowly spaced tips tending to bridge. Efficiencies are not greater than 75% 8.2 RADIAL BLADED FANS These are of two types 8.2.1 Radial Bladed Paddle Fans, e.g. firing fans. These have their main application in handling high dust burdens. They are of simple construction and therefore the wearing parts can be of substantial material or even replaceable. The impeller is unshrouded but may have a back plate in the case of fibrous dust to prevent build-up on the shaft. For corrosive applications, it is a simple matter to rubber-cover the impeller. Disadvantages of the paddle fan are its low efficiency (about 605), high speed of rotation (noisy operation) and overloading power characteristics. 8.2.2 Radial Tipped Forward Curves Fans These fans combine some of the ruggedness of the paddle fan with the higher efficiency and higher pressure development of the forward curve blade. The impeller is self-cleaning to a large extent and is used extensively for dusty flue gases. Blades themselves are not easily made removeable but replaceable wear plates and nose pieces can be fitted if dust burdens are high. Efficiencies higher than 75% are obtainable with this fan and it is therefore commonly used where high dust burdens prevent
  • 24. the use of the more efficient aerofoil or backward inclined bladed fans.
  • 25.
  • 26. 8.3 Backward Bladed Fans In the main, these fans are used for clean air applications, although the use of a reinforced nose, or leading edge, will allow low concentration of dust to be handled. Efficiencies up to 90% are obtainable with fans having aerofoil section blades (Fig.16) and therefore the fan is preferred for the larger machines. FIG.16 SECTICN THROUGH AEROFOIL For a given duty the fan must run at a higher tip speed and the higher efficiency may be partially offset by the capital cost involved in producing accurate aerofoil blading of sufficient robustness to withstand the increased rotational stress. The high efficiencies are a result of laminar flow over the blade profile and the reduction of eddies leads to quieter operation. 'Backward bladed fans are nonoverloading (see power curve on Fig.15). Summing up, the progressive use of fan types would be Forward Curved Backward Aerofoil Backward. Inclined Forward Curved Radial Tip Paddle as the dust burden increases. 9.0 FAN NOISE The main source of fan noise is caused by the formation of shock eddies as the air passes through the blades of the impeller and therefore quietness depends upon a smooth passage of the air. Fans generally operate most quietly at their highest efficiency and dampering back may cause an increase in sound level. The outlet throat velocity does not in itself control the noise level - the clearance between the impeller and the tongue being more important. The rate of expansion of the air caused by the casing shape is also of importance. An expansion that is too rapid will give a low outlet velocity but a higher noise level. The noise generated by the fan is transmitted to the casing both upstream and downstream causing resonance in ducting. Fan noise can be reduced by speed reduction (if the fan is already heavily dampered), duct and casing stiffening, or the fitting of inlet sound attenuators of the "Quietflow" type. Appendix II gives a method of predicting octave band sound pressure levels from which a dB(A) rating can be calculated.
  • 27. 10.0 OTHER PUMPING EQUIPMENT FOR GASES Table 2 shows the applications of the various types of pumps available, though the figures should be taken as a general guide only : TABLE 2 - Gas Pumping Equipment PUMP TYPE MAX. DELIVERY PRESSURE MAX. THROUGHPUT ft3/min (STP) Reciproacting compressors 4000 ats 5000 ats Rotary Compressors: High Compression Ratio 75 psi 4000 Low Compression Ratio 15 psi 5000 Centrifugal Types: Fans 40" WG 150,000 Single Stage Blowers 5 psi 12,000 Multistage Water Cooled Compressors 150 psi 50,000 (10 stage) 10.2 Other Fans Apart from the centrifugal fan, which is by far the most common in our industry, use is also made of Crossflow and Axial flow fans. 10.2.1 Cross flow fans Here the air enters and leaves radially (see Fig.17). This fan is used when a very low fan pressure is required Such as in the domestic fan heater and some hair driers.
  • 28. FIGURE 17 - CROSSFLOW FAN 10.2.2 Axial Flow fans Here, as the name implies the air enters and leaves the fan axially. They are a development of the propeller fan (e.g. Xpelair) which is only suitable for low pressures (i.e. free inlet and outlet). The modern axial fan, however, is much more versatile and has the advantage that the motor may also be mounted on the axis of the duct giving a very compact unit. The motor may be sited outside the gas stream by use of a bifurcated duct (Fig.18). The axial flow fan is generally not used in dusty conditions since the blade tip clearance tends to bridge FIGURE 18 - AXIAL FAN - EXTERNAL MOUNTING OF MOTOR
  • 29. The circular motion of the fan gives the air leaving it a rotational component which tends to increase the downstream pressure drop. For this reason guide vanes are installed either Upstream (in which case it is the fan itself that straightens out the flow) or downstream. Two fans in series (often built as a single unit) are often made contra-rotating so that guide vanes are not required. The air then leaves the second fan axially. The output of axial fans can be changed by altering the pitch of the blades, usually when the fan is stopped, though in specialized applications such in aircraft engines, the pitch may be altered while running. Typical axial fan curves are shown in Fig.15. 10.3 Reciprocating Piston Compressors These are the only machines capable of very high delivery pressures (100 pai). Their high efficiency must be balanced against the lower capital cost of a rotary compressor. 10.4 Rotary Blowers and Compressors High compression types include the sliding vane blower (Fig.19) and the Nash Hytor (Fig.20). The Hytor has an elliptical casing, partially filled with fluid which is thrown to the outside of the casing by the impeller. The spaces in vanes act as liquid pistons so that the air is sucked in, compressed and then discharged by the advancing interface. The liquid may be chosen to be inert to the gas being pumped. Neither of the above two machines are affected by tip wear. The sliding vane compressor has loose blades which simply move radially outwards as they wear whilst it can be seen that the Hytor will function provided the blade tips are always covered by liquid. For this reason these machines will maintain their efficiency over long periods. Typical of the lower compression ratio type is the Rootes Cycloidal Blower (Fig.21 ), where Increased wear and therefore back-leakage will reduce efficiency. 10.5 Centrifugal Compressors (Turbo-Blowers) -(Fig. 22) These machines are used at anything from ½ psi to 100 psi and operate at high speeds (3500 to 30,000 rpm). Water cooling may be employed in-between the stages and in-between casings which normally contain 6 or 7 stages. The axial flow compressor is used for special applications (e.g. jet engines) where its lightness is an advantage. It is a high efficiency machine but has a limited operating range, greater vulnerability to erosion and 00=03ion and susceptibility to deposits. Fig. 23 shows an axial flow compressor. The rotating element c6nsists of a single drum to which are attached several rows off aerofoil blades, of decreasing height. Between each row is a stationary row which straightens the airflow before it meets the next moving set of blades.
  • 30.
  • 31.
  • 32. 10.6 Vacuum Pumps Here, a high compression ratio is necessary and therefore the machines used include the reciprocating piston, rotary vane (Fig.24) and Nash Hytor, Piston and rotary pumps will give pressures down to 10-2 mm. Hg. but the Hytor is limited by the vapor pressure of the sealing liquid (usually water).
  • 33.
  • 34. 11.0 POWER REQUIRED FOR COMPRESSION This depends on the method of compression, which can be isothermal, adiabatic (no heat removed from system)or something in-between the two. Isothermal compression requires the least work and is therefore the most efficient so that the isothermal line on a P-V diagram should be approached as closely as possible, which means that the maximum cooling between stages should be used. Work dome on the gas in isothermal compression =         = 1 2 e 1 1 i P P log V P W Work done on gas for non-isothermal compression =           −               − = =       − 1 P P 1 n n V P W n 1 n 1 2 1 1 n where, W = work done per unit mass of gas P1 = inlet pressure P2 = outlet pressure V1 = initial volume of unit mass of gas n = value of exponent in equation of compression n PV = constant For adiabatic compression n = γ = ratio of specific heats. Compressor efficiencies (based on the isothermal power required vary between 45 (low compression rotary compressors) and 60% reciprocating piston compressor)). 12.0 FURTHER READING Most fan manufacturers have published their own technical books, the most comprehensive being 1) "Fan Engineering" Buffalo Forge Company, Buffalo, New York. (1961). A cheaper book, and perhaps a little more readable is 2) "Fans" W.C. Osborne Pergamon 1966 Of use for definitions is 3) British Standard 848 : Pt I ; 1963 (Fans)
  • 35. APPEINDIX I Effect of Speed and Temperature Change Given the fan curves shown at 50o C and 200 rpm what are the fan curves for running conditions of 0 o C and 220 rpm. Step 1 Calculate effect of speed and temperature on individual points of curve:- Speed change factor = 10 . 1 200 220 =       Density change factor = 18 . 1 273 323 =       Pressure factor = ( ) 18 . 1 10 . 1 p p n n 2 1 2 2 1 2 × =                 = 1.43 Volume factor = 10 . 1 n n 1 2 =         HP factor = 57 . 1 p p n n 1 2 3 1 2 =                 Step 2 Pick out reasonably spaced points on the original fan curves as follows:- Volume (cfm x 1000) 0 10 20 30 40 50 60 70 80 90 100 Fan Total Pressure (ins Wg) 9.7 9.8 9.9 10.0 9.7 9.2 8.2 6.8 5.1 3.4 1.5 HP 45 50 62 76 87 96 99 99 97 93 80
  • 36. Step 3 Multiply by the appropriate factors to obtain the new fan curves:- Volume 0 11 22 33 44 55 66 77 88 99 110 PT 13.9 14.0 14.2 14.3 13.9 13.1 11.7 9.7 7.4 5.0 2.1 HP 71 78 97 119 137 149 155 155 152 146 126
  • 37. APPENDIX II Prediction of Fan Noise Levels : Graham and Christie Method (1966) PWLB = PWLs + 70 Log (D) + 50 Log (S) - 259 PWLB = Octave band power level (re 10-12 watt) PWLs = Specific power level in each octave band (re 10-12 watt) D = Impeller diameter (ins) S = Fan Speed (rpm) PWLs is shown in Table 3. 5 dB should be added to the octave band containing the blade passage frequency to account for rotational noise. This frequency is given by f = N x S/60 cycles/see where N = number of blades
  • 38.
  • 39. APPENDIX III Effect of Water Injection on Fan Performance Water injection into an air stream before a fan will cool the gas and the effect will be an increase in density, the cooling effect being much larger than density reductions due to the lower molecular weight of steam. The fan curve will be changed in accordance with the fan laws, the denser gas resulting in the fan doing more work. In fact this effect is so great that water injection will result in an increased mass flow of gas measured at a point upstream of the injection point. As an example, the original Hope Preheater fan curve has been taken. The system has been simplified to that in Fig. 26.
  • 40. Example Fan curve as given (Fig.27) Initial fan gas flow = 4900 m3/min at 350 o C Gas density = 0.64 kg/m3 Gas molecular weight = 32.7 Water injection rate = 1820 gph = 137 kg/min Latest heat of steam = 540 cal/g Specific heat of steam = 0.5 cal/g o C Initial water temp. 15 o C Problem Given the above information what will the increase in kiln gas flowrate be on water injection ? Step 1 Find temperature after injection Mass of gas = 4900 x 0.64 = 3140 kg/min Let new temp = T o C Heat balance:- 3140 x 0.25 (350 – T) = 137 (540 + 85(l) + (T - 100) 0-5) T = 230 o C Step 2 : Calculate new gas density Total gas flow = 3140 kg/min at MW 32.7 Plus 137 kg/min at MW 18 = 3277 kg/min at MW 31.6 New density = 0.64 x =       ×       503 623 7 . 32 6 . 31 0.766 kg/m 3
  • 41. Step 3 : Modify fan curve Original fan curve :- Volume 0 1 2 3 4 5 6 7 Pressure 612.5 665 725 745 725 675 600 510 Power 290 410 525 630 720 790 840 Density ratio = 195 . 1 64 . 0 766 . 0 =       Therefore, volume ratio = 1.0 Pressure ratio = 1.195 Power ratio = 1.195 New fan curve:- Volume 0 1 2 3 4 5 6 7 Pressure 732 795 868 891 868 808 718 610 Power 348 490 630 755 860 945 1010 Step 4 Here, we correct the new fan curve so that the volume it applies to are at the conditions upstream of the water injection point. This is done because the system resistance line we draw in will refer to volumes at these conditions. The calculation is carried out by considering successive volumes of gas at 350 o C reaching the injection point. e.g. Volume flow = 3000 m 3 /min at 350 o C Therefore, mass flow = 3000 x 0.64 = 1920 kg/min Water required to cool.this to 230 o C = 137 x       3140 1920 = 83.8 kg/min Therefore, total mass of gas reaching fan = 83.8 + 1920 = 2003.8 kg/min
  • 42. Volume of gas at fan = =       766 . 0 8 . 2003 2616 m 3 /min Fan static pressure at this fan volume (from curve l) = 887 mm wg The following table is completed in the same way:- Volume approaching Volume.Reaching Fan Static Injection point (m 3 /min) Fan (m 3 /min) Pressure (mm wg) 0 0 732 1000 872 787 2000 1744 846 3000 2616 887 4000 3488 885 5000 4360 850 6000 5232 735 7000 6104 705 A plot of the first column against the last column is then the requiread curve (curve 1K) Step 5 : Determine System Resistance Line Assume P = kV2 (P = 680, V= 4900) is the original, operating point Hence system resistance line is:- Volume 4 5 6 7 Pressure 453 707 1019 1385 Then the new operating point will be at the intersection of the S.R.L. and the new fan curve. i.e. 5400 m3 /min (intersection with curve 1K) Step 6 : Determine gas flows Kiln gas mass flow = 5400 x 0.64 = 3460 kg/min
  • 43. Therefore, increase in gas flow = 3460 - 3140 = 320 kg/min However, we now reach a contradiction in the calculation. if the kiln gas flow increased to 3460 kg/min, then the addition of 137 kg/m, in of water would not quite cool it down to 230o C. Unfortunately, 230o C is the temperature upon which we have based our fan curve calculations. To cool the extra kiln gas down to 230o C we would need a further 137 x =       3140 320 14 kg/min of water giving a total injection rate of 151 kg/min In other words we have assumed an injection rate of 137 kg/min but found that we in fact need 151 kg/min to cool the gas down to the temperature at which we have made our calculations. Step 7 What we must do now is to make a better guess' at a water rate that we may assume to carry out our fan curve calculation, so that the total water rate, calculated at the end of Step 6, 13 equal to the true water rate. This may be done by proportion New guess = 137x =       151 137 124 kg/min We then use this value, 124, in place of 137 kg/min and repeat Steps 1-6. Step 5 then gives us an operating point of 5350 m3 /min (intersection with curve 2K) = 3420 kg/min. Extra gas = 3420 - 3140 = 280 kg/min Water required to cool extra gas down to new temperatures of 240o C =124 x =       3140 280 11 kg/min Therefore total water flow = 124 + 11 = 135 kg/min This is sufficiently close to the true water rate (137) and further iteration is not necessary.
  • 44. Power Drawn When the operating point is at 5350 m3 /min (curve 2K) the fan will handle 4700 m3 /min (curve 2). At this volume the fan shaft power will be 820 kW (curve 2P), whereas prior to water injection the fan power was 715 kW. Hence we may draw up the following table:- Before Af ter Increase Water Water Injection Injection Kiln gas flow (kg/min) 3140 3420 8.9% 'Water rate (kg/min) - 135 Fan Temp (00 350 240 - Fe.n Power (kW) 715 820 14.7% Water injection, therefore, may in some cases be used as a means of uprating the useful Volume handled by a fan. If the extra gas that is drawn were not required one would, of course, damper back along the curve 2K, or reduce the fan speed until the original kiln gas flow was attained.
  • 45.
  • 46. Blue Circle Cement PROCESS ENGINEERING TRAINING PROGRAM Module 5 Section 2 Fans and Airflows
  • 47. FANS AND AIRFLOWS 1. Principles of Fan Performance 2. Terms Used In Fan Engineering 3. Fan Blade Types 4. Fan Curves 5. Control of Fan Output 6. Unsatisfactory Fan Performance Appendix I The Fan Laws Appendix II Fan Checks
  • 48. FANS AND AIRFLOW 1. PRINCIPLES OF FAN PERFORMANCE As a first step to an understanding of fan performance it is necessary to get a clear conception as to the manner in which a fan actually operates. The purpose of a fan is to establish and maintain between the fan inlet and fan discharge a difference in pressure of the air or gas it is handling, and it is in consequence of this maintained difference in pressure that, circumstances permitting, the volume of air or gas flows through the external circuit to which the fan is connected. The fan has to set up a difference in pressure sufficient to overcome the losses in the system, including its own internal losses. 2. TERMS USED IN FAN ENGINEERING 2.1 Energy When considering fan problems it is seldom that the energies imparted to the air are referred to in terms of potential and kinetic energy. Instead the term water-gauge is used (usually expressed in inches). a) Potential Energy The air within a duct system connected to a fan will be at a different pressure to the atmosphere outside the duct and will have a capacity for doing work in virtue of that difference in pressure, i.e.,'it will have potential energy. b) Kinetic Energy The air will be in motion and will have energy by virtue of that motion, i.e., it will have kinetic energy. c) Total Energy The total energy is the algebraic sum of the potential and kinetic energies. In all fan work the pressure of the atmosphere surrounding the fan at sea level is taken as the datum, and pressures are considered positive or negative, according as they are above or below atmospheric pressure. Therefore, since atmospheric pressure is taken as a datum and since the fan produces a difference in pressure it follows that the air only possesses potential and kinetic energy by reason of being connected to the fan, and therefore the energy must be supplied by the fan. Within limits of the total energy in the air, the ratio of the amount of energy present as potential to the amount of energy present as kinetic to the amount present as kinetic is governed entirely by the external system to which the fan is connected.
  • 49. This can be summarized as a fundamental principle: “The resistance of the external system controls the volume which the fan can handle at a fixed speed." 2.2 Pressure Fig. I shows diagrammatic sketches of fan pressure terms. d) Static Pressure Static pressure is the bursting or collapsing pressure according to whether the difference in pressure between the inside and outside of the duct is positive or negative, that is, above or below atmosphere. Thus the static pressure is a measure of the capacity of the air within the duct for doing work, and is therefore a measure of the potential energy. e) Velocity Pressure Velocity pressure is a measure of the kinetic energy which the air possesses by virtue of its motion can be measured by the Pitot Tube. f) Total Pressure Total pressure is defined as the algebraic sum of the static pressure and velocity pressure. Since the velocity pressure can only be measured by impact it must always be a positive quantity, whether the flow be on the discharge side of the fan, where the static pressure is positive, or on the suction side where the static pressure is negative. It follows, therefore, that upon the discharge side of the fan the total pressure will be numerically greater than the static pressure, and upon the suction side of the fan the total pressure will be numerically less than the static pressure. g) Water-Gauge The pressures to be measured are usually quite small and it is usual to refer the pressures to the height of a column of water which could be sustained, hence the term "water-gauge". In fan work, since the pressures generated are so small, the variations in air volume, due to compression or expansion, are also quite small, and such variations are usually neglected. Variations in temperature and barometric pressure, however, produce considerable -changes in volume and air density. (A cubic foot of water weighs 62.4 lbs at 60 o F so that a cube of water with 12" sides would exert a pressure of 62.4 lbs over the area of its base which is one square foot, so that a water-gauge of 12" corresponds to a pressure of 62.4 lbs per square foot, or 1" water-gauge represents 5.2 lbs per square foot. Hence to transform a pressure expressed in terms of inches water-gauge to a corresponding pressure in lbs per square foot the water-gauge is multiplied by 5.2).
  • 50. 2.3 Manometers The apparatus which actually indicates the water-gauge is known as a manometer (Fig. 2). For use on site it generally consists of a plain glass U tube, one leg of which is connected to the duct where the pressure is to be measured, the other leg being open to atmosphere. The pressure is then obtained by the difference in the height of the water in the two legs of the tube. The correctness of the readings of a manometer will be dependent upon the accuracy with which the difference in level of the height of the columns of water in the two legs of the manometer is measured. It is essential that the U tube be held in a vertical position. The difference in the levels of the water can then be measured carefully by means of a scale attached to the gauge. The water in the tube will tend to cling to the surface of the tube, and hence the surface of separation between the air in the tube and the water will be cup-shaped, the smaller the bore of the tube the more curved will be the surface of the water. If is usual to measure to the bottom of the cup-shaped surface (or meniscus as it is called). For accurate work the cup-shaped surfaces must be the same, i.e. the tubes must be of the same bore. It is usual to specify that the bore of such a U tube shall not be less than, say, 3/8” diameter. 2.4 Measuring Fan Static Pressure Fan total pressure PT, is the difference between the total pressures at the fan inlet and outlet. Fan static pressure, PS, is the fan total pressure minus the fan velocity pressure. Fan velocity pressure, PV, is the velocity pressure corresponding to the average velocity at the fan outlet. i.e. the fan static pressure is NOT equal to the difference between the static pressures at the inlet and outlet. It should be measured as the difference between the reading of a facing tube at the inlet and a side tube at the far outlet (see Fig. 3). 3. FAN BLADE TYPES Basically, three types of impeller are used in centrifugal fans: a) Forward-Curved Blades (Blade angle 90o). b) Radial Tipped Blades (Blade tip 90o). c) Backward Sloping Blades (Blade angle 90o). Fig. 4 shows typical fan curves for each type. In general for a given impeller size, the pressure developed increases with blade angle. 3.1 Forward Curved Blades
  • 51. These fans give the highest pressure for a given size and speed. They are used for high volume applications such as ventilation, where their compact size and low speed (hence low noise) is an advantage. The low speed also allows a lighter, less expensive construction to be used. The blades are of short radial depth and consequently more of them (30 - 60 compared with 6 - 16 for other fans) are required to have an effect on the air. It can be seen (Fig. 4) that a decrease in restriction and hence an increase in airflow will cause an increase in power. The fan is said to be of the overloading type and this effect may cause the motor to trip out. These fans are unsuitable for dirty air, the narrowly spaced tips tending to bridge. Efficiencies are not greater than 75%. 3.2 Radial Bladed Fans These are of two types: 3.2.1 Radial Bladed Paddle Fans e.g. firing fans These have their main application in handling high dust burdens. They are of simple construction and therefore the wearing parts can be of substantial material or even replaceable. The impeller is unshrouded, but may have a back plate in the case of fibrous dust to prevent build-up on the shaft. Disadvantages of the paddle fan are its low efficiency (about 60%), high speed of rotation (noisy operation) and overloading power characteristics. 3.2.2 Radial Tipped Forward Curved Fans These fans combine some of the ruggedness of the paddle fan with the higher efficiency and higher pressure development of the forward curve blade. The impeller is self-cleaning to a large extent and is used extensively for dusty flue gases. Blades themselves are not easily made removable but replaceable wear plates and nose pieces can be fitted if dust burdens are high. Efficiencies higher than 75% are obtainable with this fan and it is therefore commonly used where high dust burdens prevent the use of the more efficient aerofoil or backward inclined bladed fans.
  • 52. 3.3 Backward Bladed Fans These fans are used mainly for clean air applications, although the use of a reinforced nose, or leading edge, will allow low concentration of dust to be handled. Efficiencies up to 90% are obtainable with fans having aerofoil section blades, and therefore the fan is preferred for the larger machines. The high efficiencies are a result of laminar flow over the blade profile. Backward bladed fans are non-overloading (see power curve on Fig. 4). 3.4 Blade Section Summing up, the progressive use of fan types as the dust burden increases would be:- Forward Curved Backward Aerofoil Backward Inclined Forward Curved Radial Tip Paddle 4. FAN CURVES These show the effect on fan pressure, power and efficiency of a change in fan throughput (Fig. 5). The horse-power on a fan curve is that which must be delivered to the shaft and does not take into account losses in couplings and motors. P may be either the fan static or total pressure and the efficiencies thus calculated are the static and total efficiencies respectively. 4.1 System Resistance The resistance of the system served by a fan depends upon the hardware in it and for any system there will be a relationship between flowrate and pressure drop. If the system contains only dampers, ducting, expansions etc., of fixed resistance, then the pressure drop may be taken as being approximately proportional to the square of the flowrate (V2). This would not be true in the case of a resistance (as opposed to pressure drop), that varied with flowrate, such as a cooler grate. The intersection of the system resistance line with the fan pressure/ f lowrate curve gives the point of operation of the fan. The effect of, for instance, closing a damper is to steepen the system resistance line, giving a new operating point (see Fig. 6). 5. CONTROL OF FAN OUTPUT For a centrifugal fan there are three basic methods of controlling output- a) Damper Control. b) Variable Radial Inlet Vanes - these give a beneficial swirl to the air at the fan inlet and therefore reduce the work the fan has to do.
  • 53. c) Speed Control. Initially, dampers are the cheapest to install and speed control the most expensive, but there is a reversal in a running cost comparison. However, speed control usually only becomes economic on the larger machines. Whereas variation in a damper setting will change the system resistance and not the fan curve, the other two methods produce new fan curves, that for speed control retaining the same fan efficiency. Referring to Fig. 7, the effect of turndown to ¼ V and ½ V by the three methods is shown. Graph A - Damper Control - Fan curve retained but system resistance line altered. Graph B - Variable Inlet Vanes - New fan curves produced for each vane setting. Graph D - Comparison of three methods on basis of power taken. (Note: loss due to coupling is exaggerated.) Graph C - Speed Control - New fan curves produced for each speed but efficiency retained. This graph assumes no pressure drop across a fully open damper. This is a reasonable assumption for some dampers operating in clean air, but would not be true for some ID fan dampers where build-up will occur. The expected pressure drop across an open damper, based on experienced, should be taken into account when deciding on the method of control. 6. UNSATISFACTORY FAN PERFORMANCE Fan performance may be affected by worn impellers, inefficient drives, etc., but the most common causes of fan performance being substandard are bends close to the inlet or outlet of the fan. If the bend is at the inlet, the air is pushed over to one side of the eye, whereas a manufacturer's fan curve is for uniform inlet conditions. Uniformity of flow at inlet is easily checked by a pitot traverse. The best solution to this is an inlet box which can be sized so that the air enters the impeller uniformly, failing that a square bend with cascade (Fig. 8) will give good performance. Retention of the original bend and fitting of long turning vanes will improve matters greatly (Fig. 9). Useful pressure in a fan originally comes from the acceleration of the air. As the air is again slowed down by expansion, velocity pressure is converted into static pressure. For this to be done efficiently the expansion should be slow and gentle and clearly a bend immediately on the fan outlet will not help matters. This problem is not so easy to remedy but again a cascade will help.
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  • 61. APPENDIX I THE FAN LAWS The performance of a fan may be changed by several methods, the main ones being a change in:- a) Fluid density, p. b) Impeller speed, n (rpm). c) Fan size*. Most text books only quote the fan laws for “fans of geometric similarity", so that diameter, width, outlet duct size, etc., are all increased at the same rate. Any convenient dimension may then be taken as being representative of the fan, e.g. Impeller Diameter, Dim. The effect of these on fan pressure, volume and horse-power is summarized in Table I. TABLE I - The Fan Laws Alteration In Change In Fan Pressure Volume Power Density, p α p No change α p Speed, n α n2 α n α n3 Fan Size*, Dim α Dim 2 α Dim 3 α Dim 5 Fan pressure may be either total, static or velocity. In order to help remember the above table, it is worth noting that fan pressure can be shown to be theoretically proportional to the square of the tip speed whereas volume is proportional to the swept volume of the impeller (fan size/volume is a cubic relationship since width increases as well as diameter). Power is proportional to the product of volume and pressure. APPENDIX II FAN CHECKS The following is given as a guide to minimize losses in fan & associated plant efficiencies due to mechanical/process deteriorations of fans. 1) Check to ensure that fan inlet & discharge ductwork is not restricted with material drop out ( or plates, etc. left inside following major repairs.) 2) Check to ensure that control dampers are functioning correctly, indicating correct position externally ( eg. with louvre type dampers that all louvres can move freely, and together 3) Check to ensure fan speed is correct ( eg. on belt drives check belts are not slipping. Check pulley sizes are correct). 4) Check fan power absorbed at motor. 5) Check fan inlets to ensure. a) At entrances no material deposits exist.
  • 62. b) At fan eye that there exists a close concentric fit of duct to impeller, to eliminate recirculation, within the fan casing. c) Within casing, no wear exists at nose bridge causing turbulence. 6) Check to eliminate all inleaks at duct Joints, fan inlet joints, doors, etc., on fan suction side. 7) Check to verify fan impeller dimensions ( dia. etc. ) are as designed. 8) Check to ensure there is no wear on fan blades ( and repair as necessary). 9) General inspections inside fan casing are necessary to ensure no obstructions exit, or wear of body, that turbulence could be created. The above checks are necessary at shutdowns, particularly at the beginning to assess work/repairs necessary and also before start up to ensure correct repairs have been carried out, & no extraneous materials have been left inside casings & ductwork. In addition to the above physical condition checks, Technical Departments should carry out process fan efficiency checks during normal stable fan operation. This is essentially to verify that the fans are operating on their fan curves ( ie. as designed ) and therefore corrective measures can be planned/ programmed for shutdowns if necessary. To check fan curve operation the following measurements are necessary. 1) Gas volume handled by fan, by anemometer or pitot tube measurements as applicable. ( Care should be taken to maximize accuracy of such measurements ). 2) Fan total pressure & static pressure. ( Care Should be taken to maximize accuracy and also use of this data according to fan static pressure as defined previously). 3) Fan impeller R.P.M. 4) Fan motor power. ( Note curve power is shaft power only) The above data can then be compared with curve specifications using the fan laws. Note coupling/ralley /belt efficencies compared to motor power mean that motor power is greater than fan shaft ( curve ) power. If the operating point of the fan is found to be significantly off the design curve, it is suggested that the test is repeated, for consistency, & than corrective actions/inspections can be programmed as necessary.
  • 63. 1.0 DEFINITION OF FAN PRESSURE Fan total pressure, PT, is the difference between the total pressures at the fan inlet and outlet Fan static pressure, PS, is the total pressure minus the fan velocity pressure. Fan velocity pressure, PV, is the velocity pressure         g u 2 2 corresponding to the average velocity at the fan outlet. i.e. the fan static pressure is NOT equal to the difference between the static pressures at the inlet and outlet. It should be measured as the difference between the reading of a facing tube at the inlet and a side tube at the fan outlet (see figure 1).
  • 64. These show the effect on fan pressure, power and efficiency of a change in fan, throughput (Fig.2). The horse-power on a fan curve is that which must be delivered to the shaft and does not take into account losses in couplings and motors. A calculation of the work actually done on the air will therefore give the efficiency of the fan itself. Work done on air = pressure x flowrate x       000 , 33 2 . 5 ) fm and wg (inches HP 6360 c PV       = or 9.81 PV watt (mm wg and m3/sec). P may be either the fan static or total pressure and the efficiencies thus calculated are the static and total efficiencies respectively. Because work is being done on the air, its internal energy will rise, and this, together with losses due to the inefficiency of the fan will cause the air temperature to rise.         × × = heat specific V density power fan rise e Temperatur For air at ambient conditions ( ) 0.24 V 075 . 0 779 5.2 PV × ×       × × = η efficiency fan total = η P 0.37         = η o F or about ½ o F per inch wg.
  • 65. 3.0 THE FAN LAWS The performance of a fan may be changed by several methods, the main ones being a change in:- a) fluid density, p b) impeller speed, n (rpm) c) fan size* most text books only quote the fan laws for “fans of geometric similarity”, so that diameter, width, outlet duct size, etc., are all increased at the same rate. Any convenient dimension may then be taken as being representative of the fan, e.g. Impeller Diameter, Dim. The effect of these on fan pressure, volume and horse-power is summarized in Table I. TABLE I - The Fan Laws ALTERATION IN CHANGE IN FAN PRESSURE VOLUME POWER Density, p α p No change α p Speed, n α n 2 α n α n3 Fan Size*, Dim α Dim 2 α Dim 3 α Dim 5 Fan pressure may be either total, static or velocity. In order to help remember the above table, it is worth noting that fan pressure can be shown to be theoretically proportiona1 to the square of the tip speed whereas volume is proportional. TULSA FAN Number 2 COMPARTMENT 1750 RPM 1284 RPM 1687 RPM 0.7337 0.5383 0.3950 0.964 0.9293 0.8958 Vol. I cfm x 1000 Press I “ w.g. HP 1 B. H. P. Vol. 2 cfm x 1000 Press 2 “w.g. HP 2 Vol. 3 cfm x 1000 Press 3 “w.g. HP 3 0 17.6 12 0 9.47 4.74 0 16.36 10.75 5 18.8 32 3.67 10.12 12.64 4.82 17.47 28.67 10 19.5 47 7.34 10.50 18.56 9.64 18.12 42-10 15 19.0 63 11.01 10.23 24.88 14.46 17.66 56.44 18 18.0 71 13.21 9.69 28.04 17.35 16.73 63.60 20 17.2 76 14.67 9.26 30.02 19.28 15.98 68.08 25 13.8 92 18.34 7.43 36-34 24.10 12.82 82.42 30 9.3 104 22.01 S.01 41.08 28.92 8.64 93.17 35 3.2 114 25.68 1.72 45.03 33.74 2.97 102.12
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  • 69. Blue Circle Cement PROCESS ENGINEERING TRAINING PROGRAM Module 5 Section 3 Fans and Fan Systems, Chemical Engineering March 1983
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  • 86. Blue Circle Cement PROCESS ENGINEERING TRAINING PROGRAM Module 5 Section 4 Fans FL Smith
  • 87. FANS Abstract In this lecture the most important fan types used in cement plants are presented and their advantages and disadvantages are outlined. The definition of fan efficiency and simple fan calculation methods are discussed and it is demonstrated how to save energy by optimum fan regulation and modification of existing fans with high efficient impellers. Table of contents: 1. Introduction 2. Fans functioning principles 3. Centrifugal fan types used in the cement industry 3.1 Single suction type with open overhanging impeller 3.2 Single suction type with closed overhanging impeller 3.3 Single suction type with closed impeller suspended between the bearings 3.4 Double suction type with closed impeller 4. Pan calculation and dimensioning methods 4.1 Basic definitions 4.2 Theoretical fan power and efficiency 4.3 Model fan curves 4.4 Calculations assuming imcompressible flow 4.5 Fan dimensioning assuming imcompressible flow 4.6 Calculations considering gas compressibility 4.7 Correction for dust laden gas 5. Fan drives 5.1 Direct/indirect drive 5.2 Motor type and speed regulation 5.3 Drive efficiencies 6. Capacity regulation 6.1 Damper regulation 6.2 Inlet guide vanes 6.3 Fan speed regulation 7. Fan modifications 7.1 Impeller replacement 7.2 Energy savings Fans
  • 88. 1. Introduction Fans are used in all parts of a cement plant and the operational success of nearly all main machinery in the cement plant is completely dependent on the reliability and capacity of the connected fans. The specific electrical energy consumed of the total number of fan motors included in a modern cement production unit normally amounts to 30-40 kWh/t clinker which makes the fan motors the biggest power consumers at the cement plant after the mill motors. Thus, it is very important that the fans included in a cement plant are very carefully selected and dimensioned, and the plant operators must know how to operate the fans most efficiently. 2. Fans functioning principles Very broadly speaking the function of a fan is to create a flow of gas or air in the system to which the fan is connected. In the fan the necessary energy to create the desired flow through the system is transferred to the air or gas by a rotating impeller driven by the fan drive (f.inst. an electrical motor). Fans are normally characterized as either axial or centrifugal fans. In an axial fan the flow is predominantly axial, that is in the direction of the impeller axis whereas the flow in the centrifugal fan is predominantly radial (that is perpendicular to the impeller axis) in the region where the energy transfer takes place. In the cement industry centrifugal fans are the most commonly used and in the following we will therefore only deal with centrifugal fans though much of the information, formulae etc. may also apply to axial fans. 3. Centrifugal fan types used in the cement industry A centrifugal fan can be categorized according to its various design details such as: inlet design, impeller type, blade type, impeller position etc., see table 3.1.
  • 89. Table 3.1: Design details which characterize a centrifugal fan. Impeller type: Open, paddle type Open with back-plate Closed (shrouded) Inlet type: Single axial inlet Single, with inlet box Double, with inlet boxes Blade form: Radial Forward curved Straight backwards inclined Curved backwards inclined Air foil backwards inclined Impeller position: overhanging Between bearings Bearing types: Roller bearings, grease lubricated Roller bearings, oil lubricated Journal bearings Other features: Inlet vanes Wear protection Shaft cooling Further, fans are often characterized according to their function or position in the system in which they create a flow of gas or air, namely as: 1. Blowers, i.e. fans which supply (fresh) air or gas to a system 2. Exhausters or induced draught (i.D.)fans, i.e. fans which remove air or exhaust gas from a system 3. Circulation or recirculation fans are fans which maintain a closed loop air or gas flow in a system. In the cement industry the fans are often termed by the machine system to which they are connected, f.inst.: 1. Kiln I.D. fan 2. Raw mill fan 3. Cement mill fan 4. Kiln by-pass fan 5. Precipitator or filter fan 6. Dedusting fan, etc. See also fig. 3.1.
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  • 91. Sometimes fans are also categorized as either low pressure, medium pressure or high pressure fans, according to the differential pressure rating of the fan. However, the pressure range for the different categories has not been standardized. In the following, four fan types commonly used as kiln I.D., and/or mill circulation fans in the cement industry will be presented and the advantages and disadvantages of the individual types will be explained. 3.1 Single suction type with open overhanging impeller Fig. 3.2 shows a centrifugal fan with open impeller. The air or gas is introduced in the fan in axial direction, proceeds radially through the impeller and leaves the fan through a tangential outlet in the fan casing (for the terminology used here,,-please refer to the legend of the figure). The impeller has bolted on plane, radial blades which can be changed when worn or major changes in the fan performance are required. The impeller is overhanging, that is both shaft bearings are placed on the same side of the fan casing. This provides for the use of a simple axial inlet to the fan, but the load on the bearing closest to the impeller is very big and the bearing has to have a diameter corresponding to maximum shaft diameter. Normally this fan type is designed for a fairly low specific speed, that is for a given rating the fan will be designed with a relatively large impeller running with a low rotational speed. The advantage of this type of fan is that with the radial blades the fear of dust accumulation on the backside of the impeller blades is eliminated and the low rotational speed makes the impeller less exposed to wear from dust contained in the gas being conducted through the fan. Further, should the fan blades get worn, they are easily replaceable. on the other hand this fan type exhibits a very low efficiency (65-70%) and the front side of the radial blades are often exposed to hard coating formation when operating with kiln exhaust gas. Due to its low efficiency this type of fan will probably only be used in special cases in the future, f-inst. when the gas to be transported contains high amounts-of abrasive dust.
  • 92. Fig. 3.2: Single suction fan with open overhanging paddletype impeller 1. Fan outlet 5. Blade 9. Thermal shield 12. Bearing (locating bearing) 2. Fan casing 6. Access door 10. Shaft guard 13. Shaft 3. Fan inlet 7. Cooling disc 11. Base frame 14. Foundation 4. Impeller 8. Beating
  • 93. 3.2 single suction type with closed overhanging impeller Fig. 3.3 shows a centrifugal fan with closed overhanging impeller. The impeller blades extend between the two impeller shroud plates also termed the cover plate and the back plate. Thus, the gas or air flows through the impeller in "closed" channels. This design allows for a wide variation of the blade form and inclination (angle with the impeller tangent). Normally backwards inclined blades are used as this ensures a higher fan efficiency than radial and forward curved blades. The fan in fig. 3.3 has relatively short, moderately backwards inclined S-shaped blades. This blade form increases the impeller stability, and at the same time reduces the impeller surface area exposed to dust accumulation. The efficiency of this fan type is approx. 78%. Fig. 3.3: Single suction fan with shrouded overhanging impeller with S-formed blades. 1. Fan Outlet 5. Blade 8. Bearing 11. Base Frame 2. Fan Casing 6. Access Door 9. Shaft guard 12. Shaft 3. Fan Inlet 7. Cooling Disc 10. Bearing (Locating) 13. Foundation 4. Impeller
  • 94. 3.3 Single suction type with closed impeller suspended between the bearings. For a given tip velocity an impeller with backwards inclined blades exhibits at lower pressure rating than impellers with radial or forward curved blades. Thus, a high pressure rating affords a very high impeller tip speed for which it is difficult to design the fan shaft and bearings in case the impeller is overhanging. Thus, for high pressure fans with backwards inclined blades the impeller is normally placed between the shaft bearings. Fig. 3.4 shows a high efficient fan designed according to this principle. Fig. 3.4: Single suction fan with shrouded impeller with backwards curved blades suspended between the bearings. 1. Fan outlet 5. Blade 9. Shaft 2. Fan casing 6. Cooling disc 10. Bearing bracket 3. Fan inlet 7. Beating (locating) 11. Access door 4. Impeller) 8. Thermal shield 12. Foundation
  • 95. As the bearing on the inlet side of the fan makes a simple axial gas inlet impossible the fan is provided with an inlet box integrated with the fan casing into which the gas flows radially. The fan shown in fig. 3.4 has backwards curved blades with a tip angle of 50 o . The maximum efficiency of this fan type exceeds 80% despite the inevitable pressure loss in the inlet box. It is designed for an impeller tip speed up to 180 m/s. The major draw back of this fan type is that when working with dust laden gas it is often exposed to soft coating formations on the back side of the impeller blades and/or wear of the impeller due to its high rotational speed. However, various solutions to remedy these draw backs are offered, such as pressurized air cleaning systems, hard facing of the impeller blades etc. 3.4 Double suction type with closed impeller For very large flow capacities double suction fans are used, see figure 3.5. A double suction impeller can be considered as two (laterally reversed) single suction impellers placed on the same shaft and having a common back plate. The impeller is placed in a fan casing having an inlet box on each side. Thus, the double suction fan has two inlet ports, which makes the inlet ducting to such a fan a little more complicated, but on the other hand the capacity of the double suction fan will correspond to two similar single suction fans with the same impeller diameter, and thus the double suction fan will often be the most economical solution for large capacities. Some fan suppliers provide single suction as well as double suction fans with backwards inclined airfoil blades. See figure 3.6. With this blade form a very rigid rotor is obtained even with a small plate thickness in the blades. This blade form also ensures a very high fan efficiency (up to 85%). on the other hand the thin blades will naturally be very sensitive to wear.
  • 96. Fig. 3-5: Double suction fan.(Deutsche Babcock) Fig. 3.6: Cross section of impeller with airfoil blades (Solyvent)
  • 97. 4. Fan calculation and dimensioning methods In this section some basic definitions and "fan laws" are presented and methods for "manual" calculations of fans are outlined. Normally, fans are dimensioned using a number of computer programs available at the fan suppliers, and the purpose with the discussion here is only to give an idea of the principles used for dimensioning of large fans. Further the importance of distinguishing between static and total pressure rating and different fan efficiency definitions is emphasized. 4.1 Basic definitions The operation of a fan (working with a clean gas) can be characterized by: ' 1 Q = the actual flow of gas to the inlet of the fan. ' 1 ϑ = the specific density of the gas of the inlet to the fan and ' t p ∆ = the fan pressure rating expressed as the difference between the total pressure of the gas in the fan outlet and the fan inlet, respectively: 1 t 2 t t p p p − = ∆ Refer also to fig. 4.1. Fig. 4.1 Nomenclature When using small p's for pressure we refer to the gauge pressure that is the pressure stated relatively to the ambient pressure, B (barometric pressure), at the elevation at which the fan is placed. The absolute gas pressure, corresponding to the relative pressure p, is P = p + B (p being stated as a negative pressure when below the ambient pressure)
  • 98. Pressure is measured in a great number of units. In the following mainly the SI unit Pascal is used (1 Pascal = 1 N/m2 = 0.01 mbar = 0.102 mm WG). The total pressure at a certain point in a flow of gas is the sum of static pressure, ps, and dynamic pressure, pd, of the gas at the apoint considered pt = ps + pd pt is the pressure measured through the center hole in a pitot tube directed against the gas flow direction and p is the pressure measured through the side holes of the pitot tube (perpendicular to the flow direction). The dynamic pressure is calculated as pd = pt - ps = ( ) 2 u 2 1 ϑ       in which ϑ is the gas density and u is the gas velocity. Often it is the static pressure rise ∆ps = ps2 – ps1 across the fan which has the interest. This can be calculated from the total fan pressure rating ∆pt as follows: ( )( ) ( )( )2 1 1 2 2 2 t s u 2 1 u 2 1 p p ϑ       + ϑ       − ∆ = ∆ ( ) ( ) 2 1 1 1 2 2 2 2 t A Q 2 1 A Q 2 1 p         ϑ       +         ϑ       − ∆ Herein Q1 and Q2 designate the flow through the inlet/ outlet of the fan, respectively, and A1 and A2 the cross sectional area of the inlet and outlet of the fan. The actual gas density ϑ can be calculated, if the normal density N ϑ (that is the density at 0 o C and 760 mm Hg) of the gas is known, together with the temperature t and relative (static) pressure p:       +       + ϑ = ϑ 760 p B t 273 273 N in which B (the barometric pressure) and p (with sign should be inserted in mm Hg (1 mmHg = 133.3 Pa).
  • 99. Likewise an inlet flow expressed in normal cubicmeters QN can be transformed to actual flow volume Q by         +       + =       ϑ ϑ = p B 760 273 t 273 Q Q Q N N N 4.2 Theoretical fan power and efficiency Assuming that the compression of the gas flow in the fan is an isentropic (or reversible adiabatic) process and that the gas velocity is the same in the fan inlet and outlet we can calculate the fan power input (using a normal thermodynamic nomenclature): ( ) ( ) ( ) ∫ ∫ ∫ ∫ = + = + + = = − = 2 1 2 1 2 1 2 1 1 2 is dP v m dP v ds T m dP v dv P du m dh m h h m P as ds = 0 (isentropic process) The latter expression can for an ideal gas be integrated to (see ref.(l)): ( )( ) ( ) ( ) ( )           −         ∆ +       − =           −               − = − − 1 P P 1 P 1 k k Q 1 P P P 1 k k v m P k 1 k 1 t 1 1 k 1 k 1 2 1 1 is Here k is the adiabatic constant for the ideal gas k = v p c / c Developing the last parenthesis and rearranging we find ( )( ) ( )( )( ) ( ) α ∆ =                 ∆         + + ∆       − ∆ = t 1 2 1 2 2 1 t t 1 is P Q ...... P P k 6 1 k P P k 2 1 1 P Q P The last factor is called the isotropic compressibility coefficient. The relation between α and         1 t P P for atm. air (k = 1.4) is shown in figure below (from ref. (1)):
  • 100. Fig. 4.2: Isotropic compressibility coefficient as a function of relative pressure rise. The isentropic efficiency of a fan is thus defined as ( )( )( )       α ∆ =       = η P P Q P P t 1 is is in which P is the actual power delivered to the fan shaft. As α is close to 1 when the relative pressure rise across the fan         ∆ 1 t P P is small it has been common practice to use the efficiency defined as ( )( )       ∆ = η P P Q t 1 1 for fans generally. This efficiency is also termed the inlet efficiency. It must be noticed that 1 η is not a real efficiency as 1 η would be greater than 1.0 for an ideal (high pressure) fan. One should always be aware that the fan efficiencies stated by different sources might have a different definition. 4.3 Model fan curves The dimensioning of large industrial fans is normally based on the experimental results from a performance test with a smaller fan model similar to the large fan. The results from the model performance test might be expressed as performance curves, as illustrated in figure 4.3. The "left" (full line) curve establishes the relation between the total pressure increase tm P ∆ obtained across the model fan as a function of the volume flow lm Q measured at the inlet to the fan.
  • 101. Correspondingly the "right" (full line) curve establishes the dependence between the volume flow lm Q and the fan inlet efficiency lm η defined as ( )( )         ∆ = η m tm lm lm P P Q in which m p is the measured fan power consumption (at the fan shaft) at the flow lm Q . 4.4 Calculations assuming imcompressible flow If the compressibility of gas flow through the fan can be neglected (that is assuming the volume of the gas does not change when passing the fan) the performance of a fan similar to the model fan can be calculated by means of the three basic relations: ( ) ( ) 1 Q D D n n Q m 3 m m                 = ( ) ( ) 2 P D D n n P tm 2 m 2 m m t ∆                         ϑ ϑ = ∆ ( ) 3 lm m η = η = η m Q is the flow through the model fan (= ml Q for incompressible flow) Q is the flow through the actual fan tm P ∆ is the total pressure rating of the model fan t P ∆ is the total pressure rating of the actual fan D is the actual fan diameter m D is the model fan diameter n is the rotational speed of the actual fan m n is the rotational speed of the model fan ϑ is the specific weight of the gas drawn through the actual fan m ϑ is the specific weight of the gas drawn through the model fan η is the inlet efficiency of the actual fan and m η is the efficiency of the model fan Thus, the flow/pressure curve at any rotational speed and gas density for any fan size in the same series as the model fan can be calculated from the model fan curves using the "fan laws" (1), (2) and (3).
  • 102.
  • 103.
  • 104. Likewise a flow/fan power characteristic curve can be calculated using the expression ( ) ( ) 4 Q P P t         η ∆ = based on the definition of the fan efficiency η Figure 4.4 shows actual fan performance curves for a kiln I.D.fan, which has been calculated from the model fan curves in figure 4.3. (Note the curve has been drawn by a computer which also corrects for the gas compressibility, thus the curves derivates slightly from the curves found using the above mentioned calculation methods). 4.5 Fan dimensioning assuming incompressible flow The optimal fan (that is the fan operating with maximum efficiency) for given values of t P ∆ , Q and ϑ is characterized by a diameter Dideal and a rotational speed n ideal Dideal and n ideal can be calculated from (1) and (2) by insertion: ( ) ( ) 5 Q D D n n Q mideal 3 m ideal m ideal                         = ( ) ( ) 6 P D D n n P tmideal 2 m ideal 2 m ideal m t         ∆                         ϑ ϑ = ∆ where mideal Q is the flow at which the model fan obtains maximum efficiency max m m η = η and tmideal P ∆ is the corresponding model fan pressure. Combining (3) and (4) we obtain the following expressions for direct calculation of Dideal and n ideal ( ) ( ) 7 D Q Q P P D m 5 . 0 mideal 25 . 0 t tmideal m ideal                         ∆ ∆ × ϑ ϑ = ( ) ( ) 8 n D D Q Q n m 3 ideal m mideal ideal                         = once the ideal fan diameter has been calculated a standard fan diameter close to the ideal diameter can be chosen. The corresponding rotational speed might be calculated by iteration using the model fan curves and the expressions (1) and (2), as illustrated by the following example. Example 4.1
  • 105. A raw mill fan (working with dust free air) of the HAF type should be dimensioned with max. efficiency at the following operational data: Flow at inlet s / m 100 Q 3 1 = Temperature at inlet C 100 t o 1 = Pressure at inlet Pa 9810 - WG mm 1000 p1 = − = Specific gravity of gas at inlet 3 1 kg/m 83 . 0 = ϑ Total pressure increase across the fan Pa 9810 WG mm 1000 pt = = ∆ Barometric pressure B = 720 mmHg Neglecting the flow compressibility in the model fan as well as in the actual fan the calculation goes as follows: From figure 4.3 we find that the model fan obtains max. efficiency at mideal Q = 6.5 s / m3 corresponding to a pressure rating of tmideal P ∆ = 4140 Pa. The ideal fan diameter can next be found by insertion in (7). ( ) m 17 . 3 m 1 . 1 5 . 6 100 9810 4140 2 . 1 83 . 0 D 5 . 0 25 . 0 ideal =                     × = and the ideal rotational speed (according to (8)) to ( ) rpm 932 rpm 1450 17 . 3 1 . 1 5 . 6 100 n 3 ideal =                     = Choosing the standard fan impeller HAF 315/315 and assuming a rotational speed of 940 rpm we find from (1) ( ) ( ) s / m 57 . 6 s / m 100 15 . 3 1 . 1 940 1450 Q D D n n Q 3 3 3 3 m m m =                     =                     = ( ) ( ) Pa 9820 4120 1 . 1 15 . 3 1450 940 2 . 1 83 . 0 P D D n n P 2 2 tm 2 m 2 m m t =                           =         ∆                         ϑ ϑ = ∆ which is very close to the desired t P ∆ . Thus further iteration is not necessary and a HAF 315/315 operating at 940 rpm would be the right choice.
  • 106. The inlet efficiency for the actual fan will be equal to the efficiency for the model fan at 6.57 s / m3 which is read to 81.2% thus the shaft power for the actual fan is calculated to ( )( ) ( )( ) kW 1208 812 . 0 N/m 9810 s / m 100 P Q P 2 3 1 t 1 =       =       η ∆ = With a safety factor for the fan motor of 15% the fan motor must have a minimum shaft power of 1390 kW. 4.6 Calculations considering gas compressibility In practice a fan always works with compressible flow and a treatment using the above mentioned calculation methods, which assume an incompressible medium, always introduces some error in the calculations. As long as the fan works with a pressure difference t P ∆ which is approximately the same as the pressure difference tm P ∆ read in the model fan curve used for the calculations, the error will be small. However, for accurate dimensioning of high pressure fans it will be necessary to take the flow compressibility into consideration. Such calculations are quite complicated and they are normally always performed by means of computers. Thus, we shall refrain from dealing with these here. It can be mentioned, that recalculating example 4.1 considering the-gas compressibility results in the same selection of fan diameter but the fan rotational speed is calculated at 926 rpm, and the fan power at 1170 kW rather than the 940 rpm and 1208 kW, respectively, calculated on neglecting the gas compression. 4.7 Correction for dust laden gas Up till now we have assumed that the fans to be calculated operate with clean gas. Many fans used in a cement factory operate with dust laden gases f. inst. all kiln I.D. fans and most mill air circulation fans. The dust concentration in the gas passing these fans are typically 30-60 3 m / g . Very little can be found in the literature concerning calculation of fans for gas with such a dust load. If the dust is so fine that the dust particles could be assumed to pass the fan evenly distributed in the gas, the fan could be calculated as a fan operating with (clean) gas with an equivalent (or apparent) gas density of ( ) 9 s 1 leq + ϑ = ϑ s is the dust concentration in the gas (in kg/m3) In this case the dust loading of the gas would not affect the fan power input, but the calculations would suggest the use of a slightly bigger fan operating at lower speed than in the case of the fan operating with (the same volume of) clean gas.
  • 107. However, the dust is known to concentrate along the front side of the impeller blades on its way through the rotor and the dust particles - leaving the rotor with a velocity close to the tip speed of the rotor, will at least to some extent hit the fan casing and thereby loose their kinetic energy (obtained in the rotor). According to ref. (2) it will be a good approximation to account for the dust loading simply by multiplying the fan power found for the fan operating with clean gas with the ratio between the apparent density of the dust laden gas and the clean gas density thus ( ) ( ) 10 P 1 P 1 leq dust         − ϑ ϑ = The following example will illustrate the order of magnitude of the "dust correction". Example 4.2 If the fan from example 4. 1 was to operate with gas with a dust loading of s = 40 3 m / g the approximate fan power increase according to (10) is estimated to ( ) kW 1208 1 83 . 0 04 . 0 83 . 0 Pdust       − + = = 58 kW Thus according to this calculation the fan power increases with almost 5% 5rom 1208 kW to 1266 kW when the gas is loaded with 40 3 m / g dust. 5. Fan Drives 5.1 Direct/indirect drive The fans used in the cement and similar industries are nearly always driven by an electrical motor. The motor might either be connected directly to the fan shaft via a flexible coupling or the motor power might be transmitted to the fan through a V-belt drive (up to 400 kW) or a gear unit. If direct drive by A.C. motors shall be used the fan must be dimensioned for a synchroneous or asynchroneous motor speed cf. table 5.1.
  • 108. Table 5.1 Rotational speeds for A.C. motors ( > 55 kW) Rotational Speed in RPM Frequency No. of poles Hz 2 4 6 8 10 12 Synchroneous motors 50 3000 1500 1000 750 600 500 60 3600 1800 1200 900 720 600 Asynchroneous motors ( > 55 kW ) 50 2970 1480 985 735 590 490 60 3565 1775 1180 880 705 585 In many cases a direct drive will not allow for dimensioning the fan with optimum efficiency at the actual operational conditions and in this case an indirect drive must be chosen. For large fans it will also some times be cheaper to choose an indirect drive via a reducing gear coupled to a motor with high rotational speed (few poles) than to choose a direct drive with a slow running motor (many poles). 5.2 Motor types and speed regulation To-day the following combinations of motor types, starters and speed regulation systems are used for (larger) fans: 1. Fixed speed drives: a) Squirrel cage motors with direct start b) Squirrel cage motors with Y/.t, starter. c) Synchroneous slip ring motors with rotor rheostat starter. d) Synchroneous AC motors. 2. Variable speed drive: a) Squirrel cage motors with frequency converter. b) Asynchroneous slip ring motor with rotor rheostat regulator. c) Asynchroneous slip ring motor with subsynchroneous converter cascade. d ) DC motors with thyristor converter. For smaller shaft outputs ( <100 kW) fixed speed squirrel cage motors with direct start will always be preferred due to the low price and easy maintenance of the squirrel cage motor. For larger motors the large starting current for the squirrel cage motor with direct start (4-8 times the full load current) can be a problem.
  • 109. The starting current can be reduced by using an Y/∆ starter, but also the starting torque of the motor is reduced, which results in a larger starting time which might not be permissible. In this case a slip ring motor with rotor rheostat starter has to be used. Synchroneous AC motors are relatively seldom used and only for large outputs ( >-500 kW). One advantage of this motor type is that it can act as (capacitive) phase compensator (regulating the magnitisation of the rotor). For variable speed drives slip ring motors with rotor rheostat regulator and DC motors with thyristor converter are the drive types most often used. The advantage of the slip ring motor with rotor regulator is the relative simplicity of the system, but connected to a fan the regulation range is relatively narrow (down to 60-75% of full speed) and the efficiency of the drive goes drastically down when the motor speed is reduced as the "slip power" is wasted (as heat) in the rotor rheostat (see below). The speed of the DC motor with thyristor converter (often termed thyristor drive) is controlled by regulating the armature voltage delivered by the thyristor converter. The drive system is somewhat more complicated than the slip ring motor with rotor rheostat regulation but the control range is normally very wide (down to 10% of full speed) and the drive system maintains a good efficiency even at low motor speed. The drastical reduction of the efficiency of a slip ring motor drive at reduced speed can be avoided if the slip power instead of being wasted in a simple rheostat is treated in a "slip recovery system" in which it is rectified, subsequently converted back to AC with grid frequency and then fed back to the (stator) terminals of the motor. Such a system is termed a subsynchroneous converter cascade (or just cascade regulation). However, this system is only price competitive to the DC motor drive system for large motor output (above 1000 kW). For motor outputs below 1000 kW a squirrel cage motor connected to a static frequency converter might be chosen. This drive system combines the easy maintenance of the squirrel cage motor with the wide regulation range and low regulation power loss of the DC drive system. 5.3 Drive efficiencies The electrical power used to drive a fan can be calculated if the necessary fan shaft power P is known by means of ( )( )        η η = el trans el p p 1 trans η is the efficiency of the power transmission system trans η 1.0 for direct drive and trans η 0.98-0.99 for indirect drive (gear or V-belt) el η is the efficiency of the motor inclusive a possible speed regulation system.
  • 110. For large fixed speed AC motors the efficiency el normally lies between 92% and 96%. For various variable speed drives the efficiency as a function of the relative speed reduction ratio appears from figure 5.1. Drive efficiency, % Rotational speed, % of max. Fig. 5.1: Drive efficiencies of various variable speed fan drives. a) 400 kW squirrel cage motor with frequency converter b) 1000 kW slip ring motor with rotor rheostat regulator c) 1000 kW cascade drive d).1000 kW DC motor with thyristor converter
  • 111. 6. Capacity regulation The total pressure drop t P ∆ across most flow systems in-, creases approximately with the second power of the flow through the system, that is ( )( ) ( ) ( ) 1 Q K P 2 t = ∆ At full speed the fan which transports the gas through the system is also characterized by an unique relation between total pressure rating and gas flow t P ∆ = f(Q) (2) Thus, at full speed the fan will generate a certain flow, Q., through the system which can be found from the equa,tion ( )( ) ( ) ( ) ( ) 3 Q f Q K l 2 1 = or from the fan curve, cf. figure 6.1. If it is desired to reduce the flow through the system this can be accomplished in a number of ways which are more or less economical as seen from an operational point of view. Fig. 6.1: Simple damper flow regulation
  • 112. 6.1 Damper regulation By aamper regulation the resistance coefficient in the system, K, is changed by opening or closing a damper normally placed near the inlet to the fan. Closing the damper somewhat will increase K and reduce the flow in the system according to (3) (assuming that f(Q) is a decreasing function in the relevant flow interval)), cf. also figure 6.1. In the example in figure 6.1 the flow through the fan can be reduced from 1 Q to 2 Q by closing a damper in the system so as to increase the total pressure drop in the system from 1 t p ∆ to 2 t p ∆ By this regulation method the pressure rating of the fan increases, when the flow through the fan is reduced. Therefore the power consumption of the fan only decreases slightly when the flow is reduced. Accordingly damper regulation is a fairly uneconomical flow regulation method. However, as the damper is a very simple and cheap flow control device this regulation method is normally preferred for smaller fans with limited power consumption. 6.2 Inlet guide vanes By means of inlet guide vanes placed just at the inlet to the fan the gas can be given a prerotation which - if in the rotor rotation direction - reduces the pressure rating and the power consumption of the fan for a given gas flow. Thus by changing the adjustment of the inlet guide vanes the fan curve (or the function f (Q)) can be changed and thereby the flow through the fan can be reduced. As the guide vanes not just introduce a pressure drop in the system, but also reduce the fan pressure rating and power consumption for a given flow, the flow regulation by means of guide vanes is more economical than simple damper regulation. Fig. 6.2 illustrates this. The fan with inlet guide vanes operates in a system with the system curve K. With fully open guide vanes the fan characteristic curve designated 0 o applies. The resulting flow through the fan is 1 Q , and the power consumption 1 P . If the guide vanes are turned 45 o the flow through the system is reduced to 2 Q and the fan power consumption to 2 P . If the flow had been reduced by a simple damper in the system, the considerably higher fan power consumption 2 P would have resulted. However, guide vanes are more expensive than simple dampers and they are rather sensitive to wear when operating with dust laden gas.
  • 113.
  • 114. 6.3 Fan speed regulation Reducing the speed of the fan will reduce the pressure rating of the fan at a given flow through the fan, cf. fig. 6.3. If the fan is connected to a second order flow system in which the pressure drop varies with the flow according to (1) the fan will maintain the same efficiency independent of the fan speed. That is, if the fan is designed to operate with maximum efficiency when operating with full speed the fan will still operate with maximum efficiency in the case the flow is reduced by changing the speed of the fan shaft. Thus, this form of regulation gives a loss-free regulation of the fan itself, and the operating economy depends on the losses in the speed regulation system. Various methods can be applied for speed regulation of the fan shaft. Smaller fans can be coupled to a fixed speed motor through a mechanical variator or a hydraulic or eddy current coupling. For larger fans motor speed regulation is nearly always applied to reduce the fan shaft speed. If the motor is a slip ring motor with rheostat regulation the regulation loss is considerable whereas a near to loss-free regulation is obtained with an AC motor with frequency convertors, slip ring motor with cascade regulation and DC motors with thyristor rectifiers as explained in section 5. In fig. 6.4 the fan power consumption as a function of the flow for the various flow regulation methods is compared.
  • 115. Fig. 6.3: Flow regulation by changing fan speed.
  • 116. Fig. 6.4: Fan motor power consumption curves
  • 117. Example 6.1 A raw mill fan is dimensioned for: Q = 100 s / m3 t p ∆ = 9810 Pa l ϑ = 0.83 3 m / kg It has been calculated that HAF 315/315 fan will provide the desired pressure and flow at a rotational speed of n = 926 rpm. However, the fan is supplied with a directly coupled 1600 kW, 985 rpm slip ring motor. The actual fan curves for the fan at 985 rpm and 926 rpm, respectively, are shown in fig. 6.5. A (second order) system resistance curve corresponding to the above mentioned flow and pressure for which the fan was dimensioned is also indicated on the diagram. From the diagram it can be read that the fan operating at full speed, 985 rpm, and with fully open 3 damper would produce a gas flow in the system of 106.7 s / m3 and the total pressure difference across the fan would correspond to approx. 11200 Pa. The shaft power can be read at 1425 kW. The desired flow 100 s / m3 can be obtained closing the fan damper somewhat to obtain the system resistance curve 2 K shown in the diagram fig. 6.5. From the diagram it can be read that the fan shaft power consumption hereby will decrease only slightly to 1395 kW. The pressure difference across the fan will increase to 11600 Pa of which (11600 - 9810) Pa = 1790 Pa is lost in the damper. With a motor efficiency of 0.96 the electric power consumption of the fan regulated by damper is calculated at 1450 0.96 kW 1395 Pel =       = kW If the motor speed instead is reduced to 926 rpm the desired flow would be reached with fully open fan damper. From the diagram we read in this case a fan shaft power consumption of 1175 kW. If the slip ring motor is rheostat regulated the actual motor efficiency (including losses in the slip ring rheostat) can be calculated to ( ) ( ) 90 . 0 96 . 0 985 926 n n motor nom el =             =       η         = η
  • 118. and the electrical power consumption of the fan would be       = 90 . 0 1175 Pel kW = 1305 kW If the motor was provided with cascade regulation (slip recovery system) the efficiency of the motor and speed regulator would be 0.96 - 0.015 = 0.945. In this case the electrical power consumption of the fan would be       = 945 . 0 1175 Pel kW = 1245 kW This example illustrates the importance of using speed regulation of the fan motor rather than damper regulation even when a relatively small reduction of the gas flow in a system is desired. Thus, in this case where the flow is reduced with only 6% we save 150 - 200 kW (depending on the variable speed drive type) if the flow is reduced by reducing the motor speed rather than closing the fan damper somewhat.
  • 119. Fig. 6.5: Performance curve for Fan type: HA.F 315/315 (used for example 6.1)
  • 120. 7.0 Fan modifications Rising electricity prices have created an interest in changing low efficient fans with high efficient fans. However, to change a large existing fan with a complete new fan not only involves buying the new fan but due to another design of the new fan the existing fan foundation cannot be used and it will normally also be necessary to modify the ducting around the fan which increased the installation costs and the down time. 7.1 Impeller replacement Another solution is to install a new impeller in the existing fan in order to increase its efficiency without changing the fan casing, fan drives etc. As an example we will describe the equipment which is now offered for modification of a fan with open radial impeller like the one described in section 3.1. Fig. 7.1: HP impeller
  • 121. By the modification the open impeller is replaced by a high efficient closed impeller with backwards curved blades cf. fig. 7.1. Fig. 7.2 shows a cross-section in the modified fan. Together with the new impeller an inlet cone and inlet nozzle are installed in the existing fan casing. Further, it is necessary to modify the rear plate of the fan casing somewhat and install a new shaft sealing and cooling disc (if the fan is to operate with hot gasses). In many cases the existing shaft can be reused after a machining of the end on which the new impeller is to be mounted. As the new impeller has backwards curved blades, it will normally have to operate with slightly higher tip velocity than the old impeller to obtain the required pressure rating. Normally, this problem is solved by increasing the impeller diameter in relation to the existing impeller. However, in some cases the casing does not allow for a sufficient increase in the diameter and in that case the shaft speed has to be increased introducing a new gear box or by changing the gearing ratio of an existing gear box. 7.2 Energy savings As-the operational conditions for the fan which is to be modified are normally known very well, the new impeller can be designed to operate with optimum efficiency under these conditions, and an efficiency of the modified fan exceeding 82% can normally be guaranteed. Comparing this with a typical efficiency of 68% of the existing fan with open impeller the relative power saving by installing the new impeller can be calculated at S = ( ) % 100 82 . 0 68 . 0 1       − = 17% assuming the same drive efficiency before and after the modification. If the existing fans always have been operated with reduced motor speed and/or with partially closed dampers the new impeller can be designed to allow for operation with almost full motor speed and fully open damper. In this case the relative power saving on installing the new impeller will be even greater.
  • 122. Fig. : 7.2: Fan originally provided with open impeller after modification with new high-efficient impeller etc. New parts indicated.
  • 123. If on the other hand the existing fan is operating at full capacity and a further increase in the production capacity of the system to which the fan is connected is desired, it will often be possible to design the replacement impeller for a pressure rating which allows for the desired production capacity increase. Finally, it should be mentioned, that if the existing open impeller is exposed to excessive wear, it cannot be recommended to replace it with a high efficient impeller as this will probably be exposed to even greater wear and does not allow for an easy replacement of worn impeller blades. Example 7.1 Data for an existing raw mill fan with open ¿ 4500 mm impeller: Yearly operating time 7500 h Motor power consumption 1300 kW Electricity price 0.06 USD/kWh Estimated fan efficiency 67% Budgetary quotation for new closed fan impeller etc. for modification of the fan: Impeller diameter: ¿ 4950 mm Expected efficiency: 84% Rotational speed: Same as existing impeller Price f.o.b.: 170,000 USD Fan motor power consumption after modification: 1300 kW =       84 . 0 67 . 0 1040 kW Annual electricity saving: (1300 - 1040) 7500 kWh = 1.95 x 106 kWh (1.95 x 106 ) (0.06 USD ) = 117,000 USD Estimated installation costs: 30,000 USD Total investment: 200,000 USD simple pay-back period: 200/117 years = 1.7 years. References: (1) Bruno Eck: "Ventilatoren". 5th Edition Springer Verlag, 1972 (2) "Fan engineering" - 8th Edition Buffalo Forge Company, 1982.
  • 124. Blue Circle Cement PROCESS ENGINEERING TRAINING PROGRAM Module 5 Section 5 Harleyville Preheater Exhaust Fan Performance Verification