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WHAT IS A FAN ?

A fan is a gas flow producing machine with two or more blades or

vanes attached to a rotating shaft.

Each of fan, including the impeller, converts rotational mechanical

energy, applied to their shafts, to total pressure increase of the moving

gas. This conversion is accomplished by changing the momentum of the fluid.

The fan definition is machines which increase the density of the gas

by no more than 7% as it travels from inlet to outlet. This is a rise of about 7620 Pa (30 inches of water pressure) based on standard air.

For pressure higher than 7620 Pa (30 in. WG), the air-moving

device is a compressor, or “pressure blower.”



There are three main components in a fan:

the impeller (sometimes referred to wheel or rotor),

the driving

the casing.

To forecast with reasonable accuracy the installed

performance of a fan a designer must know :

How the fan was rated and tested.The effects the air distribution system will have on the

fan’s performance.

Fans of different types, or even fans of the same typesupplied by different manufacturers, will not interact with the

system in the same way.



FAN TERMINOLOGY AND DEFINITIONS

Standard Air (SI)Dry air at 20°C and 101.325 kPa.

Under these conditions dry air has a

mass density of 1.204 kg/m3.

Water Gauge (WG)

The measure of pressure above

atmospheric expressed as the height

of a column of water in mm (inches)

(atmospheric at sea level equals10000 mm (407.1 inches) of water

(Fig. 1)



Static Pressure

The difference between

the absolute pressure at a point

in an airstream or a plenum

chamber and the absolute

pressure of ambient atmosphere

(being positive when the pressure at the point is above the

ambient pressure and negative

when below).

It acts equally in alldirections, is independent of

velocity and is a measure of the

potential energy available in an

airstream.



Velocity Pressure /Dynamic

Pressure

Is the pressure require toaccelerate air from zero velocity

to some velocity and is

proportional to the kinetic

energy of the air stream.The velocity pressure will

only be exerted in the direction

of air flow and is always

positive. (Fig. 2)



Total Pressure

The algebraic sum of static and

velocity pressure. It is a measure of the total energy available in an air

stream. (Fig.3)

TP = SP + VP

Fan Total Pressure

The algebraic difference between the

mean total pressure at the fan outlet

and the mean total pressure at the faninlet. It is the measure of the total

mechanical energy added to the air or

gas by fan. How this is measured is

show in Fig.4.



Fan Static Pressure

The fan static pressure is a defined quantity used in

rating fans and cannot be measured directly.

It is the fan total pressure minus the velocity pressurecorresponding to the mean air velocity at the fan outlet.

Note that it is not the difference between the static

pressure at the outlet and the static pressure at the inlet i.e : it

is not the external system static pressure.



Air Flow (Q)

The cubic meter per second (CMS) of air produced by a fan in

a given system is independent of the air density.

Air Horsepower (A kW)

Assuming 100% efficiency, it is the horsepower required to

move a given volume of air against a given pressure.



Brake Horsepower (B kW)

It is the actual horsepower a fan requires. It is greater than

air horsepower , because no fan is actually 100% efficient.

It may include power absorbed by V-belt drives,

accessories, and any other power requirements, in addition

to power input to the fan.



Static Efficiency (S.E.)

The static air horsepower (A kW) divided by the power

input to the fan.

Mechanical Efficiency (M.E.)

Also called total efficiency (T.E). Ratio of power output

over power input.



Blocked Tight Static Pressure (BTSP) Operating condition when the fan outlet

is completely closed, resulting in no air

flow. (Fig. 5)

Fully Open Air Flow (WOCMS)

Also called wide open CMS

(WOCMS). At this operating

condition, static pressure across the fan

is zero. (Fig. 6)



Wide Open Brake Horsepower (WOBkW)

The horsepower (kW) consumed when the fan is

operating at fully open CMS.

Frequently, fan characteristics are referred to in terms of

the percent of wide open CMS (percent WOCMS) which

is for a given fan then fixes the corresponding percent

blocked tight static pressure (percent BTSP) and percent

wide open brake horsepower WOB kW.



Application Range

The range of

operating volumes and

pressures, determined bythe manufacturer, at

which a fan will operate

satisfactorily. (Fig. 7)

Typical applicationrange for forward curved

centrifugal fan is from

30% to 80% WOCMS,

backward inclined fans is

from 40% to 85%

WOCMS



Tip speed (TS)

Also called peripheral velocity,

equals the circumference of the

fan wheel time the RPM of thefan and is expressed in m/s

(ft/min). Fig. 8



FAN LAWS

It is not practicable to test the performance of every size of fan ina manufacturer’s range at all speeds at which it may be applied.

Nor it is possible to simulate every inlet density which may be

encountered.

Fortunately, by use of the Fan Laws, it is possible to predict with

good accuracy the performance of a fan at other speeds and

densities than those of the original rating test.

It is important to note, however, that these Laws apply to a given

point of operation on the fan characteristic. They cannot be used

to predict other points on this characteristic curve.



These Laws are most often used to calculate change in flow

rate, pressure and power of a fan when the size, speed or gas

density is changed. The fan Laws will be accurate for

geometrically proportioned fans; however, because tolerancesare usually not proportioned, slightly better performance is

generally obtained when projecting from a given fan size to a

larger one.

Fan laws equations :



Change in Fan Speed

First considered are the fan laws

applying to a change only in speed

(constant system) with a given fan anda given system handling air at a given

density. (Fig.1)

Efficiency will not change.



Change in Fan Size

Fan Laws 2 account for

changes in performance due to

proportioned changes in fan

size, based on constant tip

speed, with constant speed, air

density, fan proportions and

fixed operating point. (Fig. 2)

It is used mostly by fan designers and

rarely has application in the field.



Fan Laws 3 also account for

changes in performance due to

proportioned changes in fan size but it based on constant fan speed,

with air density, fan proportions

and fixed operating point.

(Fig. 3)

It is usually used by fan manufacturers

to generate performance data for

geometrically proportioned “families”

of fans.



Change in Air Density

Considered next is the effect of

change in air density on fan

performance, three fan laws apply in

this situation.

Fan Law 4 (Fig. 4) with

constant volume, system, fan size, and

speed. The fan volume, in Q will not

change with density. A fan is a constantvolume machine and will produce the

same Q no matter what the air density

may be.



Fan law 5 (Fig. 5) with constant

pressure, system, and fan size.Variable speed.



Fan law 6 (Fig. 6) with constant

mass flow rate, constant system and

fixed fan size. Variable fan speed.

Fan laws 4 and 6 are the basis for selecting fans for other than

standard air density using the

catalogue fan tables which are

based on standard air.



Example No. 1

An air-conditioning supply fan is operating at a speed of

600 rpm against static pressure 500 Pa and requiring power

of 6.50 BkW. It is delivering 19,000 CMH at standardconditions. In order to handle an air-conditioning load

heavier than originally planned, more air is desired. In order

to increase the flow rate to 21,500 CMH, what are the new

fan speed, static pressure and power ? Using Fan Law 1(Fig. 7)





Example No.2

A fan is operating at a speed of 2715 rpm on 20°C air

against static pressure 300Pa. It is delivering 3,560 CMH

and requires 2.84 BkW. A 5 kW motor is powering thefan. The system is short capacity but the owner doesn’t

want to spend any money to change the motor. What is the

maximum capacity from his system with the existing 5

kW motor? What is the allowable speed increase? Whatwill the flow rate and static pressure be under the new

conditions? Using Fan Law 1 (Fig. 8)



Example No 3



Example No.3

A fan manufacturer wishes to project data obtained for a 400 mm-dia. fan

to a 800mm-dia. fan. At one operating point the 400 mm fan delivers 7,750

CMH of 20°C air against 100 Pa static pressure. This requires 694 rpm (tip

speed = 14.53 m/s) and 1.77 BkW. What will the projected flow rate, static pressure, power and tip speed (TS) be for a 800 mm fan at the same speed.

Using Fan Law 3 (Fig. 9)



This, plus Fan Law 1, are the fan laws used to project

catalogue data for many diameters and speeds from a test

on a single fan at one speed



Example No.4

A fan drawing air from an oven is delivering 18,620 CMH of 116°C air

against 250 Pa static pressure. It is operating at 796 rpm and requires 9.90

BkW. Assume the oven loses its heat and the air is at 20°C. What happensto the static pressure and impeller power required ?




This example illustrates why the fan motor should always be selected

on the power at the maximum density, which would be at the lowest

air temperature expected.

Example No 5



Example No.5

An engineer specifies that he wants 15,200 CMH at 200 Pa static pressure,

49°C and 300 m altitude. Determine the fan speed and power.

(There are two ways to solve this problem, Using Fan Law 4 or Fan Law 6).


In order to enter in the manufacturer’s catalogue fan tables



In order to enter in the manufacturer s catalogue fan tables

which are based on standard air, we must determine the static

pressure that would be required with standard air. From a chart

of air density ratios, we would find from the catalogue fan table,we find to deliver 15,200 CMH against 225 Pa will require

1120 rpm. The power required is 8.07 BkW. The speed is

correct at 1120, but since the fan is handling less dense air, then

:

Note also from this example that the static pressure resistance

of the system varies directly with air density.

Using Fan Law 6 (Fig 12)




In this case assume that operating condition is standard to



In this case, assume that operating condition is standard to

determine the speed and power in the catalogue. Then the

catalogue power and static pressure will be corrected according

to Fan Law 6.

The fan will deliver 13 400 CMH against 175 Pa when



The fan will deliver 13,400 CMH against 175 Pa when

operating at 988 rpm. Required power 5.55BkW. Correcting the

speed for density according to Fan Law 6, we obtain :

As would be expected, the answer comes out the

same with either solution.



Example No.6

Assume that a fan is handling 41,280 CMH at static pressure of 300 Pa, running at 418 rpm and requiring

14.99 BkW. If the speed remains constant at 418 rpm,

but an additional resistance of 100 Pa (based on

existing velocities) is placed in the system, the static pressure would be 400 Pa if the capacity, 41,280

CMH, remains the same. From the fan manufacturer’s

rating table, it is seen that the speed would have to be

increased to 454 rpm and would require 18.7 BkW.

This new fan rating must be reduced to the

predetermined speed of 418 rpm along the new duct

resistance curve by use of Fan Law 1.



This example, is useful in those cases where added

resistance, such as absolute filters, is inserted in the fan

system and thereby raises its static pressure beyond the

fan manufacturer’s catalogued ratings.



FAN PERFORMANCE CURVES

Since each type and size of fan has different characteristics,

fan performance curve must be developed by the fan

manufacturers. A fan performance curve is a graphical presentation of the

performance of a fan. Usually it covers the entire range from free delivery (no

obstruction to flow) to no delivery (an air tight system with noair flowing). One, or more of the following characteristics may be plotted

against volume flow rate (Q).

Gas density (ρ) fan size and speed (N) are usually constant for



Gas density (ρ), fan size, and speed (N) are usually constant for

the entire curve and must be stated. A typical fan performance

curve is shown in Fig. 1.

Generally these curves are determined by laboratory tests



Generally, these curves are determined by laboratory tests,

conducted according to an appropriate industry test standard,

e.g. Air Movement and Control Association International Inc.

(AMCA).

It is important to note that the test setup required by AMCA

standards is nearly ideal. For this reason, the performance

curves for static pressure and brake horsepower versus airflow,are those obtained under ideal conditions, which rarely exist in

practice.

The “Fan Laws” are used to determine the brake horsepower and performance characteristics at other speeds and fan sizes;

normally, as mentioned before, only one fan size and speed

must be tested to determine the capacity for a given “family” of

fans.

SYSTEM RESISTANCE CURVE



SYSTEM RESISTANCE CURVE

System resistance is the sum total of all pressure lossesthrough filters, coils, dampers, and duct work. The

system resistance curve (Fig. I) is simply a plot of the

pressure that is required to move air through the

system.

For fixed systems, that is, with no changes in damper

settings, etc., system resistance varies as the square of

the air volume (Q). The resistance curve for anysystem is represented by a single curve. For example.,

consider a system handling 1000 CMH with a total

resistance of 100 Pa SP .

If the Q is doubled the SP resistance will increase to 400 Pa as



If the Q is doubled, the SP resistance will increase to 400 Pa, as

shown by the squared value of the ratio given in Fig.1.This

curve changes, however, as filters load with dirt, coils start

condensing moisture, or when outlet dampers are changed in position.

The operating point (Fig. 2) at which the fan and system will



The operating point (Fig. 2) at which the fan and system will

perform is determined by the intersection of the system resistance

curve and fan performance curve. Note that every fan operates only

along its performance curve. If the system resistance designed is

not the same as the resistance in the system installed, the operating

point will change and the static pressure and volume delivers will

not be as calculated.

Note in Fig 3 that the actual system has more pressure drop than



Note in Fig.3 that the actual system has more pressure drop than

predicted in the design. Thus, air volume is reduced and static

pressure is increased.



The shape of the kW curve typically would result in a

reduction in BkW. Typically, the RPM would then be

increased and more BkW would be needed to achieve the

desired Q.

In many cases where there is a difference between actual and

calculated fan output, it is due to a change in system resistance

rather than any shortcomings of the fan or motor.

Frequently the mistake is made of taking the static pressure

reading across the fan and concluding that if it is at or abovedesign requirements, the Q is also at or above design

requirements. Fig. 3 shows why the assumption is completely

invalid.

SYSTEM SURGE FAN SURGE AND PARALLELING



SYSTEM SURGE, FAN SURGE AND PARALLELING

The three main reasons for unstable airflow in a fan systemsare (1) System surge, (2) Fan Surge and (3) Paralleling.

SYSTEM SURGE FAN SURGE AND PARALLELING



SYSTEM SURGE, FAN SURGE AND PARALLELING

System Surge

System surge occurs when the

system resistance and fan

performance curves do not

intersect at a distinct point but

rather over a range of volumes and pressures. This

situation does not occur with

backward inclined (BI),

airfoil (AF), and radial fans.However, it can occur with a

forward curve centrifugal fan

when operating, as shown in

Fig. 1.

In this situation, because the fan curve and system curve



In this situation, because the fan curve and system curve

are almost parallel, the operating point can be over a range of

airflow and static pressures.

This will result in unstable operation known as systemsurge, pulsation, or pumping.

System surge should not be confused with “paralleling,”

which can only occur when two fans are installed in parallel.

Fan Surge



Fan Surge

Fan surge is different from system surge, they may or may

not occur at the same time. (Fig.2)



For any fan, the point of minimum pressure occurs at the

center of rotation of the fan wheel and the maximum

pressure occurs just at the discharge side of the wheel. If the wheel were not turning and this pressure

differential existed, flow would be from the high pressure

point to the low pressure point. This is opposite from the

direction air normally flows through the fan. The onlything that keeps the air moving in the proper direction is

the whirling of the blades.

Stall occurs unless there is sufficient air entering the fan

wheel to completely fill the space between the blades.

This shows up in Fig 3 as fluctuation in



This shows up in Fig. 3 as fluctuation in

air volume and pressure. This surge can

both felt and heard and occurs in nearly

all fan types, to varying degrees, as

block-tight static pressure is approached.

The radial blade is a notable exception.

While the magnitude of surge varies for

different type of fans, (being greatest for

airfoil and least for forward curve), the pressure fluctuation close to block-tight

may be on the order of 10%. For

example, a fan in surge developing about

600 Pa of total static pressure might have

pressure fluctuation of 600/10 of an Pa.This explains why a large fan in surge is

in tolerable. Equipment room walls have

been cracked from the vibration of ducts

serving a fan in surge.

Selections should not be made to the left of the “surge point” on



g p

the fan curve. This point, which defines a system curve when all

operating speeds of the fan are considered, varies for different

fan installations.For instance, stable operation can be obtained much further to

the left when the fan is installed in an ideal laboratory type

situation.

These conditions, of course are seldom encountered in fieldapplications.

Consequently, most manufacturers do not catalogue operating

ranges all the way to the surge line.

However, since the catalogue cut-off point is basically one of engineering judgment, conservative catalogue performance data

will provide operating ranges, which will allow stable operation

with any reasonable field ductwork design.

PARALLELING



The third cause for unstable operation is paralleling, (Fig. 4),

which can occur only in a multiple fan installation

connected with either a common inlet or common discharge, or both in the same system, particularly when large volume of air

must be removed

The combined air flow-pressure curve in this case is obtained



p

by adding the airflow capacity of each fan at the same pressure.

(Fig. 5)

The total performance of the multiple fans



The total performance of the multiple fans

will be less than the theoretical sum it inlet

condition are restricted or the flow into the

inlets is not straight.

Some fans have a “positive” slope in the

pressure-air volume curve to the left of the peak pressure point. If fans operating in parallel are

selected in the region of this “positive” slope,

unstable operation may result.

The closed loop to the left of the peak pressure



The closed loop to the left of the peak pressure

point is the result of plotting all the possible

combinations of air volume at each pressure.

If the system curve intersects the combined air

volume-pressure curve in the area enclosed by the

loop, more than one point of operation is possible.

This may cause one of the fans to handle more of the air and could cause a motor overload if the fans are

individually driven.

This unbalanced flow condition tends to reversereadily the result that the fans will intermittently load

and unload. This “pulsing” often generates noise and

vibration and may cause damage to the fans, ductwork

This requires the installation of scroll volume (outlet volume)



This requires the installation of scroll volume (outlet volume)

dampers (Fig. 6). It serves to change the shape of the fan scroll

and thus, for each position of the damper, there is a

corresponding different performance curve.

The fan curve resulting from various positions of the



The fan curve resulting from various positions of the

outlet volume dampers is shown in Fig. 7.

The purpose is to change the fan curve sufficiently suchthat the sum of the difference curve will intersect the single fan

curve at A’ and provide stable operation.

The performance may be reduced slightly and a

corresponding increase in RPM should be made to achieve the

specified conditions. However, this is rarely done since

difference is typically negligible.



The use of axial flow fans in parallel presents very real potential



The use of axial flow fans in parallel presents very real potential

noise problems unless special measures are taken at the design

stage; add-on noise control is not normally possible.

A noise problem often encountered with fans operating in

parallel is beating. This is caused by slight difference in speed

of rotation of the two theoretically identical fans.

The resulting low frequency beating noise can be very annoying

and difficult to eliminate.

The problem can be likened to the stroboscopic effect of a

fluorescent light illuminating a rotating wheel with a slight

difference between the frequencies of rotating of the wheel and

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Cac Khai Niem Co Ban Ve Quat