Axial Fan Bearing System Vibration Analysis

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Journal of EngineeringVolume 18 February

2012

Number

2

AXIAL FAN BEARING SYSTEM VIBRATION ANALYSIS

Dr. Assim H Yousif Dr. Muawafak A Tawfik Dr. wafa Ab Sou

M!"#a$i"a% E$&i$!!ri$& D!'ar(m!$() *$i+!rsi(, of T!"#$o%o&,) Baa.

ABSTR*-T

Rotating fan shaft system was investigated experimentally and theoretically to study its

dynamic performance. The type of oil used for the bearing was taken in consideration during the

experimental program .Three types of oil were used, SAE !, SAE "! and degraded oil. #uring

the experiments, the fan blades stagger angle was changed through angles $%!&, '!&, !&, and"!&(. The shaft rotational speed also changed in the range of $!)'!!! rpm(. All these parameters

have investigated for two cases $balanced and unbalanced fan(. The performance parameters of

the fan were found experimentally by measuring the fan, volume flow rate, Reynolds andStrouhal numbers, efficiency and pressure head. Analytical part was also represented to prepare

the prediction of fan system dynamic performance. The aerodynamic forces and moments of

each blade were also predicted to obtain the rotor dynamic future. Experimentally andtheoretically the critical fan speed was obtained in the x and y direction for different lubricant oil

viscosities and shaft rotational velocities for balanced and unbalanced fan. Analysis of the

vibrational response gave important information about the dynamic performance of fan rotatingsystem. Acceptable agreement was found between analytical and experimental results.

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INTROD*-TION

n the design study of turbo)

machinery, performance and rotor dynamicconsiderations are often in compromise. The

rotor performance usually desires to

maximi-e the volume flow rate and the

pressure, and minimi-e the fluid energylosses, though a machine of limited si-e and

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weight. This generally suggests high shaftspeed, and multiple stages. All of these

features tend to create rotor dynamic

problems. The goal of increasing theefficiency of machines is essential for their

development. Engineers and Scientists workto reach ways and method to guide thedesigners to develop new designs to reach

this goal. They develop machine designs in

order to prevent the operations errors and to

predict them early.The characteri-ation of these errors and

their causes is essential in the maintenance

process in order to prevent the unexpectedfailure in these machines. These predicted

errors can be found by using simple

measuring e0uipments for vibrations andalso by analytical methods. The causes of

these vibrations can be predicted and the

pre)prevented solutions can be found, which

in fact would decrease the cost and increasethe efficiency, which is the goal, and these

are crucial in industry. This research,deals

with the prediction of these errors, in orderto reach this goal a fan rotor bearing casing

system has been built up and vibration

measuring e0uipments have been used.

The vibration in this system can bedue to one of the following variables or

combinations of them1. 2hanging the stagger angles.

%. #egradation of oil used.

'. 2hanging revolution speed.

. 3nbalance in fan or rotor shaft $crack,failure and dirt(.

n addition to that, in the current research

numerical solution, used to create, develop

and modify computer software to givethe re0uired results for the dynamic

response for each system.

ROTOR DYNAMI- F*T*RE

All fans generate some vibration.They continuously rotate and since nothing

is perfect, cyclic forces must be generated

.ts only when vibration reaches certain

amplitude that may be call 4bad4. 5ibrationmay 6ust be an indicator of some problem

with a mechanism, or it may be cause of

other problem .7igh vibration canbreakdown lubricants in the bearings and, in

addition may cause metal fatigue in the

bearings. Excessive vibration can causefasteners to loosen or can cause fatigue

failure of structurally loaded components.

Also vibration can be transmit into ad6acent

areas and interfere with precision processes,or create an annoyance for people.

Thus, experimental and numerical

results can be used to analy-e the vibratingsignal $amplitude(, and to predict the causes

of these vibrations. The design of successful

and reliable turbomachine re0uires

cooperation and compromises between thetwo descriptions. 8rom a rotor dynamic

standpoint of view, the successful design ofa turbomachine involves 9To#$)4566:.Avoiding critical speeds, if possible and

minimi-ing dynamic response at resonance,

if the critical speed is traversed. ;inimi-ingvibration and dynamic load transmitted to

machine structure through out the operating

speed range. Avoidingturbine or compressor blade tip or seal

rub, while keeping tip clearance andseals as tight as possible to increase theefficiency. Avoiding rotor dynamic

instability and avoiding torsion vibration

resonance or torsion instability of the

drive train system.

EX7ERAMENTAL SET8 *7 The axial fan used in the test rig is

usually used in airconditioning system and

industry. The fan is of ten blades aerofoil

section 9( mm, tip

cord of $=.( mm, leading edge twist angle

is $?(, trailing edge twist angle is $-ero( ?.

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d` is the torsion moment about J)axis,;*and;+are the bending moment aboutx and y axisK

e`is the axial force in J)direction,

eband 5+ are the shear force in x and y

directions, and subscripts 2 and S are

indicated casing and rotor respectively.

*NBALAN-E FOR-ES

3nbalance forces is defined as the

allowance for a circumferential change in its

position. t is a kind of excitation for the

system. The unbalance force component9M.A.Tawfik) 455=and Bar(#) 45


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A*A@ 8A< EAR has been developed to

embrace the theoretical work. The program

adopted 8igure $'( as a lumped massesdistributed uniformly with circular shapes,

negligible weight, having rigidity e0ual to

the actual systems rigidity. The programtakes into consideration the effect of,

gyroscopic moment, shear force, moment of

inertia for rotor and casing, Stiffness anddamping coefficients of 6ournal bearing and

the ranch of the fan blade. To minimi-e

the error and to enhance the accuracy of the

results, non) dimensional terms are

implemented.The fan blade branch and 6ournal

bearing system have been reduced to asingle matrix for the purpose of the

solution .The rotor shaft has been divided in

to $''( stations of masses and $'%( ofmassless sections .The program has the

ability to find the deflection, mode shape,

shear force and moment for each station

along the shaft and casing for a rang ofspeed $!)'!!!( rpm

The program was developed for therotor, casing and branch system .n order toevaluate the final matrix, the point matrices

are multiplied by the field matrices and

boundary conditions are applied for thesystem to find the state vector in three

dimensions along the system for every

rotating speed and for different stagger

angle and different oil viscosity in the6ournal bearing.

7!rforma$"! -#ara"(!ris(i" of(#! Fa$

Results were obtained for a range

of flow rate and various stagger angels ofthe fan blade at three different kinematics

viscosity of oil used in bearing system, so

that, the fan performance data could be

generated. Sample of the fan performance

characteristics is illustrated in 8igure $!( as

plots of 0ualitative performance curve for atypical fan stagger angle 9!O: and for

lubricant oil SAE ! $Pinematics 5iscosity

Q! c.s(, the efficiency curve leans more tothe rightL the head tends to decrease with

flow rate increasing. The fan efficiency

drops off rapidly for volume flow rate

higher than that at the best efficiency point.The net head curve also decreases

continuously with flow rate, although there

are some wiggles. f the fan operated belowits maximum efficiency point, the flow

might be noisy and unstable which indicates

that the fan may be oversi-ed $large than

necessary( 9Fra$k%,$) 9::9:. 8or this reasonit is usually best to run a fan at, or slightlyabove its maximum efficiency point.

n the duct fan, a vortexshedding will appear in the flow down

stream of the fan in addition to unsteady

flow which can be represented by non)dimensional group, Strouhal number$St(.

8igure $( shows a non)dimensional

numbers, Reynolds number $Re( and $St( ofthe duct fan.

$St( represents the characteristicdimension of the body $#( times the system

fre0uency of vibrations $f( divided by the

fluid free stream velocity $3(. 8or axial

fluid flow, the characteristic dimensiontaken as the outer diameter of the fan

9M"Graw8Hi%%) 9::=:, also the Re based onthe outer diameter of the fan 9Fra$k%,$)9::9:. t can be seen from this figure, thatat low stagger angle 9%!? low flow rate :,

the shedding fre0uency is fluctuating on thefan system ,due to the generation of forces

on the fan structure .The vortex will be

existed down stream of the fan 9Yu$us)9::=:, and increased as the rotor speedincreased .Cn the other hand, when using

degraded oil $kinematics viscosity Q> c.s(

at stagger angle e0uals $!O( and $"!?( thevalue of strouhal no. is relatively high. At

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the stagger angle $'!O(, the vibration is atlow values $StQ!."( and sharply increase to

higher values $StQ.( at ReQ."x!". As

Re increase the fluctuation behaviors of Stand appeared to be more stable ranging

between .'% and.". 7owever, the Stwould seem to be higher at stagger angle$ !?( and increase as the oilNs viscosity

increases, but at angle $"!O(, the vibration

still be exist. 7owever, these fluctuations

are less than that for stagger angle $%!?(, butit is still high and become higher when

using degraded oil. #ue to swirl fluid

downstream of the duct fan which leads towaste of kinetic energy and a high level of

turbulenceL this wasted kinetic energy

partially reduced the level of the turbulenceby using stator.

AM7LIT*DE IN ?X@ AND ?Y@DIRE-TION

Theoretical and experimentalamplitude in * and + directions fordifferent system rotating speed, stagger

angle and viscosity are indicated in figure

$%(. The remarkable point raised from this

8igure is that the natural fre0uency occursat the same speed in x and y directions and

the amplitude in the x)direction higher thanits values in the y)direction for the first

appear of the natural fre0uency. At some

rotating speed natural fre0uency appears in

the x directions and disappears in the ydirection. Also it can be seen that the

amplitude in x and y decreased when

stagger angle is increase. The behavior ofthe theoretical work is the same in the x and

y directions with slight differences. Theamplitude of vibration in the x)directiondecreases as the stagger angle increases.

The closeness between experimental and

theoretical results are acceptable $!)F.%G(.

The difference in values of x and ydirections amplitude demonstrate the

difference in stiffness coefficients. Hhile

8igure $'( shows that many natural

fre0uencies occurred with high viscosityfigure of oil in 6ournal bearing for the same

range of rotor speed. Also some natural

fre0uencies appear in x and disappear in ydirections.

8igure $( shows that the amplitude inx and y directions decreased as the staggerangle increased. 8igures $"( and $=( show

a comparison between the experimental and

theoretical work for x and y amplitudes

versus rotor speed for different staggerangle and oil viscosity. The main conclusion

raised from these results is that the

experimental and the theoretical results areclose to each other except at stagger angle

!&.

EFFE-T OF *NBALAN-E FOR-EON THE VIBRATION ANALYSIS

Two masses of $mQ/ gm, m%Q=gm( are replaced at a radius of =% mm from

the fan hub center line, to measure the effect

of unbalance masses on the system responseof the vibration at bearing station$%(.

8igure $>( shows the amplitude in

x)direction .t can be seen that the amplitude

of vibration at balance force is small inmagnitude and increases when the

unbalance forces are increased. Themaximum amplitudes in all the cases,

8igure $/(, occurred at a rotor speed $''F

rpm(.The maximum amplitude appears at

stagger angle %!O which is e0ual to$'.%"x!)( and decrease as the stagger

angel is increased. Hhen the rotor speed

approaches >!! rpm, the amplitude ofvibration appears with large magnitude, but

less than that for rotor speed of $''F(rpm .Hhile in y) direction, as shownin figure$/(, the amplitude is less in

magnitude, but also it is maximum at $''F

rpm(, approximately e0ual to x!)m . The

amplitude increases as the unbalance forceincreases. 7owever 8igure $/( shows that

the modulus of the amplitudes increases

drastically at and grater than the rotor speed

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8ranklyn Pelecy 4Study #emonstrates that

Simulation can Accurately Iredict 8an

Ierformance 4ournal Articles y 8luent Software

3sers. A !/, pp$)(,%!!%.

, Hattar, H.7afe-, J.Bao 4;odel ased

#iagnosis of 2haotic 5ibration signals4,A

Automatically 2leveland State3niversity, C7.%!!!

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Axial Fan Bearing System Vibration Analysis