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 $/( shows that
the modulus of the amplitudes increases
drastically at and grater than the rotor speed
/F
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8ranklyn Pelecy 4Study #emonstrates that
Simulation can Accurately Iredict 8an
Ierformance 4ournal Articles y 8luent Software
3sers. A !/, pp$)(,%!!%.
, Hattar, H.7afe-, J.Bao 4;odel ased
#iagnosis of 2haotic 5ibration signals4,A
Automatically 2leveland State3niversity, C7.%!!!
@.
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