Undamped vibration absorber

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Undamped Vibration Absorber
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ME 413: System Dynamics & ControlME 413: System Dynamics & ControlME 413: System Dynamics & ControlME 413: System Dynamics & Control
Undamped ViUndamped ViUndamped ViUndamped Vibration Absorberbration Absorberbration Absorberbration Absorber
Name: __________________________________
ID #: __________________________________
Section #: __________________________________
Due Date: __________________________________
Instructor __________________________________

ME 413: System Dynamics and Control Lab Manual
Undamped Vibration Absorber 2
UNDAMPED VIBRATION
ABSORBER
OBJECTIVES
1. To show the discomfort and dangers that result from the resonance
phenomenon.
2. To study the effect of adding an absorber to a vibratory resonant system.
Part 1: THEORY
If a single or a multi-degree of freedom system is excited into resonance (the
excitation frequency nearly coincides with the natural frequency of the system), large
amplitudes of vibration result with accompanying high dynamic stresses and noise
and fatigue problems. Excessive vibrations in engineering systems are generally
undesirable and therefore must be avoided for the sake of safety and comfort. If
neither the excitation frequency nor the natural frequency can conveniently be
altered, this resonance condition can often be successfully controlled. It is possible to
reduce the unwanted vibrations by extracting the energy that causes these
vibrations. The extraction of this energy can be established by attaching to the main
vibrating system a dynamic vibration absorber, which is simply a spring-mass
system. The dynamic vibration absorber is designed such that the natural
frequencies of the resulting system are away from the excitation frequency.
1m
)(1 tx
2/1k2/1k
tF ωο sin
Figure 1 Idealization of a machine

When we attach an auxiliary mass 2m to a machine of mass 1m through a spring of
stiffness 2k , the resulting two degrees of freedom system will look as shown in
Figure 2. The equations of motion of the masses 1m and 2m are
( )
( )
1 1 1 1 2 1 2
2 2 2 2 1
sin
0
m x k x k x x F t
m x k x x
ο ω+ + − =
+ − =
(1)
By assuming a harmonic solution,
tXtx jj ωsin)( = , j=1, 2 (2)
We can obtain the steady-state amplitude of the masses 1m and 2m as we can
obtain
2
2
2
22
2
121
2
22
1
kmkmkk
Fmk
X
−−−+
−
=
))((
)(
ωω
ω ο
(3)
2
2
2
22
2
121
2
2
kmkmkk
Fk
X
−−−+
=
))(( ωω
ο
(4)
Machine 1m
)(1 tx
2/1k2/1k
tF ωο sin
)(2 tx
2m
2k
vibration absorber
Figure 2 Dynamic vibration absorber

We are primarily interested in reducing the amplitude of the machine 1X . In order to
make the amplitude of 1m zero, the numerator of Eq. (3) should be set equal to zero.
This gives
2
22
m
k
=ω (5)
if the machine, before the addition of the dynamic vibration absorber, operates near
its resonance, 11
2
1
2
/ mk=≈ ωω . Thus if the absorber is designed such that
1
1
2
22
m
k
m
k
==ω (6)
The amplitude of vibration of the machine, while operating at its original resonant
frequency, will be zero. By defining
,
1k
F
st
ο
δ =
1
1
1
m
k
=ω
as the natural frequency of the machine or main system, and
2
2
2
m
k
=ω (7)
as the natural frequency of the absorber or auxiliary system, equations. (3) and (4)
can be rewritten as
1
2
2
2
2
11
2
21
11
1
k
k
k
k
X
st
−














−














−+






−
=
ω
ω
ω
ω
ω
ω
δ
(8)
And
1
2
2
2
2
11
2
2
11
1
k
k
k
k
X
st
−














−














−+
=
ω
ω
ω
ωδ
(9)
Figure (3) shows the variation of the amplitude of vibration of the machine stX δ/1
with the machine speed 1/ωω . The two peaks correspond to the two natural
frequencies of the composite system. As seen before, 01 =X at 1ωω = . At this
frequency, equation (9) gives

1
2
2 2
st
Fk
X
k k
ο
δ= − = − (10)
This shows that the force exerted by the auxiliary spring is opposite to the impressed
force ( )οFXk −=22 and neutralizes it, thus reducing 1X to zero. The size of the
dynamic vibration absorber can be found from equations (10) and (6):
2
2 2 2 2k X m X Fοω= = − (11)
Thus the values of 2k and 2m depend on the allowable value of 2X . It can be seen
from Figure 3 that the dynamic vibration absorber, while eliminating vibration at the
known impressed frequency ,ω introduces two resonant frequencies 1Ω and 2Ω at
which the amplitude of the machine is infinite. In practice, the operating frequency
ω must therefore be kept away from the frequencies 1Ω and 2Ω .
2Ω1Ω
21 ωω =
20
1
1
2
=
m
m
stX δ/1
1/ωω
Figure 3 Effect of undamped vibration absorber on the response of machine

NOTES
1. The primary system possess now the characteristics of a two-degrees of
freedom, it has two natural frequencies 1Ω and 2Ω . The new natural
frequencies lie in the neighborhood of the natural frequency 1ω of the primary
system alone as shown in Figure 3. It can be seen from Figure 3 that 1Ω ≤
1ω ≤ 2Ω . Thus the machine must pass through 1Ω during start-up and
stopping leading to large amplitude vibrations during these transient periods.
2. Since the dynamic absorber is tuned to one excitation frequency ω , the
steady-state amplitude of the machine is zero only at that frequency. If the
machine operates at other frequencies or if the force acting on the machine
has several frequencies, then the amplitude of vibration of the machine may
become large.
3 The preceding analysis is valid only for an undamped system. If damping is
present in the absorber it is not possible to eliminate steady state vibrations
of the original mass. The amplitude of vibration can only be reduced.
Part 1: EXPERIMENT
The above theory is applied to a simply supported beam carrying a motor with mass
unbalance at its mid-span as shown in Figure 4. In this figure, a simply supported
beam carrying a motor with mass unbalance at its mid-span is shown. The motor is
connected to a speed control unit through which the speed of rotation can be varied.
In order to measure the amplitude of vibration an accelerometer can be attached at
the beam mid-span. The output of the accelerometer is connected to a vibration
meter that will provide reading of the amplitude of vibration. Underneath the motor
assembly, the vibration absorber can be fixed.
Figure 4 Setup of the experiment (No vibration absorber attached).

Apparatus
1. The absorber
Figure 5 shows the vibration absorber clamped below the motor. It comprises two
bodies of equal mass fixed equidistant from the midpoint of the horizontal cantilever.
The distance apart of the bodies varies until the system is ‘tuned’.




l
Figure 5 The vibration absorber clamped below the motor.
2. Speed Control Unit and Exciter Motor
Figure 6 shows the speed control unit that is used in this experiment. A d.c. motor is
used for all forced vibrations experiments powered by a control unit. This
combination comprises of a control box and d.c. motor, which provides high precision
speed control of the motor up to 3000 rev/min, irrespective of the normal load
fluctuations of the motor.
The front panel of the unit contains a speed control, a fully calibrated speed meter
incorporating an automatic range switching device (there being two ranges: 0 – 1500
and 1500 – 3000 rev/min), and a power socket for:
1. Mains inputs
2. d.c. motor
3. Auxiliary output (either to stroboscope or chart recorder), sometimes
marked drum supply.

Figure 6 Speed control unit front panel layout.
3. Vibration Meter Type 2511
The Vibration Meter Type 2511 shown in Figure 7 is a wide range instrument that is
used in conjunction with piezo-electric vibration pick-up to measure mechanical
vibration and shock in terms of acceleration, velocity and displacement. It is
completely portable and is built to a high standard of accuracy and stability making it
suitable for laboratory and field conditions. The instrument is fully calibrated in both
metric and British units and has a charge amplifier input that allows the use of long
cables between the pick-up and the meter without any reduction in sensitivity.
Figure 7 Front panel of the 2511 Type Vibration Meter.

Figure 8 Rear panel of the 2511 Type Vibration Meter.
4. Transducer (Accelerometer)
In general, the transducers employed in vibration analyses convert mechanical
energy into electrical energy; that is, they produce an electrical signal which is a
function of mechanical vibration. See Figure 5.
Figure 9 Accelerometers.

Procedure
For a given frequency, the masses of the vibration absorber are adjusted along their
cantilever leaf spring so that the energy of vibration transmits to the absorber and
the amplitude of the main (primary) system, i.e. the motor and beam, is reduced to
zero.
The aim is to determine the length l , the distance of the center of each of the
bodies from the midpoint of the cantilever so that the natural frequency of transverse
vibration of this sub-system corresponds to the running speed of the main (primary)
system, i.e. the motor and beam.
IMPORTANT
For your safety and the safety of the equipment
WHEN THE SYSTEM VIBRATES AT RESONANCE CONDITIONS
DO NOT LET IT VIBRATE FOR A LONG PERIOD OF TIME
1. No Absorber
1. Turn the upper right switch of the Vibration Meter Type 2511 on displacement,
velocity or acceleration depending on what you are intending to measure.
2. Turn the lower right switch of the Vibration Meter Type 2511 for the
appropriate range (Range I: metric units and Range II: British units). Notice
that you are reading the amplitude of the quantity to be measured as shown
in Table 1.
Table 1 Reading of the Vibration Meter Type 2511
Quantity Equation of Motion
Reading Vibration Meter Type
2511
Displacement ( ) sinx t X tω= Displacement Amplitude: X
Velocity ( ) cosx t X tω ω= Velocity Amplitude: X ω
Acceleration
2
( ) sinx t X tω ω= − Acceleration Amplitude:
2
X ω−
3. Vary the motor speed and read the vibration amplitudes from the Vibration
Meter Type 2511.
4. For each speed, record your reading (vibration amplitude) in Table 2.

5. Increase the speed gradually until you get the resonance phenomenon that
results in large amplitudes of excessive vibrations. From the tabulated data
plot the response curve and determine the resonant frequency of the system.
2. With Absorber
1. With the auxiliary system (vibration absorber) attached, vary the motor speed
and record the corresponding frequency and the resulting amplitude of
vibration.
2. Repeat the above (1 to 4) steps of the previous case.
3. With the aid of the experimentally defined resonant frequency, the dynamic
vibration absorber is to be designed such that the frequency of oscillations is
equal to
3
1 3
2 2
EI
f
ml
ω
π π
= = (12)
where f is natural frequency of the auxiliary system, m is the mass of each
of the two bodies, and EI is the flexural rigidity of the double cantilever
beam. The mass m is a given constant and l is to be found from the above
formula.
4. One can easily conclude, that any three parameters of equation (12) can be
fixed, in order to determine the fourth parameter. In this experiment we will
determine the position of the mass m , at which the absorber effect is
verified. Experimentally, one can vary the position of the mass m , and excite
the system at the required excitation frequency until no vibrations of the
primary system are observed, or the position of the mass l can be
determined from equation (12), adjusted accordingly and the absorbing
effect can be verified.

REQUIREMENTS
The report should include the following:
1. Organize your measurements of the vibration amplitudes versus the rotational
speed as shown in Table 2.
2. Plot on the same graph the velocity amplitude X ω versus the rotational
speed for the cases with and without the dynamic vibration absorber. (Use
MATLAB or Excel).
3. Plot on the same graph the displacement amplitude X versus the
rotational speed for the cases with and without the dynamic vibration
absorber. (Use MATLAB or Excel)
4. Indicate resonant frequencies and the frequency at which the primary system
(no absorber included) has zero vibration
5. Measure the length l for which the amplitude of vibration is zero when the
absorber is used. Compare this by solving Equation (12) for the length l .
Use 200 GPaE = and
31
12
I bh= .
6. Give a brief discussion of your findings.
7. Give conclusions on the effect of the vibration absorber.

Table 2 Readings
With no Absorber
N
(rpm)
Disp.
Ampl.
(mm)
Vel.
Ampl.
(mm/s)
Acc.
Ampl.
(mm/s2
)
With Absorber
N
(rpm)
Disp.
Ampl.
(mm)
Vel.
Ampl.
(mm/s)
Acc.
Ampl.
(mm/s2
)

References
[1] Experiments in Vibration Using The TM 16 Universal Vibration, TQ Ltd
[2] K. Ogata, System Dynamics, Fourth Edition, Pearson Prentice Hall, 2004.
[3] S. S. Rao, Mechanical vibrations, SI Edition, Pearson Prentice Hall, 2005.
[4] http://ta.twi.tudelft.nl/nw/users/vuik/information/tacoma_eng.html
[5] http://abel.math.harvard.edu/archive/21b_fall_03/tacoma/
[6] http://perso.wanadoo.fr/olivier.granier/meca/accueil.htm
[7] http://www.kettering.edu/~drussell/Demos/absorber/DynamicAbsorber.html
[8] http://www.mfg.mtu.edu/cyberman/machtool/machtool/vibration/absorb.html
[9] http://www.kettering.edu/~drussell/Demos/absorber/DynamicAbsorber.html
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