A vibration isolation system prevents one object from affecting another. Such systems
are used extensively to isolate machinery (industrial and marine), civil
engineering structures (base isolation in building, bridges, etc.), and sensitive components
from the foundation/base. Vibration isolation schemes are to 1) reduce the propagation
of base vibration to the isolated object and 2) abate the transmission of
vibration energy of machinery to the base. Moreover, in vehicular/marine,
some industrial machines (such as mechanical presses), as well as seismic
applications, isolators are also expected to lower the impact of shock from base to isolated
object or vice-versa.
Passive vibration isolation solutions, normally provided by
elastomeric (rubber) or metal spring elements, do not address all
of the above concerns, simultaneously; their effectiveness in one of the
above areas is normally at the expense of either not being effective in or
adversely affecting other areas; see the parametric study presented below.
The engineering of passive mounting arrangments
is rather complex and involves optimizing the mounting
variables, in terms of their stiffness, damping and mass attributes so that their
performance is acceptable (but not necessarily ideal) satisfying all of the above objectives. Utilizing
the adjustability of air springs, feedback controls, and the magic of embedded computers
DEICONS has devised
Computer Controlled Air Isolation System, an uncompromising solution
satisfying all the requirements of an ideal mounting system.
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Please contact DEICON to discuss your sound and vibration isolation problem. If we do not
already have the solution for it, we can analyze the problem, customize DEICON's ideal mounting
solution or design other
vibration isolation treatment for it, and oversee its implementation.
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Table 1 Impact of mount parameter variation on vibration
isolation effectiveness
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Effect on
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Shock Isolation
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Low-Frequency Vibration Isolation
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High-Frequency Vibration Isolation
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Parameter
Variation
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Damping
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increase
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+
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-
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-
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decrease
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-
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+
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+
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Stiffness
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increase
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+
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-
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0
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decrease
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-
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+
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0
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Added
Mass
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increase
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-
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+
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+
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Evident from Table 1, no single solution enhances all the attributes of a mounting
mechanism isolating an object subject to simultanious vibration and shock loadings, e.g.,
a diesel-generator on-board a watercraft. For example, enhanced low-frequency vibration
isolation performance of a system with low damping and low stiffness is normally
achieved at the expense of excessive displacement of the mounted mass around the
resonant frequency of the system diminishing its shock isolation effectiveness.
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Only adjustable mounting schemes under the control of a computer such as
DEICON’s
Computer Controlled Air Isolation System, can address the conflicting
requirements on the attributes
of isolators, providing both shock and vibration isolation.
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DEICON’s adjustable vibration isolation system uses
air as the isolation medium. Air provides 1) the highest
degree of isolation of any type vibration isolator, 2) negligible
overall damping enhancing high-frequency vibration isolation, and most importantly 3) adjustability.
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The impact of varying the parameters of a mounted system
(damping, stiffness, and mass), one parameter at a time, on the
isolation effectiveness in a one degree of freedom system is analyzed.
The spring-mass-dashpot system of Figure 1 is frequently used
as a one degree of freedom approximation of an isolation system.
The goal is to isolate
the base from the vibration of the mass caused by the excitation force F, i.e.,
lowering the force transmitted to the base Ft , while avoiding excessive vibration
of the mass (bouncing) due to shock excitation at the base.
The spring dashpot pair, known as Voight model, is commonly used to approximate
the dynamics of commonly used viscoelastic isolators such as rubber mounts.
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Figure 1 A simple mounting system
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Figures 2(a), 3(a), and 5(a) depict the magnitudes of the frequency response
functions (FRF) mapping the vibration excitation force (F) and shock excitation
(x_base) to the transmitted force Ft and mass displacement x, commonly known as transmissibility, for
varying a) damping, b) stiffnesses, and c) mass (by adding a dead weight
to the mass), respectively. Figures 2(b), 3(b), and 5(b) depict the magnitudes of the
frequency response functions (FRF) mapping the vibration excitation force (F) to
the mass displacement x, for the same parameter variations.
![[Up]](up_botton.gif)
Damping Variation
Clear from the transmissibility plot of
Figure 2(a), lower damping does not affect neither the transmission
of force from the vibrating mass (e.g. diesel generator) to the base nor the
transmission of shock inputs at the base to the mass at low frequencies,
increase them around the resonant frequency, and lower these transmissions at
high frequencies. This is why in most isolation applications the mount is selected
to be highly underdamped with the resonant frequency of the isolated system less than
70% of the lowest vibration and shock excitation frequencies of the system. Of course,
when either or both excitations (vibration and shock) are broadband (like discontinuous
shock excitation common in marine applications), the above stated guideline will not
be effective; broadband excitation will set off the resonance of the mounted system
causing the mass to bounce. Figure 2(b) indicates that variation in damping only
influences the transmission of vibration force to the motion of the mass at resonance;
the lower the damping the more severe the resonance.
The reason practitioners do
not embrace the obvious solution to the resonance problem, i.e.,
adding damping, is that damping deteriorates the high-frequency vibration
isolation effectiveness of the mount resulting in transmission of noise.
Figure 2 Transmissibility and displacement Frequency response functions
for different damping ratios.
Stiffness Variation
Figure 3(a) indicates that the softer mounts (mounts with lower stiffness)
decrease both the transmission of force from the vibrating mass to the base and the
transmission of shock inputs at the base to the mass over a reasonably large range
in frequencies (which normally coincide with the frequency range of interest in most
mounting applications), including resonant frequency. Figure 3(b) indicates that
soft mounting adversely affects the transmission of vibration force to the motion of
the mass at lower frequencies, including at resonance; the lower the stiffness the
larger the low-frequency motion of the mass caused by vibration forces. This adverse
effect of soft-mounting would be tolerable if a) the excitation is harmonic
(not impulsive such as shock inputs) and b) the resonant frequency of the
isolated system is below the lowest excitation frequencies of the system.
Figure 3 Transmissibility and displacement Frequency response functions
for different stiffnesses.
Frequently, e.g. in marine vibration isolation applications,
the isolated object is excited by both periodic
vibration, e.g., the force of firing in a diesel engine, and impulsive shock, e.g.,
the base excitation in a marine diesel generator or any other vehicular engine
applications. The base excitation has normally a broadband spectrum, so there
is always some energy at the resonant frequency(ies) of the isolated system. If
the mounts are highly underdamped, the vibration amplitude at this (these) frequency(ies)
becomes excessive. This problem can be addressed by adding damping to the mount.
But in vibration isolators, as stated earlier, broadband damping which results in
high frequency vibration (and noise) transmission is avoided.
Adding Mass
Increasing the mass of the isolated object by adding an
auxiliary mass to it, in order to vary the attributes of the isolation
system, is a common high-frequency vibration isolation practice.
One way of realizing this scheme is to
install the vibrating machine on a massive concrete block and
isolate the block from the foundation by rubber or neoprene isolators.
A schematic of a simple mounted system with
an auxiliary mass, m, added to the isolated body (machine/equipment), M, is shown
in Figure 4.
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Figure 4 A simple mounting system with added mass (m)
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Figure 5(a) indicates that the addition of extra mass to the mounted object,
enhances the high-frequency shock and vibration isolation attributes. It does not
change the low-frequency dynamics of the system, except at resonant frequency.
It also lowers the resonant frequency of the system while making the system more
oscillatory (underdamped); note that damping ratio is inversely proportional to the
square root of the mass. Again, as in soft-mounting the adverse effect on resonane
would be tolerable if a) the excitation is harmonic (not impulsive such as shock inputs)
and b) the resonant frequency of the isolated system is below the lowest excitation
frequencies of the system.
Clear from Figure 5, the high-frequency isolation enhancement of added mass
will be realized when a sizeable amount of it, e.g., at least half the mass of the
machine being isolated, is used. To put this in perspective, for isolating a
170 KVA diesel generator weighting about 2 tons, one needs to attach a slab of steel
weighting at least 1 ton to the machine to make the effort worthwhile. The
inconvenience of implementing this solution, especially in retrofit applications,
not to mention the potential problems associated with excessive static displacement
of the mount due to the added mass is clear. In addition this solution has a
considerable weight penalty which is highly objectionable in vehicular/marine applications.
Figure 5 Transmissibility and displacement Frequency response functions
for different added masses (m).