Modal Overlap and Dissipation Effects of a Cantilever Beam with Multiple Attached Oscillators

2000 ◽  
Vol 123 (2) ◽  
pp. 181-187 ◽  
Author(s):  
M. V. Drexel ◽  
J. H. Ginsberg
Keyword(s):  
Transfer Function ◽  
Natural Frequency ◽  
Cantilever Beam ◽  
Impulse Response ◽  
Critical Issue ◽  
Harmonic Excitation ◽  
Vibration Absorber ◽  
Continuous Systems ◽  
Coupled Equations ◽  
Eigenmode Analysis

This work was prompted by a study performed by Strasberg [7] in which numerous small spring-mass-damper systems are attached to a large suspended mass representing the master structure. The isolated natural frequency of each attached system was selected to match in average the natural frequency of the isolated master structure. Strasberg found that the critical issue when an impulse excitation is applied to the master structure is the bandwidth of the isolated attached systems in comparison to the spacing between the natural frequencies of the system. Modal overlap, which corresponds to bandwidths that exceed the spacing of those frequencies, was shown to greatly influence the response of the master structure. Light damping, for which there is little or no modal overlap, corresponds to an impulse response that consists of a sequence of nearly periodic exponentially decaying pulses, and the transfer function for harmonic excitation of the master structure indicates that the substructure acts as a vibration absorber for the master structure. Increased damping leads to modal overlap, with the result that the impulse response consists of a single decaying pulse. The frequency domain transfer function indicates that the vibration absorber effect is enhanced. The present work explores these issues for continuous systems by replacing the one degree of freedom master structure with a cantilever beam. The system parameters are selected to match Strasberg’s model, with the suspended oscillators placed randomly along the beam. The beam displacement is represented as a Ritz series whose basis functions are the cantilever beam modes. The coupled equations are solved by a state-space eigenmode analysis that yields a closed form representation of the response in terms of the complex eigenmode properties. The continuous fuzzy structure is shown not to display the transfer of energy between the master structure and the substructure that was exhibited by the discrete fuzzy structure, apparently because of the asynchronous motion of the attachment points resulting from the spatial variability of the beam’s motion. The vibration absorber effect for harmonic excitation is only obtained for the heavy damping in the case of a beam.

Modal Overlap and Dissipation Effects in a Fuzzy Structure Containing a Continuous Master Element

1999 ◽  
Author(s):  
M. V. Drexel ◽  
J. H. Ginsberg
Keyword(s):  
Transfer Function ◽  
Natural Frequency ◽  
Cantilever Beam ◽  
Impulse Response ◽  
Critical Issue ◽  
Harmonic Excitation ◽  
Vibration Absorber ◽  
Continuous Systems ◽  
Coupled Equations ◽  
Eigenmode Analysis

Abstract This work was prompted by a study performed by Strasberg (1997) in which numerous small spring-mass-damper systems are attached to a large suspended mass representing the master structure. The isolated natural frequency of each attached system was selected to match in average the natural frequency of the isolated master structure. Strasberg found that the critical issue when an impulse excitation is applied to the master structure is the bandwidth of the isolated attached systems in comparison to the spacing between the natural frequencies of the system. Modal overlap, which corresponds to the bandwidths that exceed the spacing of those frequencies, was shown to greatly influence the response of the master structure. Light damping, for which there is little modal overlap, corresponds to an impulse response that consists of a sequence of nearly periodic exponentially decaying pulses, and the transfer function for harmonic excitation of the master structure indicates that the sub-structure acts as a vibration absorber for the master structure. Increased damping leads to modal overlap, with the result that the impulse response consists of a single decaying pulse. The frequency domain transfer function indicates that the vibration absorber effect is enhanced. The present work explores these issues for continuous systems by replacing the one degree of freedom master structure with a cantilever beam. The system parameters are selected to match Strasberg’s model, with the suspended sub-structure placed randomly along the beam. The beam displacement is represented as a Ritz series whose basis functions are the cantilever beam modes. The coupled equations are solved by a state-space eigenmode analysis that yields a closed form representation of the response in terms of the complex eigenmode properties. The continuous fuzzy structure is shown not to display the transfer of energy between the master structure and the substructure that was exhibited by the discrete fuzzy structure, apparently because of the asynchronous motion of the attachment points resulting from the spatial variability of the beam’s motion. The vibration absorber effect for harmonic excitation is only obtained for the heavy damping in the case of a beam.

scholarly journals Study of Magnetic Vibration Absorber with Permanent Magnets along Vibrating Beam Structure

2013 ◽  
Vol 2013 ◽  
pp. 1-5 ◽  
Author(s):  
F. B. Sayyad ◽  
N. D. Gadhave
Keyword(s):  
Cantilever Beam ◽  
Gas Turbines ◽  
Permanent Magnets ◽  
Excitation Frequency ◽  
Harmonic Excitation ◽  
Vibration Absorber ◽  
Primary System ◽  
Vibration Absorbers ◽  
Vibratory System ◽  
Electrical Generator

The vibration absorbers are frequently used to control and minimize excess vibration in structural system. Dynamic vibration absorbers are used to reduce the undesirable vibration in many applications such as pumps, gas turbines, engine, bridge, and electrical generator. To reduce the vibration of the system, the frequency of absorber should be equal to the excitation frequency. The aim of this study is to investigate the effect of magnetic vibration absorber along vibrating cantilever beam. This study will aim to develop a position of magnetic vibration absorber along the cantilever beam to adopt the change in vibratory system. The absorber system is mounted on a cantilever beam acting as the primary system. The objective is to suppress the vibration of the primary system subjected to a harmonic excitation whose frequencies are varying. It can be achieved by varying the position of magnetic vibration absorber along the length of beam. The advantage of magnetic vibration absorber is that it can be easily tuned to the excitation frequency, so it can be used to reduce the vibration of system subjected to variable excitation frequency.

Nonlinear Resonances in a Flexible Cantilever Beam

1994 ◽  
Vol 116 (4) ◽  
pp. 480-484 ◽  
Author(s):  
T. J. Anderson ◽  
B. Balachandran ◽  
A. H. Nayfeh
Keyword(s):  
Natural Frequency ◽  
Cantilever Beam ◽  
Natural Frequencies ◽  
Excitation Frequency ◽  
Continuous Systems ◽  
Modal Interactions ◽  
The Third ◽  
First Case ◽  
Second Mode ◽  
Band Limited

An experimental investigation into the response of a nonlinear continuous systems with many natural frequencies in the range of interest is presented. The system is a flexible cantilever beam whose first four natural frequencies are 0.65 Hz, 5.65 Hz, 16.19 Hz, and 31.91 Hz, respectively. The four natural frequencies correspond to the first four flexural modes. The fourth natural frequency is about fifty times the first natural frequency. Three cases were considered with this beam. For the first case, the beam was excited with a periodic base motion along its axis. The excitation frequency fe was near twice the third natural frequency f3, which for a uniform isotropic beam corresponds to approximately the fourth natural frequency f4. Thus the third mode was excited by a principal parametric resonance (i.e., fe ≈ 2f3) and the fourth mode was excited by an external resonance (i.e., fe ≈ f4) due to a slight curvature in the beam. Modal interactions were observed involving the first, third, and fourth modes. For the second case, the beam was excited with a band-limited random base motion transverse to the axis of the beam. The first and second modes were excited through nonlinear interactions. For the third case, the beam was excited with a base excitation along the axis of the beam at 138 Hz. The corresponding response was dominated by the second mode. The tools used to analyze the motions include Fourier spectra, Poincare´ sections, and dimension calculations.

Efficient Vibration Control with a Semi-Active Vibration Absorber in a Semiconductor Fab

2014 ◽  
Vol 564 ◽  
pp. 143-148 ◽  
Author(s):  
Teng Sheng Su ◽  
Chen Far Hung ◽  
Shu Hua Chang ◽  
Ting Hao Wu ◽  
Luh Maan Chang
Keyword(s):  
Control System ◽  
Vibration Control ◽  
Natural Frequency ◽  
Cantilever Beam ◽  
Active Vibration Control ◽  
Control Process ◽  
Vibration Absorber ◽  
Servo Motor ◽  
Resonant Condition ◽  
Active Vibration

In this paper a new type of semi-active vibration absorber has been developed. The vibration absorber consists of mass block, cantilever beam, magnet lock system, vibration and distance sensors, controller and servo motor. The mass block is fixed on the tip of cantilever beam, and the control process is driven by a servo motor and a transmit gears. Portion of cantilever was cut in form of gear tracks, which can be driven by servo motor through transmit gear to regulate the length of the cantilever beam, and the natural frequency of absorber will also be regulated. After the mass locates in right position (i.e. the natural frequency of absorber is in assigned condition), the magnetic lock will clamp the cantilever beam. The design has the benefit of simplified control system, and extra unknown vibration modes will be averted. A fabrication prototype of the proposed semi-active vibration absorber is constructed and tested to demonstrate the application and modeling of the new cantilever beam damper. By performing the experimental work, the semi-active vibration control system is designed not only for reduce vibration level in resonant condition, but also considered for vibration attenuation in non-resonant conditions.

Error analyses of a Multi-Input-Multi-Output cantilever beam test

2021 ◽  
pp. 107754632110337
Author(s):  
Arup Maji ◽  
Fernando Moreu ◽  
James Woodall ◽  
Maimuna Hossain
Keyword(s):  
Transfer Function ◽  
Frequency Domain ◽  
Cantilever Beam ◽  
Transfer Functions ◽  
Linear Superposition ◽  
Transfer Function Matrix ◽  
Error Analyses ◽  
Pseudo Inverse ◽  
Function Matrix

Multi-Input-Multi-Output vibration testing typically requires the determination of inputs to achieve desired response at multiple locations. First, the responses due to each input are quantified in terms of complex transfer functions in the frequency domain. In this study, two Inputs and five Responses were used leading to a 5 × 2 transfer function matrix. Inputs corresponding to the desired Responses are then computed by inversion of the rectangular matrix using Pseudo-Inverse techniques that involve least-squared solutions. It is important to understand and quantify the various sources of errors in this process toward improved implementation of Multi-Input-Multi-Output testing. In this article, tests on a cantilever beam with two actuators (input controlled smart shakers) were used as Inputs while acceleration Responses were measured at five locations including the two input locations. Variation among tests was quantified including its impact on transfer functions across the relevant frequency domain. Accuracy of linear superposition of the influence of two actuators was quantified to investigate the influence of relative phase information. Finally, the accuracy of the Multi-Input-Multi-Output inversion process was investigated while varying the number of Responses from 2 (square transfer function matrix) to 5 (full-rectangular transfer function matrix). Results were examined in the context of the resonances and anti-resonances of the system as well as the ability of the actuators to provide actuation energy across the domain. Improved understanding of the sources of uncertainty from this study can be used for more complex Multi-Input-Multi-Output experiments.

Nonlinear Normal Modes of a Continuous Cantilever Beam with Nonlinear Energy Sink Absorber

2013 ◽  
Vol 325-326 ◽  
pp. 214-217
Author(s):  
Yong Chen ◽  
Yi Xu
Keyword(s):  
Cantilever Beam ◽  
Vibration Suppression ◽  
Normal Modes ◽  
Nonlinear Energy Sink ◽  
Nonlinear Normal Modes ◽  
Harmonic Excitation ◽  
Energy Sink ◽  
Nonlinear Normal ◽  
Nonlinear Energy ◽  
Good Agreement

Using nonlinear energy sink absorber (NESA) is a good countermeasure for vibration suppression in wide board frequency region. The nonlinear normal modes (NNMs) are helpful in dynamics analysis for a NESA-attached system. Being a primary structure, a cantilever beam whose modal functions contain hyperbolic functions is surveyed, in case of being attached with NESA and subjected to a harmonic excitation. With the help of Galerkins method and Raushers method, the NNMs are obtained analytically. The comparison of analytical and numerical results indicates a good agreement, which confirms the existence of the nonlinear normal modes.

Dynamic Vibration Absorber for a Ropeway Carrier

1997 ◽  
Author(s):  
Hiroshi Matsuhisa ◽  
Osamu Nishihara
Keyword(s):  
Natural Frequency ◽  
Damping Ratio ◽  
Vibration Absorber ◽  
Moving Mass ◽  
Dynamic Vibration Absorber ◽  
Damping Effect ◽  
The World ◽  
Dynamic Absorber ◽  
First Time ◽  
Suspended Bridge

Abstract Ropeways such as gondola lifts have attracted increasing interest as a means of transportation in cities. However, swing of ropeway carriers is easily caused by wind, and usually a ropeway cannot operate if the wind velocity exceeds about 15m/s. The study of how to reduce the wind-induced swing of ropeway carriers has attracted many researchers. It had been said that it was impossible to reduce the vibration of pendulum type structures such as ropeway carriers by a dynamic absorber. But in 1993, Matsuhisa showed that the swing of carrier can be reduced by a dynamic absorber if it is located far above or below from the center of oscillation. Based on this finding, a dynamic absorber composed of a moving mass on an arc-shaped track was designed for practical use, and it was installed in chairlift-type carriers and gondola type carriers in snow skiing sites in Japan in 1995 for the first time in the world. It has been shown that a dynamic absorber with the weight of one tenth of the carrier can reduce the swing to half. The liquid dynamic absorber was also investigated. It has the same damping effect as the conventional solid absorber. It is easy to adjust the natural frequency and the damping ratio, and the structure is simple. Therefore, it will be applied for not only ropeway carriers but also ships and rope suspended bridge and others.

Transfer of Energy From High- to Low-Frequency Modes in the Bending-Torsion Oscillation of Cantilever Beams

1999 ◽  
Author(s):  
Haider N. Arafat ◽  
Ali H. Nayfeh
Keyword(s):  
Natural Frequency ◽  
Virtual Work ◽  
Low Frequency ◽  
Nonlinear Behavior ◽  
Excitation Amplitude ◽  
Harmonic Excitation ◽  
Nonlinear Bending ◽  
Torsional Mode ◽  
Bending Mode ◽  
Plane Bending

Abstract We investigate the nonlinear bending-torsion response of a cantilever beam to a transverse harmonic excitation, where the forcing frequency is near the natural frequency of the first torsional mode. We analyze the case where the first in-plane bending mode is activated by a nonresonant mechanism. We use the method of time-averaged Lagrangian and virtual work to determine the equations governing the modulations of the phases and amplitudes of the interacting modes. These equations are then used to investigate the nonlinear behavior of limit-cycle oscillations of the beam as the excitation amplitude is slowly varied. As an example, we consider the response of an aluminum beam for which the natural frequency of the first in-plane bending mode is fv1 ≈ 5.7 Hz and the natural frequency of the first torsional mode is fϕ1 ≈ 138.9 Hz.

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