AU2021102312A4

AU2021102312A4 - A system for dynamic vibration absorber and method of operation thereof

Info

Publication number: AU2021102312A4
Application number: AU2021102312A
Authority: AU
Inventors: Mahadev B. Kumbhar; Avinash V. Borgaonkar; Vishal G. Salunkhe; Pankaj B. Patil; Shital B. Potdar; Jagadeesha T.
Original assignee: Individual
Current assignee: Individual
Priority date: 2021-05-01
Filing date: 2021-05-01
Publication date: 2021-07-01
Anticipated expiration: 2029-05-01

Abstract

The present disclosure relates to a system for dynamic vibration absorber using mathematical modeling and experimental evaluation of an air spring- air damper dynamic vibration absorber. To optimize and effectively controlling the resonant amplitude ratio efficiently, a Maxwell-type air damper mathematical model is formulated for an SDOF vibrating system subjected to sinusoidal base excitation for better performance of DVA. Insight into the physics of such systems was given by mathematical modeling, design and development of an air spring air damper system for SDOF vibrating system. Experimental investigation of the air spring-air damper for the primary mass displacement amplitude to optimize and effectively control the vibration will provide a compact and right platform for DVA among the current potential market and provide the major attention of the control community as a platform for the development of low-frequency nonlinear devices. 17 C40 AMLI tw LAM

Description

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A system for dynamic vibration absorber and method of operation thereof

FIELD OF THE INVENTION

The present disclosure relates to system for dynamic vibration absorber.

BACKGROUND OF THE INVENTION

One of the most successful ways to use a dynamic vibration absorber (DVA) using air as the working medium is a semi-active air spring-air damping system with a facility to change the spring rate and damping in the system. As per the current potential market trend, to semi-active dynamic Vibration absorber are mainly MR/ER damper and pendulum type tuned absorber, air spring- air damper. Among these, MR/ER damper faces the problem regards the major cause is response time, several MR fluid drawbacks as per scenario, power supply, and a high cost of the damper. Secondly, pendulum type tuned absorber faces the problem regards, consistency, maintenance, the joints between the parts produces the friction which will disturb the vibration absorber. To overcome this air spring air damper have several merits over other types of DVAs, such as damping properties are tending to be consistent, have a low cost of manufacture, require less maintenance, working temperature, and environment, etc. It is seen that this semi-active air spring-air damping design has high performance and having a large number of practical applications such as, for automotive suspension, machinery vibration, semiconductor manufacturing industry sensitive or micro-vibration isolation, spatial vibrations of railway vehicles. The Maxwell type air damper model design is also presented for optimization of the amplitude ratio when the SDOF vibrating system is subjected to base excitation. The design determined effectively damping ratio, air spring rate ratio, and natural frequency ratio. The pneumatic automotive suspension mathematical model for air spring stiffness variation, the thermodynamic model for damping variation towards a damping model in a linearized interval is derived where the model is analyzed for load-noise isolation concepts. Pneumatic vibration isolator with its transfer function applying it in active vibration isolation as well as non-linear vibration control design against model uncertainty and sensor-actuator limitations.

To obtain optimum damping ratio, cylinder-piston and air-tank type air damper configuration are selected for the design and development of pneumatic DVA. The concept of pneumatic damper coupled with an electrical system to control the vibration in semiconductor manufacturing industries. An active vibration isolation system was developed with an electro pneumatic damper to isolate the payload in the tool and optimize settling time. It is shown that when pneumatic isolation is to be used a three-axis active system can be developed with practically no compliance by equalizing negative and positive stiffness individually and to generate infinite spring rate/s.

It examines the pneumatic system as a shock absorber model isentropic and adiabatic gas flow from space one with known parameters of actual gas state to space second, which also shows the difference between the function of actual gas flow and function of ideal gas flow at different initial temperatures and pressures of gas. A mathematical model of air spring with only one chamber with the effect of the heat transfer coefficient between wall and gas outside the chamber, for amplitude spectrum under different vibration excitation frequencies designed. The result of the pneumatic vibration isolation process shows nonlinear characteristics, when the drive frequency is near to linearized natural frequency, the system becomes more nonlinear and non-linearity produces larger amplitude. Mathematical analysis of an equivalent mass-spring damper model with piston-cylinder structure and another, nonlinear model aspects are designed for formulating, the transfer function that describes the passive dynamic characteristics of the hydro-mount system for the attenuation of engine vibration transmitted to a chassis. This concluded that a good model has a high known consistency in reality. The parametric analysis included the air spring dimensions, hoses dimensions, air pressure, and reservoir volume, and dynamic responses including acceleration, dynamic tire force, and suspension travel were then compared in the form of time and frequency domain analysis differentiates concerning the passive suspension. Mathematical modeling for determining the effect of energy conservation, valve conservation result out force, and sensitivity analysis for dynamic stiffness are discussed.

A simulation model containing the hydro-pneumatic suspension system has been developed for the rally truck. The model is validated by experimental data and close effective results show concerning static analysis, modal analysis, and harmonic analysis. Simulation results of hydro-pneumatic suspension with the inter-connection type are presented for the influence of i) accumulator parameters, ii) damping characteristics and iii) tracking controller.

Active Force Controller (AFC) based feedback controller for pneumatic actuator employed for excellent performance to absorb the random vibrations plays into the suspension system. A mathematical model of vehicle height adjustment to compensate for the "over charging", "over-discharging" and oscillation is derived by combining vehicle dynamics theory and thermodynamics theory of variable mass system with the help of Electronically Controlled Air Suspension (ECAS). A coupled hydraulically interconnected suspension and electronic controlled air spring is created to control the non-linear stiffness and low bounce frequency, which can significantly improve ride comfort.

The nonlinear mathematical model can represent the behavior of a linear and quasi-linear model of air spring for sensitivity analysis of pressure shows that the air spring model is sensitive to the effective area and model sensitivity to the thermodynamic process is small. The dynamic parameters of vibration were examined through both speeds of the damper piston and displacement. The different thermodynamic air spring models are studied concerning the geometric and physical parameters of air spring systems, for a linear spring in parallel with a linear dashpot. A mechanical model was developed for the air spring air damper to forecast the nonlinear frequency and amplitude-dependent dynamic parameters. Based on these parameters proposed air spring model has been verified through bench test results. The industry has a high demand for an accurate, simple air spring model, and a method that integrates the air spring model into a full vehicle model for ride stability and quality of comfort prediction. The increased air pressure results in a greater ratio between the upper and lower rigidity level and greater damping capacity, thereby improving the driving safety because higher damping forces act on the increased (additional load-bearing) body mass. So, the application of dynamic vibration absorber depends on the amount of vibration that absorb without affecting its functionality or structural failure.

However, an air damper consisting of cylinder piston configuration can be used to effectively control air spring stiffness and air damping properties associated with an SDOF vibrating System subjected to base excitation. Therefore, there is a need of a system for dynamic vibration absorber. SUMMARY OF THE INVENTION

The present disclosure relates to a system for dynamic vibration absorber. The system comprise, a frame structure, an air damper, an excitation module, and at least two measuring module. To optimize and effectively controlling the resonant amplitude ratio efficiently, a Maxwell-type air damper mathematical model is formulated for an SDOF vibrating system subjected to sinusoidal base excitation for better performance of DVA. Insight into the physics of such systems was given by mathematical modeling, design and development of an air spring air damper system for SDOF vibrating system. Experimental investigation of the air spring air damper for the primary mass displacement amplitude is done to optimize and effectively control the vibration.

The present disclosure seeks to provide a system for dynamic vibration absorber. The system comprises: a frame structure for holding the system, wherein the frame structure comprises of a primary mass and a primary spring for absorbing a vibration by contraction of spring; an air damper connected to a structural frame of the system for circulating the flow of air inside the support frame, wherein the air damper comprises of an auxiliary mass and an auxiliary spring for reducing an amplitude of vibration; an excitation module connected to a base of the primary spring for generating vibrations to vibrate the primary mass, wherein the vibration is generated in a specific frequency range by a cam module embodied in the excitation module upon varying a rotational speed of the cam module; and at least two measuring module connected to the excitation module for measuring the rotational speed of the excitation module.

The present disclosure also seeks to provide a method for dynamic vibration absorber. The method comprises of: holding the system using a frame structure, wherein absorbing a vibration by contraction of spring using the frame structure comprises of a primary mass and a primary spring; circulating the flow of air inside the support frame using an air damper connected to a structural frame of the system, wherein reducing an amplitude of vibration using the air damper comprises of an auxiliary mass and an auxiliary spring; generating vibrations to vibrate the primary mass using an excitation module connected to a base of the primary spring, wherein the vibration is generated in a specific frequency range by a cam module embodied in the excitation module upon varying a rotational speed of the cam module; and measuring the rotational speed of the excitation module using at least two measuring module connected to the excitation module.

An objective of the present disclosure is to develop a system for dynamic vibration absorber.

Another object of the present disclosure is to optimize and effectively controlling the resonant amplitude ratio efficiently.

Another object of the present disclosure is to investigate the air spring air damper for the primary mass displacement amplitude to optimize and effectively control the vibration.

Yet, another object of the present disclosure is to develop an air damper that is effective in controlling the resonant amplitude ratio in an optimized manner.

To further clarify advantages and features of the present disclosure, a more particular description of the invention will be rendered by reference to specific embodiments thereof, which is illustrated in the appended drawings. It is appreciated that these drawings depict only typical embodiments of the invention and are therefore not to be considered limiting of its scope. The invention will be described and explained with additional specificity and detail with the accompanying drawings.

BREIF DESCRIPTION OF FIGURES

These and other features, aspects, and advantages of the present disclosure will become better understood when the following detailed description is read with reference to the accompanying drawings in which like characters represent like parts throughout the drawings, wherein:

Figure 1 illustrates a block diagram of a system for dynamic vibration absorber in accordance with an embodiment of the present disclosure;

Figure 2 illustrates a flow chart of a method for dynamic vibration absorber in accordance with an embodiment of the present disclosure;

Figure 3 illustrates a single degree of freedom vibratory system in accordance with an embodiment of the present disclosure;

Figure 4 illustrates the specifications of an air damped 2DOF vibrating system with an optimized amplitude ratio in accordance with an embodiment of the present disclosure;

Figure 5 illustrates an air damper modeled as a Maxwell model in accordance with an embodiment of the present disclosure; Figure 6 illustrates the dimensions of an air damper in accordance with an embodiment of the present disclosure;

Figure 7 illustrates the Theoretical Frequency response plot using air Damper in accordance with an embodiment of the present disclosure;

Figure 8 illustrates the dimensions of completely designed system in accordance with an embodiment of the present disclosure;

Figure 9 illustrates the flow chart for calculating air damper parameters / dimensions in accordance with an embodiment of the presented disclosure;

Figure 10 illustrates a SDOF system in accordance with an embodiment of the presented disclosure;

Figure 11 illustrates the experimental set-up of Air Damped Dynamic Vibration Absorber in accordance with an embodiment of the present disclosure;

Figure 12 illustrates the primary setup with m, k, c, trio in accordance with an embodiment of the present disclosure;

Figure 13 illustrates the experimental frequency response of the primary system in accordance with an embodiment of the present disclosure;

Figure 14 illustrates the experimental frequency response plot in accordance with an embedment of the present disclosure;

Figure 15 illustrates the observations from the theoretical SDOF system with air damped DVA in accordance with an embodiment of the present disclosure;

Figure 16 illustrates the experimental results of optimization of amplitude ratio M in accordance with an embodiment of the present disclosure;

Further, skilled artisans will appreciate that elements in the drawings are illustrated for simplicity and may not have been necessarily been drawn to scale. For example, the flow charts illustrate the method in terms of the most prominent steps involved to help to improve understanding of aspects of the present disclosure. Furthermore, in terms of the construction of the device, one or more components of the device may have been represented in the drawings by conventional symbols, and the drawings may show only those specific details that are pertinent to understanding the embodiments of the present disclosure so as not to obscure the drawings with details that will be readily apparent to those of ordinary skill in the art having benefit of the description herein. DETAILED DESCRIPTION

For the purpose of promoting an understanding of the principles of the invention, reference will now be made to the embodiment illustrated in the drawings and specific language will be used to describe the same. It will nevertheless be understood that no limitation of the scope of the invention is thereby intended, such alterations and further modifications in the illustrated system, and such further applications of the principles of the invention as illustrated therein being contemplated as would normally occur to one skilled in the art to which the invention relates.

It will be understood by those skilled in the art that the foregoing general description and the following detailed description are exemplary and explanatory of the invention and are not intended to be restrictive thereof.

Reference throughout this specification to "an aspect", "another aspect" or similar language means that a particular feature, structure, or characteristic described in connection with the embodiment is included in at least one embodiment of the present disclosure. Thus, appearances of the phrase "in an embodiment", "in another embodiment" and similar language throughout this specification may, but do not necessarily, all refer to the same embodiment.

The terms "comprises", "comprising", or any other variations thereof, are intended to cover a non-exclusive inclusion, such that a process or method that comprises a list of steps does not include only those steps but may include other steps not expressly listed or inherent to such process or method. Similarly, one or more devices or sub-systems or elements or structures or components proceeded by "comprises...a" does not, without more constraints, preclude the existence of other devices or other sub-systems or other elements or other structures or other components or additional devices or additional sub-systems or additional elements or additional structures or additional components.

Unless otherwise defined, all technical and scientific terms used herein have the same meaning as commonly understood by one of ordinary skill in the art to which this invention belongs. The system, methods, and examples provided herein are illustrative only and not intended to be limiting.

Embodiments of the present disclosure will be described below in detail with reference to the accompanying drawings.

Figure 1 illustrates a block diagram of a system for dynamic vibration absorber in accordance with an embodiment of the present disclosure. The system 100 includes a frame structure unit 102. This unit is used for holding the system, wherein the frame structure comprises of a primary mass and a primary spring for absorbing a vibration by contraction of spring.

In an embodiment, an air damper unit 104 is connected to a structural frame of the system for circulating the flow of air inside the support frame, wherein the air damper comprises of an auxiliary mass and an auxiliary spring for reducing amplitude of vibration. To determine the effect of developed Air Damper to control the primary mass displace to a minimum value.

In an embodiment, an excitation module 106 is connected to a base of the primary spring for generating vibrations to vibrate the primary mass, wherein the vibration is generated in a specific frequency range by a cam module embodied in the excitation module upon varying a rotational speed of the cam module.

In an embodiment, a measuring module unit 108, which consist of at least two measuring modules is connected to the excitation module for measuring the rotational speed of the excitation module.

Figure 2 illustrates a flow chart of a method for dynamic vibration absorber in accordance with an embodiment of the present disclosure. At the step 202 the method 200 includes, holding the system using a frame structure, wherein absorbing a vibration by contraction of spring using the frame structure comprises of a primary mass and a primary spring.

At the step 204 the method 200 includes, circulating the flow of air inside the support frame using an air damper connected to a structural frame of the system, wherein reducing amplitude of vibration using the air damper comprises of an auxiliary mass and an auxiliary spring.

At the step 206 the method 200 includes, generating vibrations to vibrate the primary mass using an excitation module connected to a base of the primary spring, wherein the vibration is generated in a specific frequency range by a cam module embodied in the excitation module upon varying a rotational speed of the cam module.

At step 208 the method 200 includes, measuring the rotational speed of the excitation module using at least two measuring module connected to the excitation module. Dimmer-Stat is used for cam rotation speed measurement.

Figure 3 illustrates a single degree of freedom vibratory system in accordance with an embodiment of the present disclosure. A single degree of freedom vibrating system may be a simple model of any real physical system subjected to base excitation u(t) resulting in sprung mass displacement x(t). The primary system mi,ki, cthe trio becomes a 2-DOF system subjected to base excitation when associated with an auxiliary systemim 2 , k 2 , c 2 trio. The figure shows the schematic diagram of the primary system mi,ki, c1 trio associated with an auxiliary systemim 2 , k 2 , c2 the trio, associated with an 'Air Damper modeled with Maxwell concept with ka,ca parameters.

Figure 4 illustrates the specifications of an air damped 2DOF vibrating system with an optimized amplitude ratio in accordance with an embodiment of the present disclosure. The Air Damper when attached to the primary system the combined system becomes 2DOF vibrating system. This figure defines the specification of an air damped 2DOF vibrating system with an optimized amplitude ratio.

Figure 5 illustrates an air damper modeled as a Maxwell model in accordance with an embodiment of the present disclosure. The Air Damper when attached to the primary system the combined system becomes 2DOF vibrating system subjected to base excitation and theoretically the amplitude ratio M is suppressed from the resonant peak from 25 at 162 rpm cam rotation into two peaks i) M= 3.568 at 140 rpm cam rotation and ii) M= 3.611 at 198 rpm cam rotation The following equations represent the details of the Air Damper parameters.

a) Damping coefficientc = A a32

b) Eccentricity E =

c) Spring constant ka hP

d) Coefficient of discharge resistance ao_Iol 2 1

e) Nomenclature

b = rc(de + dp)/2 h = (d, - d,)/2: Radial clearance

I = Length of the piston dy, Ap: Diameter, area of the piston

d, The inner diameter of the cylinder h, Height from the bottom of the cylinder to the piston

y : Ratio of the specific heat of the air (=1.4)

po po: Atmospheric air density and pressure y 0 vo: Viscosity and kinetic viscosity of air (p1, = pOvO) For, - - =.465 = 0.1500 then kopt = 1.1307,vopt = 0.7453, -aOPt 0.3297. m, 4.31o

Figure 6 illustrates the dimensions of an air damper in accordance with an embodiment of the present disclosure. A MATLAB program is developed to estimate the Air Damper dimensions as under to provide the optimized amplitude ratio of primary mass m, displacement.

Figure 7 illustrates the Theoretical Frequency response plot using air Damper in accordance with an embodiment of the present disclosure. The frequency response plot is plotted using optimized values of kptVcpt, (aopt

Figure 8 illustrates the dimensions of completely designed system in accordance with an embodiment of the present disclosure. It is assumed that the damping coefficients c, andc 2 as zero while optimizing the air damper parameters. c, is to be evaluated experimentally and the effectof c2will be submerged into the air damper working.

Figure 9 illustrates the flow chart for calculating air damper parameters / dimensions in accordance with an embodiment of the presented disclosure. The mathematical model, reporting the physical system i.e. Air Damped dynamic vibration absorber system (a single degree of freedom system subjected to base excitation) was explained in mathematical modeling. The aim x was to optimize the amplitude ratio M - , for varying frequency ratio U

l using the optimized air damper parameters. Air spring design is based on the theoretical optimization of the Amplitude Ratio M = -, U,

The parameters (aopt,kpt and vopt as discussed in modeling and depending on the value of

mass ratio yP the theoretical analysis, considering these parameters, develops the ratiosri,

r2 and r 3 which are dimensionless parameters for the physical dimension of the Air Damper as under. r1 ,r2 =- and r3 =

Thus the procedure is developed as a flow chart.

Figure 10 illustrates a SDOF system in accordance with an embodiment of the presented disclosure. Wherein mn 1k , c i trio subjected to base excitation u(t), providing system output xi(t). As it is expected to design a laboratory model of the system, it is better to design it for a low-frequency operation i.e. the natural frequency of the primary system wi maybe about 15-20 rad/sec i.e. the cam rotation (140-190 rpm). The arrangement for the generation of vibrations of the primary mass mi must be simple and smooth in operation with the selected natural frequency say, a 1 = 18.45 rad/ sec (175.3 rpm, cam rotation).

Figure 11 illustrates the experimental set-up of Air Damped Dynamic Vibration Absorber in accordance with an embodiment of the present disclosure. Experimental set up of Air Damped Dynamic Vibration Absorber consists of,

1. Base Excitation System, a cam-follower system with variable speed drive. 2. Structural Frame to support mi,ki, c1 trio. 3. SDOF System-primary system. 4. Developed Air Damper with auxiliary spring and auxiliary mass. 5. Digital Contact Type Tachometer. 6. Dimmer-Stat for cam rotation speed measurement.

It is proposed to determine the effect of developed Air Damper to control the primary mass displacement to a minimum value. Experiments are conducted as follows, to the determination of primary coefficient of viscous damping c, and undamped natural frequency a1 and to the determination of the suppression of the sprung mass maximum displacement amplitude to an acceptable level. These experiments are carried out for sinusoidal excitation input at the base (U=i Imm).

Figure 12 illustrates the primary setup with ml, k1, c, trio in accordance with an embodiment of the present disclosure. The figure shows Experimental determination of the primary coefficient of viscous damping c, using half power point method, which shows the developed primary system mi,ki, c1 trio with the dimensions.

Figure 13 illustrates the experimental frequency response of the primary system in accordance with an embodiment of the present disclosure. The SDOF system is subjected to harmonic base excitation with an amplitude of U= 1 mm and the system responses X are observed at different cam rotation speeds. The data generated is shown in this figure in tabular form.

Figure 14 illustrates the experimental frequency response plot in accordance with an embedment of the present disclosure. The SDOF system is subjected to harmonic base excitation with amplitude of U= 1 mm and the system responses X are observed at different cam rotation speeds. The data generated is plotted as Amplitude Ratio. M=(X/U) vs. Cam Rotation Speeds in rpm.

Figure 15 illustrates the observations from the theoretical SDOF system with air damped DVA in accordance with an embodiment of the present disclosure. The developed air damper system is attached to the selected and developed SDOF system for obtaining its frequency response plot. The observations are tabulated in this figure.

Figure 16 illustrates the experimental results of optimization of amplitude ration M in accordance with an embodiment of the present disclosure. The experimental investigation indicates the following aspects.

1. The damped resonant frequency for the SDOF system is at 175 rpm cam rotation and the amplitude is about 25 i.e. )ld= 1 8 .4 2 rad/sec. 2. The primary experimental damping ratio is about =0.02. 3. The air damper is used to convert the SDOF system into a 2DOF system 4. The maximum amplitude ratio of 25 is suppressed to about 3.5 and it has two peaks i.e. 3.5 at 130 rpm and 3.0 at 182 rpm. Also, there is a minimum value of the amplitude ration of 1.75 at 155 rpm.

The maximum amplitude ratio Mmax=25 (with SDOF) suppressed to about 3 with two peaks (about 75% reduction) and its minimum value is between two peaks is 1.75 (about 93% reduction). This means that the unacceptable frequency range from 125 to 200 rpm of cam speed (SDOF system) is converted into an acceptable range using the developed air damper (2DOF system). Hence the developed air damper is quite effective in controlling the resonant amplitude ratio in an optimized manner.

The drawings and the forgoing description give examples of embodiments. Those skilled in the art will appreciate that one or more of the described elements may well be combined into a single functional element. Alternatively, certain elements may be split into multiple functional elements. Elements from one embodiment may be added to another embodiment. For example, orders of processes described herein may be changed and are not limited to the manner described herein. Moreover, the actions of any flow diagram need not be implemented in the order shown; nor do all of the acts necessarily need to be performed. Also, those acts that are not dependent on other acts may be performed in parallel with the other acts. The scope of embodiments is by no means limited by these specific examples. Numerous variations, whether explicitly given in the specification or not, such as differences in structure, dimension, and use of material, are possible. The scope of embodiments is at least as broad as given by the following claims.

Benefits, other advantages, and solutions to problems have been described above with regard to specific embodiments. However, the benefits, advantages, solutions to problems, and any component(s) that may cause any benefit, advantage, or solution to occur or become more pronounced are not to be construed as a critical, required, or essential feature or component of any or all the claims.

Claims

WE CLAIM

1. A system for dynamic vibration absorber, the system comprises of:

a frame structure for holding the system, wherein the frame structure comprises of a primary mass and a primary spring for absorbing a vibration by contraction of spring;

an air damper connected to a structural frame of the system for circulating the flow of air inside the support frame, wherein the air damper comprises of an auxiliary mass and an auxiliary spring for reducing an amplitude of vibration;

an excitation module connected to a base of the primary spring for generating vibrations to vibrate the primary mass, wherein the vibration is generated in a specific frequency range by a cam module embodied in the excitation module upon varying a rotational speed of the cam module; and

at least two measuring module connected to the excitation module for measuring the rotational speed of the excitation module.

2. The system as claimed in claim 1, wherein the at least two measuring module comprises of a dimmer module for measuring the rotational speed of the excitation module for generating vibration and a digital measuring module for measuring the rotation of the excitation module.

3. The system as claimed in claim 1, wherein the excitation module creates a sinusoidal input of the vibration.

4. The system as claimed in claim 1, where in the air damper uses a single degree of freedom, wherein the auxiliary mass is surrounded by an auxiliary spring.

5. The system as claimed in claim 1, wherein the natural frequency of the vibration is approximately 15-20 rad/sec and the cam rotation is 140-190 rpm.

6. The system as claimed in claim 1, wherein a resonant frequency for the Single Degree of freedom is at 175 rpm cam rotation and the amplitude is about 25.

7. A method for operation of dynamic vibration absorber, the method comprises of:

holding the system using a frame structure, wherein absorbing a vibration by contraction of spring using the frame structure comprises of a primary mass and a primary spring;

circulating the flow of air inside the support frame using an air damper connected to a structural frame of the system, wherein reducing an amplitude of vibration using the air damper comprises of an auxiliary mass and an auxiliary spring;

generating vibrations to vibrate the primary mass using an excitation module connected to a base of the primary spring, wherein the vibration is generated in a specific frequency range by a cam module embodied in the excitation module upon varying a rotational speed of the cam module; and

measuring the rotational speed of the excitation module using at least two measuring module connected to the excitation module.

A frame structure 102 An air damper104

An excitation module106 Measuring module 108

Figure 1

Figure 3

Figure 4

Figure 5

Height of the piston L 35mm Diameter area of the piston Dp 31.97mm The inner diameter of the Dc 32mm cylinder Height from the bottom of Hp 30mm the cylinder to the piston The outer diameter of the D 45mm cylinder Total cylinder height H 145mm Weight of Air Damper Wt 1.48kg Figure 6

Figure 7

Figure 8

Figure 9

Figure 10

4. Auxiliary 2. Structural mass Frame

4. Auxiliary spring

4. Air Damper 3. Primary mass

3. Primary spring

1. Base Excitation System 6. Dimmer 5. Tachometer

Figure 11

Figure 12

Cam rotation speed (rpm) Amplitude ratio(M= ) 0 1 50 1.2 100 1.25 150 3 162 3.5 170 7.5 175 25 180 12 200 5 225 2 250 1

Figure 13

Figure 14

Cam speed rpm Amplitude Ratio M and Theoretical Experimental µ = 0.15 Amplitude ratio Amplitude ratio 0 1 1 100 1.92 1.5 130 3.56 3.5 155 2.75 1.75 182 3.62 3 200 2.62 1.5 250 0.84 0.75

Figure 15

Figure 16