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Tuned mass damper application for submerged floating tunnel under wave and seismic excitations
Published in Joan-Ramon Casas, Dan M. Frangopol, Jose Turmo, Bridge Safety, Maintenance, Management, Life-Cycle, Resilience and Sustainability, 2022
Mooring tension associated with significant dynamic responses is regarded as a critical problem under large wave and earthquake excitations, especially for large-size and low buoyancy to weight ratio (BWR) SFTs (Jin & Kim 2018). Specifically, the resonance of SFT induced by environmental loads should be controlled. The most fundamental way to mitigate resonance-induced vibrations is to change the system’s stiffness and damping coefficients (Yanik et al. 2016). For example, by changing the BWR and mooring design, the system’s natural frequency can largely be changed to avoid resonant motions. However, ocean environments have various environmental loadings covering a wide frequency range from relatively low-frequency wind/wave loads to high-frequency seismic loads. In this case, the vibration-control system can play a critical role in mitigating resonance-induced vibrations. While various vibration control methods have been suggested in the offshore industry, e.g., damping plate (Jang et al. 2019), tuned mass damper (TMD) (Wu et al. 2016), tuned liquid column damper (Lee et al. 2006), and magnetorheological damper (Wu et al. 2011), TMDs are considered one of the most reliable vibration control devices.
Mechanical System Design (Strength and Stiffness)
Published in Seong-woo Woo, Design of Mechanical Systems Based on Statistics, 2021
The quantity M is called the ‘magnification factor’. In a lightly damped system when the forcing frequency nears the natural frequency (r ≈ 1r ≈ 1), the amplitude of the vibration can be extremely high. This phenomenon is defined as resonance. In rotor-bearing systems, any rotational speed that excites a resonant frequency is referred to as a critical speed. If resonance happens in a mechanical system, it can be very dangerous – leading to ultimate failure of the system. Thus, one of the crucial reasons for vibration analysis is to forecast when this type of resonance may happen and then to decide what steps to take to prevent it from occurring. Adding on damping can notably lessen the magnitude of the vibration. Also, the magnitude can be lessened if the natural frequency might be shifted away from the forcing frequency by modifying the mass or stiffness of the system. The magnification factor relies on the frequency ratio and the damping factor as follows.
Examples 10.1 to 10.20
Published in L. M. B. C. Campos, Classification and Examples of Differential Equations and their Applications, 2019
The deformation and buckling of elastic bodies like bars, beams, and plates lead to fourth-order differential equations and are considered in the sixth group, consisting of sections K to Q and problems 178 to 338. The seventh group, consisting of sections R to T and problems 339 to 420, concerns linear and non-linear waves in inhomogeneous and unsteady media; the solutions are obtained for linear, steady, and inhomogeneous media, when the partial differential equations reduce to ordinary differential equations with variable coefficients. The eighth group, consisting of the sections U to Z and problems 421 to 500, concerns multidimensional oscillators with several degrees-of-freedom leading to simultaneous systems of ordinary differential equations describing combined oscillations and multiple resonance. Both the seventh and eighth groups concern waves and multidimensional oscillators, which involve several modes, infinite and finite in number. Four types of resonance are considered: simple, multiple, parametric, and non-linear, each with or without damping. The problems often have mechanical electric, acoustic, and other analogues. Classifications 10.1 and 10.2 provide a summary of the contents of this book and complement each other as a quick-look guide of where to find: (a) the solution of a specific differential equation among the 500 standards in Classification 10.1; (b) the indication of some contexts in which the differential equation arises and the interpretation and application of the solutions in Classification 10.2.
Design and fabrication of a novel passive hand tremor attenuator
Published in Journal of Medical Engineering & Technology, 2021
Mehdi Masoumi, Stephen Kmanzi, Hanchuan Wang, Hadi Mohammadi
In situations where the excitation frequency of the system is extremely similar to the natural frequency of the system the system can experience resonance causing heightened amplitudes in vibrational responses. This excessive vibration can cause unwanted effects such as high dynamic stresses, noise, and component fatigue. Although these unwanted effects are usually mitigated by changing the natural frequency of the system or the provided excitation frequency, in scenarios where this is infeasible it is often possible to better control the resonance condition of the system. The resonance condition of the system can be controlled by attaching a dynamic vibration absorber to the system allowing for some of the unwanted energy to be extracted from the system, therefore decreasing the vibrational response. As a vibration absorber can be simply defined as a spring-mass system the theory as to why this is effective at vibration reduction can be quickly explained. If one was presented with a machine with a mass of m1 to which they attached an additional mass, m2, by using a spring with a spring constant of k2, they would develop the two degree of freedom system as shown in Figure 1(a). By recognising the connection between the two masses and springs within the system, one can develop the relationships as shown in Equations (1) and (2) [8–10].
Numerical method for the vibration analysis of pre-swirl stator
Published in Ships and Offshore Structures, 2021
Nikola Vladimir, Andro Bakica, Šime Malenica, Hongil Im, Ivo Senjanović, Dae-Seung Cho
VIV is an unsteady oscillating flow phenomenon induced by the interaction of body and the external flow which causes larger vortices to detach from the body surface in an periodic manner. It can cause severe structural vibrations especially when resonance happens. Structural response can be reduced by increasing the structural damping or by breaking down the wake pattern by addition of spoilers, thus affecting the VIV frequency overlap with the natural modes. All told, it is important to estimate VIV frequencies beforehand for any structure susceptible to such resonance behaviour. The simplest approach for estimating the VIV frequencies is the use of empirical formulas. Although their lack in accuracy is evident, the practicality of approach makes them frequently used in the industry. Other approach, recently frequently investigated and much more accurate is by using CFD simulations. This type of investigation requires a significantly larger amount of work and knowledge to be applied properly. It should be noted that in this work, CFD simulations are performed without a grid refinement study and with only one type of turbulence model, hence the results presented are strictly for the entire procedure illustration purpose and their detailed investigation is beyond the scope of this paper.
Resonance analysis of opposed piston linear compressor for refrigerator application
Published in International Journal of Ambient Energy, 2019
Suneeta Phadkule, Sohel Inamdar, Asif Inamdar, Amit Jomde, Virendra Bhojwani
Resonance is the system behaviour which leads to vibrations of greater amplitudes at some frequencies than others. Resonance occurs when the exciting frequency matches the natural frequency of the body resulting in a higher amplitude of vibrations. Generally, resonance in mechanical components is not desirable since it may lead to early failure. But there are certain applications which need a higher amplitude with a lower input power for increasing the efficiency of the system like the linear compressor. Thus, the linear compressor performs with a maximum efficiency under the resonance condition. The natural frequency for a spring mass system is denoted by where fn is the natural frequency of the moving object, Hz; kgas is the gas spring stiffness, N/m; kmech is the mechanical spring stiffness, N/m and me the moving mass of the piston, kg.