t t m M K = ω (8) where t M is the total moving mass and tK is the total equivalent spring stiffness. This approach is based on one degree-of-freedom system in Fig. 1a. Strictly speaking, however, this theoretical model is improper because the moving mass part is not directly connected to ground, but it is mounted on the shell through snubbers and the shell is placed on the hard floor through the grommet. That is, a linear compressor operating at a motor current frequency can be characterized by the 3rd model in Fig. 1c. Therefore, to reduce high level vibration response, one must use the 4th model in Fig. 1d, not the 2nd model in Fig. 1b. 0 20 40 60 80 100 10-8 10-6 10-4 10-2 Frequency(Hz) |FRF| 3rd model 4th model 0 20 40 60 80 100 10-8 10-6 10-4 10-2 Frequency(Hz) |FRF| 3rd model 4th model Fig. 4 Level change in frequency response functions of the 3rd and 4th models depending on excitation force frequency: (a) 3 * 2 Ω =Ω =Ω c (b) 3 * 2 Ω ≠Ω =Ω c . Conclusions In this work, dynamic absorber design scheme was investigated in terms of industrial application. The key idea in dynamic absorber design theory is to adjust the natural frequency of an additional system to excitation force frequency. However, the reason that the vibration response of the original system is reduced is that an anti-resonance frequency in a frequency response function is tuned to the excitation force frequency. Sometimes, the natural frequency adjustment of the additional system may not result in exact shift in the anti-resonance frequency. From the practical point of view, therefore, the anti-resonance frequency tuning is much more important than the natural frequency adjustment. 279
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