• Plant Journal
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• v.4 no.3
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• pp.66-72
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• 2008

The vibration of steam turbine is caused by Mass unbalance, Shaft misalignment, Oil whip and rubbing etc. But in turbine which is normally operated and maintained, the Mass unbalance component possesses the greatest portion. Our power plant has two steam turbines in capacity of 200 MW and 135 MW respectively and each turbine is supported by 6 journal bearings. However, we had many difficulties because the vibration amplitude of #3 and #4 Bearings was high during the start-up and operation mode change of steam turbine. But, with this study, we completely solved the vibration problem caused by the mass unbalance of #1 steam turbine. Until a recent date, #3 and #4 bearings which support high pressure turbine for #1 steam turbine had shown about $135{\mu}m$ in vibration amplitude (sometimes it increased to $221{\mu}m$ maximum. alarm: 6 mils, trip: 9 mils) at base load. After applying the study, they decreased to about $45{\mu}m$ maximum. It is a result from that we did not change the setting value of bearing alignment and only changed the assembly position of internal parts in Synchro clutch coupling rachet wheel which links between high pressure turbine and low pressure turbine, and increased the internal gap and machining of the Pawl cage surface. In the operation of steam turbine, if the vibration value increases by 1X, we should reduce the vibration of bearing by weight balancing. However, unless the vibration of bearing is declined by the balancing, we will have to disassemble and check the component and find the cause. In this study, we researched the way to lower mass unbalance that is 1X vibration component which has the greatest portion of vibration generated by steam turbine and we got good result by applying the findings of this study.

• Korean Journal of Applied Biomechanics
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• v.27 no.2
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• pp.83-90
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• 2017

Objective: To investigate the effect of short-term vibration frequencies on muscle force generation capabilities. Method: Six healthy participants were recruited for this study and only their dominant leg was tested. The subjects were tested under five conditions of vibration frequencies with constant amplitude: 0 Hz (no vibration), 30 Hz, 60 Hz, and 90 Hz, and the vibration amplitude was 10 mm for all frequency conditions. The vibration was applied to the rectus femoris (RF). The subjects were then instructed to maintain a steady-state isometric knee joint torque (100 Nm) for the first 6 s. After the steady-state torque production, the subjects were required to produce isometric knee joint torque by leg extension as hard as possible with a start signal within the next 3 s. The vibration was applied for ~4 s starting from 1 s before initiation of the change in the steady-state knee joint torque. Results: The results showed that the maximum voluntary torque (MVT) of the knee joint increased with the vibration frequencies. On average, the MVTs were 756.47 Nm for 0 Hz (no vibration) and 809.61 Nm for 90 Hz. There was a significant positive correlation (r = 0.71) between the MVTs and integrated electromyograms (iEMGs). Further, the co-contraction indices (CCIs) were computed, which represent the ratio of the iEMGs of the antagonist muscle to the iEMGs of all involved muscles. There was a significant negative correlation (r = 0.62) between the CCIs and MVTs, which was accompanied by a significant positive correlation (r = 0.69) between the iEMGs of the vibrated muscle (RF). There was no significant correlation between the MVTs and iEMGs of the antagonist muscle. Conclusion: The results of this study suggest that the short-term vibration on the muscle increases the level of muscle activation possibly owing to the increased Ia afferent activities, which enhances the muscle force generation capability.

• 한국전산유체공학회:학술대회논문집
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• 2008.03b
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• pp.232-235
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• 2008

The vibration of steam turbine is caused by Mass Unbalance, Shaft Misalignment, Oil Whip and Rubbing etc. but in turbine which is normally operated and maintained, the Mass Unbalance component possesses the greatest portion. Our power plant has two steam turbines in capacity of 200MW and 135MW respectively and each turbine is supported by 6 journal bearings. However, we had many difficulties because the vibration amplitude of No 3 and 4 Bearings was high during the start-up and operation mode change of steam turbine. But, with this study, we completely solved the vibration problem caused by the mass unbalance of No 1 steam turbine. Until a recent date, No 3 and 4 bearings which support high pressure turbine for No 1 steam turbine had shown about 135${\mu}$m in vibration amplitude (sometimes it increased to 221${\mu}$m maximum. alarm: 6mils, trip: 9mils) at base load. After applying the study, they decreased to about 40${\mu}$m maximum. It is a result from that we did not change the setting value of Bearing Alignment and only changed the assembly position of internal parts in Synchro Clutch Coupling Rachet Wheel which links between high pressure turbine and low pressure turbine, and increased the internal gap and machining of the Pawl stopper surface. In the operation of steam turbine, if the vibration value increases by 1X, we should reduce the vibration of bearing by weight balancing. However, unless the vibration of bearing is declined by the balancing, we will have to disassemble and check the component and find the cause. In this study, We researched the way to lower mass unbalance that is 1X vibration component which has the greatest portion of vibration generated by steam turbine and We got good result by applying the findings of this study.

• Proceedings of the Korean Society for Noise and Vibration Engineering Conference
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• 2001.11b
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• pp.1356-1360
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• 2001

A number of calibration methods are described in ISO 16063 (revision of ISO 5347) and they may be used for special purposes. However, the use of a laser interferometer is recommended for primary calibration. This paper introduces the primary vibration calibration by measuring displacement amplitude and frequency using laser interferometry.

• Wind and Structures
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• v.25 no.4
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• pp.343-360
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• 2017

The post-flutter state of streamlined steel box girder is studied in this paper. Firstly, the nonlinear aerodynamic self-excited forces of the bridge deck cross section were investigated by CFD dynamic mesh technique and then the nonlinear flutter derivatives were identified on this basis. Secondly, based on the 2-degree-of-freedom (DOF) coupling flutter theory, the torsional amplitude and the nonlinear flutter derivatives were introduced into the traditional direct flutter calculation method, and the original program was improved to the "post-flutter state analysis program" so that it can predict not only the critical flutter velocity but also the movement of the girder in the post-flutter state. Finally, wind tunnel tests were set to verify the method proposed in this paper. The results show that the effect of vertical amplitude on the nonlinear flutter derivatives is negligible, but the torsional amplitude is not; with the increase of wind speed, the post-flutter state of streamlined steel box girder includes four stages, namely, "little amplitude zone", "step amplitude zone", "linearly growing amplitude zone" and "divergence zone"; damping ratio has limited effect on the critical flutter velocity and the steady state response in the post-flutter state; after flutter occurs, the vibration form is a single frequency vibration coupled with torsional and vertical DOF.

• International Journal of Naval Architecture and Ocean Engineering
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• v.11 no.2
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• pp.875-882
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• 2019

The effects of Triangle Groove Strips (TGS) on Vortex-induced Vibration (VIV) suppression of marine riser are numerically investigated using Computational Fluid Dynamics (CFD) method. The range of Reynolds number in simulations is 4.0 × 10⁴ < Re < 1.2 × 10⁵. The two-dimensional unsteady Reynolds-Averaged Navier-Stokes (RANS) equations and Shear Stress Transport (SST) k-ω turbulence model are used to calculate the flow around marine riser. The Newmark-β method is employed for evaluating the structure dynamics of marine riser. The effect of the height ratio (ε) of TGS on VIV suppression is evaluated. The amplitude responses, frequency responses, vortex patterns and the flow around the structures are discussed in detail. With the increase of the height ratio of TGS, the suppression effect of TGS on VIV suppression is improved firstly and then weakened. When ε=0.04, the suppression effect of TGS is the best. Compared with the VIV responses of smooth marine riser, the amplitude ratio is reduced by 38.9%, the peak of the lift coefficient is reduced by 69% and the peak of the drag coefficient is reduced by 40% when Re=6.0 × 10⁴. With the increase of Reynolds number, the suppression effect of TGS on VIV suppression is improved firstly and then weakened. When the Reynolds number is 7.0 × 10⁴, the amplitude ratio can be reduced by 40.1%. As to the large-amplitude vibration cases, the TGS show nice suppression effect on VIV.

• The Journal of Korea Institute of Information, Electronics, and Communication Technology
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• v.12 no.5
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• pp.529-536
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• 2019

In this paper, the signals of the sensor for extracting characteristic parameters of the rotor are collected and the performance of the extraction technique is analyzed. To this end, a vibration test league was developed for conducting model tests to analyze the signal characteristics under normal operation. As a result, it is judged that no change in the measured the raw data amplitude will occur in the acceleration sensor with the unbalanced mass measured from the acceleration sensor. Performing FFT showed a significant increase in amplitude of the rotational frequency of 20 Hz as the unbalanced mass increased. The analysis results according to the change in the unequal mass of the speed sensor also showed a significant increase in the 1X Harmonics component, such as the acceleration sensor. There was no change in the amplitude of the acceleration sensor data when the misalignment occurred, and for the Envelope data, the amplitude of 2X (40 Hz) was increased depending on the degree of misalignment. The velocity sensor at change of misalignment also showed similar results to the acceleration sensor, and the peak was reduced at 600 Hz as the load increased in the frequency spectrum.

• Wind and Structures
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• v.21 no.1
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• pp.85-103
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• 2015

Aerodynamic forces of vortex-induced vibration and galloping are going to be coupled when their onset velocities are close to each other, which will induce the cross-wind amplitudes of the structures increased continuously with ever-increasing wind velocities. The main purpose of the present work is going to propose an empirical formula to predict the response amplitude of VIV-galloping interaction. Firstly, two typical mathematical models for the coupled oscillations, i.e., Tamura & Shimada model and Parkinson & Corless model are comparatively summarized. Then, the key parameter affecting response amplitude is determined through comparative numerical simulations with Tamura & Shimada model. For rectangular cylinders with the side ratio from 0.5 to 2.5, which are actually prone to develop the VIV and galloping induced interaction responses, an empirical amplitude prediction formula is proposed after regression analysis on comprehensively collected experimental data with the predetermined key parameter.

• Proceedings of the Korean Society for Noise and Vibration Engineering Conference
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• 2009.10a
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• pp.617-618
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• 2009

• Proceedings of the Korean Society for Noise and Vibration Engineering Conference
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• 1992.10a
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• pp.92-97
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• 1992

Undesired vibratory motion of an electric connector is controlled in an active method with piezoelectric sensors and actuators. A special electric control box is developed which is suitable for the purpose. Experimental results show that the system can effectively reduce amplitude responses connector to external forces.