Home Home
Bor Tsuen Wang, National Pingtung University of Science and Technology (Taiwan)
Rong Liang Chen, National Pingtung University of Science and Technology (Taiwan)
This paper presents the system identification technique by applying PZT actuators and PVDF sensors. An array of finite-length PVDF films is equally spaced and distributed over a cantilever beam acting as sensors. Two finite-length PZT patches are bonded symmetrically and excited 180 degrees out of phase for pure bending excitation. The theoretical base of modal analysis for system identification of piezoceramic beams is presented to derive the frequency response functions (FRFs) in conventional modal format. The physical interpretation of PZT and PVDF mode shape functions are characterized respectively. The FRFs between the PZT actuators and PVDF sensors are experimentally measured. By the operation of a series of FRFs and the application of modal parameter extraction method, the modal parameters, including system natural frequencies, mode shapes and modal damping ratios, can be obtained. Both theoretical and experimental approaches agree very well. The system information can be used for control applications and also be applicable to structural failure diagnosis.
sv970541.pdf (Scanned)
TOPMichael S. Sedov, Nizhny Novgorod Academy of Architecture and Civil Eng. (Russia)
A method aIlowing to solve the problem of wave nature of simple and complicate figures of sound oscillations of thin rectangular plates is proposed. The base of the method is presentation of transformation of wave energy. Phase raiding of propagation of free bend waves as a wave motion inside the plate with minimal energy losses has information about the spectra density of own oscillations and about the process of generation of figures at each own frequency. For instance, it was derived that in a plate with the free edges the figures with recliner and curve centers lines are formed with uniform and nonuniform bend waves. Sequence and image of Khladni figures are obtained analytically. In particular it was proved that round or cross center lines which were traditionally considered as simple forms are complicated forms. Experimental investigations confirm this statement.
sv970520.pdf (Scanned) TOP
Thomas G. Carne, Sandia National Laboratories (U.S.A.)
Clark R. Dohrmann, Sandia National Laboratories (U.S.A.)
During a modal test, the structure must be supported in some manner, in order to test it. If a model of the structure has been developed and is to be reconciled with the test data, then the support conditions must either be included in the model or accounted for. For example, they could be ignored in a simulated free test, if they do not affect the structure significantly. Frequently, a precise determination of the actual support conditions is not performed and there is uncertainly in the conditions, particularly in the damping that these condition attach to the structure. This study examines the effects of support conditions on the measured modal parameters, but more significantly, examines how uncertainties in these support conditions propagate into uncertainties in the both the measured modal frequency and damping.
sv970505.pdf (Scanned) TOP
M.H.A. Janssens, TNO Institute for Applied Physics TU-Delft (The Netherlands)
C.M. Langeveld, TNO Institute for Applied Physics TU-Delft (The Netherlands)
Jan W. Verheij, TNO Institute for Applied Physics TU-Delft (The Netherlands)
The 'pseudo-forces method' can characterize the strength of compact structure-borne sound sources. It allows for multi-dimensional behaviour of the source, whilst the measurement effort remains limited. In previous work the validity and practicability of the method was confirmed. The indirect nature of this method, however, introduces some (arbitrary) choices in substitution force location and in the transfer matrix analysis. This complicates the comparison of different experiments. In the current work an effort is made to introduce modal information of the free source, which then allows for a unique interpretation of the results and comparison of experiments using different force positions. At the same time the measurement effort needed for the characterization is kept limited. The validity and practicability of this new technique is tested by experiments on a compact source. This expansion of the method broadens the application range of the method and facilitates the translation of the results of a characterization experiment to the low noise design practice.
sv970380.pdf (Scanned) TOP
Masaaki Okuma, Tokyo Institute of Technology (Japan)
Tatsuya Oho, Tokyo Institute of Technology (Japan)
In this paper, the authors present a new experimental spatial matrix identification method that they have been developing. The method is to identify a set of the mass, damping and stiffness matrices that can represent the dynamic characteristics of an objective structure from experimental FRFs. The theory of the method is explained at first. Then, the result of an identification of a basic frame structure, which is made of L-shaped cross-sectional steel components, under the free-free boundary condition is presented. Both bending vibration modes and torsional vibration modes are located in the frequency range of the identification. The dynamic characterisitics of the specimen under a different boundary condition are estimated from the previously identified set of spatial matrices, and compared with experimental results to verify the practical validity and usefulness of the method.
sv970320.pdf (Scanned) TOP
Raouf A Ibrahim, Wayne State University (U.S.A.)
M. Hijawi, Wayne State University (U.S.A.)
The purpose of this study is to understand the main differences between deterministic and random response characteristics of a cantilever beam in the neighborhood of combination parametric resonance. The beam orientation with respect to the excitation is made in such a way that the two modes are only coupled through parametric excitation and nonlinear inertia forces. This means that both generalized and normal coordinates are the same. For sinusoidal parametric excitation the response will be determined by using the method of multiple scales, and also measured experimentally, in the neighborhood of the combination parametricresonance (Omega) = (omega)_(mu) \pm (omega)_\phi, where (Omega) is the excitation frequency and (omega)_(mu) and (omega)_\phi are the bending and torsion first mode natural frequencies, respectively, For the random excitation case the response will be predicted using Monte Carlo simulation and measurement experimentally. The center frequency of the band limited excitation is adjusted to be close to the sum of the bending and torsion mode frequencies.Under deterministic excitation, the dependence of the response amplitude on the excitation level reveals three distinct regions over which linear behavior, or jump phenomena, or energy transfer can take place. Under random excitation, the system may experience a single response, two possible responses or nonstationary responses depending on the excitation level. The response may also be Gaussian or non-Gaussian depending on the excitation level as well. The power spectra exhibit nonlinear interaction between the torsion and bending modes. Experimentally, it is possible to obtain two different responses for the same excitation level by providing some perturbation to the system.
sv970296.pdf (Scanned) TOP
Ladislav Starek, Slovak Technical University
Daniel J. Inman, Virginia Polytechnic Institute & State University (U.S.A.)
Milos Musil, Slovak Technical University
As technological demands push the performance of mechanical structures and machines, more accurate models are required. A problem of increasing importance in mechanical industries is that of producing analytical models which agree with experimental data. Several incompatibilities exist between analytical models and experimentally obtained data. For instance consider the case of finite element analysis (FEA) modelling compared with experimental modal analysis (EMA) data. This case accounts for the majority of activity in vibration modelling used in industry. In this situation the analytical model is characterised by a large number of degrees of freedom, ad hoc damping mechanism and real eigenvectors. The FEM model produces a mass, damping and stiffness matrix which is numerically solved for modal data consisting of natural frequencies, mode shapes and damping ratios. Common practice in industry is to compare this analytically generated modal data with natural frequencies, mode shapes and damping ratios obtained from EMA, The EMA data is characterised by a small number of natural frequencies, incomplete and complex mode shapes and non proportional damping. It is very common in practice for this experimentally obtained modal data to be in disagreement with the analytically derived modal data. The point of view taken is that the analytical model is in error and must be refined or corrected based on experimental data. The paper will deal with updating procedures: - for complete modal and spectral data - incomplete modal and spectral data with complete mode shapes, incomplete modal and spectral data with incomplete mode shapes. Relations for computing design parameters m, b, k from coefficient matrices M, B, K will be given too. Examples will illustrate the theory.
sv970272.pdf (Scanned) TOP