Experimental modal substructuring to couple and uncouple substructures with flexible fixtures and multi-point connections

Modal substructuring or component mode synthesis (CMS) has been standard practice for many decades in the analytical realm, yet a number of significant difficulties have been encountered when attempting to combine experimentally derived modal models with analytical ones or when predicting the effect of structural modifications using experimental measurements. This work presents a new method that removes the effects of a flexible fixture from an experimentally obtained modal model. It can be viewed as an extension to the approach where rigid masses are removed from a structure. The approach presented here improves the modal basis of the substructure, so that it can be used to more accurately estimate the modal parameters of the built-up system. New types of constraints are also presented, which constrain the modal degrees of freedom of the substructures, avoiding the need to estimate the connection point displacements and rotations. These constraints together with the use of a flexible fixture enable a new approach for joining structures, especially those with statically indeterminate multi-point connections, such as two circular flanges that are joined by many more bolts than required to enforce compatibility if the substructures were rigid. Fixture design is discussed, one objective of which is to achieve a mass-loaded boundary condition that exercises the substructure at the connection point as it is in the built up system. The proposed approach is demonstrated with two examples using experimental measurements from laboratory systems. The first is a simple problem of joining two beams of differing lengths, while the second consists of a three-dimensional structure comprising a circular plate that is bolted at eight locations to a flange on a cylindrical structure. In both cases frequency response functions predicted by the substructuring methods agree well with those of the actual coupled structures over a significant range of frequencies.     

Effects of Precise FRF Measurements for Frequency Based Substructuring

Impedance modeling is often utilized to generate frequency response functions of an assembled system using modal characteristics of individual components. Sometimes slight inaccuracies are manifested in the characteristics of the components. Although those inaccuracies may seem minor in regards to the component for a traditional modal model, the inaccuracies can be amplified when impedance modeling is performed. Frequency based substructuring involves inversion of the frequency response functions and therefore requires considerable accuracy for the measurement, particularly in the area of anti-resonances.
For this study, some aspects of frequency based substructuring were explored, namely difficulties that might be encountered in experimental testing related to drive point measurements. The majority of studies were performed using analytical models to minimize contamination of data. The study focuses on connection point measurements where the actual location might be inaccessible due to physical constraints such as hardware, holes, etc. Hence the effect of introducing slight inaccuracies to the location of the drive point measurement was studied, along with other aspects related to the accuracy of the critical connection measurement.     

IDENTIFICATION OF MODAL PARAMETERS FROM MEASURED OUTPUT DATA USING VECTOR BACKWARD AUTOREGRESSIVE MODEL

In this paper, a modal identification system that is based on the vector backward autoregressive (VBAR) model has been developed for the identification of natural frequencies, damping ratios and mode shapes of structures from measured output data. The modal identification using forward autoregressive approach has some problems in discriminating the structure modes from spurious modes. On the contrary, the VBAR approach provides a determinate boundary for the separation of system modes from spurious modes, and an eigenvalue filter for the selection of physical modes is existent in the proposed method. For convenience of application, the backward state equation established from VBAR model is transformed into a forward state equation, which is termed as transformed VFAR model in this paper. In addition, the extraction of equivalent system matrix of state equation of motion for structures from the transformed VFAR model has been developed, and then the normal modes can be calculated from the identified equivalent system matrix. Two examples of modal identification are carried out to demonstrate the availability and effectiveness of the proposed backward approach: (1) Numerical modal identification for a three-degree-of-freedom dynamic system with noise level in 20% of r.m.s of measured output data; (2) experimental modal identification of a cantilever beam. Finally, to show the advantage of the proposed VBAR approach on the selection of physical modes, the modal identification by stochastic subspace method was performed. The results from both methods are compared.     

Eliminating the modal truncation problem encountered in frequency responses of viscoelastic systems

Increasing the number of degrees of freedom used in finite element analysis for mechanical and structural systems with viscoelastic damping, the need to consider the modal truncation problem of viscoelastic systems is more than ever before. The higher modes may be unnecessary to obtain in dynamic analysis for engineering applications. For viscoelastic systems, the modal truncation problem may be more frequently encountered since the nonviscous modes are difficult or even impossible to be found accurately even if a small-scaled problem is considered for some eigensolution methods. This study aims at eliminating the influence of the higher modes on the frequency responses of viscoelastically damped systems. A method is presented by making the equilibrium equations of motion into a subspace equation spanned in terms of the columns of a projection basis obtained by considering the use of the contribution of the lower modes and the first two terms of the Neumann expansion of the contribution of the unavailable modes. Finally, three example studies are provided to illustrate the effectiveness of the derived results. It is shown that the proposed method can reduce the modal truncation error significantly.     

Inverse substructure method for model updating of structures

Traditional model updating of large-scale structures is usually time-consuming because the global structural model needs to be repeatedly re-analyzed as a whole to match global measurements. This paper proposes a new substructural model updating method. The modal data measured on the global structure are disassembled to obtain the independent substructural dynamic flexibility matrices under force and displacement compatibility conditions. The method is extended to the case when the measurement is carried out at partial degrees-of-freedom of the structure. The extracted substructural flexibility matrices are then used as references for updating the corresponding substructural models. An orthogonal projector is employed on both the extracted substructural measurements and the substructural models to remove the rigid body modes of the free–free substructures. Compared with the traditional model updating at the global structure level, only the sub-models at the substructural level are re-analyzed in the proposed substructure-based model updating process, resulting in a rapid convergence of optimization. Moreover, only measurement on the local area corresponding to the concerned substructures is required, and those on other components can be avoided. The effectiveness and efficiency of the proposed substructuring method are verified through applications to a laboratory-tested frame structure and a large-scale 600 m tall Guangzhou New TV Tower. The present technique is referred to as the inverse substructuring model updating method as the measured global modal data are disassembled into the substructure level and then the updating is conducted on the substructures only. This differs from the substructuring model updating method previously proposed by the authors, in which the model updating is still conducted in the global level and the numerical global modal data are assembled from those of substructures. That can be referred to as the forward substructuring model updating method.     

Direct mode-shape expansion of a spatially incomplete measured mode by a hybrid-vector modification

A direct estimation method for expanding incomplete experimental mode shapes is presented. The approach adopts a hybrid vector which includes measured data at master degrees of freedom (dofs) and constant values at slave dofs. The constant values are refined by a set of mode-correction factors. Modelling errors between the analytical model and tested structure are also considered by introducing a series of model-correction factors. Initial-guess values of the mode-correction factors are used to decouple the coupled constructed equations, and an iterative technique for solving these equations is proposed. The results from a five-degree-of-freedom mass–spring system indicate that the proposed approach provided a better performance than the commonly used existing expansion methods and can reliably estimate unmeasured components of mode shapes, even in cases with limited modal measurements and severe measurement noise. The performance of the proposed method was also investigated using real measurements from a steel cantilever-beam experiment. Experimental data were measured by 20 accelerometers mounted at the cantilever beam: among these accelerometers, three of these were assumed to be measured, and the others were used to check the estimation accuracy of the proposed method. The results show that the unmeasured components in the mode shapes were properly estimated by implementing the proposed method, even for high-frequency modes.     

An inertia-capacitance beam substructure formulation based on the bond graph method with application to rotating beams

In this paper, a novel inertia-capacitance (IC) beam substructure formulation based on the IC-field presentation from the bond graph method is developed. The IC beam provides a modular, systematic and graphical approach to beam modeling. These features allow the modeler to focus more on the modeling and less on the mathematics. As such, the IC beam is proposed as an alternative to the many existing types of beam models available in the literature. The IC beam is formulated in the center of mass body fixed coordinate system allowing for easy interfacing in a multibody system setting. This floating frame approach is also computationally cheap. Elastic deformations in the IC beam are assumed to be small and described by modal superposition. The formulation couples rigid body and elastic deformations in a nonlinear fashion. The formulation is also compact and efficient. Detailed derivations for a two-dimensional planar IC beam with bending modes are presented. A modal acceleration method based on the decoupling of bending modes is proposed for use in the IC beam. The rotating beam spin-up maneuver problem is solved. The Karnopp-Margolis method is applied to ensure complete integral causality for an efficient numerical system. Geometric substructuring technique is applied to model large deflections. The IC beam is shown to be capable of solving the rotating beam problem accurately and efficiently.     

The mass normalization of the displacement and strain mode shapes in a strain experimental modal analysis using the mass-change strategy

The classic experimental modal analysis (EMA) is a well-known procedure for determining the modal parameters. The less frequently used strain EMA is based on a response measurement using strain sensors. The results of a strain EMA are the modal parameters, where in addition to the displacement mode shapes the strain mode shapes are also identified. The strain EMA can be used for an experimental investigation of a stress–strain distribution without the need to build a dynamical model. It can also be used to determine the modal parameters when, during modal testing, a motion sensor cannot be used and so a strain sensor is used instead. The displacement and strain mode shapes that are determined with the strain EMA are not mass normalized (scaled with respect to the orthogonality properties of the mass-normalized modal matrix), and therefore some dynamical properties of the system cannot be obtained. The mass normalization can be made with the classic EMA, which requires the use of a motion sensor. In this research a new approach to the mass normalization in the strain EMA, without using a motion sensor, is presented. It is based on the recently introduced mass-change structural modification method, which is used for the mass normalization in an operational modal analysis. This method was modified in such a way that it can be used for the mass normalization in the strain EMA. The mass-normalized displacement and strain mode shapes were obtained using a combination of the proposed approach and the strain EMA. The proposed approach was validated on real structures (beam and plate).     

Control of a rotating cantilever beam using a torque actuator and a distributed piezoelectric polymer actuator

A torque actuator and a distributed piezoelectric polymer (PVDF) actuator are utilized for control of a rotating cantilever flexible beam. The torque control contains proportional and derivative (PD) feedback for rigid motion control and a PVDF actuator control for vibration damping. Unlike previous approaches in the literature in which the angular velocity feedback was utilized, in this study we propose to use the linear velocity feedback (L-type) in our controller design for feasible implementation and avoiding modal truncation. The stability of the system with the L-type control has been analyzed, using the concept of a virtual joint model. The advantage of the proposed scheme lies in easy implementation, avoidance of modal truncation, efficient suppression of the dominant mode of vibration, and allowing high-speed motions. Numerical examples demonstrate the effectiveness of the proposed approach.     

Active vibration suppression of a cantilever wing

A method for the active vibration suppression of a cantilever wing is presented. The approach is based on modal control, in which a modal feedback control law relating the motion of the control surfaces to the controlled modes is implemented. Modal displacements and velocities required for feedback are extracted from sensor measurements by means of modal filters. A numerical example is presented.     

A MATHEMATICAL MODEL FOR WIND TURBINE BLADES

A mathematical model for an elastic wind turbine blade mounted on a rigid test stand is derived and compared with experimental results. The linear equations of motion describe small rotations of the test stand, blade lateral deflections and rotation of the chord. Warping, extension and tilt of the cross-sections are slaved to the dependent minimal co-ordinates in order to reduce the number of state variables. Using the principle of virtual work, a procedure is employed which combines the volume discretization of general “solid”, or shell-type finite elements (FE), with the approach of global form functions (stretching over the whole blade length). The equations of motion are solved as an eigenvalue problem and the results are compared with an experimental modal analysis of a 19 m long blade. The computed eigenfrequencies fit well, but the mathematical model underestimates the pitch motion of the blade chord. Parameter studies show the effect of warping. Despite the few degrees of freedom and uncertainties in the model parameters, the mathematical model approximates the measured blade dynamics well.     

Reverberation time measurements in non-diffuse acoustic field by the modal reverberation time

The increasing presence of low frequency sources and the lack of acoustic standard measurement procedures make the extension of reverberation time measurements to frequencies below 100 Hz necessary. In typical ordinary rooms with volumes between 30 m³ and 200 m³ the sound field is non-diffuse at such low frequencies, entailing inhomogeneities in space and frequency domains. Presence of standing waves is also the main cause of bad quality of listening in terms of clarity and rumble effects. Since standard measurements according to ISO 3382 fail to achieve accurate and precise values in third octave bands due to non-linear decays caused by room modes, a new approach based on reverberation time measurements of single resonant frequencies (the modal reverberation time) has been introduced. From background theory, due to the intrinsic relation between modal decays and half bandwidth of resonant frequencies, two measurement methods have been proposed together with proper measurement procedures: a direct method based on interrupted source signal method, and an indirect method based on half bandwidth measurements. With microphones placed at corners of rectangular rooms in order to detect all modes and maximize SNRs, different source signals were tested. Anti-resonant sine waves and sweep signal turned out to be the most suitable for direct and indirect measurement methods respectively. From spatial measurements in an empty rectangular test room, comparison between direct and indirect methods showed good and significant agreements. This is the first experimental validation of the relation between resonant half bandwidth and modal reverberation time. Furthermore, comparisons between means and standard deviations of modal reverberation times and standard reverberation times in third octave bands confirm the inadequacy of standard procedure to get accurate and precise values at low frequencies with respect to the modal approach. Modal reverberation time measurements applied to furnished ordinary rooms confirm previous results in the limit of modal sound field: for highly damped modes due to furniture or acoustic treatment, the indirect method is not applicable due to strong suppression of modes and the consequent deviation of the acoustic field from a non-diffuse condition to a damped modal condition, while standard reverberation times align with direct method values. In the future, further investigations will be necessary in different rooms to improve uncertainty evaluation.     

A modal disturbance estimator for vibration suppression in nonlinear flexible structures

This paper presents a control strategy for the suppression of vibration due to unknown disturbance forces in large, nonlinear flexible structures. The control action proposed, based on the modal approach, consists of two contributions. The first is the well-known Independent Modal-Space Control, which increases system damping and improves its behavior close to the resonance frequencies. The second is a disturbance estimator, which calculates the modal components of the external forces acting on the system and compensates for them using actuator forces. The system modal coordinates, required by both logics, are estimated through a modal state observer.The proposed control logic is tested on a flexible boom. The paper reports the numerical and experimental results both for the linear and nonlinear (large motion) boom configuration.     

Transient vibration phenomena in deep mine hoisting cables. Part 2: Numerical simulation of the dynamic response

A simulation model is presented which investigates the dynamic response of a deep mine hoisting cable system during a winding cycle. The response, namely the lateral motions of the catenary cable and the longitudinal motion of the vertical rope with conveyance is observed on the fast time scale, and the slow time scale is introduced to monitor the variation of slowly varying parameters of the system. The cable equivalent proportional damping parameters, and periodic excitation functions resulting from the cross-over cable motion on the winder drum are identified. Subsequently, the model is solved numerically using parameters of a double-drum multi-rope system. Since the system eigenvalues are widely spread and the problem is of stiff nature, the numerical simulation is conducted using a stiff solver. The results of the simulation demonstrate various transient non-linear resonance phenomena arising in the system during the wind. The nominal ascending cycle simulation results reveal adverse dynamic behaviour of the catenary largely due to the autoparametric interactions between the in- and out-of-plane modes. Principal parametric resonances of the lateral modes also occur, and conditions for autoparametric interactions between the lateral and longitudinal modes arise. Additionally, a transition through a number of primary longitudinal resonances takes place during the wind. The adverse dynamic motions in the system promote large oscillations in the cable tension which must be considered significant with respect to fatigue of the cable. It is noted that a small change in the winding velocity may cause large changes in the dynamic response due to the resonance region shifts. Consequently, the resonance modal interactions can be avoided, to a large extent, if the winding velocity is increased to an appropriate level.     

Experimental identification of viscous damping in linear vibration

This paper is concerned with the experimental evaluation of the performance of viscous damping identification methods in linear vibration theory. Both existing and some new methods proposed by the present authors [A.S. Phani, J. Woodhouse, Viscous damping identification in linear vibration, Journal of Sound and Vibration 303 (3–5) (2007) 475–500] are applied to experimental data measured on two test structures: a coupled three cantilever beam with moderate modal overlap and a free–free beam with low modal overlap. The performance of each method is quantified and compared based on three norms and the best methods are identified. The role of complex modes in damping identification from vibration measurements is critically assessed.     

Analysis of nonlinear steady state vibration of a multi-degree-of-freedom system using component mode synthesis method

In this paper, an analytical approach for nonlinear forced vibration of a multi-degree-of-freedom system is proposed using the component mode synthesis method. The whole system is divided into some components and a nonlinear modal equation of each component is derived using the free-interface vibration modes. The modal equations of all components and the conjunction conditions are solved simultaneously, and then the modal responses of components are derived. Finally, the dynamic responses of the whole system can be obtained. The degrees of freedom of modal equations can be reduced when the lower vibration modes are only adopted in each component. As a numerical example, a nine-degree-of-freedom system is considered, in which all spring have cubic type nonlinearity. As a result, it is shown that when there are no rigid modes in components, the compliance by the proposed method agrees very well with the exact one even if the lower vibration modes of components are only adopted. The other hand, in the case with rigid modes in components, the compliance has a little error compared with the exact result. It is recognized that the method proposed is very effective in the case without rigid modes in components for the actual application.     

ANALYSIS OF THE FLEXURAL VIBRATION OF A THIN-PLATE BOX USING A COMBINATION OF FINITE ELEMENT ANALYSIS AND ANALYTICAL IMPEDANCES

Many practical built-up thin-plate structures, e.g., a modern car body, are essentially assemblies of numerous thin plates joined at their edges. The plates are so thin that they invariably support the weight of the structure and machinery using their substantial in-plane stiffness. Consequently, vibrational power injected into the structure from sources mounted at these stiff points is controlled by high impedance long-wavelength in-plane waves in the plates. As the long in-plane waves propagate around the structure, they impinge upon the numerous structural joints at which short-wavelength flexural waves are generated in adjoining plates. These flexural waves have much lower impedance than the in-plane waves. Hence, the vibration of thin-plate structures excited at their stiff points develops into a mixture of long in-plane waves and short flexural waves. In a previous paper by the same authors, a numerically efficient finite element analysis which accommodated only the long in-plane waves was used to predict the forced response of a six-sided thin-plate box at the stiff points. This paper takes that finite element analysis and, drawing on theory developed in two additional papers by the same authors, couples analytical impedances to it in order to represent the short flexural waves generated at the structural joints. The parameters needed to define these analytical impedances are identified. The vibration of the impedances are used to calculate estimates of the mean-square flexural vibration of the box sides which compare modestly with laboratory measurements. The method should have merit in predicting the vibration of built-up thin-plate structures in the so-called “mid-frequency” region where the modal density of the long waves is too low to allow confident application of statistical energy analysis, yet the modal density of the short flexural waves is too high to allow efficient finite element analysis.     

Mode shape expansion using perturbed force approach

A new approach for expanding incomplete experimental mode shapes is presented which considers the modelling errors in the analytical model and the uncertainties in the vibration modal data measurements. The proposed approach adopts the perturbed force vector that includes the effect of the discrepancy in mass and stiffness between the finite element model and the actual tested dynamic system. From the developed formulations, the perturbed force vector can be obtained from measured modal data and is then used for predicting the unmeasured components of the expanded experimental mode shapes. A special case that does not require the experimental natural frequency in the mode shape expansion process is also discussed. A regularization algorithm based on the Tikhonov solution incorporating the generalized cross-validation method is employed to filter out the influence of noise in measured modal data on the predictions of unmeasured mode components. The accuracy and robustness of the proposed approach is verified with respect to the size of measured data set, sensor location, model deficiency and measurement uncertainty. The results from two numerical examples, a plane frame structure and a thin plate structure, show that the proposed approach has the best performance compared with the commonly used existing expansion methods, and can reliably produce the predictions of mode shape expansion, even in the cases with limited modal data measurements, large modelling errors and severe measurement noise.     

Analytical and experimental investigations of an autoparametric beam structure

This paper discusses theoretical and experimental investigations of vibrations of an autoparametric system composed of two beams with rectangular cross sections. Different flexibilities in the two orthogonal directions are the specific features of the structure. Differential equations of motion and associated boundary conditions, up to third-order approximation, are derived by application of the Hamilton principle of least action. Experimental response of the system, tuned for the 1:4 internal resonance condition, are performed for random and harmonic excitations. The most important vibration modes are extracted from a real mechanical system. It is shown that certain modes in the stiff and flexible directions of both beams may interact, and, intuitively unexpected out-of-plane motion may appear. Preliminary numerical calculations, based on the mathematical model, are also presented.     

Acoustic attenuation analysis of network systems by using impedance matrix method

An impedance matrix method is proposed to predict the acoustic attenuation characteristics of network systems. The system may contain several silencer modules and each module may be composed of complex components such as multiply connected tubes, portions with any-shaped cross-section and dissipative parts. The technique of substructuring is adopted and the system is divided into several substructure modules. Three strategies: boundary element method (BEM), numerical point collocation approach and numerical mode matching technique are introduced and the impedance matrix of each module may be computed by a certain appropriate methodology according to the dimensions and geometry of the substructure. Impedance matrix synthesis is employed to obtain the resultant impedance matrix and then transmission loss may be calculated. All the calculation results are verified by experimental measurements and 3-D BEM predictions.