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Vibration of pump shaft is damped at impeller--where vibration-induced deflections are greatest--by shroud and seal. Damping reduces vibrational motion of shaft at bearings and load shaft places on them. Flow through clearance channel absorbs vibration energy.
Vibration-prone parts in fast-flowing liquids or gases damped by simple provision: drill or cast small holes in them and fill holes with particles. Particles absorb vibration energy without obstructing flow. Damping holes add little to cost of manufacturing parts. Reduces masses of parts because masses of inserted particles less than materials removed to make holes.
Non-Obstructive Particle Damping (NOPD) technology is a passive vibration damping approach whereby metallic or non-metallic particles in spherical or irregular shapes, of heavy or light consistency, and even liquid particles are placed inside cavities or attached to structures by an appropriate means at strategic locations, to absorb vibration energy. The objective of the work described herein is the development of a design optimization procedure and discussion of test results for such a NOPD treatment on honeycomb (HC) composite structures, based on finite element modeling (FEM) analyses, optimization and tests. Modeling and predictions were performed and tests were carried out to correlate the test data with the FEM. The optimization procedure consisted of defining a global objective function, using finite difference methods, to determine the optimal values of the design variables through quadratic linear programming. The optimization process was carried out by targeting the highest dynamic displacements of several vibration modes of the structure and finding an optimal treatment configuration that will minimize them. An optimal design was thus derived and laboratory tests were conducted to evaluate its performance under different vibration environments. Three honeycomb composite beams, with Nomex core and aluminum face sheets, empty (untreated), uniformly treated with NOPD, and optimally treated with NOPD, according to the analytically predicted optimal design configuration, were tested in the laboratory. It is shown that the beam with optimal treatment has the lowest response amplitude. Described below are results of modal vibration tests and FEM analyses from predictions of the modal characteristics of honeycomb beams under zero, 50% uniform treatment and an optimal NOPD treatment design configuration and verification with test data.
The vibration of an experimental flexible space truss is controlled with internal control forces produced by several proof mass actuators. Four candidate control law strategies are evaluated in terms of performance and robustness. These control laws are experimentally implemented on a quasi free-free planar truss. Sensor and actuator dynamics are included in the model such that the final closed loop is self-equilibrated. The first two control laws considered are based on direct output feedback and consist of tuning the actuator feedback gains to the lowest mode intended to receive damping. The first method feeds back only the position and velocity of the proof mass relative to the structure; this results in a traditional vibration absorber. The second method includes the same feedback paths as the first plus feedback of the local structural velocity. The third law is designed with robust H infinity control theory. The fourth strategy is an active implementation of a viscous damper, where the actuator is configured to provide a bending moment at two points on the structure. The vibration control system is then evaluated in terms of how it would benefit the space structure's position control system.
Bronze wool pads, impregnated with silicon rubber, meet the requirement for a thermally conductive, shock and vibration absorbing material. They serve as spacers in equipment mounting and are resistant to high temperatures.
A steady state analysis of the shaft and the bearing housing motion was made by assuming synchronous precession of the system. The conditions under which the support system would act as a dynamic vibration absorber at the rotor critical speed were studied; plots of the rotor and support amplitudes, phase angles, and forces transmitted were evaluated by the computer, and the performance curves were automatically plotted by a CalComp plotter unit. Curves are presented on the optimization of the support housing characteristics to attenuate the rotor unbalance response over the entire rotor speed range. The complete transient motion including rotor unbalance was examined by integrating the equations of motion numerically using a modified fourth order Runge-Kutta procedure, and the resulting whirl orbits were plotted by the CalComp plotter unit. The results of the transient analysis are discussed with regards to the design optimization procedure derived from the steady-state analysis.
The distribution pattern of the amplitudes of vibrations, measured at various points on the surface of a shell, stimulated by a point source of harmonic force in the sound frequency range was studied experimentally. This pattern permits a valid spacing estimate to be carried out of the average structural characteristics of the vibrations, which are used for both estimation of the vibration excitability of the structure and for estimation of the effectiveness of use of a vibration-absorbing covering.
The dynamic unabalance response and transient motion of the single mass Jeffcott rotor in elastic bearings mounted on damped, flexible supports are discussed. A steady state analysis of the shaft and the bearing housing motion was made by assuming synchronous precession of the system. The conditions under which the support system would act as a dynamic vibration absorber at the rotor critical speed were studied. Plots of the rotor and support amplitudes, phase angles, and forces transmitted were evaluated by the computer and the performance curves were plotted by an automatic plotter unit. Curves are presented on the optimization of the support housing characteristics of attenuate the rotor synchronous unbalance response.
A coupled rotor/bifilar/airframe analysis was developed and utilized to study the dynamic characteristics of the centrifugally tuned, rotor-hub-mounted, bifilar vibration absorber. The analysis contains the major components that impact the bifilar absorber performance, namely, an elastic rotor with hover aerodynamics, a flexible fuselage, and nonlinear individual degrees of freedom for each bifilar mass. Airspeed, rotor speed, bifilar mass and tuning variations are considered. The performance of the bifilar absorber is shown to be a function of its basic parameters: dynamic mass, damping and tuning, as well as the impedance of the rotor hub. The effect of the dissimilar responses of the individual bifilar masses which are caused by tolerance induced mass, damping and tuning variations is also examined.
Vibration absorbers are introduced into an asymmetric configuration of thin cylinders and tori enclosing an acoustic medium. The absorbers consist of thin axial strips bonded to the cylinder with a thin viscoelastic layer. The constrained layer dissipates the energy of relative motions between strip and cylinder. The absorber is most effective on response modes with two or more circumferential waves. The use of transfer matrices is extended to the coupled cylinder-absorber system.
A flexible structure was modeled and actively controlled by using a single space realizable linear proof mass actuator. The NASA/UVA/UB actuator was attached to a flexible planar truss structure at an optimal location and it was considered as both passive and active device. The placement of the actuator was specified by examining the eigenvalues of the modified model that included the actuator dynamics, and the frequency response functions of the modified system. The electronic stiffness of the actuator was specified, such that the proof mass actuator system was tuned to the fourth structural mode of the truss by using traditional vibration absorber design. The active control law was limited to velocity feedback by integrating of the signals of two accelerometers attached to the structure. The two lower modes of the closed-loop structure were placed further in the LHS of the complex plane. The theoretically predicted passive and active control law was experimentally verified.
This paper presents a survey of the state-of-the-art in the field of structural optimization when applied to vibration reduction of helicopters in forward flight with aeroelastic and multidisciplinary constraints. It emphasizes the application of the modern approach where the optimization is formulated as a mathematical programming problem, the objective function consists of the vibration levels at the hub, and behavior constraints are imposed on the blade frequencies and aeroelastic stability margins, as well as on a number of additional ingredients that can have a significant effect on the overall performance and flight mechanics of the helicopter. It is shown that the integrated multidisciplinary optimization of rotorcraft offers the potential for substantial improvements, which can be achieved by careful preliminary design and analysis without requiring additional hardware such as rotor vibration absorbers of isolation systems.
This paper presents a review of the state-of-the-art in the field of structural optimization when applied to vibration reduction of helicopters in forward flight with aeroelastic and multidisciplinary constraints. It emphasizes the application of the modern approach where the optimization is formulated as a mathematical programming problem and the objective function consists of the vibration levels at the hub and behavior constraints are imposed on the blade frequencies, aeroelastic stability margins as well as on a number of additional ingredients which can have a significant effect on the overall performance and flight mechanics of the helicopter. It is shown that the integrated multidisciplinary optimization of rotorcraft offers the potential for substantial improvements which can be achieved by careful preliminary design and analysis without requiring additional hardware such as rotor vibration absorbers or isolation systems.
The Second Generation Comprehensive Helicopter Analysis System employs the present formulation of the general-purpose nonlinear rigid body mass finite element, which represents the hub masses, blade tip masses, and pendulum vibration absorbers. The rigid body mass element has six degrees of freedom, and accounts for gravitational as well as dynamic effects. A consequence of deriving the element's equations from various physical principles is that, prior to the transformation which couples the rigid body mass element to the rotor blade finite element, the forces obtained for each element are fundamentally different; this is true notwithstanding the degrees-of-freedom of each element are parameterized using the same coordinates.
The properties of virtual passive controllers, i.e., active vibration absorbers, and of the closed-loop systems achievable by means of such compensation are analyzed. It is shown that these controllers introduce additional transmission zeros into the closed-loop system. Therefore, they provide an extra degree of design freedom over that provided by state feedback, which cannot alter the zeros in any way.
A device for mounting a data transmission apparatus to a rotating, tapered, and instrumented shaft is provided. This device permits attachment without interfering with shaft rotation or the accuracy of data output, and prevents both radial and axial slippage of the data transmission apparatus. The mounting device consists of a sleeve assembly which is attached to the shaft by means of clamps that are situated at some distance removed from the instrumented area of the shaft. The data transmission device is secured to the sleeve such that the entire assembly rotates with the shaft. Shim adjustments between sleeve sections assure that a minimum compressive load is transferred to the instrumented area of the shaft and a rubber lining is affixed to a large portion of the interior surface of the sleeve to absorb vibration.
Passive damping of structural dynamics using piezoceramic electromechanical energy conversion and passive electrical networks is a relatively recent concept with little implementation experience base. This report describes an implementation case study, starting from conceptual design and technique selection, through detailed component design and testing to simulation on the structure to be damped. About 0.5kg. of piezoelectric material was employed to damp the ASTREX testbed, a 500kg structure. Emphasis was placed upon designing the damping to enable high bandwidth robust feedback control. Resistive piezoelectric shunting provided the necessary broadband damping. The piezoelectric element was incorporated into a mechanically-tuned vibration absorber in order to concentrate damping into the 30 to 40 Hz frequency modes at the rolloff region of the proposed compensator. A prototype of a steel flex-tensional motion amplification device was built and tested. The effective stiffness and damping of the flex-tensional device was experimentally verified. When six of these effective springs are placed in an orthogonal configuration, strain energy is absorbed from all six degrees of freedom of a 90kg. mass. A NASTRAN finite element model of the testbed was modified to include the six-spring damping system. An analytical model was developed for the spring in order to see how the flex-tensional device and piezoelectric dimensions effect the critical stress and strain energy distribution throughout the component. Simulation of the testbed demonstrated the damping levels achievable in the completed system.
A Timoshenko beam continuum model is developed to determine the transient response of a beam-like latticed large space structure, subjected to a unit impulse. It is demonstrated that an increase in diagonal stiffness, on account of the stiffness of the vertical girder, leads to a rise in the transverse shear rigidity. This results in higher natural frequencies and reduction in peak displacement. In addition, in an asymmetrical truss configuration, coupling between the extensional and shear modes raises the maximum peak displacement compared with that obtained for a symmetrical truss. The model is modified to investigate the introduction of passive damping in the form of several dynamic vibration absorbers. A unique absorber parameter optimization procedure, based on the classical steady-state criteria, is used to tune the absorbers, having a total allocated mass budget of 10%, to the first two frequencies of the latticed structure. It is found that inclusion of transverse shear rigidity as a design parameter in damping augmentation studies reduces settling time for predetermined maximum peak displacements.