Means of Protection Against Vibrations of Load Lifting Devices Working With Vibration Exciters

You can access other parts of this book here: Vibro-Engineering and the Technology of Piling and Boring Work.

Editor’s Note

This is a topic which was of great interest to Vulcan in the development of its vibratory pile drivers and extractors.

As was the case with the Russians, all earlier vibratory drivers (including the Uraga units Vulcan distributed) used steel coil springs in the suspension. These were expensive, subject to breakage and had little internal damping, something the Russians attempted to remedy.

MKT introduced the “Lord” type rubber shear fenders in its suspensions. When Vulcan produced its first vibratory in 1984, it opted to use the “Morse” shear fenders that ICE used. These were very soft and, although they were a challenge to run hoses to and from the exciter, provided excellent damping. Unfortunately four years later a change in manufacturing made them vulnerable to separation of the rubber spring from the metal plate mount. When Vulcan came out with the “A” models in 1991, they opted to use the “Lord” type fenders. Although the damping isn’t quite what it is with the Morse fenders, it was acceptable.

The fact that we have aftermarket solutions to this problem shows that we have not “arrived” on this issue.

Vulcan also developed a hydraulic braking system for rapid stops of the vibratory hammers to mitigate resonant conditions during stopping of the eccentrics. Although Vulcan never went through the extensive testing that the Russians performed, the results were similar in nature.

For additional information on the development of Vulcan’s vibratory hammer line from a theoretical standpoint, including the stiffness of the suspension and its effect on the power requirements of the machine, it is discussed in detail in the paper Development of a Parameter Selection Method for Vibratory Pile Driver Design with Hammer Suspension.

When extracting sheet piles and other elements using a vibratory driver, the metal structures of the crane are subjected to simultaneous static and dynamic loads, which causes reasonable concerns for the strength and durability of cranes. In order to develop vibration isolation tools for extracting the sheet pile with a V-401 vibrodriver, VNIIGS together with VNIIstroidormash conducted experimental studies to determine dynamic loads in the crane metal structures. These loads can be subdivided into those arising in transient modes of operation of the vibratory mechanism – during its starting and stopping, and during the steady-state (operating) mode of operation in the process of extraction.

Currently, there are the following main directions for reducing the vibration impact:

introduction of elastic shock absorbers into the system;
introduction of attenuation into the system;
dynamic vibration damping;
increasing the rate of passage of the system through the resonance.

The above methods can be used separately or together.

The introduction of elastic shock absorbers into the system is associated, as a rule, with the operation of vibration machines in the steady state, when due to the rational choice of the rigidity of the elastic elements, the vibration effect on the object being isolated is minimized.

The spring dampers in the system are adjusted to the operating mode of the vibration units, and the value of the vibration isolation coefficient, which is the ratio of the dynamic force transmitted to the insulated object, to the amplitude value of the driving force is set from 0.2 and below, i.e., the shock absorbers should provide a sufficiently low intrinsic frequency and low attenuation. It should be noted that the possibility of further reduction of the vibration isolation coefficient in the existing designs of shock absorbers is limited due to the use of cylindrical springs, which reduce the useful lifting height of the crane hook due to significant geometric dimensions.

At present, the improvement of spring shock absorbers is proceeding along the path of using springs with an adjustable characteristic, which provides, with resonant vibrations in the starting and stopping mode of the vibrator, a certain decrease in its displacements and the force transmitted to the isolated object.

In addition, rubber-cord shells are widely used as elastic elements, with the help of which it is possible to reduce the natural frequency of the shock absorber.

Direct damping of oscillations is associated with the conversion of part of the mechanical energy of oscillations into dissipated thermal energy using various dampers directly connected between the oscillating mass and the object being isolated. Such a dissipation of energy always takes place also because of the imperfection of the elastic elements.

Viscous friction makes it possible to reduce the amplitude of oscillations when passing through resonance, but at the same time, the vibration isolation coefficient in the operating mode is somewhat reduced, since the dynamic force is transmitted to the supporting structure not only through the elastic element, but also through the viscous friction damper. In this case, energy is dissipated in the damper both when passing through resonance, mainly in the operating mode, and this consumes up to 20% of the energy necessary for the operation of the machine.

The main advantage of the direct inclusion of damping elements is simplicity, and the main disadvantage is the decrease in the vibration isolation coefficient and relatively high energy losses in the operating mode.

Dynamic vibration dampers are usually a mass, either pressed by a spring to the body, the vibrations of which must be damped, or installed with a gap relative to this body. With significant amplitudes developing in the resonant mode, the inertial force acting on the mass of the dynamic absorber begins to exceed the elastic force of the spring, as a result of which collisions of the mass with the oscillating body occur, leading to a decrease in the amplitude of its oscillations.

The considered types of absorbers operate in a narrow frequency range, which is a great difficulty if it is necessary to isolate vibration systems with several resonances, therefore, the main proposals for the modernization of shock absorbers are related to expanding the operating frequency range. At the moment of starting or stopping the vibrator, the amplitudes of oscillations that develop when passing through resonance may not reach a large value even with a very small attenuation of the system. This will happen if the frequency of the driving force changes so quickly that the installation leaves the resonant zone before the system has time to swing.

From the theory of the question of passage through resonance (A. M. Kats, 1947; B. G. Korenev, 1957; I. S. Sheinin, 1961) it is known that the maximum amplitude $A_{max}$ is directly proportional to the natural frequency $\theta_0$ and inversely proportional to indicator of the degree of speed $\lambda$ , with which the frequency of the driving force changes.

In the simplest case of passing through the resonance of a vibrator with a constant eccentric moment, $\lambda$ has an exponent equal to 0.5, and the maximum amplitude is directly proportional to the value $\frac{\theta_0}{\sqrt{\lambda}}$ .

Mechanical and electrical methods of braking are known, one of which is countercurrent braking. So, for example, the Mekhanobr Institute proposed a device for electric braking using a typical time relay. The developed device was tested on a laboratory screen. Studies have shown that as a result of the application of electric braking, the resonant amplitude of vibrations decreased by a factor of 2.5.

The circuit for braking the engine with countercurrent was also developed at TsNIISK (B. G. Korenev, N. A. Pikulev, I. S. Sheinin, 1969,) where instead of a time relay, an angular velocity direction relay was introduced. Experimental studies of the described scheme have shown that during electric braking of vibrosets, the maximum amplitude of the stopping resonance becomes even less than with the starting resonance. In this case, the overload of the electrical network by countercurrent does not exceed its overload by the starting current. Another way to increase the speed of passage through resonance is to automatically change the natural frequencies of the system.

To reduce the resonant amplitudes during the start-up of vibrators, a method has been developed, consisting in the fact that when the frequency of the driving force of the vibrator reaches a value close to the frequency of the natural vibrations of the machine, the vibrator motor is turned off. At the same time, the energy of the rotational movement of the vibrator increases due to the energy of the oscillatory movement (and the amplitude of the oscillations decreases) and, accordingly, the angular velocity of rotation of the eccentrics increases (it passes through the resonant mode,) after which the electric motor is switched on again, providing a further increase in the frequency of the driving force, but already in the range-of-resonance region.

The above description of the main directions of development of vibration isolation means shows that in order to successfully solve the problems of reducing dynamic loads in the metal structures of load-lifting cranes working with vibratory drivers, it is necessary to use complex methods, ensuring the reduction of these loads both in steady state and in transient modes.

Research at the laboratory stand of VNIIGS found that when using a self-propelled jib crane-vibration pile driver complex in vibratory extraction of a metal sheet pile, in order to reduce the number of loading cycles and the amplitude value of dynamic loads transmitted to the metal structure of the crane from the vibration mechanism, it is necessary to:

ensure a high speed of passage through the resonance when the vibration exciter runs out due to the use of dynamic braking of the electric motor; and
reduce the total (static and dynamic) loads on the crane load-handling device in the steady-state operation of the vibration exciter to values not exceeding 80-90% of the crane’s rated load capacity at a given reach by including shock-absorbing devices in the system.

In addition, studies have shown that with an elastic shock absorber, the vibration exciter in the starting mode for the existing characteristics of the system has a dynamic effect on the load gripping device, which not only does not exceed, but even somewhat less than in the steady state.

Based on the results of research on a laboratory stand, a vibration isolation system for self-propelled jib cranes was developed, including electrical equipment with a mounted dynamic braking system for the V-401 vibratory driver, and an additional shock absorber installed between the vibration driver and the crane hook.

Experimental work was carried out at one of the objects during the vibration extraction of the sheet pile, which included determining the magnitude and nature of the dynamic loads that occur in the metal structure of the jib crane when working with the V-401 vibratory driver, as well as checking the effectiveness of the additional shock absorber and dynamic braking system.

Sheet piling Larsen-IV and Larsen-V 10-12 m long were extracted with a V-401 vibratory driver with a driving force frequency of 16 Hz and a static mass moment of unbalances of 1000 kg – cm. An MKG-20 crane was used as a load-lifting mechanism. A tensometric dipamometer was installed between the load gripping body of the crane and the suspension of the vibratory driver to record the force effects on the crane during the extraction of the sheet pile. To measure the stresses in the metal structure of the boom in two of its sections at a distance of 0.2l and 0.55l (l = the full length of the boom,) strain gauges were fixed on the upper and lower chords. The same sensors were installed in two sections of the bipedal rack. The results of the experiments were recorded on an oscillogram.

The experimental work consisted of two stages: without the use of systems that reduce the dynamic loads transferred to the crane, and with the use of an additional shock absorber and dynamic braking of the vibrator electric motor, carried out by supplying direct current to the motor stator, using a specially designed electrical circuit.

The methodology for conducting research under production conditions consisted in determining the loads on the hook and the stresses in two sections of the boom and the two-legged crane stand during starting, steady state and stopping of the vibratory driver at the beginning and after extraction by 1.4 and 10 m.

The conducted studies have established that in the current mode of operation of the vibratory driver, the frequency range of dynamic loads transmitted to the crane hook is from 4 to 80 Hz. Therefore, when determining the steady-state operating mode of the crane-vibrator system, it is necessary to focus not on the constancy of frequencies that change from cycle to cycle, but on the constancy of the amplitude values of the dynamic loads transmitted to the crane, and design the means of vibration isolation accordingly.

As a result of processing the oscillograms, the stresses in the metal structure of the boom and the two-legged crane stand were determined, which occur under the action of both static extraction force and dynamic loads transmitted to the crane boom during the operation of the vibration exciter. It has been established that the stresses in the boom arising from the static extraction force do not exceed 50-60 MPa. The stresses from the action of dynamic loads on the boom are up to 40% at start-up, and in operating mode – up to 15% of the stresses arising under the action of static force.

When the vibration exciter runs out, the dynamic component of the stresses in the boom is two times or more higher than the stresses from the static force of the crane. The introduction of a dynamic braking system during the stopping of the vibration exciter electric motor makes it possible to reduce the stresses from the dynamic load by 2.5-3 times. The change in the ratio of stress in one of the boom chords from dynamic loads $\sigma_a$ to stresses from static force $\sigma_{st}$ , when the vibration exciter runs out, is shown in Figure 55.

Figure 55. Change in the ratio of stresses in the boom from dynamic $\sigma_a$ AND static $\sigma_{st}$ loads during the run-out of the vibration exciter 1 – with the use of dynamic braking; 2 – without braking

Figure 55. Change in the ratio of stresses in the boom from dynamic $\sigma_a$ AND static $\sigma_{st}$ loads during the run-out of the vibration exciter 1 – with the use of dynamic braking; 2 – without braking

During vibration extraction of the tongue, the sum of static and dynamic stresses in the bipedal stand was no more than 35 MPa, while the dynamic stresses did not exceed 10% of the static ones. Thus, the measurement of stresses in the bipedal stance also showed. that they are not dangerous for the crane.

The measured levels of noise and vibration at the driver’s workplace during the operation of the MKG-20 crane from the electric drive with the V-401 vibratory driver equipped with a dynamic braking system do not exceed the permissible ones in accordance with GOST 12.1.003-83 and GOST 12.1.012-78.

The developed vibration isolation system does not require changes in the design of the crane and the vibrodriver, it is a separate device that can be equipped with each V-401 vibrodriver.