WO2005056201A1 - Vibrateur de battage de pieux pour des articles battus - Google Patents

Vibrateur de battage de pieux pour des articles battus Download PDF

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Publication number
WO2005056201A1
WO2005056201A1 PCT/EP2004/014145 EP2004014145W WO2005056201A1 WO 2005056201 A1 WO2005056201 A1 WO 2005056201A1 EP 2004014145 W EP2004014145 W EP 2004014145W WO 2005056201 A1 WO2005056201 A1 WO 2005056201A1
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WIPO (PCT)
Prior art keywords
mass
excitation
frequency
vibrator according
working
Prior art date
Application number
PCT/EP2004/014145
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German (de)
English (en)
Inventor
Hubert Bald
Original Assignee
GEDIB Ingenieurbüro und Innovationsberatung GmbH
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Publication of WO2005056201A1 publication Critical patent/WO2005056201A1/fr

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Classifications

    • B—PERFORMING OPERATIONS; TRANSPORTING
    • B06—GENERATING OR TRANSMITTING MECHANICAL VIBRATIONS IN GENERAL
    • B06B—METHODS OR APPARATUS FOR GENERATING OR TRANSMITTING MECHANICAL VIBRATIONS OF INFRASONIC, SONIC, OR ULTRASONIC FREQUENCY, e.g. FOR PERFORMING MECHANICAL WORK IN GENERAL
    • B06B1/00—Methods or apparatus for generating mechanical vibrations of infrasonic, sonic, or ultrasonic frequency
    • B06B1/18—Methods or apparatus for generating mechanical vibrations of infrasonic, sonic, or ultrasonic frequency wherein the vibrator is actuated by pressure fluid
    • B06B1/183—Methods or apparatus for generating mechanical vibrations of infrasonic, sonic, or ultrasonic frequency wherein the vibrator is actuated by pressure fluid operating with reciprocating masses
    • E—FIXED CONSTRUCTIONS
    • E02—HYDRAULIC ENGINEERING; FOUNDATIONS; SOIL SHIFTING
    • E02D—FOUNDATIONS; EXCAVATIONS; EMBANKMENTS; UNDERGROUND OR UNDERWATER STRUCTURES
    • E02D7/00—Methods or apparatus for placing sheet pile bulkheads, piles, mouldpipes, or other moulds
    • E02D7/18—Placing by vibrating

Definitions

  • the invention relates to a ramming vibrator for ramming material, such as sheet piles or piles, which are driven into or pulled out of the ground by oscillating movements of the ramming vibrator.
  • ramming vibrator for ramming material, such as sheet piles or piles, which are driven into or pulled out of the ground by oscillating movements of the ramming vibrator.
  • the predominantly vertical vibratory movements are derived from the centrifugal forces of rotating imbalances, which are synchronized in such a way that the horizontal centrifugal force components cancel each other and the vertical centrifugal force components add up.
  • Gear transmissions are predominantly used for the synchronization of the unbalances and possibly also for their drive.
  • Modern unbalance vibrators of this type are required to be adjustable or adjustable with regard to their vibration path amplitude while the unbalance is rotating, which also includes the vibration path amplitude with the value zero.
  • these unbalanced ram vibrators can also be operated in such a way that (at existing contact of the pile material with the soil) during the acceleration of the unbalance rotation movement from zero speed to no natural frequencies excitable in the soil below the desired operating frequency. This takes place in the case of the amplitude-controllable unbalance ram vibrators in that the oscillation amplitude is kept at zero while the unbalance rotary motions are ramping up from zero until the (high) operating frequency is reached and a predetermined amplitude is only set after the operating frequency has been reached.
  • a ram vibrator is described in EP 1 167 632 A2, which has a linear dual-mass oscillator that can be excited by at least one excitation actuator designed as a hydraulic cylinder.
  • This comprises a cantilever mass in the form of a support frame, a working mass in the form of an outer frame with a holder for the pile and a spring system coupling both masses, the two masses oscillating essentially in phase and in opposite directions.
  • the piston of the hydraulic cylinder is acted upon by a control rotor with rotary slides for alternating reciprocating movement of the working mass.
  • the work to be performed is the sum of the work performed and the vibration loss required to force the vibrations without performing any work.
  • the power loss is from Difference between the natural frequency of the dual mass oscillator and the excitation frequency, which can be changed here by changing the speed of the control rotor.
  • the vibration loss power is inevitably relatively large in this ramming vibrator, since work must be carried out in frequency ranges which ensure a controllable vibration path amplitude.
  • the object of the invention is therefore to provide a ram vibrator which has a lower power requirement. This object is achieved in accordance with the features of claim 1.
  • the invention therefore relates to a ramming vibrator for ramming material with a linear dual-mass oscillator which can be excited by at least one excitation actuator and which has a cantilever mass, a working mass and a spring system coupling both masses, the two masses oscillating essentially in phase and in opposite directions and one of the Mass carries a holder for the pile, a control circuit for the excitation power, which comprises a controller, the at least one excitation actuator and a sensor device for direct or indirect measurement of the vibration path amplitude, is provided with which the vibration path amplitude in the frequency range adjacent to the main natural frequency of the dual mass vibrator a predetermined value can be regulated, the excitation frequency being predeterminable in the frequency range adjacent to the main natural frequency of the dual mass oscillator.
  • 1a shows the course of an oscillating movement of a working mass and a free-oscillating mass over time t for a two-mass oscillator.
  • 1b symbolically shows a device comprising a dual-mass oscillator for compacting as an oscillating model.
  • 1c shows the frequency-dependent course of the oscillation travel amplitude and excitation power of a dual-mass oscillator in the region of its main natural frequency.
  • 2a schematically shows an embodiment of a ramming vibrator.
  • Fig. 2b shows a hydraulic circuit diagram for the ram vibrator of Fig. 2a.
  • 3 shows schematically a further embodiment of a ram vibrator.
  • the oscillation frequency is chosen (high) so that the effect of the softly set springs of the vibration isolation (the isolation device for suspending the vibrators on a suitable carrier device) on the vibration path amplitude of the vibrators is negligible or only insignificant.
  • the idle vibration mode is by no means a fictitious one Operating mode, since it is carried out approximately at the beginning of a ramming process and at the end of a pulling process in practical application.
  • the mass that vibrates along with the pile is usually referred to as "m dyn " for the "dynamic mass” in the case of the unbalanced pile vibrators and subsequently as "ma” for the "working mass” in the case of a pile vibrator according to the invention.
  • the unbalance ram vibrator carries out a so-called harmonic, ie a sinusoidal vibration in idle mode, which is characterized on the one hand by the harmonic course of the resulting centrifugal forces, but also by the conversion of kinetic energy from one energy form into another energy form.
  • harmonic ie a sinusoidal vibration in idle mode
  • the kinetic energy of the dynamic mass zero, the kinetic energy becomes of the rotating masses increased to a maximum by increasing the rotational speed.
  • the vibration path amplitude and, using the vibration frequency, the maximum kinetic energy of the vibrating dynamic mass converted twice per vibration period and a so-called vibration power P su of the unbalance ram vibrator can be calculated. If this oscillating power P su could not be continuously drawn from the kinetic energy of the rotating masses, it would have to be supplied from another power source.
  • the same vibration power P su as in idle mode must be implemented with the same working frequency and the same size of the vibration path amplitude during work, here over the pile in addition, a useful power P N (mainly as a friction power) is introduced into the ground, which useful power is also to be implemented via the unbalance excitation system.
  • a useful power P N (mainly as a friction power) is introduced into the ground, which useful power is also to be implemented via the unbalance excitation system.
  • the amount of vibration power R su in the case of unbalance ram vibrators common in practice is generally considerably larger than the useful power P N that can be implemented on the ramming material via the ramming material.
  • the ratio P su / P N may well exceed a value of 1.5.
  • the vibrating ram vibrator is understood as a two-mass vibrator which is forced to vibrate by an excitation system with a cantilever mass mf, a working mass ma and a spring system which forcibly connects both masses in both vibration directions, the working mass ma also taking into account the mass of the rammed material.
  • a vibration power P sz of the vibrating cantilever mass and the working mass can also be defined for the ram vibrator, in which case the ratio of vibration power P sz to useful power P N can reach a value P sz / P N of greater than 1.5 .
  • a working mass ma is shown connected to the ground 100 by means of a spring 102, as a result of which an elastic behavior of the soil adhering to a pile by friction is illustrated.
  • Damping element 104 is intended to indicate that, with a set damping dimension D> 0, a damping power can be delivered via the working mass ma as a compression power from a dual-mass oscillator.
  • a periodic excitation force f (t) 108 of an excitation system is simultaneously applied to a cantilever mass mf and to the working mass ma and is indicated to act in both oscillation directions.
  • the excitation power of the excitation system may have the same amount as the amount of the dissipated damping power when the two-mass oscillator is excited with an excitation frequency equal to the natural frequency (resonance mode).
  • a spring system 106 is provided, the resultant spring constant of which, together with the amounts of the masses ma and mf, determines the (main) natural frequency fn of the dual-mass oscillator.
  • the spring system 106 simultaneously introduces (positive and negative) acceleration forces derived from spring deformation forces into the cantilever mass mf and into the working mass ma during the execution of the oscillating movements. Because of the damping power emitted by the dual-mass oscillator, the working mass ma and the cantilever mass mf must be forced to perform vibrations by the excitation system. However, the dual-mass oscillator can also be forced to oscillate frequencies greater or less than the natural frequency within certain limits by means of a suitable excitation frequency. Then the oscillation path amplitudes of the masses ma and mf turn out to be smaller with comparable excitation power.
  • the oscillation path amplitudes of the working mass ma and the free oscillation mass mf are designated Aa and Af, the corresponding double amplitudes with Ha and Hf.
  • the oscillation path amplitude Af is assumed twice as large in the drawing as the oscillation path amplitude Aa.
  • the masses ma, mf of the dual-mass oscillator oscillate in phase and in opposite directions. The largest distance from the center of mass is designated Smax and the smallest distance is designated Smin.
  • Neglecting the low damping energy that is also emitted in practice in idle vibration operation the following can be determined: At the reversal points for the amplitudes Af and Aa for Smax and Smin, the kinetic energies previously contained in the masses mf and ma at the greatest vibration speed are completely in the deformation energy of the spring system and are converted completely into kinetic energy during the subsequent swinging movements.
  • the vibration power P sz if it has been applied once by the excitation system, is preserved in a similar way to the vibration power P S) U in an unbalance ram vibrator while maintaining constant vibration path amplitudes and does not have to be continuously supplied by the excitation system (even when ramming is in operation) become.
  • the excitation system supplies the oscillation system with a specific excitation power which is just as great as the damping power which is emitted predominantly by the power loss of the pile material vibrating in the ground and is characterized by a specific damping measure D.
  • the curve Va characterizes the oscillation path amplitudes A that can be achieved with a constant excitation force amplitude with a cooperating (average) working mass ma with a variable excitation frequency fe. It can be seen that the maximum vibration path amplitude A max occurs at the natural frequency fn of the vibration system. Within a resonance frequency range Dfa (with a given identical excitation force amplitude) at least oscillation path amplitudes of the amount A 0 can be achieved at all excitation frequencies.
  • a constant oscillation path amplitude A 0 can of course also be set in this range for all excitation frequencies lying between the frequencies fu and fo, which can be achieved by reducing the excitation force amplitude (by means of a control device).
  • the curve Pa represents the excitation powers associated with the oscillation travel amplitudes A with an active (mean) working mass ma, the excitation power at the natural frequency fn having a minimum value Pn.
  • the curve Pa also shows that with an excitation power Pe, which is somewhat larger than Pn, with a prevailing damping dimension D with a predetermined maximum vibration path amplitude A 0, the ram vibrator can be operated in a resonance frequency range Dfa. However, the power amount Pe minus Pn is not recoverable.
  • This procedure also realizes a procedural advantage in that it excludes a soil resonance range Dfs below the natural frequency fn, in which the latter could be excited to form resonance vibrations when working with the pile in the soil.
  • the avoidance of dangerous resonance vibrations of the ground is also the main goal of modern high-frequency unbalance ram vibrators with adjustable vibration path amplitude between zero and a maximum amount. Since the soil resonance range Dfs to be avoided generally extends to approximately 34 Hz, such unbalance ram vibrators are operated as so-called high-frequency vibrators with a working frequency of usually greater than 30 Hz.
  • the amount of the (average) working mass ma can fluctuate between a minimum value mal and a maximum value ma2 through the use of ramming goods of different masses.
  • the natural frequencies fm-1 and fm-2 are assigned vibration path amplitudes according to curves Vm-1 and Vm-2 and excitation powers according to curves Pm-1 and Pm-2, respectively, with the excitation force amplitudes set accordingly, cf. Fig. 1c. If, what is preferred, the dual mass transducer excited with an excitation frequency that corresponds to the natural frequency dependent on the working mass (e.g. fm-1 or fm-2), you can perform the vibration operation (the pile driving) with the least amount of excitation power. If the natural frequency should change slightly as the pile driving progresses due to the changing influence of the soil, the excitation frequency can also be adjusted in this regard with suitable control means.
  • the ram vibrator can be operated practically like an unbalanced high-frequency ram vibrator , So that a dual-mass oscillator and thus the ramming vibrator for performing forced vibrations in a frequency range Dfa, Dfm-1 or Dfm-2 is provided at or near its natural frequency, it applies that if the supply of excitation power is suddenly switched off, the dual-mass oscillator with a frequency in the frequency range Dfa, Dfm-1 or Dfm-2 continues to oscillate with a free damped oscillation with decreasing oscillation path amplitude.
  • a cantilever mass 200 includes the parts of a crossbar 206, a spring cylinder housing 208 including bearing cover 210 and a hydraulic exciter actuator 212 with cylinder housing 214, cylinder cover 216, sensor holder 218 and sensor-1 220, which are firmly connected to one another.
  • the exciter actuator 212 is together with cylinder housing 214, Piston 222, upper piston rod 224 and lower piston rod 226 are constructed as a synchronous cylinder and have an upper displacement work chamber 228 and a lower displacement work chamber 230.
  • the two displacement work chambers 228, 230 can be operated with the aid of the servo directional control valve 232, to which they are each connected via a line are alternately charged with an oil volume with higher pressure and with an oil volume with lower pressure in time with the excitation frequency, whereby - in conjunction with the respective relative movements of the piston 222 - a predetermined excitation energy per oscillation period enters the two-mass oscillator 234 to be led.
  • a pump 244 which is adjustable with regard to its delivery volume and / or output pressure represents the original pressure and volume flow source for the excitation actuator 212.
  • the pump output to which a pressure accumulator 242 is connected in parallel, is connected to the input of a pressure regulator 240 which can be continuously controlled or regulated with respect to its output pressure via an input signal 246, the so-called manipulated variable.
  • a volume flow with a pressure regulated according to a predetermined value passes via a line 236 to the servo directional control valve 232 and from there in an alternating manner caused by the servo directional control valve 232 into one and the other displacement work spaces 228 and 230 des Exciter actuators 212.
  • the displacement workrooms 228 and 230 could optionally also be connected to a low-pressure pressure source (not shown) with a check valve each, in order to avoid cavitation phenomena.
  • the pump 244 is designed by means of an adjusting device, not shown, in such a way that its outlet pressure is set as a function of the outlet pressure of the pressure regulator 240.
  • the pressure at the outlet of the pump 244 could also be set or regulated (in a manner not shown) by means of a suitable adjusting device of the pump 244 as a function of the input signal 246, in which case the pressure regulator 240 could be omitted.
  • the spring system 204 shown in FIG. 2a which uses the compressibility of the hydraulic oil as a spring principle, is designed with the spring cylinder housing 208, the bearing cover 210, a spring piston 250, an upper piston rod 252 and a lower piston rod 254 as a synchronous cylinder.
  • an upper compression space 256 and a lower compression space 258 are each resiliently compressed by an amount He / 2.
  • Both compression spaces 256, 258 can be connected in a symmetrical design via lines 260 or 262 to a further compression chamber 264 and 266, the latter being able to be excluded from the compression process by a shut-off valve 269.
  • a first spring constant and thus a first natural frequency can be defined.
  • a second, lower natural frequency can be created if necessary.
  • Non-return valves 268, 270 connect both compression chambers 264, 266 to a pressure source 272, 272 'which is adjustable with respect to their (relatively low) pressure.
  • the upper piston rod 252 is connected to the lower piston rod 226 of the excitation actuator 212 and the lower piston rod 254 is connected to a yoke 276.
  • the working mass 202 also comprises a holder 278, approximately in the form of a collet and ramming material 280 which is clamped with the aid of the collet.
  • a spring support 286 is for the purpose of maintaining a predetermined central position of the spring piston 250 relative to the spring cylinder housing 208 and the piston 222 relative to the associated cylinder housing 214 provided with which a certain relative position between the cantilever mass 200 and the working mass 202 is also determined at the same time.
  • the spring support 286, which is provided twice in a symmetrical design, consists in each case of a support body 288 and two spring elements 290 and 292, which are fastened to the cross member 206 or to the yoke 276.
  • the springs of the spring supports 286 naturally contribute a certain proportion to the determination of the spring constants of the hydraulic spring system 204.
  • a sensor-1 220 is attached to a sensor holder 218 and is designed such that it displaces the upper piston rod 224 or the piston 222 or the spring piston 250 and thus also the displacement of the working mass 202 relative to the cantilever mass 200 (e.g. the The sum He of the double amplitudes Ha + Hf shown in FIG.
  • a sensor 2 291 is attached to the yoke 276 and is provided for generating measurement signals of the actual value of the vibration acceleration of the working mass 202.
  • Measurement signals for the vibration path for example for the vibration path amplitude Aa or double amplitude Ha of the working mass 202, can be obtained in the control 233 from the acceleration signals.
  • the size of the double amplitude Hf of the cantilever mass mf can thus also be determined from the difference between the physical quantities He and Ha.
  • the support bracket 294 is fastened to the yoke 276 by means of two (soft) spring elements 296, which serve to isolate vibrations, and is used to suspend the entire ram vibrator on a carrying device (not shown), for example on a crane.
  • 2a, 2b is as follows:
  • the controller 233 is used to carry out general control tasks, for example actuation of the servo directional control valve 232 with an alternating frequency of the alternating pressurization and pressure relief of the displacement work spaces 228 and 230 corresponding to the excitation frequency that can be predetermined.
  • the control 233 also regulates the vibration path amplitude Aa of the working mass 202 and / or the vibration path amplitude Af of the cantilever mass 200 via processing of predetermined setpoints and measured actual values this parameter.
  • These physical vibration quantities are regulated by influencing the energy portions supplied to the excitation actuator 212 per oscillation period or half-period, which portions are formed by means of the servo directional control valve 232 from the volume flow or energy flow supplied via the line 236.
  • the size of the energy portions or their energy content is determined (with a predetermined oscillation path amplitude) via the supply pressure for the servo directional control valve 232, which can be adjusted dynamically at the output of the pressure regulator 240 (or alternatively via a direct regulation of the output pressure of the pump 244 by means of an integrated therein pressure regulator).
  • the input signal 246 or the manipulated variable of the pressure regulator is derived from the output signal of the regulator.
  • the pressure regulator is thus the so-called actuator, which intervenes in the energy flow.
  • the size of the working mass and thus the natural frequency of the dual mass vibrator can change within certain limits depending on the size of the mass of the pile being used.
  • a special device with a special control algorithm is expediently provided within the control 233, with which the control 233 automatically searches the current natural frequency and adjusts the excitation frequency in accordance with the natural frequency, starting from a specific predefinable reference frequency. This can be done, for example, in such a way that the reference frequency is initially adjusted slightly in order to find out whether it is possible to reduce the excitation power with the adjusted excitation frequency. If appropriate confirmatory information can be obtained, further adjustment steps are carried out until the excitation frequency is found at which a minimum of excitation power is required.
  • the influence of a damping power that is to be emitted due to the ramming operation is of course also compensated for.
  • the oscillation path amplitudes Af and Aa of the free-oscillating mass 200 and the Working mass 202 become increasingly smaller, the oscillation path amplitudes Aa and Af can also be regulated according to a predetermined desired value with great dynamics, even without the cooperation of the pressure regulator 240, with an increasing or decreasing need for excitation power with increasing or decreasing damping performance through an adjustment of the delivery volume and / or the outlet pressure of the pump 244 can be taken into account, and in this case the adjustment of the pump 244 with the lower dynamic characteristic of this adjustment is permissible.
  • the preload pressures build up automatically from the rest position of the spring piston 250 (corresponding to its central position) until the predetermined oscillation path amplitude or the intended force stroke He of the excitation actuator 212 is reached:
  • the ram vibrator shown in Fig. 3 is constructed very similar to the ram vibrator shown in Figs. 2a and 2b and can be operated just like this.
  • the individual assemblies or Components which perform the same functions as those shown in FIGS. 2a and 2b are given the same reference numerals in FIG. 3, while the different features are identified by reference numerals in the form of 3-digit numbers with the number 3 at the beginning ,
  • the difference from FIG. 2a is as follows: In contrast to FIG. 2a, where a hydraulic exciter actuator 212 is connected with its cylinder housing to the cantilever mass, FIG.
  • FIG. 3 shows two exciter actuators 312 and 312 'of the same type and operated in parallel, and their cylinder housings 314 are firmly connected to the working mass 202.
  • the upper piston rods 324 are screwed together with angle components 389, which in turn are fastened to the bearing cover 210, and in this way transmit the excitation forces acting between the cantilever mass 200 and the working mass 202.
  • the two spring supports 386 are designed differently in that support bodies 388 are now attached directly to the yoke 276 and thus belong to the working mass 202.
  • the support members 388 are connected to the cross member 206 via spring elements 390.
  • the sensor-1 220 attached to a sensor holder 218 can detect the displacement of the upper piston rod 252 and thus also the displacement of the working mass 202 relative to the cantilever mass 200 and pass it on to the controller 233 as a measurement signal.
  • the dashed lines used in FIGS. 2a and 3 indicate fastening means for the fixed connection of different components.
  • an electrical linear actuator or an electrical torque motor can also be provided, the latter preferably being designed as three-phase AC motors.
  • a hydraulic Swing motor or an electric torque motor would be the conversion of the motor swiveling movement into a linear drive movement preferably to be realized by a rack and pinion gear.
  • a rotor directional valve can also be provided. In such a rotor directional control valve, the previously described function of a servo directional control valve with a control spool that moves back and forth in time with the excitation frequency to be specified is replaced by a control rotor rotating at the same frequency.
  • control rotor must be driven by a rotary motor, for example by a small hydraulic axial piston motor, this rotary motor preferably being controllable or regulatable in terms of its speed.
  • a rotary motor for example by a small hydraulic axial piston motor, this rotary motor preferably being controllable or regulatable in terms of its speed.
  • electrical actuators the advantage that the free oscillating mass is smaller than the working mass, the advantage that a longer output working path of the exciter actuator results can be used.
  • the spring system can also be implemented with mechanical springs, preferably using double leaf springs.
  • slide guides are formed by the piston itself or by cylindrical bodies arranged coaxially to the central axis of the piston.
  • a spring support which is effective between the cantilever mass and the working mass, is expediently provided for a central standstill position of the exciter actuator, with an intermediate position of the output member of the exciter actuator relative to the mass in which the exciter actuator is accommodated, or in the case of a hydraulic spring system, at least when the ram vibrator is at a standstill a middle position of the spring piston relative to its cylinder housing is predetermined by an effective spring support between the cantilever mass and the working mass, which is used by the spring system itself when using mechanical spring elements and which is realized by a special spring support in a hydraulic design of the spring system, whereby due to the special or separate spring support, the resulting spring constant is also determined by the spring system.
  • the ramming vibrator advantageously works in resonance mode at or near its main natural frequency of the dual mass oscillator. This results in low energy losses.
  • the vibration performance of the cantilever mass and the working mass can always be almost completely recovered and reused.
  • the at least one exciter actuator can be dimensioned smaller as a result of a smaller excitation power to be implemented and can operate or be controlled with a higher dynamic and accuracy.
  • the ramming vibrator can be operated safely in the resonance range in the vicinity of its natural frequency, since the actual value of a physical vibration variable, e.g.
  • the amplitude of the vibration path sf or sa of the cantilever mass or the working mass is measured by a measuring device and processed by means of a controller in such a way that the physical vibration quantity is regulated with great dynamics and accuracy according to the specified value.
  • That the ramming vibrator which is designed as a dual-mass oscillator for carrying out forced vibrations, is located at its main natural frequency or can be operated in a frequency range (Dfa), (Dfm-1), (Dfm-2) which is symmetrical to its main natural frequency, can be demonstrated in that when the supply of excitation power to the at least one excitation actuator is suddenly switched off, the oscillating movements of the dual mass oscillator with it Main natural frequency can be continued with decreasing vibration amplitude.
  • a predetermined physical vibration quantity for example the amplitude of the vibration path sf or sa
  • a pressure regulator with adjustable output pressure Thanks to the use of a spring system with which the spring energy required for resonance operation can be stored, a high loss rice force can be developed at the reversal point of the oscillating movement of the working mass, even at a high working frequency, without any time delay.
  • a cantilever mass smaller than the working mass can be used, which leads to an overall lower mass of the dual-mass oscillator and, at the same time, to a desirable longer output working path of the exciter actuator.

Abstract

L'invention concerne un vibrateur de battage de pieux destiné à des articles battus et comprenant un oscillateur linéaire à deux masses (234) pouvant être excité par au moins un actionneur excitateur (212) de façon à générer des vibrations contraintes et présentant une masse à oscillation libre (200), une masse de travail (202) ainsi qu'un système de ressort (204) couplant les deux masses. Les deux masses oscillent pratiquement en phase et en sens opposé et l'une des masses porte un élément de fixation (278) pour les articles battus (280). Une boucle de régulation pour la puissance d'excitation comprend un régulateur, (233), l'actionneur excitateur (212) et un dispositif capteur (220, 291) pour la mesure directe ou indirecte de l'amplitude de course d'oscillation (298) et sert à réguler à une valeur prédéfinie l'amplitude de course d'oscillation dans la gamme de fréquence voisine de la fréquence propre principale de l'oscillateur à deux masses, la fréquence d'excitation pouvant être prédéfinie dans la gamme de fréquence voisine de la fréquence propre principale de l'oscillateur à deux masses.
PCT/EP2004/014145 2003-12-14 2004-12-13 Vibrateur de battage de pieux pour des articles battus WO2005056201A1 (fr)

Applications Claiming Priority (4)

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DE10358808.6 2003-12-14
DE10358808 2003-12-14
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DE102004005839 2004-02-06

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WO2007141075A1 (fr) * 2006-06-07 2007-12-13 Siemens Aktiengesellschaft Procédé pour surmonter le frottement par adhérence, procédé pour déplacer une charge utile ainsi que machine de chantier, machine de production ou robot
CN102444126A (zh) * 2010-09-30 2012-05-09 爱科昇株式会社 混凝土打桩系统以及混凝土打桩方法
EP3101179A1 (fr) * 2015-06-03 2016-12-07 ABI Anlagentechnik-Baumaschinen-Industriebedarf Maschinenfabrik und Vertriebsgesellschaft mbH Appareil de travail, notamment pour un engin de chantier
WO2020234639A1 (fr) * 2019-04-07 2020-11-26 Resonance Technology International Inc. Circuit d'attaque résonnant unidimensionnel à masse unique

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DK3417951T3 (da) * 2017-06-19 2022-07-04 Eurodrill Gmbh Anordning og fremgangsmåde til generering af slagimpulser eller svingninger til en byggemaskine

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DE2008059A1 (de) * 1969-09-04 1971-09-09 Gunther Neumann Hydraulischer pneumatischer und mechaniser Antrieb für oszillierende Bewegungen
DE4434679A1 (de) * 1993-09-29 1995-03-30 Hubert Bald Verdichtungssystem zum Formen und Verdichten von Formstoffen zu Formkörpern in Formkästen
WO2001047698A1 (fr) * 1999-12-24 2001-07-05 GEDIB Ingenieurbüro und Innovationsberatung GmbH Dispositif de compression pour effectuer des operations de compression sur des corps moules a base de matieres granuleuses
EP1167632A2 (fr) * 2000-06-23 2002-01-02 ThyssenKrupp AG Vibrateur avec unité d'entraínement linéare

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DE2008059A1 (de) * 1969-09-04 1971-09-09 Gunther Neumann Hydraulischer pneumatischer und mechaniser Antrieb für oszillierende Bewegungen
DE4434679A1 (de) * 1993-09-29 1995-03-30 Hubert Bald Verdichtungssystem zum Formen und Verdichten von Formstoffen zu Formkörpern in Formkästen
WO2001047698A1 (fr) * 1999-12-24 2001-07-05 GEDIB Ingenieurbüro und Innovationsberatung GmbH Dispositif de compression pour effectuer des operations de compression sur des corps moules a base de matieres granuleuses
EP1167632A2 (fr) * 2000-06-23 2002-01-02 ThyssenKrupp AG Vibrateur avec unité d'entraínement linéare

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WO2007141075A1 (fr) * 2006-06-07 2007-12-13 Siemens Aktiengesellschaft Procédé pour surmonter le frottement par adhérence, procédé pour déplacer une charge utile ainsi que machine de chantier, machine de production ou robot
CN102444126A (zh) * 2010-09-30 2012-05-09 爱科昇株式会社 混凝土打桩系统以及混凝土打桩方法
EP3101179A1 (fr) * 2015-06-03 2016-12-07 ABI Anlagentechnik-Baumaschinen-Industriebedarf Maschinenfabrik und Vertriebsgesellschaft mbH Appareil de travail, notamment pour un engin de chantier
WO2020234639A1 (fr) * 2019-04-07 2020-11-26 Resonance Technology International Inc. Circuit d'attaque résonnant unidimensionnel à masse unique
US11338326B2 (en) 2019-04-07 2022-05-24 Resonance Technology International Inc. Single-mass, one-dimensional resonant driver

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