WO2005056201A1

WO2005056201A1 - Piling vibrator for material that is to be rammed

Info

Publication number: WO2005056201A1
Application number: PCT/EP2004/014145
Authority: WO
Inventors: Hubert Bald
Original assignee: GEDIB Ingenieurbüro und Innovationsberatung GmbH
Priority date: 2003-12-14
Filing date: 2004-12-13
Publication date: 2005-06-23
Also published as: DE102004059938A1

Abstract

The invention relates to a piling vibrator for material that is to be rammed. Said piling vibrator comprises a linear dual-mass oscillator (234) which can be excited by at least one exciting actuator (212) so as to generate forced vibrations and is provided with a free-swinging mass (200), a working mass (202), and a spring system (204) that couples the two masses. The two masses oscillate essentially in phase and in opposite directions. One of the masses supports a fixture (278) for the rammed material (280). A control loop comprising a controller (233), the at least one exciting actuator (212), and a sensor device (220, 291) for directly or indirectly measuring the vibration amplitude (298) is provided for controlling the excitation power. Said control loop allows the vibration amplitude to be regulated to a predefined value lying within the frequency range next to the principal eigenfrequency of the dual-mass oscillator, the excitation frequency being predefinable in the frequency range next to the principal eigenfrequency of the dual-mass oscillator.

Description

Ram vibrator for rammed material

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. In the practical ram vibrators used to date, 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. If such an adjustability of the vibration path amplitude is given, 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. Apart from a few ramming vibrators that work with unbalance and can be adjusted with regard to the vibration path amplitude without using a gear transmission, which in turn have special disadvantages, the majority of ramming vibrators in practical use have the vibration path amplitudes from zero to zero during the rotation of the unbalance a suitable value can be adjusted greater than zero, the synchronization of the unbalance and also the adjustability of the vibration path amplitudes by using gear mechanisms heavily loaded by mass forces. However, this type of ram vibrator leads to very high noise emissions and high wear due to the gear drives. In order to counter the noise emissions, 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. This makes it possible to achieve the desired work output and to use the resonance range without inadmissible vibration displacement amplitude values, i.e. to work at or in the vicinity of the main natural frequency of the dual-mass oscillator, where the power required to operate the ramming vibrator is considerably reduced, since in this operation the Spring system stores the energy used for vibration in one direction and is then largely released again for vibration in the opposite direction. Apart from this, the frequency range up to the working frequency can be traversed with little or no vibration amplitude, so that the ground resonance range of approx. 30 Hz can be traversed without any problems. Further embodiments of the invention can be found in the following description and the subclaims. The invention is explained in more detail below on the basis of exemplary embodiments illustrated in the attached figures. 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. In connection with FIGS. 1a-c, the principle of largely imitating the kinetics of the unbalance ram vibrators, which is implemented here, is explained in general below by the use of a dual-mass oscillator vibrating at its location or in the vicinity of its natural frequency in "idle oscillation operation". A comparison of the conditions in the case of unbalance vibrators and ram vibrators according to the invention is most clear when both types of vibrator are considered in that operating state, here called idle vibration mode, in which the vibrators vibrate freely with a certain mass that vibrates with the rammable material and with a predetermined frequency. without any useful work being put into the ground. 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. During the oscillation movement, there is a continuous conversion of the kinetic energy of the dynamic mass and a portion of the kinetic energy of the rotating masses, which takes place twice per oscillation period: At the reversal point of the oscillation movement, where 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. In the middle of the swinging movement, where the kinetic energy of the dynamic mass reaches a maximum at maximum speed, the kinetic energy of the rotating masses is reduced to a minimum by reducing the rotational speed. From the data of the dynamic mass m _dyn and the static moment of the unbalance, 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. As can be shown, 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. With regard to the advantageous properties of an unbalance ram vibrator, it is still important that at the moment of reversal of the centrifugal force the maximum resulting centrifugal force, which is greater than the average due to the sinusoidal shape, is available. When vibrating the pile in the ground, depending on the frictional relationships between the pile and the soil, the pile can have a tendency to get stuck in the ground (static friction). Since the law also applies here that the friction of rest is greater than the friction during movement, it is very advantageous in practice that in this case the required (detaching) vibration stroke can be started under the influence of the maximum resulting centrifugal force , In the present case, 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 . According to FIG. 1b, 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. On 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 effect of a spring and the damping of a device for vibration isolation for suspending the ramming vibrator, for example on a crane, are not taken into account here because the corresponding spring must be set very softly and thus hardly influences the vibration behavior in the range of the natural frequency fn. 1 a shows the course of the vibrations of the dual-mass vibrator in the idle vibration mode. The abscissa axes could also represent the common phase angle instead of the common time profile t. The masses ma and mf are - symbolically reduced to one point - each in the shown upper and lower reversal of the swinging movements. 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 general relationship can be derived with n = Af / Aa: mf = ma / n. Accordingly, in the case of n = 2 shown, the ratio applies: mf = ma / 2. As can be seen from Fig. 1a, 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. Relative to the case Smin = 0, the maximum relative displacement of the mass centers is therefore Smax_0 = Ha + Hf and with Hf = n * Ha: Smax_0 = Ha * (1 + n). With a selected value n = 2, an excitation actuator of the excitation system can cover a power stroke He of He = 3 * Ha. In the sense of reducing the mass of the dual-mass oscillator and favorably increasing the excitation stroke of the excitation actuator, it is advantageous to provide a mass ratio ma / mf greater than 1. 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. Strictly speaking, however, the principle of perfect energy conservation only applies to oscillation operation with an oscillation frequency at the location of the natural frequency fn of the dual-mass oscillator. From the explanation of Figures 1a and 1b can also it can be seen that the acceleration forces for accelerating the masses mf and ma are greatest at the reversal points of the oscillating movements. 1c shows the course of vibration path amplitudes A (ordinate A) and excitation powers P (ordinate P) over the excitation frequency fe in the vicinity of the natural frequency and the vibration system of the frame vibrator in the working vibration mode. It is assumed that 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. Would you at a given maximum Vibration path amplitude A ₀ want to operate the ram vibrator with a frequency outside the resonance frequency range Dfa, so one would have to reckon with an increasing excitation power loss or vibration power loss with increasing distance from the natural frequency fn. This also means that when the oscillation operation of the resonance ram vibrator starts, the excitation frequency cannot be continuously increased from zero to the working frequency (in the resonance frequency range Dfa). Rather, with the start of an oscillation process, the oscillation excitation is started immediately with the planned working frequency (with the maximum excitation force amplitude applied). Initially, only oscillation path amplitudes A smaller than A _Q are initially achieved, which, however, continue to increase with each half-period until after a corresponding "build-up" of the oscillatory movements (accumulation of stored kinetic energy) the predetermined oscillation path amplitude A ^ is reached after several half-periods , Conversely, reducing the oscillatory movements to zero at the end of a ramming or pulling process also requires a few half-periods in order to remove any oscillating energy from the oscillating system by means of damping. This procedure, however, 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 natural frequency fn is of course determined by the constructive choice of the masses mf and ma and the spring constant of the spring system, so that a ramming vibrator Working with working frequencies outside of a predetermined resonance frequency range Dfa (at a constant value of the spring constant) is not to be assumed with normal delivery of working damping power. If one sets the natural frequency fn sufficiently high above the earth resonance range Dfs approximately fn = 35 Hz or higher in the resonance ram vibrator according to the invention working as a dual mass vibrator, then one has a counterpart to the unbalanced high frequency vibrator. Depending on the work task, 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. A ratio q = ma2 / ma1 of approximately q = 1, 7 usually satisfies all the usual requirements. One can easily calculate the natural frequency fn of the ramming vibrator with good approximation if one considers the same as a so-called "unbound two-mass oscillator", that is, if one considers the spring 102 and the damping element 104 as not present in FIG. 1b, which is true in the idling mode is also the case. In this case, the following applies to the main natural frequency fn:

If one defines the natural frequencies to be assigned to the masses mal or ma2 (ma2> mal) with fm-1 or fm-2 (fm-2 <fm-1), then from the above relationship for an amount of q = 1, 7 and deduce with n = 3 that the difference Dfm = fm-1 minus fm-2, based on the amount of fm-1, is about 10%. If one uses smaller or larger working masses times or ma2 instead of a mean working mass ma with assigned natural frequency fn, larger or smaller natural frequencies fm-1 or fm-2 result automatically. 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. Also when performing working vibrations with a working mass deviating from an average working mass ma, e.g. with a working mass ma2 with a natural frequency fm-2, it is also true that with a slightly increased excitation power, working frequencies in a frequency range Dfm-2 close to the natural frequency fm -2 are feasible. If one sets the expected lowest natural frequency fm-2 by a corresponding design of the resulting spring constant in such a way that it is still above the earth resonance range Dfs, then 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. To determine that the ramming vibrator actually vibrates in the vicinity of the main natural frequency of the dual mass oscillator, one can proceed as follows, for example with a view to the curve Va in FIG. 1c: With a change in the excitation frequency by a few percent (e.g. 5%) it must be kept constant Excitation power also noticeably change the vibration path amplitude Af or Aa (eg by> 1%), or by changing the excitation frequency by a few percent (eg 5%) the excitation power requirement must also change noticeably (eg by> 1%) if the vibration path amplitude is kept constant. 2a, a cantilever mass 200, a working mass 202 and a spring system 204 denote the main components of the dual mass oscillator 234 of a ramming vibrator. The 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. According to FIG. 2b, 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. From the output of the pressure regulator 240, 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. Those that are also ejected from the displacement work spaces 228, 230 in an alternating manner at low pressure Volumes of fluid, the output of which is controlled by servo-way valve 232, are returned to a tank via line 238. 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. Alternatively, however, 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. When the spring piston 250 moves upward or downward relative to the cantilever mass 200, 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. With the size of the compression chamber 264, together with the size of the compression space 258, a first spring constant and thus a first natural frequency can be defined. By connecting the compression chamber 266, 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. By the one emerging from the pressure source 272 or 272 ' Oil volume flow at the moment of the greatest pressure relief of the compression spaces 264, 266, the lost leakage oil can be replaced. Through a different design of the total volume ina of the compression spaces 256, 258 including the connected compression chambers 264, 266 it can be achieved that differently large spring constants are used for different vibration directions in both vibration directions. In this way it can be achieved, for example, that the downward stroke of the spring piston 250 or the working mass 202 takes place at a higher speed than the upward stroke. In the central area of the spring piston 250, which is expediently operated in the spring cylinder housing 208 with a gap seal (without special sealing elements). there is an annular groove 274 in the spring cylinder housing 208 which is connected to the tank. As a result, leakage oil can be discharged, which at the same time deliberately flushes out the compression spaces 256, 258 for the purpose of removing the heated hydraulic oil. 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. In addition to the yoke 276 and the components of the spring system 204 and the exciter actuator 212 connected to it, 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. In a bore of the spring cylinder housing 208 and in a bore of the Bearing covers 210 are pockets for hydrostatic bearings 282 and 284 of the upper piston rod 252 and the lower piston rod 254 (instead of hydrostatic, hydrodynamic bearings can be used). The means of generating pressure and the supply lines for the hydrostatic bearings are not shown. The pressure generation could also be derived from the compression pressures in the compression spaces 256, 258. 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. 1 a can be detected and passed on as a measurement signal to a controller of a controller 233. 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. When regulating the oscillation path amplitudes, the influence of a damping power that is to be emitted due to the ramming operation is of course also compensated for. By taking advantage of the effect that, with an increasing deviation of the excitation frequency from the unique frequency of the two-mass oscillator with constant excitation power supplied, 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. In the compression spaces 256 and 258 of the spring system 204, after reaching a predetermined oscillation path amplitude Aa (symbolized by the double arrow 298) or Af (symbolized by the double arrow 299), there is a preload pressure in a central position of the spring piston 250, which only arises in the end positions of the spring piston 250 that low pressure drops which corresponds to the pressures of the two pressure sources 272, 272 ', which can also be set differently. At the start of the oscillation movements, 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: In the case caused by the excitation forces of the excitation actuator 212 , from the start of continuously increasing displacements of the spring piston 250 from the central position, the additional volumicina required in the compression spaces 252, 254 from the pressure sources 272 and 272 'are sucked in by the cooperation of the check valves 268. This continues until the intended force strokes He are reached by the accumulation of spring energy in the compression spaces 252, 254. Thereafter, smaller oil volumes are fed from pressure sources 272, 272 'only for the purpose of substituting leakage oil. 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. For the sake of simplicity, 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. 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 moving parts of the excitation actuators 312, 312 'consist of pistons 322, lower piston rods 326 and upper piston rods 324. Together with cylinder housings 314 and cylinder covers 316, they form upper displacement working spaces 328 and lower displacement working spaces 330 which, as in FIG. 2a, via lines to the servo -Way valve 232 are connected. 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. In FIG. 3, the two spring supports 386, designated 286 in FIG. 2a, 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. In the case of FIG. 3, 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. Instead of a hydraulic linear actuator or a hydraulic swivel motor, an electrical linear actuator or an electrical torque motor can also be provided, the latter preferably being designed as three-phase AC motors. When using 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. Instead of a servo directional valve, 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. In this case, the 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. When using 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. A predetermined physical vibration quantity, for example the amplitude of the vibration path sf of the cantilever mass mf and / or the amplitude of the vibration path sa of the working mass ma or its sum ss = sf ± sa, or a time derivative sf, sf "or sa ', sa" or "ss', ss" or a maximum compression pressure of a hydraulic spring cylinder used can be regulated by means of the regulator by influencing the excitation power supplied and / or the excitation frequency, the actual value signal of the predefined physical vibration variable and possibly one measured by a measuring device being regulated by the regulator When using a hydraulic spring system with a piston and a cylinder housing, it is preferred that 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. When using a hydraulic exciter actuator, a predetermined physical vibration quantity, for example the amplitude of the vibration path sf or sa, can be dynamically and precisely controlled by the use of 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. With the ram vibrator, 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. When working with a pile in a resonance-prone soil with an oscillation frequency fz above the earth resonance range with the highest frequency fr, which oscillation frequency fz is in the range of the natural frequency of the dual mass oscillator, even with a sudden failure of the excitation energy source or with a sudden disturbance no resonance frequency in the ground can be excited by the control, since the ramming vibrator continues to vibrate with the main natural frequency fn above the frequency fr as the vibration amplitude decreases.

Claims

claims

1 . Ramming vibrator for ramming material with a linear two-mass oscillator (234) that can be excited to forced vibrations by at least one excitation actuator (212) and which has a cantilever mass (200), a working mass (202) and a spring system (204) coupling both masses, the two masses oscillate essentially in phase and in opposite directions and one of the masses carries a holder (278) for the pile material (280), characterized in that a control circuit for the excitation power, a controller (233), the at least one excitation actuator (212) and a sensor device (220, 291) for direct or indirect measurement of the vibration path amplitude (298) is provided, with which the vibration path amplitude in the frequency range adjacent to the main natural frequency of the dual-mass oscillator can be regulated to a predetermined value, the excitation frequency in the frequency range adjacent to the main natural frequency of the dual-mass oscillator being specifiable ,

2. A ram vibrator according to claim 1, characterized in that the spring system (204) is dimensioned such that the deformation energy which can be stored in it is sufficient for the continuous implementation of vibrations in both directions (298) even with the largest vibration path amplitudes (Ao) provided for the cantilever mass (200) and / or the working mass (202), on the condition that the damping power to be dissipated and the excitation power to be supplied are each zero.

3. A ram vibrator according to claim 1 or 2, characterized in that the oscillation frequency within a resonance frequency range (Dfa, Dfm-1, Dfm-2), which is substantially symmetrical to the main natural frequency (fn) of the dual-mass oscillator (234) by adjustment the excitation frequency (fe) is adjustable relative to the main natural frequency.

4. A ram vibrator according to one of claims 1 to 3, characterized in that the working mass (202) for changing the main natural frequency (fn) is variable.

5. ram vibrator according to one of claims 1 to 4, characterized in that the maximum deviation (fu, fo) of the working frequency (fe) from the main natural frequency (fn) at the maximum permissible vibration amplitude (Ao) of the working mass by an amount of 30% , based on the respective main natural frequency (fn).

6. A ram vibrator according to one of claims 1 to 5, characterized in that the spring system (204) comprises at least one hydraulic spring based on the compressibility of a hydraulic fluid volume in a compression space (256, 258).

7. ram vibrator according to claim 6, characterized in that the compression space (258) of the hydraulic spring between a piston (250) and a piston (250) receiving the cylinder housing (208) is arranged.

8. A ram vibrator according to claim 7, characterized in that at least one additional, the same compression pressure as the compression space (258) in the cylinder housing (208) subjected to the compression chamber (264, 266) is provided.

9. A ram vibrator according to claim 8, characterized in that the at least one compression chamber (266) can be switched on.

10. Ramming vibrator according to one of claims 6 to 9, characterized in that at the latest after reaching a vibration operation with a predetermined vibration path amplitude in the at least one compression space (256, 258) after the pressure drop when a predetermined vibration path amplitude is reached a certain minimum pressure (272, 272 ') is maintainable.

1 1. A ram vibrator according to one of claims 6 to 10, characterized in that the sealing of the at least one compression space (256, 258) between the piston (250) and the cylinder housing (208) is effected by a gap seal.

12. A ram vibrator according to one of claims 1 to 1 1, characterized in that the spring system comprises mechanical spring elements.

13. A ram vibrator according to claim 12, characterized in that the mechanical spring elements are double leaf springs.

14. A ram vibrator according to one of claims 1 to 13, characterized in that at least one linear sliding guide (282, 284) are provided for straight guidance of the working and cantilever mass relative to one another.

15. A ram vibrator according to one of claims 1 to 14, characterized in that the at least one excitation actuator (212) is selected from the group comprising a hydraulic and an electrical excitation actuator.

16. A ram vibrator according to one of claims 1 to 15, characterized in that an effective spring support (286) between the cantilever mass and the working mass is provided for a standstill middle position of the exciter actuator (212).

17. Ramming vibrator according to one of claims 1 to 16, characterized in that it is designed as a high-frequency vibrator for working frequencies substantially above 30 Hz.

18. A ram vibrator according to one of claims 1 to 17, characterized in that the cantilever mass (200) is smaller than the working mass (202), the mass of the pile (280) and the mass of the holder used for the pile (280) 278) is considered as part of the working mass (202).

19. Ramming vibrator according to one of claims 1 to 18, characterized in that the pressure of the inflow volume flow (240) is adjustable via the controller (233) in a hydraulic exciter actuator.

20. A ram vibrator according to one of claims 1 to 19, characterized in that in the case of an electrical exciter actuator the excitation power supplied to it can be regulated by means of the controller (233) in such a way that the exciter actuator has one during each period or half period of the oscillation changeable, certain amount of electrical energy is supplied.

21. A ram vibrator according to one of claims 1 to 20, characterized in that the regulation of the oscillation path amplitude can be carried out by changing the excitation frequency from its amount prior to influencing the rule.

22. A method of operating a ramming vibrator according to one of claims 1 to 21, characterized in that the oscillating operation is started and carried out with the highest frequency as possible while avoiding the generation of oscillating frequencies within an earth resonance range: With the participation of the exciter actuator (212) The vibration excitation takes place from the beginning with an excitation frequency above the uppermost frequency fr with a predetermined, possibly also maximum excitation force, the oscillation path amplitudes of at least the cantilever mass (200) due to the amount of spring energy stored in the spring system in an increasing manner each time the oscillation movement is reversed enlarged, - when the amount approaches a predetermined setpoint amount for the steady state of a predetermined physical quantity of the oscillating movement, which is continuously measured with respect to its actual value with a sensor device, an adjustment takes place ng of the excitation force or the excitation energy per oscillation period or half-period in such a way that the setpoint amount is set.

23. The method according to claim 22, characterized in that in the case of a working mass changed compared to an assumed mean working mass, the excitation frequency is automatically adjusted such that it corresponds to the changed main natural frequency.

24. The method according to claim 23, characterized in that the excitation frequency is sought and set, in which the lowest excitation power is required to set a predetermined physical vibration quantity.

25. The method according to any one of claims 22 to 24, characterized characterized in that when idling with or without an additional mass acting as a ramming, the oscillation is carried out with a predeterminable, regulated amount of a physical vibration variable, whereby in the case of the participation of a ramming material as an additional mass, this gives no or no appreciable damping power to the ground.