EP1332028B1 - Compaction device for compacting moulded bodies from granular substances and a method for using said device - Google Patents

Abstract

Apparatus and method for carrying out compaction operations on molded bodies that consist of granular materials and are placed on pallets, the compaction being achieved by impact of a vibrating table on the underside of the pallet. The vibrating table, together with a spring system forms a mass-spring system, which acts as a vibrator capable of oscillation that is excited by an excitation device to produce forced vibrations. The spring system, together with the system mass, is designed to develop at least one individual frequency within the range of the compaction frequency, whereby it is also possible to adjust the individual frequency gradually or continuously. This, together with the fact that the excitation frequency can be adjusted, allows the vibrator to be operated partially or completely in resonance mode over the whole frequency range of the compaction.

Description

The invention relates to a device for compaction according to claim 1 and a use of the device according to claim 26. In use, the molding material is before the Verdichtungsvörgang in the mold cavities first as a volume of mass loosely adhering granular components, which only during the compaction process by the action be formed of compactors on the top and bottom to solid moldings. The volumetric mass, when the compactor is used in machines for producing concrete finished products (e.g., paving stones), e.g. consist of wet concrete mortar. In vibrating compactors for producing finished concrete products, one can distinguish three known genera which are suitable for the description of the prior art of interest, and which have in common that the molding box and the molding material are on top of a pallet during the compaction process or a base plate are arranged. In this case, during the main compression, there is a pressing plate on the upper side of the molding material, which can be moved by a pressing device in the vertical direction and can be driven to exert a predetermined pressing pressure.

In the first genus is the widespread and known in the art "conventional nature" of collision compression, in which with respect to its vibration amplitude adjustable vibrating table of a vibrator at each oscillation period once encountered from below against the pallet. This genus represents the closest prior art described by EP 0 515 305 B1 , Also in the second genus, whose compression device works considerably differently than in the first genus, the compression energy originally generated by the vibrator is introduced via impact processes in the molding material. In this case, the pallet and the molding box are firmly clamped during the compression process with the vibrating table, so that their masses are included in the mass of the vibrating system and resonate with it. The definable by the collision of different masses at different speeds joint here is at the top and bottom of the molding material itself, wherein during the compaction, an air gap between the lower mold base and the pallet on the one hand and the mold body top and the press plate on the other. This second genus, described by the DE 44 34 679 A1 , is most aptly called a compactor to carry out a "shaking compaction". In the third genus, evidenced by the EP 0 870 585 A1 , the masses of the molding material, the molding box, the pallet and the swinging table together form a mass system, which represents the oscillating mass of a mass-spring system working with harmonic (sinoid) oscillating movements. The introduced at the top and bottom of the molding dynamic forces derived from the vibration accelerations of the resonating masses produce a likewise sinoid dynamic compression pressure (harmonic compression). Some information of interest to the prior art according to the EP 0 515 305 B1 and the EP 0 870 585 A1 can also be found in an article in the journal " BFT ", Sept. 2000 edition, pages 44-52 , Publisher: Bauverlag GmbH, Am Klingenweg 4a, D-65396 Walluf.

All of these three genres are based on different philosophies about the physical effects occurring during compaction. In this case, even seemingly minor feature differences of the physical effects used can be of importance, such as. the formation of one and the same static moment of unbalanced bodies of unbalanced vibrators with larger or smaller center of gravity distances associated with smaller or larger masses. All three genera have in common that the operation of the compression devices strives to operate the oscillating systems such that the highest possible compression accelerations in the molding material at the highest possible vibration frequencies (possibly up to about 70 Hz) achieved, the accelerations and the frequencies even after be set to predetermined values. In each case, the vibration acceleration of the vibrating table always involved, depend on the next to the compaction result, the loads of the components involved, a linear function of the oscillation amplitude and a quadratic function of the oscillation frequency.

• Die oberste Schwingfrequenz wird in der Praxis wegen der zu berücksichtigenden Dauer-Belastungsgrenze in der Regel auf 50 Hz eingeschränkt, wobei die Grenzbelastung vor allem bei den Wälzlagerungen der Unwuchtwellen und bei den mitschwingenden Kardanwellen erreicht wird. Hierzu siehe auch den oben zitierten Fachzeitschrift-Artikel auf Seite 45, mittlerer Abschnitt und auf Seite 47, mittlerer Abschnitt.
• Durch die ständig umzusetzende Blindleistung und durch die bei hohen Fliehkräften erzeugten hohen Lagerreibungs-Leistungen treten hohe Verlustleistungen auf. Da die hohen Verlustleistungen auch in den Antriebsmotoren der Unwuchtwellen umgesetzt werden müssen, werden die Motoren und deren Ansteuergeräte mit Bezug auf die reine Verdichtungsleistung unnötig groß dimensioniert.
• Bedingt durch die zu überwindenden Trägheitsmassen der Motoren und Unwuchtkörper und bedingt durch die Tatsache, daß mit einer Veränderung des Phasenwinkels sogleich auch immer eine Veränderung des ebenfalls mit auszuregelnden Blindleistungs-Drehmomentes verbunden ist, können die Werte der als Regelgröße vorgegebenen Phasenwinkel (statisches Moment) durch die elektronische Regelung (oder auch durch alternative mechanische Regelungen) nur mit groben Toleranzen geregelt werden, was zu entsprechenden Ungleichförmigkeiten des Schwingwegverlaufes des Schwingtisches während des über viele Schwingungsperioden ablaufenden Verdichtungsvorganges und damit zu einer schlechten Reproduzierbarkeit der Verdichtungsqualität führt. Hinzu kommt hier der Nachteil, daß von den groben Toleranzen der Regelgröße "Phasenwinkel" die relative Winkellage von insgesamt 4 Unwuchtkörpern betroffen ist, die üblicherweise mit ihren Rotationsachsen in einer Ebene liegen und deren Anodnung sich über einen großen Teil der Längsausdehnung des Schwingtisches erstreckt. Die Ungleichheiten der relativen Winkellagen führt zu ungleichen Beschleunigungen bezogen auf die ganze Tischoberfläche. Dies führt wiederum zu ungleichen Verdichtungsergebnissen an unterschiedlichen Orten der Tischoberfläche.
• Die für die Verdichtungswirkung maßgebliche Schwingwegamplitude des Schwingtisches ist nur indirekt und träge über den verstellbaren Phasenwinkel regelbar.
• Die Regelung des Phasenwinkels wird abgesehen von den Trägheitsmassen prinzipiell erschwert durch die Tatsache, daß bei dem Stoß des Schwingtisches gegen die Palette die Rotations-Geschwindigkeit der Unwuchtwellen stets eine ruckartige Veränderung erfährt, wobei wegen der vom Phasenwinkel abhängigen Relativlage der Unwuchtkörper während des Stoßes die Geschwindigkeits- und damit Drehwinkel-Veränderungen unterschiedlich ausfallen.
• Die Regelung des Phasenwinkels geschieht dadurch, daß die Drehgeschwindigkeit der Unwuchtwellen relativ zueinander geregelt wird. Dies bedeutet, daß eine gleichzeitige Regelung von Phasenwinkel und Schwingungsfrequenz praktisch nicht gleichzeitig und nur schwer zu erreichen ist.
• Es ist erwünscht, ein Verfahren anwenden zu können, bei dem während des Vorganges der Hauptverdichtung ein vorgegebener Bereich der Verdichtungsfrequenz bis hin zu höchsten Frequenzen mit vorgegebenen Werten für die Schwingwegamplitude des Schwingtisches durchfahren wird. Bei diesem Verfahren können die in dem Formstoff enthaltenen und durch die unterschiedlichen Korngrößen definierten Mikro-Schwingsysteme mit unterschiedlichen Eigenfrequenzen zu Resonanzerscheinungen angeregt werden, wodurch die Verdichtung verbessert wird. Das Durchfahren des Frequenzbereiches muß dabei in ca. 3 Sekunden durchführbar sein. Beim Stand der Technik wird die Durchführung dieses Verfahrens behindert durch die Begrenzung der Schwingungsfrequenzen des Schwingtisches und durch die schlechte gleichzeitige Regelbarkeit von Schwingfrequenz und Schwingwegamplitude.

The publication EP 0 515 305 B1 describes a relative to the oscillation amplitude (amplitude here relevant to the compression acceleration) and the oscillation frequency adjustable directional vibrator with 4 unbalanced shafts of a compression device according to the first class. The 4 unbalanced shafts are each driven by their own drive and adjustment motor via cardan shafts. The adjustment of the oscillation path amplitude defining phase angle is done exclusively via appropriately adjusted engine torques, which generate reactive power at a deviating from the value 0 ° or 180 ° phase angle (as for example in the DE 40 00 011 C2 is described). As disadvantages in such an unbalanced vibrator and compacting method, the following features should be mentioned:

• The upper oscillation frequency is limited in practice because of the duration load limit to be considered usually to 50 Hz, the limit load is achieved especially in the rolling bearings of the unbalanced shafts and the mitschwingenden cardan shafts. See also the above-cited journal article on page 45, middle section and on page 47, middle section.
• Due to the reactive power to be constantly implemented and the high bearing friction generated at high centrifugal forces high power losses occur. Since the high power losses must be implemented in the drive motors of unbalanced shafts, the motors and their control units are unnecessarily large dimensions with respect to the pure compaction performance.
• Due to the inertia masses of the motors and unbalanced body to be overcome and due to the fact that a change in the phase angle is always accompanied by a change in the reactive power torque which is also to be corrected, the values of the phase angle (static torque) given as the controlled variable can pass through the electronic control (or by alternative mechanical controls) are controlled only with rough tolerances, resulting in corresponding non-uniformities of the vibration path of the vibrating table during the course of many periods of vibration compression process and thus to a poor reproducibility of the compression quality. Added to this is the disadvantage that of the coarse tolerances of the controlled variable "phase angle" the relative angular position of a total of 4 unbalanced bodies is concerned, which usually lie with their axes of rotation in a plane and the Anodnung extends over a large part of the longitudinal extent of the vibrating table. The inequalities of the relative angular positions leads to unequal accelerations relative to the entire table surface. This in turn leads to unequal compaction results at different locations of the table surface.
• The decisive for the compaction effect vibration amplitude of the vibrating table is only indirectly and sluggish adjustable over the adjustable phase angle.
• The control of the phase angle is apart from the inertial masses in principle complicated by the fact that in the collision of the rocking table against the pallet, the rotational speed of the unbalanced shafts always undergoes a sudden change, and because of the dependent phase angle of the relative position of the unbalanced body during the impact of the speed - And turn angle changes vary.
• The control of the phase angle is achieved in that the rotational speed of the unbalanced shafts is controlled relative to each other. This means that a simultaneous control of phase angle and vibration frequency is practically not simultaneous and difficult to achieve.
• It is desirable to be able to use a method in which, during the process of main compression, a predetermined range of the compression frequency up to the highest frequencies is traversed with predetermined values for the oscillation path amplitude of the oscillation table. In this method, the micro-vibration systems contained in the molding material and defined by the different particle sizes with different natural frequencies can be stimulated to resonance phenomena, whereby the compression is improved. The passage through the frequency range must be feasible in about 3 seconds. In the prior art, the implementation of this method is hampered by the limitation of the vibration frequencies of the vibrating table and by the poor simultaneous controllability of oscillation frequency and vibration amplitude.

By the mentioned publications DE 44 34 679 A1 respectively. EP 0 870 585 A1 The present invention is therefore not suggested, because here quite differently working compression devices (Schüttelverdichtung or harmonic compression) are described with different compression mechanisms. That in the DE 44 34 679 described spring system of the rocking table, as far as a power transmission is provided by the springs in both directions of vibration, not serve as a model, since in the described spring system
Spring elements 116 are provided, which also work as compression springs and tension springs. This means a twice as high stress on the springs due to stresses compared to a design in which springs are loaded only on pressure. In addition, the power connection of a spring loaded with pressure and tension at its ends with the frame (or the foundation) of the compression device on the one hand and with the swing table on the other hand is very problematic and not permanently executable at a highly dynamic operation provided here. The in the DE 44 34 679 shown hydraulic excitation actuators must also take over the function of a linear guide of the vibrating table with. Since the vibrating table tends to constantly change inclinations under the pallet during a collision operation, this means a high mechanical load on the excitation actuators due to the function of the linear guide assigned to them, which is further increased by the inclination occurring in this case with two existing linear guides for clamping.

The by the publication EP 0 870 585 described compression device can also have no role model function with respect to the following functions: The hydraulically formed system spring can exert a spring action only in a downward swinging motion and the use of the same fluid medium for the hydraulic exciter and for the hydraulic spring demonstrably leads to significant energy losses even when exercising the spring function. As can be seen from column 2, lines 25 to 30, the spring constant should obviously be changeable only for the purpose of adjusting the compression method to the differently sized products to be compacted occurring different masses to the fixed predetermined natural frequency of the mass-spring system restore. A change in the natural frequency during the compression process is not provided.

Out DE-A-2 041 520 a device according to the preamble of claim 1 is known.

The object of the invention is to eliminate or reduce the disadvantages described above in the prior art, in which the compression energy is introduced predominantly by impacts of the vibrating table from below against the pallet in the molding. High shock frequencies should be applicable and the compression device should work with a in a wide range (even during the compression process) adjustable compression frequency up to highest frequencies of 75 Hz and higher with long life of the components involved and with low energy consumption. At the same time, with the means of the invention, the repeating accuracy of the generation of the compaction acceleration by the impacts on the pallet or on the underside of the moldings themselves and the uniformity of the distribution of the compaction acceleration over the entire surface of the pallet are to be improved.

The solution of the problem is described in the independent claims 1 and 26. Further advantageous embodiments of the invention are defined in the subclaims.

The invention uses, inter alia, the following principle: In the conventional generation of oscillatory movements of the vibrating table with the use of springs, which serve only the vibration isolation and are therefore set soft, the acceleration forces to be applied to the oscillating masses, predominantly by directed centrifugal forces of the unbalanced body generated. When generating the oscillating movements according to the invention, the acceleration forces are applied at least in that case, where they must reach the highest values at the highest vibration frequencies, predominantly by spring forces and only to a smaller extent by the excitation forces of the excitation means. This is achieved by utilizing the effect of resonance enhancement. In a further embodiment of the invention, this effect is exploited even better that it is provided to be able to produce at least a second natural frequency of the mass-spring system in the operatively covered range of vibration frequencies in addition to lying in the region of highest vibration frequencies natural frequency. This leads, as shown in Fig. 6, that the necessary excitation forces can be further reduced, which also facilitates, inter alia, the use of marketable AC linear motors and also the ability to vary the compression frequency over a wide frequency range during a compression process.

To store the entrained in the upward swinging motion of the vibrating table kinetic energy of the system mass and spring elements may be included in the spring system, the spring force from above the pallet acting, including such spring forces count, which are applied via the press plate with , If these are spring forces which are not passed over the press plate, as e.g. In the case of the springs 124 in FIG. 1, these contribute to the fact that the vibration displacement amplitude of the vibrating table or the mold can be regulated according to predetermined values even when the compression system oscillates during idling or in the pre-compression. The kinetic energy storing spring elements of the system spring have to save compared to the soft set insulation springs in the conventional compression systems a much higher amount of energy. Not only in the interest of their life (risk of self-destruction by heat) but also in order to avoid unnecessary energy losses, the spring elements of the system spring are therefore preferably made of steel or a low-damping elastomer material or are embodied by a (low-attenuation) liquid compressible medium.

The use of adjustable with respect to their static moment unbalance vibrators as excitation actuators in the invention makes quite a sense, since even here higher than conventionally achievable exciter frequencies all here interesting properties of the vibrator static static moment can be kept lower than in a due to the use of resonance amplification Vibration excitation only by the centrifugal forces of an unbalance vibrator. This means: Smaller bearing forces of the imbalance shafts, with lower bearing forces turn bearings can be used with higher allowable limit speeds. Smaller moments of inertia of the unbalance body itself and the drive motors of imbalances, with smaller moments of inertia improve the controllability of the phase angle. Smaller bearing friction power losses and smaller reactive powers, where the reactive power depends on the square of the magnitude of the static torque. Possible closer arrangement of the unbalanced shafts, this feature due to the improved central attack of the centrifugal forces leads to less irregularities in the acceleration of the vibrating table due to incorrect rotational positions of the unbalanced body.

The following definitions apply to the terms "hard" and "soft" springs used in connection with the spring system: A soft spring is used to isolate the acceleration effect of oscillating masses. The value of the "magnification function" Φ calculated according to a known formula (eg shown in Diagram 6.3-5 on page 300 of the " Physikhütte, Volume 1 ", 29th edition, Verlag Wilhelm Ernst & Sohn, Berlin, Munich, Dusseldorf ) must be Φ ≤ 1 for soft springs. This value is reached when the ratio η = f _E / f _N ≥ 1.41, where f _{E is} the excitation frequency and f _{N is} the natural frequency. However, for a reasonable isolation, at least one value of η = f _E / f _N ≥ 2 is generally required. In other words, the exciter frequency f _E (= compression frequency) must always be between the value f _E = 0 and the value f _E = 1.41 * f _N , optimally in the range f _E = f for a spring set hard for the purpose of using the resonance effect _N are. The exciter frequency f _E must always have a value of f _E = greater than 2 * f _N for a spring set soft for the purpose of insulation. A hard-set system spring in the case of the present invention means that the effect of the magnification function Φ should be claimed for values Φ> 1. The statement in claim 1, that the system spring is set hard at least for the downward swinging motion, states that a system spring can also be constructed such that in both directions of vibration different spring constants are effective. Example of hard and soft set springs: According to a known relationship q = 248.5 / f _N ² and q (in mm), the deflection q of a mass stored on a spring with the natural frequency f _N (in Hz) can be determined under its own weight become. If the natural frequency of a "hard" system spring is at least 30 Hz (or higher), the deflection q below the system mass can be calculated as: q = 0.27 mm (or smaller). If, at a lowest permissible excitation frequency of a compression device with soft insulating springs, the insulating springs are correctly selected, then the natural frequency achievable with their spring constant should be at most 15 Hz. In this case, the value would be q = 1.1 mm.

By the proposed possibility of regulating the amplitude of the oscillation travel s of the vibrating table is resorted to the proven in practice in the prior art influencing this physical quantity by controlling the phase angle in the sense of influencing the compression intensity. The phase angle also indirectly determines the value of the oscillation travel amplitude s, which physically is the actual measure of the compression intensity actually to be controlled. The metrological determination of the phase angle, which is defined by the relative angular position of rotating unbalanced bodies, is complicated and subject to noticeable measurement errors. Different to In the prior art, however, in the invention when using linear motors as exciter actuators, the value of the oscillation amplitude s is not influenced indirectly by the detour of another variable to be controlled, but it is directly controlled (and measured directly), which together with the fact that not at the same time a changing reactive power torque is to control, leads to a more precise controllability of the compression intensity. When using hydraulic or electric linear motors, these can be subjected to such a force that even if several linear motors are used with parallel action, their force development is precisely symmetrical, so that only because of their multiple arrangement no asymmetrical accelerations occur on the vibrating table.

It is desirable that, when the value of the oscillation travel amplitude s is influenced, the oscillation frequency can also be changed in a predeterminable manner. This object is made possible in the present invention by the good controllability of the oscillation amplitude s in combination with the possibility given in the invention that not a rotational speed must be changed, but only a repetition frequency in the dosage of certain amounts of excitation energy per oscillation period, which Case of hydraulic linear motors is very low in inertia and in the case of electric linear motors can be done almost inertia.

The application of electric (three-phase AC) linear motors is very advantageous because they are a "clean" and operating with low power losses solution. However, the marketable electric linear motors are not readily usable for the intended task, since they are provided with their series-produced actuators to perform linear movements with a predetermined course and speed curve and thereby automatically generate those forces that are used for the acceleration of the moving masses or which are required for overcoming the linear displacement opposing forces (usually machining forces). The typical application for such linear motors is given in machine tools. The normally purchasable control devices must therefore be replaced by a special control device. The main differences in the use of the linear motors in the invention in comparison with the conventional tasks are given in the following features: The acceleration and deceleration of the oscillating masses, including the mass of the resonant motor part of the linear motor, in the compacting device predominantly, in particular, when the excitation frequencies are in the vicinity of the natural frequencies, determined by the forces of the system spring (in resonance mode). Therefore, a usual in the linear motors control device for Generation of a programmed sequence of movements is not used because it does not know the spring forces and can not influence and because the engine forces alone are not sufficient for the accelerations to be generated by far.

In contrast, in the present invention task has the linear motor per oscillation period (after once started up vibration) pass on only those amounts of energy to the system mass, which deprived the oscillating system mass by friction or by the compression energy delivered at the impact become. In the case of a vibration path amplitude which is to be kept constant, therefore, it is important to reintroduce with each oscillation period the oscillating system mass that energy portion which is required to maintain the given oscillation path amplitude. The power development on the linear motor does not have to follow a size determined by the time of oscillation time function (eg rectangular or sine function), since only the (per period) transmitted energy portion is crucial, of course, the timing of the beginning and end of the force development also play a role and have to be determined by the controller. The control device must also be able to take into account the phenomenon of the occurrence of a phase shift angle γ and the change of its value which occurs automatically as the compression process progresses (the phase shift angle γ defines the angle by which the swing path amplitude lags the exciter force amplitude), which incidentally also affects the hydraulic linear motor Control applies. Since the time of measurement of the physical variable s, s ', s "or f, f', f" to be controlled, and the time of conversion of the value derived therefrom by a control algorithm for the manipulated variable y (for determining the size of the next zu transmitted energy portion) is not identical, measured values and / or derived values must be temporarily stored temporarily.

It is advantageous to limit the vibrating table in its three-dimensional freedom of movement not only by the system spring, but to force the same to enforce a rectified acceleration of all parts of the vibrating table by a single central linear guide. The linear guide, which is optimally a cylindrical guide, has to absorb all horizontal acceleration forces, which may result, for example, from the impact. Such a linear guide can also be dispensed with when using an electric linear motor if the air gap existing in the motors between the fixed part and the movable part is still able to absorb the horizontal deviations of the vibrating table. When using a hydraulic linear motor and when using hydraulic cylinders of conventional design should not be waived on a linear guide, however, unless hydraulic cylinders and linear guide are integrated by appropriate design measures in a unit. A linear guide has not only the advantage that it ensures a uniform distribution of the shock accelerations, but it also results in a reduction of mold wear.

The particular advantages of the invention can be summarized as follows: elimination or reduction of the mentioned disadvantages of vibration displacement amplitude adjustable unbalance vibrators, associated with an increase in the quality of the compression process by greater reproducibility of the result in the conversion of the kinetic oscillatory energy in compression energy. High achievable vibration frequencies. Lower necessary excitation power. Especially when using linear motors as excitation actuators, the excitation energy is converted directly into compression energy and energy is saved by eliminating the reactive power and the storage friction power. Continuous fast adjustability of the compression frequency with simultaneous control of the oscillation travel amplitudes.

Special advantages arise when using an electric linear motor instead of a hydraulic linear motor by the following features: The electric linear motor work virtually wear-free. The development of the excitation forces is particularly low in inertia feasible, which is why these linear motors are also more dynamic and accurate control. The force curve does not need to be sinoidal, as is the case with the hydraulic linear motor, in practice by the use of servo valves. The collision of the swinging table against the pallet results in high damaging pressure peaks in a hydraulic linear motor. The electric linear motor is advantageous in this respect, because the force jumps in the elastic field of the air gap are effective and because electrical surge voltages can be absorbed by electrical means.

The invention will be explained in more detail with reference to 6 drawings. Fig. 1 shows schematically a compacting device of the first type, in which the vibrating table abuts against the pallet once from below during each oscillation period. In Fig. 2 in the upper part of the drawing the same swinging table as shown in Fig. 1, but connected to another system spring, wherein the lower spring system shown in Fig. 1 is replaced with respect to the spring constant adjustable spring system with a single leaf spring resilient element. Fig. 3 shows details of another variant of the compression device according to Fig. 1, which is about additional switched on and off spring elements.

In Fig. 4, other possibilities of development of a compression device according to Fig. 1 are shown. 5 shows a diagram with the profile of the oscillation travel amplitude A over the excitation frequency f _{N of} the system mass of a compression device according to the invention with a single natural frequency for explaining possible amplitude control. FIG. 6 shows a diagram similar to that of FIG. 5, wherein the advantage of an additional natural frequency of the oscillating system is explained.

In Fig. 1 , 100 is the frame of the compacting device which stands on the foundation 102 and by which the forces to be transmitted by the pressing device 104 and the exciter device 106 are supported against each other. The frame may in this case be firmly connected to the foundation, which is represented symbolically by the lines 190, but with small mass of the frame considerable excitation forces are to be transmitted to the foundation. The molded body 110 enclosed in the molding recess of the molding box 108 lies with its underside on a pallet 112. The pallet itself rests on an impact bar 114 secured to the frame 100 (and indicated by hatching for the sake of clarity), which is provided with recesses 116 through which the bumpers 118 of the vibrating table 120 pass and in the oscillating motion of the vibrating table after overcoming of the air gap 122 may collide against the underside of the pallet. The resting on the pallet molding box 108 is pressed by springs 124 which are supported by lugs 126 against the frame, firmly on top of the pallet 112. In this way, the mold box maintains a firm connection with the pallet even in the case where the pallet is pushed upwards by the bumpers 118 and thereby can lift off the baffle bar 114. However, the molding box could also be clamped tightly to the pallet (by a clamping device, not shown). The oscillating table 120 forms with its mass the main part of the system mass of the oscillatable mass-spring system 140, whose vibration forces are absorbed or generated primarily by the associated system spring 142.

The system spring consists of an upper spring system 144, through which at least a portion of the kinetic energy entrained in the upward swinging motion is stored, and a lower spring system 146, through which the majority of the maximum kinetic energy entrained in the downward swing motion is stored. The upper spring system 144 and the lower spring system 146 consists of a plurality of spring elements 148 and 150, which may also be variable or adjustable with respect to their spring constants, which is symbolically indicated by the arrows 152. The spring elements 148 and 150 may be formed as compression springs, thrust springs, torsion springs or torsion springs and are braced in the case of FIG. 1 against each other, that they still have a residual spring deformation even with the largest vibration amplitudes of the system mass to be performed. The forces of the spring elements 148 and 150 are clamped at one end between parts of the frame 100 and supported at the other ends against a power connector 154, which is part of a power transmission member 156, with which the forces of the upper and lower spring system on the system mass be transmitted. It is advantageous to transmit the forces of the spring elements of the spring system at least at those ends at which the forces of the springs are transmitted into the system mass by pressure forces and / or shear forces in the power connection parts, since these points with respect to the reliability and durability critical Points are, which quickly fail when connecting the spring elements to the power connector parts at predominant application of tensile forces at this point.

The exciter device 106 comprises an excitation actuator 170, comprising a fixed actuator part 172 connected to the frame 100, a movable actuator part 174 connected to the system ground, and a drive device 196, which also includes a regulator 198. By means of the control device, the energy transfer means (electric current or hydraulic volume flow) are shaped or controlled such that when applying a predetermined constant or variable excitation frequency by the movable actuator 174 at each half-period or full period of the vibration excitation forces and thus excitation energy portions on the Mass-spring system are transmitted, whereby this is forced to carry out vibrations and for the delivery of impact energy for the compression process. Depending on the size of the set air gap 122 (which may also be set to the value zero or a negative value), the oscillation travel amplitudes A are to be generated with such a size that sufficient impact energy is transferred for the compaction taking place in a manner known per se can. Preferably, the physical vibrational quantity defining the transferable compaction energy, e.g. the oscillation travel amplitude A, be controlled or regulated, even at constant held oscillation frequency.

The pressing device 104 comprises a fixed part 182, a movable part 184, to which the pressing plate 180 is connected and a (not shown in the drawing) control part for carrying out a direction indicated by the arrow 186 vertical adjustment movement of the pressing plate.
The forces of the upper and lower spring system receiving parts of the frame 100 could also be separated together with the forces of the exciter 106 receiving parts of the frame of the frame 100 and together on one of the Foundation 102 separately existing, special (not shown in the drawing) foundation part angeodnet, which foundation part in this case (serving as damping mass) preferably (not shown in the drawing) insulation springs against the foundation 102 would be supported. The exciter device 106 with its exciter actuator 170, which is required to be able to transmit variable amounts of energy into the oscillating system together with a drive device even when the exciter frequency is kept constant, can be embodied in different variants. The exciter actuator may be an unbalance directional vibrator that is controllable with respect to the static torque, or a linear motor that is hydraulically or electrically operated with respect to the convertible excitation energy portions. To measure the vibration amplitude A to be controlled, a measuring device is provided which consists of a part 192 fixed to the frame and of a part 194 connected to the vibration table. The signal of the measured magnitude is supplied to the controller 198 for processing (not shown).

Hydraulic or mechanical springs are provided in the upper spring system 144 and / or in the lower spring system 146, the spring constants of which are constant in the simplest case and given a resultant system spring whose natural frequency is at a particular location, e.g. may be located in the middle of the frequency range of the exciter frequency, whereby at this point a resonance point is formed. Although the resonant effect of the amplitude amplification to be exploited at the resonance point is greatest at the resonance point, the rerender effect should also be above and / or below in the dimensions then attenuated in accordance with the resonance curve (with the possibility of continuously passing through the exciter frequency through a predetermined frequency range) Be used resonance point. Due to the resonance effect, the vibration acceleration of the system mass takes place predominantly with the assistance of the spring forces or with the assistance of the energy quantities stored in the springs. This has the advantage that these forces and the energy amounts to be assigned to them no longer have to be generated by the excitation device, which has a significant effect on the size of the exciter device and on the size of the lost energy converted in this. In the ideal case of the equality of excitation frequency and natural frequency of the excitation device only the vibration system by the friction losses and the vibration system extracted as compression energy loss energy must be implemented.

It can be seen that it must be of great advantage if each exciter frequency within the frequency range of the adjustable exciter frequency could be assigned a natural frequency of the system spring. This ideal case is intended according to the invention with a continuous adjustable natural frequency of the system spring can be achieved, with the adjustment of the excitation frequency f _E, the natural frequency f _N can be mitverstellt simultaneously while maintaining any value for η = f _E / f _N. Alternatively, instead of a continuously adjustable natural frequency with less effort, a stepwise adjustment of the natural frequency come into question.

The spring constant of the system spring is always to be understood as a resulting spring constant C _R , which results from the spring constants of all spring elements participating in the system spring. The resulting spring constant C _R can be defined by determining the resulting natural frequency along with the system mass. In a gradual change in the resulting spring constant (during standstill or during compression) can be provided, for example, that one or more springs are constantly fully in use or turned on and that these springs are constantly switched on gradually other springs in addition to the power transmission of Vibrational forces are involved. This can be done, for example, that springs of different spring constants are switched so that their deformation completely coincides with the vibration path of the system mass, or even such that their deformation path only makes up a predictable and adjustable portion of the vibration path of the system mass. In the latter case, it is then an adjustment of the "progression" of the spring characteristic of the resulting spring constants. When using a gradually adjustable or variable progression operating system spring it should also be possible according to the invention, caused by the changes in the resulting spring constant change in the physical variables of the oscillating system (eg vibration amplitude A) using a specially equipped for this purpose control device for the exciter to smooth out or correct the influence parameter of the excitation energy to be supplied or removed in the sense of keeping the physical quantities constant. An on and off switchable spring is explained in more detail in FIG.

If the lower or upper spring system is designed as a spring system with respect to its resulting spring constant and the resulting spring constant of the lower or upper spring system is determined by at least one non-adjustable and at least one switchable adjustable spring, it can be achieved thereby reducing the effort that the Adjustment range of the natural frequency starts at a certain frequency upwards. This is sufficient for the needs of the practice, where, for example, an adjustment of the natural frequency can be provided from about 30 Hz to 75 Hz.

An adjustable mechanical spring element is described below in FIG. 2. An adjustable hydraulic spring element can be provided in that a spring element of the system spring is embodied by a compressible pressure fluid volume (hydraulic oil) clamped at least partially in a cylinder body by a spring piston and in that the spring rate is variable by changing the size of the pressure fluid volume. either by the fact that the size of the pressure fluid volume is formed by a plurality of separable from each other by switchable check valves sub-volumes, or in that a portion of the pressure fluid volume is clamped in a cylinder whose cylinder space is variable by one in the cylinder after predetermined Way and preferably continuously displaceable piston, wherein the displacement of the piston z. B. is performed by a threaded spindle drive.

FIG. 2 shows a variant of the vibratory mass-spring system shown in principle in FIG. 1 with the system mass and with the system spring of another type. An exciter device is not shown for the sake of simplicity and it could be imagined in the form of two serving as excitation actuators linear motors in addition to the vibrating table 120 attacking. In the upper part of Fig. 2, the components whose reference numbers begin with the numeral 1, identical to the components of the same name in Fig. 1. The vibration-transmitting terminal body 202 may be identical to the frame 100 shown in FIG. The system spring has in this case an upper spring system 144, consisting of compression springs 124 and a lower spring system 244, which has a respect to their spring constant adjustable and predominantly subjected to bending leaf spring 282. The dynamic mass forces (or spring forces) to be exchanged between the leaf spring 282 of the lower spring system and the oscillating table 120 at a vibration of the system mass in the direction of the double arrow 230 in a downward swinging motion are guided via the vibrating punch 280, which is located at the top of the vibrating table 120 is attached and at the lower end has a curve, with which it snuggles into the curve 284 of the leaf spring, the lower end acts as a force introduction element of the first kind on the center of the force Fm under exclusive generation of compressive forces at the force introduction point 209 center is inserted into the leaf spring. A (preferably provided) even with the largest vibration displacement amplitudes A still existing bias on the springs 124 and the leaf spring 282 ensures that the contact between vibrating punch 280 and leaf spring 282 is never lost. The dynamic forces acting on the leaf spring at this attacking mass forces Fm are on the at equal intervals L1 below the leaf spring at the force introduction points 211, 211 'arranged roller-shaped force introduction elements second type 210, 210 'in half under exclusive generation of compressive forces as supporting forces Fa transmitted.

The main extension direction of the leaf spring is symbolized by the double arrow 240. The roller-type force introduction elements of the second type 210, 210 'are mounted in roller carriers 212 and 212'. The double arrows 216 and 216 'indicate that the roller carriers can be displaced in both directions and incidentally also under the pulsed load by the supporting forces Fa. During their displacement, the force introduction elements of the second type 210 and 210 'are also allowed to rotate, which is indicated by the double arrows 218, 218'.

The displacement of the roller supports 212 and 212 'in respectively opposite directions is performed synchronously, which is effected by a threaded spindle 220 with counter-rotating thread. The threaded spindle 220 is driven by a motor-driven drive unit 222, which in turn is controlled by a (not shown) control. By means of the control and the drive unit 222, the roller carriers 212, 212 'and thus the introduction points of the second kind 211, 211' for the support forces Fa can be brought into any predeterminable positions, for example in order to achieve e.g. to make the distances L1 or L2. The roller carriers brought into the positions L2 are indicated by dashed lines. The distances L1 and L2 refer to the introduction point of the first kind 209. It is obvious that with the arbitrarily adjustable positions for the introduction points of the second type 211, 211 '(within certain limits) arbitrarily and continuously adjustable spring constants of the leaf spring are connected.

Fig. 3 shows a variation of the compression device according to Fig. 1, wherein two similar additional spring systems 300 and 300 'are shown with additional switched on and off spring elements which are arranged between the vibrating table 120 and the foundation 102 to transmit power. In a power transmission part of the second type 302, two spring elements 304 and 306 designed as compression springs and also in the switched-off state are arranged such that they transmit their spring forces to a lower cantilever part of a first type of power transmission part 308. The power transmission part of the first type is connected via an upper Kragteil fixed to the vibrating table and destined to transfer the resulting force resulting from the deformation of the spring elements on the vibrating table. The power transmission part of the second type 302 is fixedly connected to a piston 312 of a hydraulic switching device 310, whereby it is capable of depending on the switching state of the switching device resulting from the deformation of the spring elements resulting force on the fixed to the foundation To transfer or not to transfer cylinder 314 on the foundation 102. The piston 312 can be moved up and down in the cylinder 314 at a first switching state, almost without transmitting any force, or it can be firmly clamped by the fluid medium in a second switching state in the cylinder. The switching states of the switching device 310 are determined by the position of the valve 320. In the illustrated position, the cylinder chambers 316 and 318 of the cylinder 314 are connected via the valve, so that the piston in the cylinder can move up and down without constraining forces. In a second position of the valve, the cylinder chambers are closed, so that the force of the power transmission part of the second type 302 is transmitted directly to the foundation.

4 , other possibilities of development of the invention are shown, wherein the different functions in the compression device of FIG. 1 can be arranged and are connected on the one hand to the rocking table 120 and the other with the frame 100 (or the foundation 102).

The oscillating table 120 is fixedly connected to a central guide cylinder 412, the center axis of which passes through the center of gravity of the oscillating table and which is freely movable with its outer cylinder in the inner cylinder of a cylinder sliding guide 414. As a result, a linear guide 410 is formed, which constitutes a forced guidance of the oscillating table for executing the oscillating movement on a straight line only in a double direction with a guide part arranged centrally and mirror-symmetrically on the oscillating table. As exciter actuators two identical linear motors 420 are provided, which can be acted upon by a special drive means, not shown, so that they generate excitation forces in the vertical direction. Each linear motor 420 consists of a fixed motor part 422 and a movable motor part 424, both of which are separated by an air gap 426. The movable motor part 424 is fixedly connected to the swing table 120 via a support part 428, while the fixed motor part 422 is fixed directly to the frame 100. The linear motors 420, which are preferably designed as three-phase AC motors, are controlled via the special control device such that a physical variable of the oscillation profile of the vibrating table 120 or the mold 108 (in FIG. 1) is controlled according to predetermined values, and thus indirectly also the course of the compacting process or regulated.

430 shows a spring system which, at least in the pre-compression, optionally together with the spring elements 124 shown in FIG. 1, represents the system spring. This system spring develops in this case with its special, from a Elastomeric material produced thrust spring 434 spring forces in two directions for the storage of in both directions of vibration through the system mass entrained kinetic energy quantities. The thrust spring 434, which in this case is designed as a hollow cylinder, is externally connected to a spring ring 432 and internally to a cylinder 436, which is fastened to the guide cylinder 412. The spring ring 432 is supported in terms of strength over two holders 438 firmly against the damping mass 450, but the support could also be made against the foundation 102 or the frame 100. It can be seen from the arrangement of the spring system 430 that this could also take over the task of the linear guide 410 at the same time. In other words, a spring system with a thrust spring, which can develop spring forces in both directions of vibration, can also be provided as a linear guide and exercise the function of a forced operation to perform the oscillating motion of the vibrating table in a double direction, provided that the spring forces with a centrally arranged on the vibrating table guide member be transmitted.

With 440 an additional and disconnectable additional mass is referred to, with which the size of the system mass can be changed in order to change the natural frequency of the mass-spring system can. Within the additional mass, a hydraulic cylinder 442 is housed, in which there is a piston 444 which is fixedly connected to the cylinder 436 and thus to the system ground. By the piston 442 two displacement chambers are formed in the hydraulic cylinder, which can be individually shut off via a switchable valve 446 or connected to each other. In the event that the displacement chambers are interconnected, the piston 444 can move freely in the cylinder 442 up and down, without that the additional mass would be moved along. If the displacement chambers are shut off individually, the additional mass 440 is forced to resonate in synchronism with the system ground. In this case, the springs 448 are only small forces transmitted to the damping mass (or the foundation), since they are designed as soft springs, which only have to keep the additional mass at a certain height, if it is not resonant. Unlike in Fig. 1, where the system spring 142 is supported in terms of strength against the frame 100, the system spring 430 is supported in Fig. 4 against a particular damping mass 450, which in turn again over soft set springs 452 against the frame 100 and Foundation 102 supported. With this measure, depending on the dimensioning of the additional mass achieved that derived from the system spring 432 vibration forces, which can reach peak values of about 20 tons at 70 Hz, eg at a system mass of 1000 kg and a vibration amplitude of 1 mm, only reduced can get into the foundation.

Fig. 5 shows a diagram with the course of the vibration amplitude A over the excitation frequency f _{N of} the system mass of a compression device according to the invention (eg Fig. 1) with a single, located at about 70 Hz natural frequency and with a certain damping D1 for the Curve K1. In this diagram, a sinusoidal excitation force with a constant exciting force amplitude over the entire range of the excitation frequency is provided. Damping D1 takes into account the frictional losses and the energy losses of the oscillating system due to the compression energy delivered. The curve K1 represents the known resonance curve. The excitation force is able to produce an amplitude of A = 0.36 mm in the range of very low frequencies. In the natural frequency range, the same excitation force generates an amplitude of A = 1.8 mm, which corresponds to an amplitude amplification (resonance amplification) of Φ = 5. If one wanted to achieve the same amplitude of 1.8 mm at lower excitation frequencies, for example at 58 Hz, then the value of the exciter force amplitude would have to be increased by a factor of 1.8 in this case. 5 shows two different methods of controlling the amplitude A according to a predetermined value at a given natural frequency of 70 Hz:
In a first method (which of the in the publication DE 44 34 679 A1 The method is similar to the above-mentioned method, although the oscillation path amplitude A is not to be regulated there), the force excitation is carried out by an unbalance directional vibrator which is not controllable with respect to its static torque and which should operate with a nominal excitation frequency of 63 Hz, the centrifugal forces then developed ( the exciter force amplitude is set to = 100%) produce an amplitude of A = 1.4 mm (point Q on the curve K1). Increasing the excitation frequency from 63 Hz to 70 Hz will increase the amplitude to A = 1.8 mm (and reduce the excitation frequency to 58 Hz to reduce the amplitude to A = 1 mm). As can be seen, this first method involves changing the excitation frequency in order to change the amplitude A. Conversely, the amplitude A changes automatically when passing through a certain range of the excitation frequency.

In a second method, the force excitation is generated by a variable in its excitation force amplitude linear motor whose excitation frequency is set to 63 Hz and its excitation force amplitude to 100%. The achievable vibration displacement amplitude in this case is also A = 1.4 mm. However, the change in the amplitude A is achieved here by changing the exciter force amplitude (a) while maintaining the exciter frequency (of 63 Hz). In order to adjust the amplitude A to a value of A = 1.8 mm, the exciting force amplitude (a) must be increased so that a very different Resonance curve K2 is generated whose intersection with the 63 Hz line reaches the value of A = 1.8 mm. In order to set an amplitude of A = 1 mm at 63 Hz, a different type of resonance curve K3 must be generated by reducing the exciting force amplitude (a). It can be seen that, in contrast to the first method, an arbitrarily predeterminable amplitude A can be achieved independently of the excitation frequency. At the same time, the application of the second method also makes it possible to change the excitation frequency within a predetermined frequency range as desired (also continuously) according to a predeterminable time function and additionally to generate arbitrarily predefinable amplitudes A. The second method is that used in the present invention. When using this second method, the periodic excitation force does not necessarily have to be generated following a sinusoidal function. Decisive for the generation of a specific amplitude A at a predetermined damping D is the amount of energy supplied via the exciter device per oscillation period. The temporal course of the excitation force could also follow a rectangular function instead of a sine function, it being possible to deduce from the amount of energy converted per period to a substitute exciter force amplitude (a *) for a sinusoidal course of the excitation force.

Fig. 6 shows a diagram similar to that of Fig. 5, wherein the curve K1 corresponds to the curve K1 shown in Fig. 5 and indicates a mass-spring system having a natural frequency at about 70 Hz. A second curve K4 represents the resonance curve of the same mass-spring system, in which case, however, the natural frequency (by varying the resulting spring constant of the system spring) is switched to a different value of about 46 Hz. The force excitation of the associated mass-spring system should, as in the second method described in FIG. 5, be achieved by generating the exciting force amplitude (a or a *) using a controllable linear motor, wherein the application of force to the exciter actuator should be controlled by a special control device, wherein the amount of energy to be converted should also be able to be influenced for regulating a predetermined value for the amplitude A (assuming a suitable measuring device for measuring the size of A). For curve K4, an equal excitation force amplitude was assumed to be K1, but an attenuation value D4 doubled compared to D1. As a result of the lower value of the spring constant, an amplitude of A = 0.78 mm is achieved even at a very low exciter frequency. The diagram shows that when using the rocking properties of both curves over a range of the excitation frequency of 27 to 78 Hz, a vibration amplitude of 1.1 mm can be achieved. This means, in comparison to the possibility given by curve K1 alone, an extension of that frequency range within which at least one equal amplitude can be set. This is for the present invention Used in a compression process, the excitation frequency, which in this case is identical to the compression frequency, (in the example of this diagram) is traversed by a value of 27 Hz up to a value of 78 Hz, the amplitude by the regulation of the pro Period of the excitation energy to be converted to a value of A = 1 mm. In a compression process, in practice, the damping value D changes continuously from a higher value (D4) to a lower value (D1). During the execution of the compression at continuously increasing excitation frequency is switched at a certain frequency to the natural frequency of 70 Hz corresponding spring constant. If the natural frequency can be adjusted in more than one step, optimally continuously, the described method can be further optimized by also adjusting the natural frequency with an altered excitation frequency, wherein the amplitude is simultaneously regulated according to a predetermined value for A. In such a method one could reach the given values for A with a significantly lower excitation energy in comparison to the oscillation excitation of conventional type.

Applies to all drawings of Figures 1 to 4, that fixed connections of two components are shown symbolically by dotted lines.

Claims (28)

1. Device for compacting granular materials to form moulded bodies by means of pre-compaction and main compaction in moulds (108), a pallet (112) or base plate being provided as a rest for the moulding material (110) and a pressing plate (180) being provided for subjecting the moulding material (110) to a pressing force on the upper side, and at least part of the overall compaction energy being able to be introduced from a vibrating table (120) into the moulded material (110) by impact processes, which can be generated by instances of impact of the oscillating vibrating table (120) from below against the pallet (112), characterized in that

   - the vibrating table (120) is part of an oscillatory mass-spring system (140) with a system spring (142), which is a spring set "hard", at least for the downwardly directed oscillating movement, and has a system mass, the main mass component of which is the vibrating table (120) with its associated co-oscillating members (156, 174),

   - the system spring (142) is a means of storing energy of at least part of the kinetic energy taken along as a maximum in the upward oscillating movement by the system spring (142), in the downward oscillating movement the system spring (142) having "hard"-set spring elements (150) being a means for storing the main component of the kinetic energy of the system mass taken along as a maximum,

   - the combination of the values of the resulting spring constant of the system spring (142) and the system mass has the effect that at least one natural frequency of the mass-spring system (140) which is in the range of the upper compaction frequency used in practice for the pre-compaction and/or the main compaction can be set,

   - the mass-spring system (140) can be driven by means of an exciter device (106), operating with periodic exciter force generation, to produce enforced oscillating movements, with at least one exciter frequency which can be given, the exciter energy that can be transferred by the exciter device being able to be set by a regulating device (196, 198) in such a way that, at least during idling of the compacting system, i.e. without moulding material (110) and without the pressing plate (180) resting in place, or at least during the operation of pre-compaction, i.e. without the pressing plate (180) resting on the moulding material 110, the physical variable of the upper or lower amplitude of the oscillating excursion s of the vibrating table (120) or of the oscillating excursion f of the mould or a variable derived from it of the oscillating velocity or oscillating acceleration s', f or s", f" can be directly or indirectly regulated or controlled according to a value which can be given, and

   - provided for the exciter device (106) are one or more exciter actuators (172/174), which are designed in the form of electrical linear motors (422/424) or in the form of hydraulic linear motors, or in the form of unbalanced vibrators which can be adjusted with respect to their static moment and the resulting directed centrifugal forces of which are at least 20% smaller than the accelerating forces required on the system mass for carrying out the intended oscillating excursion amplitudes with the intended maximum frequency.
2. Device according to Claim 1, characterized in that the spring elements of the system spring (430) storing the kinetic energy are produced from steel or a low-damping elastomer material (434) or are embodied by a liquid medium, which is preferably a hydraulic oil, securely enclosed in a compression chamber.
3. Device according to either of Claims 1 and 2, characterized in that, with involvement or non-involvement of the pressing plate in the transfer of compacting forces, co-operating in the resilient effect of the system spring (142) equipped with mechanical spring elements are:

   - an upper spring system (144), with one or more upper spring elements (148) which are predominantly subjected to compression and by which at least part of the kinetic energy of the system mass taken along as a maximum in the upward oscillating movement is stored for a short time, and a lower spring system (146), with one or more lower spring elements (150), which are predominantly subjected to compression and by which the main part of the kinetic energy of the system mass taken along as a maximum in the downward oscillating movement is stored for a short time, the forces of the upper and lower spring systems acting on the system mass,

   - and/or a spring system (430) with one or more spring elements (434), which are subjected to bending, torsion or thrust, so that both at least part of the kinetic energy of the system mass taken along as a maximum in the upward oscillating movement and the main part of the kinetic energy of the system mass taken along as a maximum in the downward oscillating movement is stored by the same spring element or elements (434), the forces developed during the energy storage acting on the system mass.
4. Device according to one of Claims 1 to 3, characterized in that part of the kinetic energy taken along in the downward oscillating movement while the previous impact process is being carried out can be stored by upper-lying spring elements (124), the spring forces of which are effective from above on the pallet (112), in this case the upper-lying spring elements (124) constituting part of the upper spring system (144).
5. Device according to one of Claims 1 to 4, characterized in that an adjustable mechanical spring element is a leaf spring (282) subjected to bending, in that a spring-effective spring length (L1, L2) is defined between a point of force introduction (209) of an introduced force Fm and a point of force introduction (210, 210') of a supported force Fa = Fm/2, and in that the adjustment is brought about by a variation of the spring-effective spring length (L1, L2), preferably using an auxiliary motor drive (222).
6. Device according to one of Claims 1 to 5, characterized in that, when the system spring is equipped with a hydraulic spring as the spring element, said spring is adjustable by changing the compressible spring volume in a compression chamber.
7. Device according to one of Claims 1 to 6, characterized in that the exciter energy that can be transferred by the exciter device (106) can be influenced by a regulating device (198) in such a way that, as an alternative to or at the same time as the operation of pre-compaction, also during the operation of main compaction the physical variable of the upper or lower amplitude of the oscillating excursion s, A of the vibrating table (120) or of the oscillating excursion f of the mould or a variable derived from it of the oscillating velocity or oscillating acceleration s', f or s", f' is regulated according to a value which can be given.
8. Device according to one of Claims 1 to 7, characterized in that a physical variable s, s', s" or f, f, f" is regulated according to a constant or variable value which can be given, for constant or variable exciter frequencies which can be differently given.
9. Device according to one of Claims 1 to 8, characterized in that the electrical linear motor or motors (170, 420) provided as exciter actuators (70) are AC motors, preferably three-phase AC motors, which are equipped with permanent-magnet excitation or designed as asynchronous motors and which have a fixed motor part (422) and a linearly movable motor part (424), and in that a physical variable s, s', s" or f, f', f" is regulated by the variable apportioning of the portions of energy supplied or removed in an oscillating period.
10. Device according to one of Claims 1 to 9, characterized in that, in the case of the linear motors (170, 420) designed as three-phase AC motors, the magnetizing current and the current forming the thrust force can be set as separate components.
11. Device according to one of Claims 1 to 10, characterized in that the electrical linear motors are three-phase AC motors with a special activating device (196, 198), which is designed for the generation of specific and influenceable portions of exciter energy per oscillating period.
12. Device according to Claim 11, characterized in that the following functions are alternatively or simultaneously executed by the special activating device (196/198) for the electrical linear motors (170, 420),

    - the beginning and end of the development of the motor exciter force and the magnitude of the motor exciter force are determined or calculated by the special activating device (196/198) once or twice within the oscillating period (of 360°) in time with an exciter frequency which can be given,

    - for the purpose of controlling the phenomenon of the occurrence of a phase shifting angle γ and the changing of the phase shifting angle γ automatically occurring under the influence of certain parameters, a special algorithm is used by the special activating device (196/198), which has the effect that the measured value of the physical variable s, s', s" or f, f, f' to be regulated and/or of the value derived from it by the control algorithm for the manipulated variable y for fixing the magnitude of the next portion of energy to be transferred is buffer-stored for a short time.
13. Device according to one of Claims 1 to 12, characterized in that, apart from the feeding of exciter energy into the oscillatory system via the exciter actuators, energy can also be extracted from the oscillatory system for delaying the oscillation process after an overshooting regulating process or for rapidly stopping the oscillation process.
14. Device according to one of Claims 1 to 13, characterized in that the at least one settable or set natural frequency of the mass-spring system is not greater than about 30% of the upper compacting frequency used in practice for the pre-compaction or the main compaction and/or in that the at least one settable or set natural frequency of the mass-spring system is above a value of about 30 Hz.
15. Device according to one of Claims 1 to 14, characterized in that, when electrical or hydraulic linear motors (420) are used as exciter actuators, the vibrating table (120) is guided in a constrained manner in its oscillating movement by a single central linear guide (410), to absorb horizontal forces on the vibrating table and to ensure a co-directed acceleration at all the parts of said vibrating table.
16. Device according to one of Claims 1 to 15, characterized in that, for the purpose of adjusting the natural frequency of the oscillatory mass-spring system, one or more additional masses (440) can be connected to and disconnected from the system mass by a switching operation, in such a way that, with the additional mass connected, this mass is co-oscillating synchronously together with the system mass, it being preferred for the switching operation to be carried out using a hydraulically actuated component (442/444).
17. Device according to one of Claims 1 to 16, characterized in that, for the purpose of changing the resulting spring constant of the spring system, the co-operation of one or more spring elements (304/306) can be additionally connected or disconnected during the operation of storing the oscillating energy, the spring elements to be switched being firmly connected to a first force transferring part (308), by which the spring force is transferred to the system mass, and connected to a second force transferring part (302), by which the spring force is transferred to the foundation (102) or to a special damping mass (450), the second force transferring part being able to be coupled to the foundation or to the damping mass by a switching operation of a switching device (310) operating with mechanical or hydraulic means, and, when one or more switchable second force transferring parts are used, changing of the resultant spring constant of the spring system also being carried out in one or more steps with different exciter frequencies.
18. Device according to one of Claims 1 to 17, characterized in that, for the purpose of changing the resulting spring constant of the spring system, one or more spring elements (150, 282) are adjustable with respect to their own spring constant continuously or in steps.
19. Device according to one of Claims 16 to 18, characterized in that, while passing through a range of the exciter frequency during the compaction, either the adjustment has taken place in steps for one or more assigned exciter frequencies which can be given, in the case of step-by-step adjustability of the natural frequency of the mass-spring system, or the adjustment of the natural frequency has taken place simultaneously with the adjustment of the exciter frequency, in the case of continuous adjustability of the natural frequency.
20. Device according to one of Claims 1 to 19, characterized in that the system spring of the mass-spring system is connected in a force-transferring and rigid manner to a damping mass (450) for the purpose of transferring the dynamic spring forces to the latter, the mass of which is at least 20 times greater than the system mass, the damping mass either being part of the foundation to which the frame of the compacting device is likewise connected in a force-transferring manner, or else representing a mass of its own, which is preferably supported by means of isolating springs (452) in a soft manner against the foundation.
21. Device according to one of Claims 1 to 20, characterized in that the exciter device, as an exciter actuator, comprises one or more rotational motors with a connected movement-converting gear mechanism for generating a linear exciter movement derived from the rotational movement, in which arrangement, if at least two rotational motors are provided, they are connected to a common movement-converting gear mechanism in such a way that an adjustment of the relative angle of rotation of the two motors causes the generation of a resulting drive output movement which is adjustable in its movement stroke.
22. Device according to one of Claims 1 to 21, characterized in that an unbalanced vibrator which can be regulated with respect to the rotational speed, but not with respect to its static moment, is provided for the exciter device as an exciter actuator, and in that the physical variable of the upper or lower amplitude of the oscillating excursion s of the vibrating table or of the oscillating excursion f of the mould or a variable derived from it of the oscillating velocity or oscillating acceleration s', f or s", f" is regulated by a regulating device according to a value which can be given, in such a way that the excess exciter energy transferred by the exciter device is extracted from the oscillatory mass-spring system by a damping device influenced by the regulating device, the extracted energy being transferred by the oscillating movement of the mass-spring system and the damping device being hydraulic, for example, operating with a conversion of motional energy into thermal energy.
23. Device according to one of Claims 1 to 22, characterized in that a measuring system (192/194) is provided, by which the actual values of the physical variables s', s" or f, f" to be regulated are determined.
24. Device according to one of Claims 1 to 23, characterized in that the system spring of the vibrating table is set hard for both directions of oscillation.
25. Device according to one of Claims 1 to 24, characterized in that hydraulic linear motors are provided only on condition that a constrained guidance is at the same time provided for executing the oscillating movement of the vibrating table in a double direction and with a guide part arranged centrally on the vibrating table.
26. Use of the device according to one of Claims 1 to 25 for carrying out compacting operations on moulded bodies (110) of granular materials, such as dry concrete mortar for example, characterized in that, when carrying out the compacting operation, the oscillating excitation takes place by an exciter device (106) with the exciter frequency passing through a given range with increasing values for the exciter frequency.
27. Use according to Claim 26, characterized in that, while passing through the frequency range of the exciter frequency, changing of the natural frequency takes place, in that an adjustment of the value of the spring constant of the system spring (142) and/or an adjustment of the value of the system mass (440) is carried out.
28. Use according to Claim 26 or 27, characterized in that at least one pre-compaction, in which the moulded body (110) cannot be brought into connection with the pressing plate (180), can be carried out.

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