EP2504130A1

EP2504130A1 - Variation of the natural frequency of vibratory means in electric tools

Info

Publication number: EP2504130A1
Application number: EP10767980A
Authority: EP
Inventors: Gerd Schlesak; Carsten Diem; Willy Braun; Hardy Schmid; Holger Ruebsaamen; Michael Weiss; Jan Koalick; Peter Loehnert
Original assignee: Robert Bosch GmbH
Current assignee: Robert Bosch GmbH
Priority date: 2009-11-25
Filing date: 2010-10-06
Publication date: 2012-10-03
Also published as: DE102009047106A1; US20120279741A1; WO2011064028A1; CN102666027A; CN102666027B; RU2012126307A

Abstract

The present invention relates to an electric tool (1) having a vibratory means (58) which is provided for imparting a counteracting vibration (103) which counteracts a housing vibration (100) of the electric tool (1), wherein a vibration-relevant characteristic (Wo, 51, Kf, 112-115, 106, 52) of the vibratory means (58) can be adapted during the operation of the electric tool (1) in such a way that the amplitude (104), the phase position and/or the frequency (1/T) of the counteracting vibration (103) varies with a change in the vibration-relevant characteristic. The present invention also relates to an electric tool (1) which is in particular in accordance with the invention and whose vibratory means (58) have a natural frequency which can be varied by variation means of the electric tool (1). The present invention also relates to a method for compensating housing vibrations (100), in particular of the electric tool (1) according to the invention, having a vibratory means (58) which is provided for imparting a counteracting vibration (103) which counteracts a housing vibration (100) of the electric tool (1), wherein the amplitude, the phase position and/or the frequency of the counteracting vibration (103) is varied during the operation of the electric tool (1).

Description

Description title

Variation of the natural frequency of vibration in power tools prior art

The present invention relates to a power tool with a vibration means, which is arranged to compensate for housing vibrations in the power tool and a method for compensating housing vibrations of a power tool.

As the legal requirement to use daily power to couple the daily allowable workload to the physical load applied to the operator, the subject of vibration is becoming more and more important for power tools, especially for drilling and impact hammers.

When hammering and chiselling a hammer, a very large physical load for the operator comes from the housing vibration generated by the impact mechanism. Especially with large drilling and impact hammers, the vibrations are very pronounced due to the high impact energy. For operators of such machines, the permitted working time is reduced considerably without further measures. As a result, in the development of increasingly worked solutions in which vibrations of power tools are reduced. This can ensure that you can continue to work with these devices without restrictions.

FIG. 6 shows a typical housing oscillation 100 which arises when the housing of drilling and impact hammers 7 is vibrated. 8 is caused, in which the racket 121 is driven by an eccentric piston drive 12. On the horizontal axis 101, the rotation angle [in °] is shown, on the vertical axis 102, the deflection [in mm] of the housing. The vibration generating case vibration 100 is composed of a plurality of frequency components. The main frequency is derived from the periodic acceleration of the racket 121. However, FIG. 6 shows that the deflection, which is caused by the periodic acceleration of the racket 121, is superimposed on further frequency components from other sources of vibration, for example from the impact and recoil processes of the impact chain and from unbalanced mass forces of the drive. Because the

Housing vibration 100 does not substantially sinusoidally with the main frequency, but the sinusoidal waveform with main frequency more frequency components are superimposed. Since nonlinear systems work with only partially harmonic motion sequences, the individual vibration components are superimposed in a complex manner. By playing between the individual components, by nonlinear elasticity courses, by the non-linear impact processes and by the only approximately harmonic reaction forces from the impact mechanism, inharmonious housing vibrations of complex order result.

An optimal reduction of the housing vibration is achieved when a vibration reduction system counteracts the housing vibration shown in FIG. 6 as accurately as possible.

In practice, the generation of counter-forces, which counteract the housing vibrations, for example by means of absorbers occurs.

A damper is a spring-mass system with a fixed resonant frequency that can achieve significant vibration reduction only in a small region near the resonant frequency. Therefore, the Tilgereigenfrequenz is selected as close to the greatest disturbing vibration frequency of the housing, so that the absorber acts as effectively as possible in this frequency range. However, during operation of a power tool, the oscillations that occur regularly have different sources. Due to their superimposition, these case vibrations cause different and variable frequencies.

For example, when changing the load parameters and / or the operating parameters of the power tool, in particular by changing the rotational speed of the drive motor of the power tool or during processing of a workpiece made of different materials, which regularly takes place during operation of the power tool, the effective range of a damper can be exceeded and Therefore, eradication will be ineffective.

Therefore, additional measures are required to improve the effect of the absorber in the operating and load conditions occurring or to achieve the greatest possible reduction in vibration.

Disclosure of the invention

The object of the invention is therefore to provide a power tool that is better adapted to the changing requirements in the power tool, so that the housing vibration of the power tool is effectively reduced, and a method for reducing the housing vibration of the power tool.

The object is achieved with a power tool having a vibration means, wherein the vibration means is provided for exerting a counter vibration, which counteracts a housing vibration of the power tool, wherein a vibration-relevant property of the vibration means during operation of the power tool is adaptable so that the amplitude, the phase position and / or the frequency of the countervibration changed when changing the vibration relevant property.

Since, according to the invention, the amplitude, the phase position and / or the frequency of the counter-vibration of the vibration means are changed by adaptation of the vibration-relevant property of the vibration means during operation If necessary, the countervibration is dynamically adapted to the vibration conditions in the power tool. As a result, the frequency range in which the vibration means can be used effectively to compensate for the housing vibration is increased. The compensation of the housing oscillation of the power tool is therefore effectively possible in a larger frequency range.

In a preferred embodiment, the power tool has varying means with which the amplitude, the phase angle and / or the frequency of the countervibration during operation of the power tool is variable. Thereby, the counter vibration of the vibrating means during operation of the power tool is dynamically adaptable to the housing vibration, so that the counter vibration with respect to their amplitude, their phase position and / or their frequency can be changed so that it counteracts the housing vibration more accurately, even in the case of unexpected Changes in the housing vibration, for example, by material changes in the workpiece. The housing oscillation can therefore be counteracted even better with changing operating and environmental parameters.

Preferably, the housing oscillation can be compensated both as a function of the instantaneous operating state of the power tool and also independently of the operating point of the power tool. Therefore, the power tool according to the invention allows both the operating settings and operating parameters of the power tool as well as changes in the machined workpiece or the behavior of the operator are taken into account.

In a preferred embodiment, which also solves the problem, the vibration means has a natural frequency, which is variable by means of the change means. The natural frequency of the vibrating means is a vibration-relevant property. It is known to the person skilled in the art that the natural frequency is the frequency of a vibration means with which the vibration means would vibrate after a single impact, if the vibration were not damped by friction, and if no exciting forces force the vibration of the mass. It is also known to the person skilled in the art that the natural frequency of a mass-spring system is calculated according to the formula: ω ₀ = k _F / m

Here, k _{F is} the spring constant of the spring, m the weight of the mass and ω ₀ the natural frequency of the mass-spring system.

When the frequency of a vibration exciting the vibrator is close to the natural frequency of the vibrator, the vibrator vibrates with a very large amplitude. If the counter vibration acts as accurately as possible against the housing oscillation, is therefore with a frequency of the

Oscillating means near its natural frequency, a substantially maximum amount of the housing oscillation compensated. By changing the natural frequency of the oscillating means, it is primarily the amplitude and, at least to a small extent, the phase position of the countervibration that can be changed. When changing the mass of the vibrating means and the frequency of the counter-vibration changes.

In order to change the natural frequency of the vibration means, the vibration means preferably has a mass which is changeable. The mass is provided to a free, the housing vibration or the vibration causing the housing vibration counteracting vibration. In a preferred embodiment, the mass comprises at least two partial masses, which can be reversibly coupled to one another by means of the changing means. Thereby, the weight of the oscillating mass of the vibrating means is variable, wherein the change of the weight of the oscillating mass leads to the change of the natural frequency. Namely, as the weight of the mass increases, the natural frequency of the vibrator is shifted toward lower frequencies. The mass of the vibrating means is therefore a vibration-relevant property.

Preferably, the vibration means to change its natural frequency to a spring constant, which is variable by means of the change means. Particularly preferably, the vibration means has a spring characteristic which is non-linear. The spring constant or the spring characteristic of the spring of Vibrating means are therefore vibration-relevant properties of the vibrating means.

In a preferred embodiment, the mass is arranged on at least one spring, in particular a spiral spring, a helical compression spring or a leaf spring. In this embodiment, the vibrating means is a absorber.

In a further preferred embodiment, the oscillation means comprises a plurality of springs which are interconnected such that the spring characteristic of the oscillation means is non-linear. In a particularly preferred embodiment, the vibration means comprises the spring on which the mass is arranged, and at least one second spring, which cooperates with the spring in dependence on the amplitude of the counter-vibration. The second spring is preferably connected in parallel to the spring so that the spring constant is increased. In a further preferred embodiment, both springs with linear and springs with nonlinear spring characteristic are interconnected.

One skilled in the art will understand that the spring constant of the vibrating means is the spring constant of the spring, or the spring constant resulting from the series and / or parallel connection of the plurality of springs of the vibrating means. A person skilled in the art knows that a spring characteristic reflects the course of the spring constant, which results from the quotient of the magnitude of the spring-stretching force, which is also called spring preload, and the extension caused by the stretching force. The spring characteristic of the vibration means is therefore also the spring characteristic of the spring of the vibration means, or resulting from the series and / or parallel connection of the springs of the vibration means spring characteristic. The change in the spring constant causes a change in the natural frequency of the vibrating means. Namely, as the spring constant increases, the natural frequency of the vibrating means is shifted toward higher frequencies.

In a likewise preferred embodiment, the spring bias of the vibration means can be changed by means of the change means. Both springs with linear and springs with non-linear spring characteristic are preferred. The skilled person understands that the spring bias of the vibrating means is the spring bias of the spring, or the spring bias resulting from the series and / or parallel connection of the plurality of springs of the vibrating means. The change in the spring bias causes in particular a change in the amplitude of the countervibration. The spring preload is therefore a vibration-relevant property of the vibrating means.

Preferably, the spring of the vibrating means is mounted on a bearing point, wherein by means of the changing means for changing the spring bias of the bearing point of the spring is displaceable. By moving the bearing point of the spring, the bias of the spring is variable, so that in particular the amplitude of the counter-vibration is changed.

The changing means comprise in a preferred embodiment, an electrical control means which cooperates with the vibration means, in particular with the mass and / or the spring. Particularly preferably, the electrical control means interacts directly with the mass and / or the spring. Or it also preferably cooperates indirectly with the mass and / or the spring, for example by activating or actuating a further changing means, which interacts directly with the mass and / or the spring. As a result, the change in the amplitude, frequency and / or phase position of the countervibration is electrically controlled or regulated. The skilled person understands that another form of control or regulation is applicable, for example, a mechanical control or regulation.

The electrical control means is preferably an actuator, or also preferably the control means comprises an actuator, in particular a servomotor, a linear motor or an electromagnet.

In a preferred embodiment, the power tool further comprises a detection means for detecting the housing oscillation of the power tool, the speed and / or the rotational speed of a drive motor of the power tool, the countervibration of the mass, and / or other vibration-relevant variables, so that the amplitude, the phase position and / or the free Frequency of the countervibration of the mass in response to these vibration relevant variables is variable. For example, acceleration sensors and / or displacement measuring sensors are used as detection means.

Preferably, the power tool further comprises an evaluation unit which is connected to the detection means to evaluate the vibration-relevant quantities, and to provide the control means, depending on the vibration-relevant variables output signal. Such an evaluation unit preferably comprises a logic with which the vibration-relevant variables can be converted into the output signal. Preferably, the analysis of the vibration-relevant variables is carried out by comparison with standard sizes. However, as an intelligent control or regulation is also preferably used, in particular an adaptive control. The evaluation unit is, for example, a processor-controlled unit. However, it can also be designed as an electrical circuit, in particular an integrated circuit, for example as an ASIC (Application Specific Integrated Circuit).

The control or regulation of the countervibration of the vibration means by means of an electrical control means and in particular as a function of vibration relevant variables allows a targeted to the vibration level dynamic adjustment of the countervibration, both as a function of the current operating state of the power tool as well as of known dynamic behavior of the power tool in its different operating modes as well as the behavior of an operator or the machining or the properties of the workpiece.

The object is further achieved with a method for compensating housing oscillations of a particular power tool according to the invention, with a vibration means that is provided to exert a counter vibration of the housing vibration, wherein the amplitude, the phase angle and / or the frequency of the counter-vibration during operation of the power tool to be changed. The change in the amplitude, phase position and / or frequency of the countervibration is preferably effected by an adaptation of a vibration-relevant property of the vibration means.

As a result, the effective frequency range of the vibration medium is increased. In addition, the active changing of the amplitude, phase position and / or frequency of the countervibration allows an adjustment of the countervibration to the changing during the operation of the power tool housing vibrations. The counter-vibration is therefore optimizable during operation of the power tool so that it more accurately counteracts the housing vibration and thus better compensates for it.

In the following the invention will be described with reference to figures. The figures are merely exemplary and do not limit the general idea of the invention.

FIGS. 1 to 5 schematically show various embodiments of a power tool according to the invention,

Fig. 6 shows a housing vibration of a power tool and a counter vibration of a vibrating means, and

Fig. 7 shows spring characteristics of differently designed springs.

FIGS. 1 to 5 schematically show various embodiments of a power tool 1 according to the invention.

As a power tool 1, a hammer drill is shown here by way of example, which comprises a striking mechanism assembly 3. In the striking mechanism assembly 3, a racket 121 is provided which is linearly driven via a connecting rod 12 which is eccentrically mounted by means of an eccentric pin 1 1 on an eccentric disc 10 which rotates about an eccentric axis 9. The eccentric disc 10 is drivable by means of a likewise rotatable about the eccentric axis 9 gear 23, which is in engagement with a drive pinion 22 which is non-rotatably arranged on a drive shaft 21 of a drive motor 20 of the power tool 1. Upon rotation of the eccentric disk 10 in a direction of rotation 8 about the eccentric axis 9, the racket 121 of the striking mechanism assembly 3 is moved back and forth in a longitudinal direction 4.

However, the present invention is not limited to power tools 1 with impact mechanism assembly 3, but also for other power tools 1 used, for example, on drills, jigsaws or the like.

In the following, the term rotary hammer is used synonymously for the power tool 1. According to the invention, a vibration-relevant property ω _0, 51, k _F, 1 12 -

1 15, 106, 52 of a vibration means 58 during operation of the electric tool 1 adapted so that the amplitude 104, the phase angle φ and / or the frequency 1 / T of a counter-vibration 103 executed by the vibration means 58 changes. Vibration-relevant properties ω _0, 51, k _F, 1 12 - 1 15, 106, 52 of the vibration means 58 are, for example, their natural frequency ω ₀ , the spring preload 106, spring constant k _F or spring characteristic 1 12 - 1 15 of their spring 52 and their mass 51st Here, the vibration-relevant properties ω _0, 51, k _F, 1 12 - 1 15, 106, 52 of the vibration means 58 in a plurality of interconnected springs 52, 521-524 and / or masses 51, 51 1, 512 resulting from the Series and / or parallel connection of the plurality of springs 52, 521-524 and / or masses 51, 51 1, 512 resulting vibration relevant properties ω _0, 51, k _F, 1 12- 1 15, 106, 52nd

In the embodiment of FIG. 1, a vibrating means 58 is provided which comprises a mass 51. The vibration means further comprises a first spring 521 and a second spring 522, wherein the mass 51 between the first spring 521 and the second spring 522 is arranged. The vibrating means 58 is therefore a absorber. As first and second springs 521, 522, spiral springs are present here. seen. In the description of FIG. 1, therefore, the term spiral spring is used synonymously for the term spring 521, 522.

In the power tool 1 changing means 98, 90, 56 are provided, with which the spring bias of the vibrating means 58 is variable. Namely, a slide 90 is provided in a gate 98 as a change means, wherein the first spring 521 is mounted on the slide 90, so that it forms a bearing point of the first spring 521. The terms bearing point and slide 90 are therefore used interchangeably in the description of FIG.

As a further changing means 98, 90, 56, a centrifugal weight assembly 56 is provided, which cooperates with the slide 90. The centrifugal weight assembly 56 is rotatably connected to the eccentric axis 9, so that the centrifugal weight assembly 56 is driven by rotating the eccentric axis 9. In this case, the slide 90 is moved in an extension direction 91 along the backdrop. In the present case, the extension direction 91 is the extension direction 91 of the first and second springs 521, 522, so that the slide 90 is displaced in or against the force of the first and second coil spring 521, 522, and thus the spring bias of the two springs 521, 522 changes.

The displacement of the bearing point 90 causes a change in the spring bias 106 of the vibration means 58, so that in particular the amplitude 104 of a counter-vibration 103 of the vibration means 58 changes. In this embodiment, the spring preload 106 is dependent on the

Speed of the drive motor 20 increases, and that is when using springs 521, 522 with increasing spring characteristic, the spring bias 106, the greater, the faster the drive motor 20 rotates. Thereby, the amplitude 104 of the counter vibration 103 is reduced. By using a spring 521, 522 with a nonlinear spring characteristic (see Fig. 7), preferably with a progressive spring characteristic curve 1 15, it is furthermore possible to prevent the mass 51 from striking its mechanical end position. The mechanical end position is reached when the springs 521, 522 are no longer compressible. The stop to the mechanical end position would lead to an impairment of the function and the life of the vibrating means 58.

An embodiment is also conceivable in which an electrical control means 54 (see, for example, Figures 3, 4) is used to drive the centrifugal weight assembly 56.

In the embodiment of FIG. 2, the abutment with the mechanical end position is prevented by suspending the mass 51 of the oscillating means 58 between a first coil spring 521 and a second coil spring 522, wherein the first coil spring 521 is a third coil spring 523 and the second coil spring 522 a fourth spiral spring 524 are arranged in parallel, which cooperate in dependence on the amplitude 104 of the counter-vibration 103 with the first and second coil spring 521, 522. At high amplitude 104 of the counter-vibration 103, either the first coil spring 521 and the third coil spring 523 are connected in parallel, so that the spring constant k _F52 i, k _F 522 of the springs 521, 523 add and the spring constant k _{F of} the vibrating means 58 therefore increases , As a result, the natural frequency ω _{0 of} the vibration means 58 shifts to larger frequencies. Or, the second coil spring 522 is disposed in parallel with the fourth coil spring 524 with the same sequence. A shift of the natural frequency oo ₀ to larger frequencies causes a reduction in the amplitude 104 of the counter-vibration 103rd

In this embodiment, no further changing means are provided in the power tool 1.

In the embodiment of FIG. 3, a mass 51 of a vibrating means 58 is arranged on a leaf spring 52 of the oscillating means 58. In the context of the figure description of Fig. 3, therefore, the terms leaf spring and spring 52 are used synonymously.

Here, the hammer drill 1 has a detection means 61 for detecting vibration-relevant variables E1, with which the housing vibration 100 of the hammer drill 1 can be detected. The detection means 61 is therefore, for example, a acceleration sensor or a distance measuring sensor. Alternatively or additionally, however, it is possible to detect other vibration-relevant variables E1 by means of the detection means 61 or further detection means E1, for example the rotational speed and / or the angle of rotation of the drive motor 20 of the hammer drill 1, whereby conventional rotational speed and / or rotational angle sensors can be used for this purpose are, for example commutation, tachometer, resolver, position encoder and others. Further vibration-relevant variables E1 are, for example, also the current movement of the mass 51 and / or settings which can be changed by the operator.

The detected vibration-relevant quantities E1 are transmitted to an evaluation unit 7, which is connected to the detection means 61, for evaluation. The evaluation unit 7 comprises a logic with which the vibration-relevant quantities E1 can be converted into an output signal A, which is provided to an electrical control means 54. In this embodiment of the power tool 1, therefore, the electrical control means 54 is provided as a change means 54, so that the counter-vibration 103 of the vibration means 58 here is actively adaptable to the requirements in the power tool 1. In the embodiment of FIG. 3, a servomotor is provided as the control means 54. This is also the case in the embodiment of FIG. 4, so that in these FIGS. 3, 4 the terms control means 54 and servomotor are used synonymously.

In FIG. 3, a bearing point 90 of the leaf spring 52, here a clamping point 90, is displaceable by means of the servomotor 54. In Fig. 3, therefore, the terms

Bearing point 90 and clamping point 90 used synonymously. At one end of the leaf spring 52, the mass 51 is arranged, while the leaf spring 52 is mounted with its other end on the housing 33 of the power tool 1. The mass 51 is therefore provided so that with it a counter-vibration 103, which counteracts the housing vibration 100 and this at least partially compensated, executable.

The servomotor 54 drives a gear 531, which cooperates with a toothed slide 53. As the gear wheel 531 rotates, the slider 53 is released. long an extension direction 91 of the leaf spring 52 is moved. On the slider 53, a clamping means 532 is arranged, which forms the clamping point 90 for the leaf spring 52, so that when moving the slider 53 of the clamping point 90 of the leaf spring 52 moves.

In the embodiment of FIG. 3, therefore, the servo motor 54 does not interact directly with the mass 51 and / or the leaf spring 52, but there are further change means 53, 531, 532 provided, here a gear 531, a slider 53 and a clamping means 532 that interact with the leaf spring 52.

By changing the Einspannpunktes 90, the spring constant k _{F of} the leaf spring 52 and thus the natural frequency ω _{0 of} the vibration means 58 changes, so that above all the amplitude 104 of the counter-vibration 103 changes. Since the effective length of the leaf spring 52 changes here, the displacement also causes a change in the frequency 1 / T of the counter-vibration 103rd

In the embodiment of Fig. 4, in contrast to the embodiment of Fig. 3, the mass 51 is suspended between a first coil spring 521 and a second coil spring 522, wherein the first coil spring 521 on a first bearing means 901 and the second coil spring 522 on a second bearing means 902 is stored. The first bearing means 901 and the second bearing means 902 are reciprocally slidable along a spindle 99 in the extending direction 91 of the first and second coil springs 521, 522.

By means of the servo motor 54, the spindle 99 is rotatable, so that the first and the second bearing means 901, 902 along the extension direction 91 shifts. Embodiments are also possible in which the first and second bearing means 901, 902 are displaceable separately from each other.

By displacing the bearing means 901, 902, the spring bias 106 (see Fig. 7) of the spiral springs 521, 522 changes, so that in particular the amplitude 104 of the counteroscillation 103 changes. In this embodiment, the bearing means 901, 902, the spindle 99 and the servomotor 54 are changing means 54, 99, 901, 902. By using springs 521, 522 with nonlinear and in particular progressive spring characteristic curve 15 (see Fig. 7), it is also possible in this case to prevent the impact of mass 51 on its mechanical end position.

In addition, the use of nonlinear spring characteristic springs 521, 522 (see Fig. 7) causes a dynamic change in the spring constant k _F and thus a change in the natural frequency ω _{0 of} the oscillator 58. Thus, the oscillator 58 is effectively usable for a wider frequency band.

In the embodiment shown in Fig. 5 is provided as a control means 55, an electromagnet. Therefore, in this Fig. 5, the terms control means 55 and solenoid used synonymously.

In this embodiment, a first partial mass 51 1 suspended between a first coil spring 521 and a second coil spring 522. Furthermore, a second partial mass 512 is provided, which is arranged in the region of the first partial mass 51 1. The second sub-mass 512 extends, for example, at least partially along the first sub-mass 51 1 or is arranged, for example, around it. Between the first partial mass 51 1 and the second partial mass 512, a magnetorheological fluid 57 is arranged, for example in a gap (not shown here).

In the area of the partial masses 51 1, 512, an electromagnet 55 is arranged as adjusting means 55 such that when the electromagnet 55 is switched on, the magnetorheological fluid 57 causes the second partial mass 512 to be coupled to the first partial mass 51 1, so that the weight of the partial mass 51 Mass 51 changes. Namely, the weight is substantially the weight of the first partial mass 51 1, and with the solenoid 55 is substantially the weight of the first partial mass 51 1 plus the weight of the second partial mass 512, when the solenoid 55 is not switched on, so that the mass 51 in the first case is first part of mass 51 1, and so that the mass 51 is formed in the second case of the first part of mass 51 1 and the second part of mass 512. This embodiment therefore has the electromagnet 55 and the magneto-theological fluid 57 as a change means 55, 57, so that here, the counter-vibration 103 of the vibration means 58 is actively controlled or regulated.

The greater weight causes a shift of the natural frequency ω _{0 of} the vibration means 58 to lower frequencies, so that both the amplitude 104 and the phase position φ and the frequency 1 / T of the counter-vibration 103 of the vibration means 58 change.

For a control or regulation of the amplitude 104, phase angle φ and / or frequency 1 / T of the counter-vibration 103 of the vibration means 58 and mechanical controls are conceivable. For example, it is possible to arrange the sub-masses in such a way that a coupling of the sub-masses 51 1, 512 by means of a bolt which engages in receiving openings of the sub-masses 51 1, 512 takes place as a function of a vibration-relevant variable, so that they are coupled with sub-masses 51 1 coupled to one another 512, the mass 51 of the vibrating means 58 form, while in non-coupled masses 51 1, 512 only one of the two sub-masses 51 1, 512, the mass 51 of the vibrating means 58 forms.

6 shows a housing oscillation 100 of a power tool 1 and a countervibration 103 of a mass 51 which is provided in the power tool 1 for compensating the housing oscillation 100. Since a housing vibration 100 is caused by a variety of vibration sources, e.g. due to the impact of a percussion unit 3, the shock and recoil operations of the percussion chain, unbalanced mass forces of the drive, and others, the housing vibration 100 is not substantially sinusoidal. But the housing oscillation 100 is composed, as shown in FIG. 6, of a multiplicity of sinusoidal oscillations of different amplitudes, phase positions and frequencies.

By means of a substantially sinusoidal countervibration 103, therefore, the housing oscillation 100 can only be partially compensated. However, the effectively usable frequency range of a vibration means 58 by changing the natural frequency oo _{0 of} the vibration means 58 or the accuracy with which the counter-vibration 103 of the housing oscillation 100 counteracts by adjusting the phase position φ, amplitude 104 and / or frequency 1 / T of the counter-vibration 103 are optimized so that a more effective and better compensation of the housing oscillation 100 is possible.

6 shows, by way of example, a sinusoidal counteroscillation 103, which is carried out, for example, by a mass 51 suspended on a spring 52 (see FIG. In addition, here the amplitude 104 of the countervibration 103, its frequency 1 / T are represented by their period T and their phase position φ relative to the housing oscillation 100.

Fig. 7 shows spring characteristics 1 1 1 - 1 15 of variously executed springs 52, 521 - 524, which can be optionally used in the vibratory means 58 of the invention Elektor- rowerkzeugs 1.

On the vertical axis 106 is the amount of force [in N] with which the spring 52, 521-524 is stretched, and on the horizontal axis 105 the elongation [in mm] caused by the extension is plotted.

The spring characteristic 1 1 1 shows a linear course, the spring characteristic 1 12 shows a constant course, the spring characteristic 1 13 shows a discontinuous increase, the spring characteristic 1 14 shows a degressive and the spring characteristic 1 15 shows a progressive course.

An unsteady increase 1 13 can be achieved, for example, by the parallel connection of two springs 521 - 524 shown in FIG. 2, or by corresponding leaf springs. A degressive or progressive course 1 14, 1 15 can be achieved for example by a corresponding winding of the springs 52, 521-524.

In the case of the oscillation means 58, the amplitude 104, the phase angle φ and / or the frequency 1 / T of the counteroscillation 103 executed by the oscillation means 58 are variable, so that the effective frequency range of the Oscillating means 58 is increased. In the case of the power tool 1 according to the invention, a change or adaptation of the countervibration of the vibration means 58 provided for compensating the housing vibration 100 is provided during operation of the power tool 1, namely by the amplitude 104, the phase angle φ and / or the

Change frequency 1 / T of the counter-vibration 103 of the vibrating means 58, the compensation of the housing vibration 100 of the power tool 1 is very effectively possible. Furthermore, the countervibration 103 of the oscillation means 58 of the power tool 1 according to the invention is also actively dynamically adaptable, so that it can be changed both as a function of the instantaneous operating state of the power tool 1 and independently of the operating point of the power tool 1.

Claims

claims

1 . Electric tool (1) having a vibration means (58) which is provided for exerting a counter-vibration (103) which counteracts a housing oscillation (100) of the power tool (1),

characterized in that

a vibration-relevant property (ω _0, 51, k _F, 1 12 - 1 15, 106, 52) of the vibration means (58) during operation of the power tool (1) is adaptable so that the amplitude (104), the phase position ( φ) and / or the frequency (1 / T) of the countervibration (103) when changing the vibration-relevant property (ω _0, 51, k _F, 1 12 - 1 15, 106, 52) changed.

2. Power tool (1) according to claim 1, characterized in that the power tool (1) has change means (53, 531, 532, 54, 55, 56, 57, 90, 901, 902, 98, 99), with which the Amplitude (104), the phase angle (φ) and / or the frequency (1 / T) of the countervibration (103) during operation of the power tool (1) is variable.

3. Power tool (1), in particular according to one of the preceding claims, characterized in that the housing vibration (100) both in response to an instantaneous operating state of the power tool (1) and independently of an operating point of the power tool (1) is compensated.

4. Power tool (1), in particular according to one of the preceding claims, characterized in that the vibration means (58) has a natural frequency (ω ₀ ) by means of changing means (53, 531, 532, 54, 55, 56, 57, 90 , 901, 902, 98, 99) of the power tool (1) is variable.

5. Power tool (1) according to one of the preceding claims, characterized in that the vibration means (58) has a mass (51, 51 1, 512) which is variable.

6. Power tool (1) according to one of the preceding claims, characterized in that the mass (51) at least two partial masses (51 1, 512) summarizes, by means of the changing means (53, 531, 532, 54, 55, 56 , 57, 90, 901, 902, 98, 99) are reversibly coupled to each other.

7. Power tool (1) according to one of the preceding claims, characterized in that the oscillation means (58) has a spring constant (k _F ), which by means of the changing means (53, 531, 532, 54, 55, 56, 57, 90, 901, 902, 98, 99) is changeable.

8. Power tool (1) according to one of the preceding claims, characterized in that the vibration means (58) has a spring characteristic (1 12 - 1 15), which is non-linear.

9. Power tool (1) according to one of the preceding claims, characterized in that the mass (51, 51 1, 512) on at least one spring (52, 521, 522) is arranged.

10. Power tool (1) according to one of the preceding claims, characterized in that the vibration means (58) a plurality of springs (52, 521 - 524), which are interconnected so that the spring characteristic (1 12 - 1 15) of the vibrating means (58) is nonlinear.

1 1. Power tool (1) according to one of the preceding claims, characterized in that the vibration means (58) comprises the spring (52, 521, 522) on which the mass (51) is arranged, and at least one second spring (523, 524) which cooperates with the spring (52, 521, 522) in dependence on the amplitude (104) of the countervibration (103).

12. Power tool (1) according to one of the preceding claims, characterized in that by means of the changing means (53, 531, 532, 54, 55, 56, 57, 90,

901, 902, 98, 99), the spring bias (106) of the vibration means (58) is variable.

13. Power tool (1) according to one of the preceding claims, characterized in that the spring (52, 521, 522) of the vibration means (58) at a bearing point (90, 901, 902) is mounted, wherein by means of the changing means (53, 531, 532, 54, 55, 56, 57, 90, 901, 902, 98, 99) of the bearing point (90, 901, 902) of the spring (52, 521, 522) is displaceable.

14. Power tool (1) according to one of the preceding claims, characterized in that the changing means (53, 531, 532, 54, 55, 56, 57, 90, 901,

902, 98, 99) comprise an electrical control means (54, 55) cooperating with the oscillating means (58), in particular with the mass (51, 51 1, 512) and / or the spring (52, 521, 522) ,

15. Power tool (1) according to one of the preceding claims, characterized in that it comprises a detection means (61) for detecting the housing oscillation (100) of the power tool (1) and / or further vibration-relevant variables (E1).

16. A method for compensating housing oscillations (100), in particular of the electric tool (1) according to one of the preceding claims, comprising a vibration means (58) for operating a counter-vibration (103) acting against a housing oscillation (100) of the electric tool (1). is provided,

characterized in that

the amplitude (104), the phase angle (φ) and / or the frequency (1 / T) of the countervibration (103) during operation of the power tool (1) is changed.