Patent Application
20070234578
The invention relates to a portable handheld work apparatus such as a chain saw, cutoff machine, brushcutter or the like.
During operation of work apparatus of this kind, vibrations occur which are excited by a driven tool of the work apparatus. Additional vibrations are excited especially where the drive motor of the work apparatus is in the form of an internal combustion engine because of the moving masses of the engine. In general, these engines are single cylinder engines and have an engine running which is comparatively rough and burdened with vibrations. The vibrations, which are generated at the engine end, cannot be completely eliminated by balancing the moving engine parts. In total, oscillations caused by the tool and engine lead to vibrations which are disturbingly noticeable at the handles of the work apparatus. The handle end vibration can only be reduced to a limited extent with additional measures such as a vibration decoupling of the handles from the engine housing by means of antivibration elements.
U.S. Pat. No. 4,836,297 discloses a portable handheld work apparatus driven by an internal combustion engine wherein imbalance weights are mounted in a crankshaft assembly of the drive motor. An imbalance is deliberately caused by the imbalance weights on the crankshaft web and/or on the fan wheel. The imbalance is so dimensioned with respect to magnitude and phase position that the imbalance, as vibration suppressor, forms a balance or compensation for operation-caused translatory vibrations.
The targeted imbalance of the vibration suppressor results from the imbalance masses which are defined in accordance with phase angle and magnitude. The targeted imbalance of the vibration suppressor can be designed to an optimum of the equivalent oscillation value in order to reduce the vibration level at the handle locations. The imbalance operates to reduce specific oscillation forms from the handle system and from the antivibration system. The equivalent oscillation value results from the values of the representative operating conditions. These values are defined, for example, in motor-driven chain saws as idle rpm values, full-load rpm values and maximum rpm values. It has been shown that a vibration suppressor, which is optimized to the equivalent oscillation value, exhibits an effect which is, under some circumstances, insufficient in the above-mentioned individual operating states.
It is an object of the invention to provide a portable handheld work apparatus having a vibration suppressor which is so improved that an improved suppression effect is ensured over a large operating parameter range.
The portable handheld work apparatus of the invention includes: a vibration suppressor for suppressing translatory vibrations occurring during operation of the work apparatus; a drive motor driving the vibration suppressor; the vibration suppressor defining a rotational axis and including a suppression mass mounted at a radius from the rotational axis for generating an imbalance and, as a consequence of the imbalance, the suppression mass generating an rpm-dependent translatory vibration; resilient biasing means for applying a resilient biasing force to the suppression mass in opposition to an rpm-dependent centrifugal force applied to the suppression mass during the rotation; the suppression mass being mounted so as to be radially movable along a path under the action of the forces; the suppression mass defining first and second equilibrium positions along the path at first and second radii from the rotational axis corresponding to first and second rpms of the vibration suppressor; and, the biasing force and the centrifugal force conjointly defining a resultant force for effecting an rpm-dependent position transfer of the suppression mass between the first and second equilibrium positions in both directions.
An arrangement is suggested wherein at least one suppression mass is mounted so as to be radially movable under the force of its rpm-dependent centrifugal force and an opposing spring force. For at least two different rpms of the vibration suppressor, an equilibrium position of the at least one suppression mass is provided in each case with a different radius to the rotational axis. A total force acts on the suppression mass and results from the centrifugal force and the spring force. This total force is provided for an rpm-dependent position change between the two equilibrium positions in both directions. The radial displacement utilizes the situation that the centrifugal force, which acts on the suppression mass, is also directed in the radial direction. In this way, the centrifugal force and the spring force, which acts radially inwardly and in the opposite direction, are used to bring about the rpm-dependent automatic position displacement of the suppression mass without external energy supply. An arrangement is provided which is self acting and adapted to the different operating conditions. This arrangement functions without a separate control unit, without active actuating elements or the like and overall without external intervention. The at least one suppression mass generates a defined imbalance at a first rpm or within a first rpm range. This defined imbalance can effectively suppress translatory vibrations generated at other locations. For a deviating rpm, it was observed that the excitation vibrations to be suppressed change in magnitude and/or phase. However, here the resulting total force, which acts on the suppression mass, changes and moves the suppression mass into a deviating equilibrium position. The changed radius and possibly also the changed phase angle of this additional equilibrium position is so dimensioned that the automatically changed imbalance generates a changed translatory vibration. This vibration is adapted to the rpm-dependent changed excitation vibration in such a manner that both translatory vibrations at least approximately mutually suppress each other. The rpm-dependent position transition between both equilibrium positions takes place in both directions so that an adapted suppression action takes place for rpms which are caused by operation and repeatedly varied, that is, increase and decrease.
In an advantageous embodiment, a radially inner stop is provided for a radially inner equilibrium position of the at least one suppression mass. Alternatively, or in addition, it is advantageous to provide a radially outer stop for a radially outer equilibrium position of the at least one suppression mass. The stop(s) effect a limiting of the movement of the suppression mass. Here, the total force, which acts on the suppression mass, is made up of the centrifugal force and the spring force and also the contact force of the stop. In a specific rpm range, the suppression mass remains fixed in its position. In this way, a fixed non-varying base match of the suppression effect is adjusted within the above-mentioned rpm range.
In a practical embodiment, equilibrium positions are provided which are uniformly distributed in radial direction in addition to or alternatively to the above-mentioned stops and equilibrium positions. These uniformly distributed equilibrium positions variably adjust in dependence upon the rpm because the centrifugal force and the opposing force are there in equilibrium. Without the action of the stops or the like, the radial deflection of the suppression mass changes continuously with the changing rpm. With increasing rpm, the radius of the suppression mass continuously increases whereas the radius continuously decreases with falling rpm. With the selection of a suitable spring characteristic line, a linear or even a nonlinear relationship can be established between rpm and radial deflection of the suppression mass depending upon the operating conditions. Each rpm is assigned a specific position of the suppression mass and therefore also a specific imbalance. At least section wise, a continuous rpm-dependent adaptation of the suppression action can be achieved on the excitation vibration which likewise changes in dependence upon rpm.
It can be practical to permit only a radial deflection of the at least one suppression mass. The rpm-dependent automatic adaptation of the suppressor is limited to a change of the imbalance magnitude. Advantageously, the suppression mass can be additionally arranged with a changing phase angle. The phase angle also changes for an rpm-dependent radial deflection. In this way, the situation can be accounted for that the excitation frequency, which is to be suppressed, can change not only with respect to its magnitude but also with respect to its phase in dependence upon rpm for specific arrangements. With an rpm-dependent phase change of the suppression mass adapted thereto, the suppression action can be further improved.
In a preferred embodiment, the suppression mass is displaceably guided in a translatory guide. Such a translatory guide can be configured in a simple manner, for example, by a simple radial bore in which the suppression mass is slideably held against the force of a spring element. With minimum manufacturing complexity, a precise and reliable arrangement is found which is protected against outside influences. Furthermore, almost any desired number of matching possibilities can be found. The translatory guide can, for example, be exactly radially arranged whereby an exclusively radial guidance of the suppression mass is provided. It is, however, also possible to arrange the translatory guide with radial and tangential directional components, that is, inclined to the radial direction. In this way, a tangential deflection of the suppression mass is additionally provided which is coupled to the radial deflection whereby the imbalance and the suppression action resulting therefrom are changeable not only with respect to their magnitude but also with respect to their phase. It is understood that the translatory guide is not limited to a linear configuration. A curve-shaped displacement path can also be practical which makes possible a nonlinear phase change in dependence upon the radial displacement path. Furthermore, the possibility is present to utilize rubber-elastic pressure-spring elements or the like which have a nonlinear spring characteristic line. The suppression mass can assume intermediate positions in radial and possibly also in tangential direction for specific rpms wherein the centrifugal force and the countering spring force are in equilibrium. Via a targeted adaptation of the nonlinear spring characteristic, a nonlinear displacement path of the suppression mass can also be adjusted in dependence upon the rpm or the centrifugal force resulting therefrom.
In an advantageous embodiment, the suppression mass, which is changeable with respect to its position, is journalled by means of a pivot arm on the vibration suppressor. The pivot arm permits a precise, low wear and robust guidance of the suppression mass.
Advantageously, at least two suppression masses, which are changeable in their positions, are each provided with different ratios of centrifugal force to spring force. The different mass, spring force and/or spring pretensioning are so selected that the individual suppression masses change their positions sequentially in a cascading manner as a function of the rpm. A finely stepped, rpm-dependent displacement of the resulting imbalance in magnitude and phase is also possible which facilitates a finely stepped adaptation to the excitation frequency characteristic.
It is practical to provide one stationary suppression mass and at least one suppression mass which is moveable with respect to its position. A base matching can be achieved with the fixed suppression mass. The suppression masses, which are changeable with respect to their positions, function only to provide the adaptation to rpms which deviate from the base matching. The suppression masses, which are changeable in their positions, can be configured to be correspondingly small whereby a reliable, precise guidance is simplified even at high rpm levels.
In a practical embodiment, the suppression mass, which is changeable in its position, is mounted angularly offset to the stationary suppression mass. Even a radial displacement of the individual suppression masses effects a shift of the total mass center of gravity of the vibration suppressor in magnitude and phase whereby an adaptation of the suppression performance is made possible with kinematically simple means.
The vibration suppressor of the invention can be mounted at different component assemblies of the work apparatus which are rotatably driven. In one embodiment of the drive motor as an internal combustion engine, the vibration suppressor is advantageously mounted on a crankshaft assembly and especially on a fan wheel for generating a cooling air flow. The fan wheel is part of the crankshaft assembly. The coupling of the vibration suppressor to the crankshaft assembly ensures that the vibration suppressor operates with identical rpm or frequency as the excitation oscillations at least of the engine without the constructively provided phase position between excitation vibration and suppressor oscillation being able to change. A permanent suppression action is ensured. The fan wheel has a comparatively large diameter wherein correspondingly small suppression masses can be accommodated without additional need for space.
The invention will now be described with reference to the drawings wherein:
In the embodiment shown, the drive motor 1 has a single cylinder 15 wherein a piston 17 is guided so as to reciprocate in the longitudinal direction. The piston 17 is connected to a crankshaft 19 by a connecting rod 18 for generating a rotational movement about a rotational axis 3.
The saw chain 29 runs along the edges of a guide bar 30. A guide wheel 32, which is rotatable about an axis 31, is provided at the end of the guide bar 30 facing away from the clutch 22 for changing the direction of the saw chain 29. In the region of the end of the guide bar 30 close to the engine, the saw chain 29 engages around a clutch 22 which is attached to an end of the crankshaft 19. The saw chain 29 is driven via the clutch 22 starting at a pregiven rpm of the crankshaft 19.
A fan wheel 14 is at the end of the combustion engine 1 and lies opposite the clutch 22. The fan wheel 14 is for cooling the engine especially in the region of the cylinder 15 and is driven by the crankshaft 19. The fan wheel carries an ignition magnet 23 which passes by a housing-fixed ignition coil 24, which is radially on the outside, with the rotation of the fan wheel. In the ignition coil 24, an ignition voltage is generated for a spark plug 21 mounted in the cylinder 15 whereby an air/fuel mixture in the interior of the cylinder 15 is ignited. Spark plug 21, ignition magnet 23 and ignition coil 24 are parts of an ignition system 20.
The clutch 22, the crankshaft 19 and the fan wheel 14 are fixedly connected to each other. They form a crankshaft assembly 13 with a uniform rpm during operation. The drive motor 1 with its crankshaft assembly 13 is mounted in a motor housing 25. The clutch 22 is covered by a clutch cover 26. Forward and rearward handles (27, 28) are attached to the motor housing 25 for guiding the chain saw 16.
In the embodiment shown, the vibration suppressor 2 includes overall three suppression masses (4, 5, 6) for generating a targeted imbalance. The suppression masses (4, 5, 6) are arranged at a radius to the rotational axis 3. With a rotation of the vibration suppressor 2, this imbalance generates an rpm-dependent translatory vibration which is provided to suppress another translatory vibration. Such a translatory vibration, which is to be suppressed can, for example, be brought about by the saw chain 29 (
The first suppression mass 4 lies fixed on the fan wheel 14. The two additional suppression masses (5, 6) are pivotally journalled on vibration suppressor 2 (that is, the fan wheel 14) by means of respective pivot arms (7, 8). Springs (9, 10) act on the pivot arms (7, 8), respectively, and pull the corresponding pivot arm (7, 8) with the corresponding suppression mass (5, 6) with a spring force under pretension radially inwardly into the position shown. The suppression masses (5, 6) are supported by radially inner stops (39, 40) radially inwardly against the pretensioning force of the springs (9, 10).
The suppression masses (4, 5, 6) generate centrifugal forces with the rotation of the illustrated arrangement at idle rpm and in a mid rpm range. The centrifugal forces are indicated by respective arrows (35, 36, 37) and are directed radially outwardly from the rotational axis 3. The centrifugal forces (36, 37) are not sufficient to overcome the opposing spring forces of the springs (9, 10). Both movable suppression masses (5, 6) are each in a radial inner equilibrium position when contacting against the radially inner stops (39, 40). In these equilibrium positions, the centrifugal forces (36, 37), which spring forces act effectively on the suppression masses (5, 6), and the contact forces at the stops (39, 40) are in equilibrium with each other.
An arrow 38, which shows the resultant centrifugal force, can be formed from a geometric addition of the arrows (35, 36, 37). The suppression masses (4, 5, 6) are shown angularly offset with respect to each other and effect a center of gravity shift of the balanced fan wheel 14 away from the rotational axis 3 radially outwardly in the direction of the arrow 38. It is in this direction of arrow 38 that the resulting imbalance or centrifugal force also acts.
As a consequence of the rotation of the arrangement shown, a translatory oscillation arises which, in magnitude and phase, is so matched to the excitation oscillation of the work apparatus of
Above constructively predetermined limit rpms, the moveably supported suppression masses (5, 6) can move radially outwardly along arcuately-shaped displacement paths (33, 34). The displacement paths (33, 34) are limited outwardly by assigned stops (11, 12), respectively. When contacting the radial outer stops (11, 12), a further deviating radially outer equilibrium position adjusts wherein the following are in equilibrium with each other: the centrifugal forces of
The suppression masses (5, 6), the corresponding springs (9, 10) and their stiffnesses, pretensionings and geometric relative arrangement are so matched to each other that a different effective spring pretensioning results at the two suppression masses (5, 6). The effective spring pretensionings are so selected that the centrifugal force, which acts on the suppression mass 5, is sufficient in order to overcome the pretensioning of the assigned spring 9. The total force, which acts on the suppression mass 5, is directed radially outwardly. This total force results from the assigned centrifugal force and the countering spring force. The pivot arm 7 pivots automatically because of the action of the resulting total force in common with the suppression mass 5 from the radial inner equilibrium position into the radial outer equilibrium position identified by reference numeral 5′. This radial outer equilibrium position is radially outwardly delimited by the stop 11. The suppression mass 5′ is displaced with a radial deflection (a) and a phase angle changed by Δα compared to its position shown in
The rpm increased relative to
A geometric addition of the arrows (36′, 35 and 37) leads to a resultant centrifugal force or unbalance force (shown by arrow 38′) which is changed by a phase change angle Δφ and a radius Δr relative to the arrow 38 of
In the absence of an external load, a further rpm increase can occur up to a maximum rpm. In this situation, a configuration of the vibration suppressor 2 of
The pivot arm 8 with the suppression mass 6 is deflected radially outwardly up to the position delimited by the stop 12 and identified by reference numeral 6′. Compared to its original position identified by reference numeral 6, the suppression mass 6′ is displaced by a radial deflection path (b) as well as by a deflection angle Δβ. A centrifugal force, which is shown by arrow 37′, acts on the suppression mass 6′. This damping force, when geometrically added to arrows 36′ and 35, leads to a resultant centrifugal force 38″. The phase change angle Δφ relative to the original position of arrow 38 of
The return movement of the suppression masses (5, 6) of
In the embodiment of FIGS. 2 to 4, at least equilibrium positions are defined, namely, respective outer and inner equilibrium positions of the suppression masses (5, 6) via the stops (11, 12, 39, 40). For a corresponding design of the springs (9, 10), many equilibrium positions can be generated which are distributed in radial direction or positioned between these stops (11, 12, 39, 40). For specific rpms, the suppression masses (5, 6) can assume intermediate positions wherein, without the action of a stop, the effective centrifugal force and the counter spring force are in equilibrium. For the occurring imbalance, the above applies in the same manner with respect to magnitude and phase.
The embodiment shown has a suppression mass 4, which is fixed on the vibration suppressor 2, and two additional suppression masses (5, 6) which change with respect to their positions. Another number of changeable suppression masses (5, 6) can be practical. Likewise, it can be advantageous to do without a fixed suppression mass 4 and, in total, provide at least one suppression mass (5, 6) changeable with respect to its position.
In the embodiment shown, the suppression masses (5, 6) are so pivotally guided that they change their positions with respect to radius and phase angle in dependence upon the occurring rpm. As a result, a change of the resulting imbalance adjusts with respect to magnitude and phase. A comparable effect can also be obtained with a displacement of the suppression masses (5, 6) which is exclusively radial or exclusively tangential. Leaf springs for supporting and holding the suppression masses (5, 6) can be practical in lieu of the pivot arms (7, 8) and their springs (9, 10). The springs (9, 10) can have any desired configuration. In addition to metal helical springs, leaf springs or spiral springs, also elastic spring bodies made of plastic and especially made of elastomer can be considered.
The bore 44 is configured as a blind bore. The radial inner end of the blind bore faces toward the rotational axis 3 and defines a radially inner stop 41 for the suppression mass 5. The radial outer end of the bore 44 is closed with a plug 45 at the peripheral contour of the vibration suppressor 2. A pressure spring 43 is arranged between the stop 45 and the suppression mass 5. The spring force of pressure spring 43 presses the suppression mass 5 radially inwardly toward the inner stop 41. The spring characteristic line of the pressure spring 43 and its pretensioning are matched in such a manner to the suppression mass 5 that the suppression mass 5 remains pressed against the radial inner stop 41 below a lower limit rpm. Here, a first equilibrium position adjusts which is made up of the acting centrifugal force, the spring force and the contact force at the stop 43.
When the lower limit rpm is exceeded, the centrifugal force, which acts on the suppression mass 5, becomes so great that the spring force, which acts in the first equilibrium position, is overcome and the stop force vanishes. The total force, which adjusts, is directed radially outwardly and brings about a displacement path of the suppression mass 5 radially outwardly against the spring force. As a consequence of the radial displacement path, the spring force of the pressure spring 43 increases. For a suitable matching of its spring characteristic line, different equilibrium positions 5″ of the suppression mass adjust wherein the centrifugal force and the countering spring force are in equilibrium. The radial position of the equilibrium position of the suppression mass 5″ increases continuously with increasing rpm or drops continuously with decreasing rpm.
As soon as an upper limit rpm is reached or exceeded, a radial displacement path of the suppression mass 5 adjusts wherein the suppression mass in its radially outer equilibrium position 5′ is pressed against a radial outer stop 42 via the action of the centrifugal force. It can also be practical that the pressure spring 43 is pressed together to the length of a block. The pressure spring 43, which is pressed together to the length of the block, then forms the radial outer stop 42 for the suppression mass. A further advantageous option for all embodiments can be to omit entirely radial inner and/or radial outer stops (11, 12, 39, 40, 41, 42). Radial inner or radial outer equilibrium positions of the suppression mass 5 adjust via the force equilibrium between centrifugal force on the suppression mass 5 and the counter spring force.
The pressure spring 43 is, for example, configured as a metal helical pressure spring. Rubber elastic pressure spring elements, tension spring elements or the like with comparative spring action can also be used. The pressure spring 43 shown has a linear spring characteristic line by way of example. However, a configuration having a nonlinear spring characteristic line can be practical. In this way, an equilibrium position of the suppression mass 5″ can adjust which is adapted to the particular operating conditions and is nonlinear but changes continuously with the rpm.
In the embodiment shown, the bore 44 exhibits an inclination in the peripheral direction in addition to its radial alignment whereby a tangentially directed component of the translatory displacement path is formed. Accordingly, a phase angle change Δφ with any desired intermediate values occurs from the radial displacement path of the suppression mass 5 with the radius difference ΔR between the radial outer equilibrium position and the radial inner equilibrium position. In the case of a linear bore 44, a linear relationship is present between the radial difference ΔR and the phase change angle Δφ. As required, also a nonlinear relationship can be established via a curved translatory guide corresponding to the arcuately-shaped slot 46. A further advantageous option can be to arrange the bore 44 or another suitable translatory guide exclusively in radial direction corresponding to the bore 44′ shown in phantom outline. Here, no phase change angle Δφ results between the different equilibrium positions of the suppression mass 5.
With respect to the remaining features, reference numerals and information as to operation, the embodiment of
It is understood that the foregoing description is that of the preferred embodiments of the invention and that various changes and modifications may be made thereto without departing from the spirit and scope of the invention as defined in the appended claims.
| Number | Date | Country | Kind |
|---|---|---|---|
| 10 2004 056 919.3 | Nov 2004 | DE | national |
This is a continuation-in-part application of U.S. patent application Ser. No. 11/286,398, filed Nov. 25, 2005, and claims priority of German patent application no. 10 2004 056 919.3, filed Nov. 25, 2004, the entire contents of which are incorporated herein by reference.
| Number | Date | Country | |
|---|---|---|---|
| Parent | 11286398 | Nov 2005 | US |
| Child | 11802351 | May 2007 | US |
