Power tool

It is an object of the invention to provide a technique for further improving the vibration reducing performance in the power tool, while avoiding complicating the construction of the power tool. According to the present invention, a representative power tool may comprise a striker, a tool bit and a vibration reducer. The vibration reducer serves to reduce vibration on the striker by reciprocating in a direction opposite to the reciprocating direction of the striker. The path of the center of gravity of the vibration reducer is arranged to coincide with a path of the center of gravity of the striker. With such construction, because rotating moment is not exerted onto the reciprocating cylinder during the operation of the power tool, vibration reduction can be performed in a stable manner.

BACKGROUND OF THE INVENTION

1. Field of the Invention

The present invention relates to a power tool, and more particularly, to a technique of reducing and alleviating vibration in a power tool, such as a hammer and a hammer drill.

2. Description of the Related Art

Japanese non-examined laid-open Patent Publication No. 52-109673 discloses a hammer with a vibration reducing device. The known hammer includes a vibration-isolating chamber provided in the region under the body housing of the hammer. A dynamic vibration reducer is housed in the vibration-isolating chamber and serves to reduce and alleviate strong vibration developed in the axial direction of the hammer during the operation.

However, the vibration-isolating chamber is separately formed within the body housing and components parts of the dynamic vibration reducer are incorporated therein. Therefore, the construction and assembling operation are complicated and the weight of the entire hammer is increased. Further, because the space for housing the dynamic vibration reducer must be ensured, the appearance of the hammer is impaired.

SUMMARY OF THE INVENTION

Accordingly, it is an object of the present invention to provide a technique for further improving the vibration reducing performance in the power tool, while avoiding complicating the construction of the power tool.

According to the present invention, a representative power tool may comprise a striker, a tool bit and a vibration reducer. The striker reciprocates by pressure fluctuations within a cylinder. The tool bit performs a predetermined operation by a striking force of the striker. The vibration reducer serves to reduce vibration on the striker by reciprocating in a direction opposite to the reciprocating direction of the striker. The path of the center of gravity of the vibration reducer is arranged to coincide with a path of the center of gravity of the striker. With such construction, the vibration reducer can be closely associated with the striker without requiring any vibration-isolating chamber, it can be avoided to complicate the construction of the power tool with a vibration reducing function. Further, because the paths of the center of gravity of the striker and the vibration reducer coincide to each other and thus rotating (turning) moment is not exerted onto the reciprocating cylinder during the operation of the power tool, vibration reduction can be performed in a stable manner.

Other objects, features and advantages of the present invention will be readily understood after reading the following detailed description together with the accompanying drawings and the claims.

DETAILED DESCRIPTION OF THE INVENTION

According to the present invention, a representative power tool may comprise a striker, a tool bit and a vibration reducer. The striker reciprocates by pressure fluctuations within a cylinder. The striker may directly collide with the tool bit by pressure fluctuations within the cylinder. Alternatively, the striker may be driven by pressure fluctuations within the cylinder and caused to collide with another impact force transmitting element such as an impact bolt, which in turn is caused to collide with the tool bit. The tool bit performs a predetermined operation by a striking force of the striker. The vibration reducer serves to reduce vibration on the striker by reciprocating in a direction opposite to the reciprocating direction of the striker. The path of the center of gravity of the vibration reducer is arranged to coincide with a path of the center of gravity of the striker. With such construction, because rotating (turning) moment is not exerted onto the reciprocating cylinder during the operation of the power tool, vibration reduction can be performed in a stable manner.

In the power tool of the present invention, the cylinder may preferably reciprocate in a direction opposite to the reciprocating direction of the striker such that the reciprocating cylinder functions as a counter weight that reduces the vibration caused by the striker. In order to cause the cylinder to reciprocate, typically, a crank mechanism that converts a rotating output of a driving motor to linear motion may be used.

Because a power tool such as a hammer inherently includes a cylinder to drive the striker and such an existing cylinder can be utilized as a vibration reducer, the design of the power tool with a vibration reducing function can be simplified. Thus, the power tool can be simpler in construction and can be manufactured at reduced costs, having a lighter weight and better appearance.

The striker and the cylinder may be separately caused to reciprocate by a first crank and a second crank which respectively convert a rotating output of a driving motor to linear motion. In other words, a crank for driving the striker to reciprocate and a crank for driving the cylinder to reciprocate may be separately provided. Further, in an actual operation of the power tool, the striker typically starts to strike the tool bit with a certain time delay after the movement of the piston that causes pressure fluctuations within the cylinder. Therefore, the first crank and the second crank may preferably be driven with a different timing so that the cylinder reciprocates in a direction opposite to the reciprocating direction of the striker. The striker and the cylinder may preferably be driven via the first and the second crank mechanisms by using a common driving motor.

Instead of utilizing the cylinder as a vibration reducer, the vibration reducer may comprise a counter weight disposed along the entirety or part of the outer circumferential surface of the cylinder. In such case, the counter weight reciprocates to alleviate an impact force during hammering operation, thereby performing vibration reduction against the impact force. In utilizing such counter weight, a rotation preventing mechanism may preferably be disposed between the body and the counter weight in order to prevent the counter weight from moving in the circumferential direction of the cylinder. Further, an air vent may be provided in the cylinder such that outside air can be introduced into the cylinder when the pressure within the cylinder decreases. The air vent may be opened and closed when the counter weight reciprocates on the cylinder.

Further, the power tool may comprise first crank mechanism to drive the striker by reciprocating a driver within the cylinder and second crank mechanism to reciprocate the counter weight. The first and second crank mechanisms may be supported by first and second bearings. By such construction, the driver and the counter weight can be driven with stability.

Each of the additional features and method steps disclosed above and below may be utilized separately or in conjunction with other features and method steps to provide improved power tools and devices utilized therein. Representative examples of the present invention, which examples utilized many of these additional features and method steps in conjunction, will now be described in detail with reference to the drawings. This detailed description is merely intended to teach a person skilled in the art further details for practicing preferred aspects of the present teachings and is not intended to limit the scope of the invention. Only the claims define the scope of the claimed invention. Therefore, combinations of features and steps disclosed within the following detailed description may not be necessary to practice the invention in the broadest sense, and are instead taught merely to particularly describe some representative examples of the invention, which detailed description will now be given with reference to the accompanying drawings.

FIRST REPRESENTATIVE EMBODIMENT

First representative embodiment of the present invention will now be described with reference to the drawings. As shown inFIG. 1, an electric hammer101as a representative embodiment of the power tool according to the present invention comprises a body103, a tool holder117connected to the tip end region of the body103, and a hammer bit119detachably coupled to the tool holder117. The hammer bit119is a feature that corresponds to the “tool bit” according to the present invention.FIG. 2shows the electric hammer101in plan view.

The body103includes a motor housing105, a gear housing107and a handgrip109. The motor housing105houses a driving motor111. The gear housing107houses a first motion converting mechanism113, a second motion converting mechanism213and a striking mechanism115. The first motion converting mechanism113is adapted to convert the rotating output of the driving motor111to linear motion and then to transmit it to the striking mechanism115. As a result, an impact force is generated in the axial direction of the hammer bit119via the striking mechanism115.

Further, the second motion converting mechanism213is adapted to convert the rotating output of the driving motor111to linear motion and then to transmit it to a cylinder129that defines a vibration reducing mechanism201. As a result, the cylinder129is caused to reciprocate in its axial direction as to correspond to the impact force by the striking movement of the hammer bit119. Thus, vibration caused in the hammer101can be alleviated or reduced. The hammer101may be configured such that it can be switched over by the user to a hammer drill mode and a hammer-drill mode.

FIG. 2shows a detailed construction of the first and second motion converting mechanisms113,213of the electric hammer101. The first motion converting mechanism113includes a driving gear121, an intermediate gear122, a driven gear123, a first crank disc124, a first eccentric shaft (crank pin)125and a first connecting rod126. The driving gear121is rotated in a vertical plane by the driving motor111. The intermediate gear122rotates together with the driving gear121and the driven gear123engages the intermediate gear122. The first crank disc124rotates together with the driven gear123. The first eccentric shaft125is eccentrically disposed in a position displaced from the center of rotation of the first crank disc124. One end of the first connecting rod126is loosely connected to the first eccentric shaft125and the other end is loosely connected to a driver in the form of a piston128via a first connecting shaft127. The first crank disc124, the first eccentric shaft125and the first connecting rod126form a first crank mechanism. The first crank mechanism is a feature that corresponds to the “first crank” according to the present invention.

Further, as shown inFIG. 1, a striking mechanism115includes a striker131and an impact bolt133. The striker131is slidably disposed within the bore of the cylinder129together with the piston128. The impact bolt133is slidably disposed within the tool holder117and is adapted to transmit the kinetic energy of the striker131to the hammer bit119.

As shown inFIG. 2, the cylinder129is disposed within a barrel108connected to the gear housing107and can slide in the axial direction. The cylinder129functions as a counter weight for reducing vibration during hammering operation by reciprocating in a direction opposite to the sliding direction of the striker131. In other words, the cylinder129that reciprocates in a direction opposite to the sliding direction of the striker131defines the vibration reducing mechanism201in the barrel108.

InFIG. 2, a path of the center of gravity of the cylinder129reciprocating within the barrel108is shown by reference symbol “P”, while a path of the center of gravity of the piston128as well as the striker131reciprocating within the cylinder129is shown by reference symbol “Q”. The path P of the center of gravity of the cylinder129is arranged substantially to coincide with the path Q of the center of gravity of the piston128and the striker131.

As shown inFIG. 2, the second motion converting mechanism213that causes the cylinder129to reciprocate includes a second crank disc221, a second eccentric shaft (crank pin)223and a second connecting rod225. The second eccentric shaft223is eccentrically disposed in a position displaced from the center of rotation of the second crank disc221on the edge portion of the second crank disc221. One end of the second connecting rod225is loosely connected to the second eccentric shaft223and the other end is loosely connected to the cylinder129via a second connecting shaft227. The second crank disc221, the second eccentric shaft223and the second connecting rod225form a second crank mechanism. The second crank mechanism is a feature that corresponds to the “second crank” according to the present invention.

The second crank disc221is arranged such that its axis of rotation substantially coincides with the axis of rotation of the first crank disc124of the first motion converting mechanism113. The second crank disc221is loosely connected to the first eccentric shaft125in a position displaced from its axis of rotation. As shown inFIG. 3, this connection is achieved by the fact that a U-shaped engaging portion221aof the second crank disc221loosely engages with a small-diameter portion125aof the first eccentric shaft125. Thus, power is taken out from the power transmission path of the first motion converting mechanism113driven by the driving motor111and such power is utilized to drive the second motion converting mechanism213. The second connecting rod225is connected to the cylinder129via a joint ring229fitted around the axial end of the cylinder129and the second connecting shaft227fitted in the joint ring229.

A phase difference is provided between the reciprocating movement of the striker131and the reciprocating movement of the cylinder129. By such phase difference, the cylinder129reciprocates in a direction opposite to the reciprocating direction of the striker131. The striker131is driven by the action of an air spring caused within the cylinder129by means of sliding movement of the piston128. The striker131therefore moves with a predetermined time delay with respect to the movement of the piston128. As shown inFIG. 3, a phase difference (delay with respect to the piston128) between a point of connection of the second connecting rod225to the second crank disc221via the second eccentric shaft223and a point of connection of the first connecting rod126to the first crank disc124via the first eccentric shaft125is about 270° in the rotational direction (counterclockwise direction as viewed inFIG. 3) of the first and the second crank discs124and221. Therefore, the second motion converting mechanism213is arranged to drive the cylinder129with a delay of about 270° in terms of a crank angle with respect to the first motion converting mechanism113.

FIG. 3schematically shows a relative positional relationship of the piston128, the cylinder129and the first and the second connecting rods126and225when the hammer101is in the state shown inFIG. 2. InFIGS. 2 and 3, the piston128is shown at a non-compression side dead point (sliding end when slid toward the driving motor111, or retracting end).

Operation of the hammer101constructed as described above will now be explained. When the driving motor111(shown inFIG. 1) is driven, the rotating output of the driving motor111causes the driving gear121(shown inFIG. 2) to rotate. When the driving gear122rotates, the first crank disc124rotates via the intermediate gear122and the driven gear123. Then, the first eccentric shaft123on the first crank disc124revolves, which in turn causes the first connecting rod126to swing. The piston128on the end of the first connecting rod126then slidingly reciprocates within the cylinder129. When the piston128slides toward the hammer bit119from the non-compression side dead point, a force of moving the striker131toward the hammer bit119acts on the striker131by the action of the air spring function as a result of the compression of the air within the cylinder147between the striker and the impact bolt. Thus, the striker131reciprocates within the cylinder129at a speed higher than the piston128in the same direction and collides with the impact bolt133. The kinetic energy (striking force) of the striker131caused by the collision with the impact bolt133is transmitted to the hammer bit119. Thus, the hammer bit119slidingly reciprocates within the tool holder117and performs a hammering operation on the workpiece.

FIG. 1shows the state in which the striker131has transmitted the striking force to the hammer bit119via the impact bolt133, while the piston128that drives the striker131has retracted to the non-compression side dead point after the compression process of the air spring. The actual sliding movement of the striker131including collision with the impact bolt133occurs with a predetermined time delay after the sliding movement of the piston128in relation to the time required for the air spring to act on the striker131and the inertial force of the striker131.

On the other hand, within the second motion converting mechanism213, the second crank disc221rotates as the first eccentric shaft125is caused to revolve by rotation of the first crank disc124. Then, the second eccentric shaft223on the second crank disc221revolves, which in turn causes the second connecting rod126to swing. The cylinder129then slidingly reciprocates within the barrel108.

At this time, the cylinder129slides in a direction opposite to the sliding direction of the striker131when the striker131slides toward the impact bolt133. This is because, in the hammer, certain time is necessary to drive the striker131after the piston128starts to compress the air within the air spring chamber129afor increasing the pressure within the air spring chamber129a. Therefore, a phase difference is provided such that the cylinder129reciprocates in a direction opposite to the reciprocating direction of the striker131with an appropriate timing with respect to the reciprocating movement of the striker131(specifically, a phase difference of about 270° is provided between the point of connection of the second connecting rod225to the second crank disc221and the point of connection of the first connecting rod126to the first crank disc124). According to this embodiment, the cylinder129functions as a “counter weight” by actively reciprocating in a direction opposite to the reciprocating direction of the striker131. As a result, vibration caused in the hammer101when the striker131collides with the impact bolt133can be reduced.

When the piston128slides away from the compression side dead point, a force of moving the striker131away from the hammer bit119acts on the striker131by the action of the air spring upon the inflation side (the side opposite to the piston128). When the piston128slides to the non-compression side dead point, the striker131starts to slide away from the hammer bit119. This sliding movement of the striker131continues even if the piston128reaches the non-compression side dead point and starts to slide in the reverse direction toward the compression side dead point. During the retracting movement of the striker131away from the hammer bit119, the cylinder129also slides in a direction opposite to the sliding direction of the striker131. Thus, the vibration reducing mechanism effectively functions with the actively driven cylinder129. The weight of the cylinder129that functions as a counter weight may appropriately be selected such that a vibration reducing force to be obtained by the cylinder129can be maximized. When the cylinder129slides within the barrel108, the capacity of the space within the housing which faces the axial end of the cylinder129fluctuates. Preferably, said space may be configured to communicate with the outside in order to reduce pressure fluctuations which are caused by such capacity fluctuations and thus to prevent the capacity fluctuations from interfering with the sliding movement of the cylinder129.

According to the embodiment, as shown inFIG. 3, the path “P” of the center of gravity of the cylinder129substantially coincides with the path “Q” of the center of gravity of the piston128and the striker131. If, for example, the counter weight is disposed in a position displaced from the path of the striker, a rotating moment will be exerted on the cylinder and that may cause another vibration. According to this embodiment, such problem is eliminated and vibration reduction can be performed in a stable manner.

As shown inFIG. 1, the hammer101according to this embodiment is constructed as a relatively large-sized hammer including a handgrip109on the both right and left sides of the body103and mainly used for chipping floors. In a normal manner of using the hammer101of this type, the hammer bit119is pressed against the workpiece or the floor surface under the own weight of the hammer101, so that a load is applied to the hammer bit119. The vibration reducing mechanism201is especially useful for such type of hammer because the hammer of this type is normally driven under loaded condition and therefore vibration reducing is always required. Otherwise, if the hammer is driven under unloaded condition, the cylinder129that always reciprocates during the operation may uselessly cause vibration.

While, in this embodiment, the striking force of the striker131is transmitted to the hammer bit119via the impact bolt133, the present invention can also be applied to the configuration in which the striker131directly collides with the hammer bit119.

SECOND REPRESENTATIVE EMBODIMENT

Second representative embodiment of the present invention is now explained in greater detail in reference toFIGS. 4 to 8. In explaining the second embodiment, features having substantially the same constructions with the respective features utilized in the above-explained first embodiment are shown with same reference numbers in the drawings. As shown inFIGS. 4 and 5, the cylinder129of the second representative embodiment is fixedly disposed within the barrel108that is connected to the gear housing107. Further, a cylindrical counter weight231is disposed between the outer circumferential surface of the cylinder129and the inner circumferential surface of the barrel108. The cylindrical counter weight231can slide in the axial direction of the hammer bit119so as to function as a vibration reducing weight during hammering operation by reciprocating in a direction opposite to the sliding direction of the striker131. A cylindrical accommodation space233for accommodating the counter weight231is defined between the outer circumferential surface of the cylinder129and the inner circumferential surface of the barrel108. The accommodation space233has an axial length long enough to allow the counter weight231to slide in its axial direction.

InFIG. 4, a path of the center of gravity of the counter weight231that reciprocates within the barrel108is shown by reference symbol “P”, while a path of the center of gravity of the piston129as well as the striker131reciprocating within the cylinder129is shown by reference symbol “Q”. The path P of the center of gravity of the counter weight231substantially coincides with the path Q of the center of gravity of the piston128and the striker131.

As shown inFIGS. 4 and 5, the second motion converting mechanism213is provided in order to cause the counter weight231to reciprocate. The mechanism213includes a second crank disc221, a second eccentric shaft (crank pin)223and a second connecting rod225. The second eccentric shaft223is eccentrically disposed in a position displaced from the center of rotation of the second crank disc221on the edge portion of the second crank disc221. One end of the second connecting rod225is loosely connected to the second eccentric shaft223and the other end is loosely connected to the counter weight231via a second connecting shaft227. The second crank disc221, the second eccentric shaft223and the second connecting rod225forms a second crank mechanism. The counter weight231reciprocates via the second crank mechanism between the advancing end nearest to the hammer bit119and the retracting end remotest from the hammer bit119.

The second crank disc221is arranged such that its axis of rotation substantially coincides with the axis of rotation of the first crank disc124of the first motion converting mechanism113. The second crank disc221is loosely connected to the first eccentric shaft125in a position displaced from its axis of rotation. As shown inFIG. 6, this connection is achieved by the fact that a U-shaped engaging portion221aof the second crank disc221loosely engages with a small-diameter portion125aof the first eccentric shaft125. The second crank disc221is rotatably supported by a second bearing229.

Further, as shown inFIG. 7, a rotation preventing mechanism235is provided in the mounting area of the second connecting shaft227. Via the shaft227, the counter weight231is connected to the second connecting rod225. The rotation preventing mechanism235prevents the counter weight231from moving in its circumferential direction. The rotation preventing mechanism235comprises a guide groove237and an engaged sliding portion239. The guide groove237is formed in the inside of a portion of the barrel108that bulges outside. The engaged sliding portion239is formed in the shaft mounting portion on the outer circumferential surface of the counter weight231so as to bulge outside. The guide groove237extends in a direction parallel to the moving direction of the counter weight231. The engaged sliding portion239slidably engages in the guide groove237. The counter weight231is prevented from moving in its circumferential direction by the engaged sliding portion239being in contact with the wall surface of the guide groove237in the circumferential direction. In order to achieve smooth sliding movement of the engaged sliding portion239along the guide groove237, a slide plate241is disposed on the sliding surface between the guide groove237and the engaged sliding portion239. The guide groove237and the engaged sliding portion239form an engaged sliding structure along the entire extent of movement of the counter weight231.

In this embodiment, a phase difference is provided between the reciprocating movement of the piston128and the reciprocating movement of the counter weight231such that the counter weight231reciprocates in a direction opposite to the reciprocating direction of the striker131that applies an impact force to the hammer bit119via the impact bolt133. As shown inFIG. 6, a phase difference between a point of connection of the second connecting rod225to the second crank disc221via the second eccentric shaft223and a point of connection of the first connecting rod126to the first crank disc124via the first eccentric shaft125is about 260° in the rotational direction (counterclockwise direction as viewed inFIG. 6) of the first and the second crank discs124and221.

As shown inFIGS. 4 and 5, a slide ring243is provided on the inner circumferential surface of the counter weight231on its both ends in the sliding direction in order to achieve smooth sliding movement of the counter weight231. As particularly shown inFIG. 8, the slide ring243has a C-ring shape with a notch243ain a circumferential portion. The slide ring243is fitted in a groove231aformed in the inner circumferential surface of the counter weight231. The slide ring243is formed of a synthetic resin, such as polyacetal, which is slippery and highly resistant to wear.

Further, as shown inFIGS. 4 and 5, an air vent245for controlling the pressure within the air spring chamber129ais formed in the cylinder129. The air vent245communicates the air spring chamber129awith the outside (the crank chamber) via a clearance247, communication holes249, passages251. The clearance247is defined between the outer circumferential surface of the cylinder129and the inner circumferential surface of the counter weight231. Communication holes249are formed in the counter weight231. Passages251(seeFIG. 7) are formed between the outer circumferential surface of the counter weight231and the inner circumferential surface of the barrel108. The passages are arranged at predetermined intervals in the circumferential direction. As to the above-explained slide rings243, the rear one (right one as viewed in the drawings) opens and closes the air vent245. Specifically, the rear slide ring243comprises an opening-and-closing valve for opening and closing the air vent245. The rear slide ring243will be hereinafter referred to as an opening-and-closing valve.

The opening-and-closing valve243is in sliding contact with the outer circumferential surface of the cylinder129while exerting a predetermined biasing force on it. Then, when the air vent245is closed, the inside is kept airtight. The opening-and-closing valve243closes the air vent245in a predetermined region (in the range of about 160 to 200° by the crank angle of the second crank mechanism, taking the position of the retracting end as 0° (360°)) in the neighborhood of the advancing end within the range of movement of the counter weight231(seeFIG. 6), while it opens the air vent245in the other region. In other words, the opening-and-closing valve243closes the air vent245in an effective compression region (in the range of about 60 to 100° by the crank angle of the first crank mechanism) in obtaining a strong striking force of the striker131in the process of compression by the piston128, while it opens the air vent245in a region other than the effective compression region.

Operation of the hammer101constructed as described above will now be explained. When the driving motor (not particularly shown in the drawings) is driven, the rotating output of the driving motor causes the first crank disc124(shown inFIG. 4) to rotate. As a result, the first eccentric shaft123on the first crank disc124revolves, which in turn causes the first connecting rod126to swing. The piston128on the end of the first connecting rod126then slidingly reciprocates within the cylinder129to drive the striker131.

On the other hand, as to the second motion converting mechanism213, the second crank disc221rotates as the first eccentric shaft125is caused to revolve by rotation of the first crank disc124. Then, the second eccentric shaft223on the second crank disc221revolves, which in turn causes the second connecting rod126to swing. The counter weight231then slidingly reciprocates along the outer circumferential surface of the cylinder129. The counter weight231slides in a direction opposite to the sliding direction of the striker131when the striker131slides toward the impact bolt133. This is because a phase difference is provided such that the counter weight231reciprocates in a direction opposite to the reciprocating direction of the striker131with an appropriate timing with respect to the reciprocating movement of the striker131.

According to the second representative embodiment, the counter weight231is caused to reciprocate in its axial direction with such timing as to correspond to the impact force by the striking movement of the hammer bit119. In this manner, vibration caused in the hammer101can be alleviated.

When the piston128moves toward the compression side dead point and reaches the intermediate region (in the range of about 60 to 100° by the crank angle of the first crank mechanism), the air spring chamber129ais in the optimum compression region, and when it is in a position of about 100° by the crank angle, it is in the maximum compression state (seeFIG. 5). At this time, the counter weight231which is driven with a delay of about 260° with respect to the piston128is located in a region (in the range of about 160 to 200° by the crank angle of the second crank mechanism) in the neighborhood of the advancing end nearest to the hammer bit119. In this region, the opening-and-closing valve243on the counter weight231closes the air vent245. This means that the opening-and-closing valve243closes the air vent245when the air spring chamber129ais in the optimum compression region. Therefore, communication of the air spring chamber129awith the outside is interrupted, so that air within the air spring chamber129ais prevented from flowing out to the outside. As a result, loss the compression efficiency within the cylinder can be improved and the striker131can produce a stronger striking force.

When the piston128slides away from the hammer bit119from the compression side dead point, the counter weight231is moved in the retracting direction from the advancing end. At this time, the opening-and-closing valve243opens the air vent245, so that the air spring chamber129acommunicates with the outside. Thus, the outside air is introduced into the air spring chamber129aand the suction force within the cylinder is weakened. As a result, the striker131is prevented from moving toward the piston128beyond its proper position.

In regard to the timing for the opening-and-closing valve243to open and close the air vent245, in this embodiment, it closes the air vent245in the range of about 160 to 200° by the crank angle of the second crank mechanism. However, this timing can be appropriately set by adjusting the width (ring width) of the opening-and-closing valve243in the moving direction, in consideration of the effectiveness of preventing outflow of the air within the air spring chamber129aand the optimization of the return movement of the striker131.

Further, when the counter weight231slides along the outer circumferential surface of the cylinder129, the capacity of the accommodation space233which faces the axial end of the counter weight231fluctuates. In this embodiment, however, the accommodation space233communicates with the crank chamber via the passages251that comprise grooves formed in the inner circumferential surface of the barrel108. Therefore, pressure fluctuations caused within the accommodation space233by the capacity fluctuations can be reduced and thus, the counter weight231can smoothly slide.

In this embodiment, the counter weight231is disposed between the barrel108and the outer circumferential surface of the cylinder129and serves to reduce vibration on the striker131by reciprocating in a direction opposite to the reciprocating direction of the striker131. For this purpose, the accommodation space233for the counter weight231is provided between the outer circumferential surface of the cylinder129and the barrel108. By such construction, a space for accommodating the counter weight231can be ensured without substantial change in the appearance of the barrel108.

Further, in this embodiment, a path P of the center of gravity of the counter weight231substantially coincides with the path Q of the center of gravity of the piston128and the striker131. As a result, vibration reduction can be performed in a stable manner.

When the second crank mechanism is driven, the counter weight231may possibly receive a force (rotational force) to move the counter weight231in its circumferential direction via the second connecting shaft227. According to the second embodiment, as shown inFIGS. 4 and 7, the rotation preventing mechanism235bears such rotational force so that the counter weight231is prevented from moving in its circumferential direction. Therefore, in spite of the above mentioned rotational force, stable reciprocating movement of the counter weight231can be ensured. In addition, unintentional torsion can be prevented from acting on the second connecting shaft227, the second connecting rod225and the second eccentric shaft223so that the counter weight231can move with stability.

In this embodiment, as shown inFIGS. 4 and 5, the first crank disc124of the first motion converting mechanism113is rotatably supported by a first bearing120. The second crank disc221of the second motion converting mechanism213is rotatably supported by a second bearing229. Further, the first crank disc124is connected to the second crank disc221via the first eccentric shaft125. With this construction, the first crank disc124, the first eccentric shaft125and the second crank disc221are supported as one integral rigid body by the first and the second bearings120,229. As a result, such rotation driving mechanism can be driven with stability.

Further, in this embodiment, the axial length (length in the moving direction) of the counter weight231is designed to be larger than the outer diameter of the cylinder129. As a result, the counter weight231is prevented from tilting with respect to the axis of the cylinder129due to the existence of a clearance between the cylinder and the counter weight. As a result, the stability of the reciprocating movement of the counter weight231along the cylinder129is improved.

Although, in the second embodiment, the driving force of the counter weight231is inputted from one side (upper side as viewed inFIGS. 4 and 5) of the axis of movement of the counter weight231, it may be inputted from the both sides. For this purpose, a motion converting mechanism (crank mechanism) similar to the second motion converting mechanism213may be provided symmetrically on the opposite side of the first motion converting mechanism113with respect to the second motion converting mechanism213. Specifically, inFIG. 4, a crank disk may be provided on the opposite side (lower side as viewed inFIG. 4) of the bearing123athat supports the shaft of the driven gear123, with respect to the driven gear123. In such case, one end of a connecting rod may be rotatably connected to the crank disc via an eccentric shaft, while the other end may be rotatably connected to the counter weight231via a connecting shaft. With such modification, the driving force of the counter weight231can be inputted parallel to each other from the both sides of the axis of movement of the counter weight231. Thus, the counter weight231can slide with stability. Further, the rotation preventing mechanism can be omitted.

DESCRIPTION OF NUMERALS