Abstract:
It is an object of the invention to provide a technique for a reduction of an impact force cased by rebound of a tool bit after its striking movement in an impact power tool. The representative impact power tool includes a tool body, a hammer actuating member, a striker, a weight and an elastic element. A reaction force is transmitted from the hammer actuating member to the weight and the elastic element is elastically deformed when the weight moves ward by the reaction to absorb the reaction force. The invention is characterized in that the mass of the weight is set to about 40% or more of the mass of the striker.

Description:
BACKGROUND OF THE INVENTION 
     1. Field of the Invention 
     The present invention relates to an impact power tool for performing a linear hammering operation on a workpiece, and more particularly to a technique for cushioning a reaction force received from the workpiece during hammering operation. 
     2. Description of the Related Art 
     Japanese non-examined laid-open Patent Publication No. 8-318342 discloses a technique for cushioning an impact force caused by rebound of a tool bit after its striking movement in a hammer drill. In this known hammer drill, a rubber ring (cushion member) is disposed between the axial end surface of a cylinder on the body side and an intermediate element in the form of an impact bolt which strikes the tool bit When the tool bit receives a reaction force from the workpiece and rebounds after striking movement of the tool bit, the impact bolt collides with the rubber ring. At this time, the rubber ring cushions the impact force by elastic deformation. Further, the rubber ring also functions as a member for positioning the hammer drill body with respect to the workpiece during hammering operation. During the striking movement of the tool bit, the tip end of the tool bit is held pressed against the workpiece (the tool bit is held in its striking position) by application of the user&#39;s forward pressing force to the hammer drill body. The cylinder on the body side receives the pressing force via the rubber ring. 
     As described above, the known rubber ring has a function of cushioning the impact force caused by rebound of the tool bit and a function of positioning the hammer drill. It is advantageous for the rubber ring to be soft in order to absorb the rebound of the tool bit. On the contrary, it is advantageous for the rubber ring to be hard in order to improve the positioning accuracy. In other words, two different properties are demanded of the known rubber ring. It is difficult to provide the rubber ring with a hardness that satisfies the both functional requirements. In this point, further improvement is required. 
     SUMMARY OF THE INVENTION 
     Accordingly, it is an object of the present invention to provide a technique that contributes to reduction of an impact force caused by rebound of a tool bit after its striking movement in an impact power tool. 
     In order to solve the above-described problem, the representative impact power tool according to the present invention includes a tool body, a hammer actuating member and a striker. The hammer actuating member is disposed in a tip end region of the tool body and performs a predetermined hammering operation on a workpiece by reciprocating in its axial direction. The striker performs a striking movement on the hammer actuating member by reciprocating in the longitudinal direction of the tool body. The “predetermined hammering operation” in this invention includes not only a hammering operation in which the hammer actuating member performs only a linear striking movement, but a hammer drill operation in which it performs a linear striking movement and a rotation in the circumferential direction. The “hammer actuating member” in this invention typically comprises a tool bit and an impact bolt that transmits a striking force in the state of contact with the tool bit. 
     The impact power tool of this invention further includes a weight and an elastic element. When the hammer actuating member performs a hammering operation on the workpiece, a reaction force is transmitted from the hammer actuating member to the weight in a reaction force transmitting position in which the weight is placed in direct contact with the hammer actuating member or in which the weight is placed in contact with the hammer actuating member via an intervening member made of hard metal. When the weight is caused to move rearward on the reaction force transmitting position by the reaction force transmitted to the weight and pushes the elastic element, the elastic element elastically deforms and thereby absorbs the reaction force. Further, in a preferred aspect of the present invention, the mass of the weight is set to about 40% or more of the mass of the striker. The “weight” in this invention typically comprises a cylindrical member, but it may comprise a plurality of elements separated from each other in the circumferential direction. Further, the “elastic element” typically comprises a spring, but it may comprise a rubber. 
     During hammering operation, the hammer actuating member is caused to rebound by receiving the reaction force of the workpiece after striking movement. According to this invention, with the construction in which the reaction force is transmitted from the hammer actuating member to the weight in the reaction force transmitting position in which the weight is placed in direct contact with the hammer actuating member or in which the weight is placed in contact with the hammer actuating member via an intervening member made of hard metal, the reaction force is nearly 100% transmitted. In other words, the reaction force is transmitted by exchange of momentum between the hammer actuating member and the weight. By this transmission of the reaction force, the weight is caused to move rearward in the direction of action of the reaction force. The rearward moving weight elastically deforms the elastic element, and the reaction force of the weight is absorbed by such elastic deformation. Specifically, according to this invention, the reaction force caused by rebound of the hammer actuating member can be absorbed by the rearward movement of the weight and by the elastic deformation of the elastic element which is caused by the movement of the weight. As a result, vibration of the impact power tool can be reduced. 
     The hammering operation using the impact power tool is performed under loaded conditions in which the tip end of the hammer actuating member is pressed against the workpiece by the user&#39;s pressing force applied forward to the tool body (i.e. in the state in which the impact power tool is positioned with respect to the workpiece). At this time, the hammer actuating member is held in a position to be driven by the driving mechanism, or in a striking position in which the striker strikes the hammer actuating member. The “reaction force transmitting position” in this invention refers to a position in which the reaction force received from the workpiece by the hammer actuating member is transmitted from the hammer actuating member to the weight when the hammer actuating member is driven by the driving mechanism, whether the hammer actuating member is in direct contact with the weight or in contact with the weight via an intervening member. Therefore, the reaction force transmitting position generally coincides with the above-described striking position. 
     According to the invention, the mass of the weight is set about 40% or more of the mass of the striker. As a result, the peak acceleration generated by the reaction force of rebound when the striking movement is performed can be advantageously reduced. 
     As one aspect of the invention, high vibration reducing function is performed when the mass of the weight is set in the range of the lower limit of about 40% of the mass of the striker to the upper limit of about 200% of the mass of the striker. Particularly, when the mass of the weight is about 80% of the mass of the striker, the vibration reducing effect can be further enhanced. Further, when the mass of the weight is about 200% of the mass of the striker, the vibration reducing effect can be practically maximized. Further, this vibration reducing effect can also be maintained with the weight having a further increased mass over 200%. However, the mass of the weight may preferably be set to about 200% or below of the mass of the striker due to the balance between the mass ratio of the weight and the entire mass of the hammer drill. 
     As described above, during hammering operation by the hammer actuating member, the weight is caused to move rearward by the reaction force caused by rebound of the hammer actuating member. At this time, the elastic element elastically deforms and absorbs the reaction force transmitted to the weight. The weight is then returned by the restoring force of the elastic element to the reaction force transmitting position in which the reaction force was transmitted from the hammer actuating member to the weight. However, when the striker performs the next striking movement the hammer actuating member in a midway region by the time the weight is returned to the reaction force transmitting position after the weight is caused to move rearward from the reaction force transmitting position by receiving the reaction force, the weight and the elastic element do not function properly. 
     Having regard to this problem, according to one aspect of the invention, a resonance frequency defined under the assumption that the weight and the elastic element are models of a spring mass system may be set over half of the period of striking which is performed on the hammer actuating member by the striker. With such a construction, the weight can be returned to the initial reaction force transmitting position by the time the striker performs the next striking after the weight is caused to move rearward by receiving the reaction force from the hammer actuating member. Therefore, the weight and the elastic element can reliably function for each stroke of the striker. Thus, the vibration reducing performance can be increased. 
     Further, as one aspect of the invention, the elastic element comprises a coil spring, and a spring constant of the coil spring is set to satisfy that k&gt;π 2 mfo 2 , wherein the spring constant is taken as k, the pi is π, the mass of the weight is m, and the frequency of striking which is performed on the hammer actuating member by the striker is fo. By setting the spring constant k of the coil spring to such a value that satisfies the above-mentioned equation, an impact absorbing mechanism can be provided in which the resonance frequency defined under the assumption that the weight and the elastic element are models of a spring mass system is set over half of the period of striking which is performed on the hammer actuating member by the striker. 
     Further, as one aspect of the invention, a viscoelastic member may be disposed between the weight and the elastic element and serves to absorb a stress wave of the weight when the reaction force of the hammer actuating member is transmitted to the weight The viscoelastic member may typically comprise a rubber. 
     During hammering operation, a reaction force caused by rebound of the hammer actuating member is transmitted to the weight and produces a stress wave in the weight. With such construction, the stress wave produced in the weight can be absorbed by deformation of the viscoelastic member. Therefore, when the elastic element comprises a spring, the spring can be prevented from surging which may be caused by transmission of the stress wave to the spring. Thus, the spring can be protected. 
     According to the invention, a technique is provided which contributes to reduction of an impact force caused by rebound of a tool bit after its striking movement in an impact power tool. 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. 
    
    
     
       BRIEF DESCRIPTION OF THE DRAWINGS 
         FIG. 1  is a sectional side view schematically showing an entire electric hammer drill according to an embodiment of this invention, under loaded conditions in which a hammer bit is pressed against a workpiece. 
         FIG. 2  is an enlarged sectional view showing an essential part of the hammer drill. 
         FIG. 3  is a sectional plan view showing the hammer drill under loaded conditions in which the hammer bit is pressed against the workpiece. 
         FIG. 4  is a sectional plan view showing the hammer drill during operation of a weight and a coil spring. 
         FIG. 5  is a graph showing the change of rebound acceleration (reaction force) with respect to the mass of the weight. 
         FIG. 6  shows the acceleration wave form in the absence of the weight and the coil spring. 
         FIG. 7  shows the acceleration wave form when the mass of the weight is 50 g (the mass ratio of the weight to the striker is 0.36). 
         FIG. 8  shows the acceleration wave form when the mass of the weight is 110 g (the mass ratio of the weight to the striker is 0.79). 
         FIG. 9  shows the acceleration wave form when the mass of the weight is 280 g (the mass ratio of the weight to the striker is 2.0). 
     
    
    
     DETAILED DESCRIPTION OF THE INVENTION 
     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 and manufacture improved impact power tools and method for using such impact 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 drawing. 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. 
     An embodiment of the present invention is now described with reference to  FIGS. 1 to 9 .  FIG. 1  is a sectional side view showing an entire electric hammer drill  101  as a representative embodiment of the impact power tool according to the present invention, under loaded conditions in which a hammer bit is pressed against a workpiece. As shown in  FIG. 1 , the hammer drill  101  of this embodiment includes a body  103 , a hammer bit  119  detachably coupled to the tip end region (on the left side as viewed in  FIG. 1 ) of the body  103  via a tool holder  137 , and a handgrip  109  that is connected to the rear end region (on the right side as viewed in  FIG. 1 ) of the body  103  and designed to be held by a user. The body  103  is a feature that corresponds to the “tool body” according to the present invention. The hammer bit  119  is held by the tool holder  137  such that it is allowed to reciprocate with respect to the tool holder  137  in its axial direction and prevented from rotating with respect to the tool holder  137  in its circumferential direction. In the present embodiment, for the sake of convenience of explanation, the side of the hammer bit  119  is taken as the front side and the side of the handgrip  109  as the rear side. 
     The body  103  includes a motor housing  105  that houses a driving motor  111 , and a gear housing  107  that houses a driving mechanism in the form of a motion converting mechanism  113 , a striking mechanism  115  and a power transmitting mechanism  117 . The motion converting mechanism  113  is adapted to appropriately convert the rotating output of the driving motor  111  to linear motion and then to transmit to the striking mechanism  115 . As a result, an impact force is generated in the axial direction of the hammer bit  119  via the striking mechanism  115 . Further, the speed of the rotating output of the driving motor  111  is appropriately reduced by the power transmitting mechanism  117  and then transmitted to the hammer bit  119 . As a result, the hammer bit  119  is caused to rotate in the circumferential direction. The handgrip  109  is generally U-shaped in side view, having a lower end and an upper end. The lower end of the handgrip  109  is rotatably connected to the rear end lower portion of the motor housing  105  via a pivot  109   a , and the upper end is connected to the rear end upper portion of the motor housing  105  via an elastic spring  109   b  for absorbing vibration. Thus, the transmission of vibration from the body  103  to the handgrip  109  is reduced. 
       FIG. 2  is an enlarged sectional view showing an essential part of the hammer drill  101 . The motion converting mechanism  113  includes a driving gear  121  that is rotated in a horizontal plane by the driving motor  111 , a driven gear  123  that engages with the diving gear  121 , a crank plate  125  that rotates together with the driven gear  123  in a horizontal plane, a crank arm  127  that is loosely connected at one end to the crank plate  125  via an eccentric shaft  126  in a position displaced a predetermined distance from the center of rotation of the crank plate  125 , and a driving element in the form of a piston  129  mounted to the other end of the crank arm  127  via a connecting shaft  128 . The crank plate  125 , the crank arm  127  and the piston  129  form a crank mechanism 
     The power transmitting mechanism  117  includes a driving gear  121  that is driven by the driving motor  111 , a transmission gear  131  that engages with the driving gear  121 , a transmission shaft  133  that is caused to rotate in a horizontal plane together with the transmission gear  131 , a small bevel gear  134  mounted onto the transmission shaft  133 , a large bevel gear  135  that engages with the small bevel gear  134 , and a tool holder  137  that is caused to rotate together with the large bevel gear  135  in a vertical plane. The hammer drill  101  can be switched between hammer mode and hammer drill mode. In the hammering mode, the hammer drill  101  performs a hammering operation on a workpiece by applying only a striking force to the hammer bit  119  in its axial direction. In the hammer drill mode, the hammer drill  101  performs a hammer drill operation on a workpiece by applying a striking force in the axial direction and a rotating force in the circumferential direction to the hammer bit  119 . This construction of the hammer drill  101  is not directly related to the present invention and therefore will not be described in further detail. The workpiece is not shown here in the drawings. 
     The striking mechanism  115  includes a striker  143  that is slidably disposed together with the piston  129  within the bore of the cylinder  141 . The striker  143  is driven via the action of an air spring of an air chamber  141   a  of the cylinder  141  which is caused by sliding movement of the piston  129 . The striker  143  then collides with (strikes) an intermediate element in the form of an impact bolt  145  that is slidably disposed within the tool holder  137  and transmits the striking force to the hammer bit  119  via the impact bolt  145 . The impact bolt  145  and the hammer bit  119  are features that correspond to the “hammer actuating member” according to this invention. The impact bolt  145  includes a large-diameter portion  145   a , a small-diameter portion  145   b  and a tapered portion  145   c . The large-diameter portion  145   a  is fitted in close contact with the inner surface of the tool holder  137 , while a predetermined extent of space is defined between a small-diameter portion  145   b  and the inner peripheral surface of the tool holder  137 . The tapered portion  145   c  is formed in the boundary region between the both diameter portions  145   a  and  145   b . The impact bolt  145  is disposed within the tool holder  137  in such an orientation that the large-diameter portion  145   a  is on the front side and the small diameter portion  145   b  is on the rear side. 
     The hammer drill  101  includes a positioning member  115  that positions the body  103  with respect to the workpiece by contact with the impact bolt  145  when the impact bolt  145  is pushed rearward (toward the piston  129 ) together with the hammer bit  119  under loaded conditions in which the hammer bit  119  is pressed against the workpiece by the user&#39;s pressing force applied forward to the body  103 . The positioning member  151  is a unit part including a rubber ring  153 , a front-side hard metal washer  155  joined to the axially front surface of the rubber ring  153 , and a rear-side hard metal washer  157  joined to the axially rear surface of the rubber ring  153 . The positioning member  151  is loosely fitted onto the small-diameter portion  145   b  of the impact bolt  145 . 
     When the impact bolt  145  is pushed rearward, the tapered portion  145   c  of the impact bolt  145  contacts the front metal washer  155  of the positioning member  151  and the rear metal washer  157  contacts the front end of the cylinder  141 . Thus, the rubber ring  153  of the positioning member  151  elastically connects the impact bolt  145  to the cylinder  141  that is fixedly mounted to the gear housing  107 . The front metal washer  155  has a tapered bore. When the impact bolt  145  is pushed rearward, the tapered surface of the front metal washer  155  closely contacts the tapered portion  145   c  of the impact bolt  145 . Further, the rear metal washer  157  has a generally hat-like sectional shape, having a cylindrical portion of a predetermined length which is fitted onto the small-diameter portion  145   b  of the impact bolt  145  and a flange that extends radially outward from the cylindrical portion. The rear surface of the flange is in contact with the axial front end of the cylinder  141  via a spacer  159 . 
     In order to absorb the impact force (reaction force) that is caused by rebound of the hammer bit  119  after the striking movement of the hammer bit  119  during hammering operation on the workpiece, the hammer drill  101  according to this embodiment includes a hard metal cylindrical weight  163  that contacts the impact bolt  145  via the front metal washer  155  and a coil spring  165  that normally biases the cylindrical weight  163  toward the impact bolt  145  (forward). The cylindrical weight  163  and the coil spring  165  form an impact absorbing mechanism which is also referred to as an impact damper. The cylindrical weight  163 , the coil spring  165  and the front metal washer  155  are features that correspond to the “weight”, the “elastic element” and the “intervening member”, respectively, according to this invention. Further, a rubber ring  164  is disposed between the cylindrical weight  163  and the coil spring  165  and serves to absorb a stress wave of the cylindrical weight  163 . The rubber ring  164  is a feature that corresponds to the “viscoelastic member” according to this invention. 
     The cylindrical weight  163  is disposed between the outer surface of the positioning member  151  and an inner surface of the tool holder  137  and can move in the axial direction of the hammer bit. The movement of the weight  163  is guided along the inner surface of the tool holder  137 . Specifically, the cylindrical weight  163  and the positioning member  151  are arranged in parallel in the radial direction and in the same position on the axis of the hammer bit  119 . The cylindrical weight  163  extends further rearward from the outer peripheral region of the positioning member  151  to the outer front region of the cylinder  141 . The rubber ring  164  is disposed on the rear end of the weight  163 , and the coil spring  165  is elastically disposed between the rubber ring  164  and the tool holder  137  under a predetermined initial load. Thus, the cylindrical weight  163  is biased forward and its front end is normally in contact with a control member in the form of a stepped position control stopper  169  formed in the tool holder  137 , so that the weight  163  is prevented from moving forward beyond its striking position. In other words, the biasing force (elastic force) of the coil spring  165  that biases the weight  163  forward is controlled to be prevented from substantially acting forward beyond the striking position of the weight  163 . The striking position here refers to a position in which the striker  143  collides with (strikes) the impact bolt  145 . This striking position coincides with a position in which the reaction force from the impact bolt  145  is transmitted to the weight  163 . This striking position is a feature that corresponds to the “reaction force transmitting position” according to this invention. 
     Under loaded conditions in which the impact bolt  145  is pushed rearward together with the hammer bit  119 , the axial front end of the cylindrical weight  163  is in surface contact with the radially outward portion of the rear surface of the front metal washer  155  of the positioning member  151 . Specifically, the cylindrical weight  163  is in contact with the impact bolt  145  via the front metal washer  155 . Therefore, when the hammer bit  119  and the impact bolt  145  are caused to rebound by receiving a reaction force from the workpiece after striking movement, the reaction force from the impact bolt  145  is transmitted to the cylindrical weight  163  which is in contact with the impact bolt  145  via the front metal washer  155 . The front metal washer  155  forms a reaction force transmitting member and has a larger diameter than the outside diameter of the rubber ring  153 . Thus, the axial front end of the cylindrical weight  163  is in contact with an outer region of the front metal washer  155  outward of the outer surface of the rubber ring  153 . The rubber ring  164  disposed between the cylindrical weight  163  and the coil spring  165  elastically deforms by a stress wave transmitted from the impact bolt  145  to the cylindrical weight  163 . Thus, the rubber ring  164  absorbs the stress wave and prevents transmission of the stress wave to the coil spring  165 . Specifically, the rubber ring  164  mainly serves as a member for absorbing a stress wave. When the cylindrical weight  163  is moved rearward by receiving a reaction form from the impact bolt  145 , the coil spring  165  is pushed via the rubber ring  164  by the cylindrical weight  163 . As a result, the coil spring  165  elastically deforms and absorbs the reaction force. One axial end of the coil spring  165  is held in contact with the axial rear end surface of the cylindrical weight  163  and the other axial end is in contact with a spring receiving ring  167  fixed to the tool holder  137 . 
     Operation of the hammer drill  101  constructed as described above will now be explained When the driving motor  111  (shown in  FIG. 1 ) is driven, the rotating output of the driving motor  111  causes the driving gear  121  to rotate in the horizontal plane. When the driving gear  121  rotate, the crank plate  125  revolves in the horizontal plane via the driven gear  123  that engages with the driving gear  121 . Then, the piston  129  slidingly reciprocates within the cylinder  141  via the crank arm  127 . The striker  143  reciprocates within the cylinder  141  and collides with (strikes) the impact bolt  145  by the action of the air spring function within the cylinder  141  as a result of the sliding movement of the piston  129 . The kinetic energy of the striker  143  which is caused by the collision with the impact bolt  145  is transmitted to the hammer bit  119 . Thus, the hammer bit  119  performs a striking movement in its axial direction, and the hammering operation is performed on a work-piece. 
     When the hammer drill  101  is driven in hammer drill mode, the driving gear  121  is caused to rotate by the rotating output of the driving motor  111 , and the transmission gear  131  that engages with the driving gear  121  is caused to rotate together with the transmission shaft  133  and the small bevel gear  134  in a horizontal plane. The large bevel gear  135  that engages with the small bevel gear  134  is then caused to rotate in a vertical plane, which in turn causes the tool holder  137  and the hammer bit  119  held by the tool holder  137  to rotate together with the large bevel gear  135 . Thus, in the hammer drill mode, the hammer bit  119  performs a striking movement in the axial direction and a rotary movement in the circumferential direction, so that the hammer drill operation is performed on the work-piece. 
     The above described operation is performed in the state in which the hammer bit  119  is pressed against the workpiece and in which the hammer bit  119  and the tool holder  137  are pushed rearward as shown in  FIGS. 1 to 3 . The impact bolt  145  is pushed rearward when the tool holder  137  is pushed rearward. The impact bolt  145  then contacts the front metal washer  155  of the positioning member  151  and the rear metal washer  157  contacts the front end of the cylinder  141 . Specifically, the cylinder  141  on the body  103  side receives the force of pushing in the hammer bit  119 , so that the body  103  is positioned with respect to the workpiece. In this state, a hammering operation or a hammer drill operation is performed. At this time, as described above, the front end surface of the cylindrical weight  163  is held in contact with the rear surface of the front metal washer  155  of the positioning member  151 . 
     After striking movement of the hammer bit  119  upon the workpiece, the hammer bit  119  is caused to rebound by the reaction force from the workpiece. This rebound causes the impact bolt  145  to be acted upon by a rearward reaction force. At this time, the cylindrical weight  163  is in contact with the impact bolt  145  via the front metal washer  155  of the positioning member  151 . Therefore, in this state of contact via the front metal washer  155 , the reaction force of the impact bolt  145  is transmitted to the cylindrical weight  163 . In other words, momentum is exchanged between the impact bolt  145  and the cylindrical weight  163 . By such transmission of the reaction force, the impact bolt  145  is held substantially at rest in the striking position, while the cylindrical weight  163  is caused to move rearward in the direction of action of the reaction force. As shown in  FIG. 4 , the rearward moving cylindrical weight  163  elastically deforms the coil spring  165 , and the reaction force of the weight  163  is absorbed by such elastic deformation. 
     At this time, the reaction force of the impact bolt  145  also acts upon the rubber ring  153  which is kept in contact with the impact bolt  145  via the front metal washer  155 . Generally, the transmission rate of a force of one object is raised according to the Young&#39;s modulus of the other object placed in contact with the one object. According to this embodiment the cylindrical weight  163  of the impact damper  161  is made of hard metal and has high Young&#39;s modulus, while the rubber ring  153  made of rubber has low Young&#39;s modulus. Therefore, most of the reaction force of the impact bolt  145  is transmitted to the cylindrical weight  163  which has high Young&#39;s modulus and which is placed in contact with the metal impact bolt  145  via the hard front metal washer  155 . Thus, the impact force caused by rebound of the hammer bit  119  and the impact bolt  145  can be efficiently absorbed by the rearward movement of the cylindrical weight  163  and by the elastic deformation of the coil spring  165  which is caused by the movement of the cylindrical weight  163 . As a result, vibration of the hammer drill  101  can be reduced. At this time, the rubber ring  164  disposed between the cylindrical weight  163  and the coil spring  165  elastically deforms and thereby absorbs a stress wave transmitted from the impact bolt  145  to the cylindrical weight  163 . Thus, the rubber ring  164  prevents transmission of the stress wave of the cylindrical weight  163  to the coil spring  165 . As a result, the rubber ring  164  can prevent the coil spring  165  from surging and can protect it. 
     Thus, according to this embodiment, most of the reaction force that the hammer bit  119  and the impact bolt  145  receive from the workpiece after the striking movement is transmitted from the impact bolt  145  to the cylindrical weight  163 . The impact bolt  145  is placed substantially at rest as viewed from the striking position. Therefore, only a small reaction force acts upon the rubber ring  153 . Accordingly, only a slight amount of elastic deformation is caused in the rubber ring  153  by such reaction force, and a subsequent repulsion is also reduced. Further, the reaction force of the impact bolt  145  can be absorbed by the impact damper  161  which includes the cylindrical weight  163  and the coil spring  165 . Therefore, the rubber ring  153  can be made hard. As a result, such rubber ring  153  can provide correct positioning of the body  103  with respect to the workpiece. 
     Further, in this embodiment, the stopper  169  controls the biasing force of the coil spring  165  such that the biasing force is prevented from substantially acting forward beyond the striking position. Therefore, during striking movement, when the user applies a pressing force forward to the body  103  to hold the hammer bit  119  and the impact bolt  145  in the striking position, even with a provision of the coil spring  165  for absorbing the reaction force, unnecessary force for holding the hammer bit  119  and the impact bolt  145  is not required. Unlike the construction, such as an idle driving prevention mechanism, in which a forward spring force normally acts upon the hammer bit  119  and the impact bolt  145  during striking movement, an efficient mechanism can be realized in which the adverse effect of the elastic force for absorbing a reaction force can be reduced. 
     Further, according to this embodiment, the forward position of the cylindrical weight  163  is mechanically controlled by the stopper  169 . Thus, in this state in which the biasing force of the coil spring  165  is applied to the cylindrical weight  163 , the cylindrical weight  163  is controlled to be prevented from moving beyond the striking position. Therefore, the condition settings for absorption of the reaction force, including the settings of the biasing force of the coil spring  165  or the weight of the cylindrical weight  163 , can be facilitated. 
     Further, according to this embodiment, the reaction force from the workpiece is transmitted to the cylindrical weight  163  via the hammer bit  119  and the impact bolt  145 . Thus, the reaction force from the workpiece can be transmitted in a concentrated manner to the cylindrical weight  163  without being scattered midway on the transmission path. As a result, the efficiency of transmission of the reaction force to the cylindrical weight  163  is increased, so that the impact absorbing function can be enhanced. 
     Further, in this embodiment, the cylindrical weight  163  and the positioning member  151  are arranged in parallel in the radial direction and in the same position on the axis of the hammer bit  119 . Thus, an effective configuration for space savings can be realized. Further, the impact bolt  145  contacts the cylindrical weight  163  and the rubber ring  153  via a common hard metal sheet or the front metal washer  155 . Therefore, the reaction force of the impact bolt  145  can be transmitted from one point to two members via a common member, that is, from the impact bolt  145  to the cylindrical weight  163  and the rubber ring  153  via the front metal washer  155 . Further, the structure can be simplified. 
     Inventor conducted an impact test on the hammer drill  101  having the cylindrical weight (hereinafter referred to simply as “weight”)  163  and the coil spring  165  in order to verify the relationship between the mass of the weight  163  and the vibration reducing effect, assuming that the mass of the weight  163  affects the reaction force absorbing effect or the vibration reducing effect. The impact test was conducted under the conditions in which the mass of the testing device is 5.85 kg, the pressing force of the testing device is 100N, the mass of the striker is 140 g, the speed of the striker is 9.65 m/s (average), the drill diameter is φ20, and the low-pass filter is 1 kHz. Further, a plurality of weights  163  varying in mass in the range of 20 to 560 g were used in the impact test. The impact test was conducted several times for each weight  163  having a different mass. 
       FIG. 5  shows the test results.  FIG. 5  shows the change of rebound acceleration (reaction force) with respect to the mass of the weight  163 . The abscissa indicates the mass ratio of the weight  163  to the striker  143 , and the ordinate indicates the rebound peak acceleration ratio which is taken as 100% in the absence of the weight  163  and the coil spring  165 . The test results showed that the peak acceleration by the reaction force of rebound during striking is reduced about 10% when the mass ratio of the weight  163  to the striker  143  is about 0.4. Further, the peak acceleration by the reaction force of rebound during striking is reduced about 50% when the mass ratio of the weight  163  to the striker  143  is about 0.8. Further, it was also shown that when the mass ratio of the weight  163  to the striker  143  is about 2.0, the peak acceleration by the reaction force of rebound during striking is reduced about 60% and a higher vibration reducing effect can be obtained. In this test, it was also shown that, when the mass ratio exceeds such a value that can obtain the higher vibration reducing effect, the peak acceleration does not substantially charge and the higher vibration reducing effect can be maintained. 
       FIGS. 6 to 9  show the specific test results for verifying the vibration reducing effect from the mass ratio of the weight  163  and the peak acceleration as described above.  FIGS. 6 to 9  show acceleration wave forms by mass ratio of the weight  163 . Specifically,  FIG. 6  shows the acceleration wave form in the absence of the weight  163  and the coil spring  165 .  FIG. 7  shows the acceleration wave form when the mass of the weight  163  is 50 g (the mass ratio of the weight  163  to the striker  143  is 0.36).  FIG. 8  shows the acceleration wave form when the mass of the weight  163  is 110 g (the mass ratio of the weight  163  to the striker  143  is 0.79).  FIG. 9  shows the acceleration wave form when the mass of the weight  163  is 280 g (the mass ratio of the weight  163  to the striker  143  is 2.0). 
     According to the test results, when the mass ratio of the weight  163  is 0 in the absence of the weight  163  and the coil spring  165 , as shown in  FIG. 6 , the acceleration is as high as about 240 m/s 2 . When the mass ratio is 0.36, as shown in  FIG. 7 , the acceleration is reduced to about 170 m/s 2 . Further, when the mass ratio is 0.79, as shown in  FIG. 8 , the acceleration is reduced to about 100 m/s 2 . Further, when the mass ratio is 2.0, as shown in  FIG. 9 , the acceleration is reduced to about 60 m/s 2 . 
     Having regard to the above-described, a high vibration reducing function can be performed when the mass of the weight  163  is set in the range of the lower limit of about 400% of the mass of the striker  143  to the upper limit of about 200% of the mass of the striker  143 . Particularly, when the mass of the weight  163  is about 80% of the mass of the striker  143 , the vibration reducing effect can be further enhanced. Further, when the mass of the weight  163  is about 200% of the mass of the striker  143 , the vibration reducing effect can be practically maximized. Further, this vibration reducing effect can also be maintained with the weight  163  having a further increased mass. However, it was also found to be practically preferable that the mass of the weight  163  is about 200% or below of the mass of the striker  143  due to the balance between the mass ratio of the weight and the entire mass of the hammer drill  101 . 
     In hammering operation by the hammer bit  119 , as described above, the weight  163  is caused to move rearward by the reaction force caused by rebound of the impact bolt  145 . At this time, the coil spring  165  elastically deforms and absorbs the reaction force. The weight  163  is then returned by the restoring force of the coil spring  165  to the reaction force transmitting position in which the reaction force was transmitted from the impact bolt  145  to the weight  163 . However, when the striker  143  performs the next striking movement on the impact bolt  145  in a midway region by the time the weight  163  is returned to the reaction force transmitting position after the weight  163  is caused to move rearward by receiving the reaction force, the weight  163  and the coil spring  165  do not function properly. 
     Therefore, in this embodiment, the resonance frequency defined under the assumption that the weight  163  and the coil spring  165  are models of the spring mass system is set over half of the frequency of striking which is performed on the impact bolt  145  by the striker  143 . In other words, the spring constant of the coil spring  165  is set such that the resonance period defined under the assumption that the weight  163  and the coil spring  165  are models of the spring mass system is set below half of the period of striking which is performed on the impact bolt  145  by the striker  143 . In this manner, the weight  163  and the coil spring  165  can function properly, Specifically, the weight  163  and the coil spring  165  can reliably absorb the impact for each stroke of the striker  143 . 
     The condition to be satisfied by the spring constant of the coil spring  165  in order for the weight  163  and the coil spring  165  to properly function for each stroke of the striker  143  is mathematically obtained as follows:
 
 fo =1/ To,   (1)
 
     wherein fo [Hz] and To [s] are the striking frequency and the striking period of the striker  143 , respectively. 
     Further, under the assumption that the weight  163  and the coil spring  165  are models of the spring mass system, the angular velocity ω during resonance of the spring-mass system models is obtained as follows:
 
ω=√( k/m )=2π/T [rad/ s],   (2)
 
     wherein the mass of the weight  163  is taken as m[kg], the spring constant of the coil spring  165  is k[N/m], and the resonance frequency of the spring-mass system models is T[s]. 
     Further, from the relationship between the resonance period of the spring-mass system models and the striking period of the sinker  143 ,
 
 T/ 2&lt; To    (3)
 
     Substituting T=2π√(m/k) from Equation (2) into Equation (3) yields:
 
π√( m/k )&lt; To    (4)
 
     Squaring Equation (4), wherein the striking period To, the spring constant k and the mass m are all positive numbers,
 
π 2   m/k&lt;To   2    k&gt;π   2   m/To   2 =π 2   mfo   2   (5)
 
     Therefore, the condition to be satisfied by the spring constant of the coil spring  165  is:
 
k&gt;π 2 mfo 2   (6)
 
     By setting the spring constant of the coil spring  165  to such a value that satisfies Equation (6), it can be constructed such that the weight  163  and the coil spring  165  function properly. 
     Further, in this embodiment, the viscoelastic member in the form of the rubber ring  164  is disposed between the cylindrical weight  163  and the coil spring  165  and serves to absorb a stress wave of the cylindrical weight  163 . The mass of the rubber ring  164  is extremely smaller than the mass of the cylindrical weight  163 . Further, although the rubber ring  164  deforms by the stress wave of the cylindrical weight  163 , the amount of such deformation is extremely smaller than the amount of deformation of the coil spring  165 . Therefore, in setting the above-described spring constant of the coil spring  165 , the rubber ring  164  can be considered as part of the weight  163  and practically has little adverse effect. 
     Further, in the hammer drill  101  according to this embodiment a dynamic vibration reducer, which is not shown, may be mounted in the body  103  and can be used together with the impact absorbing mechanism having the weight  163  and the coil spring  165 . In this case, a passive vibration reducing function can be performed on periodic vibration which is caused in the body  103  in the longitudinal direction of the body  103  during hammering operation. Thus, the vibration of the body  103  can be effectively reduced. Further, the pressure within the crank chamber that houses the crank mechanism fluctuates when the hammer drill  101  is driven. Therefore, it can be constructed such that the fluctuating pressure is introduced into the dynamic vibration reducer and a weight forming a component part of the dynamic vibration reducer is actively driven. In other words, a forced vibration method can be employed. In this case, the dynamic vibration reducer functions as an effective vibration reducing mechanism by forced vibration of the weight. Thus, the vibration caused in the body  103  during hammering operation can be further effectively reduced. 
     In the above-described embodiment, the hammer drill  101  was described as a representative example of the impact power tool. However, the present invention can also be applied to a hammer. Further, in the above embodiment, the reaction force was described as being transmitted via a path from the impact bolt  145  to the cylindrical weight  163 , it may be configured such that the reaction force is transmitted via a path from the hammer bit  119  to the cylindrical weight  163 . Further, the cylindrical weight  163  may have a shape other than a cylindrical shape. 
     Further, in the above embodiment, the crank mechanism was described as being used as the motion converting mechanism  113  for converting the rotating output of the driving motor  111  to linear motion in order to linearly drive the hammer bit  119 . However, the motion converting mechanism is not limited to the crank mechanism, but, for example, a swash plate that axially swings may be utilized as the motion converting mechanism. Further, in the above embodiment, the stopper  169  serves to prevent forward movement of the cylindrical weight  163  so that the biasing force of the coil spring  165  is controlled to be prevented from substantially acting forward beyond the striking position. However, instead of provision of control by the stopper  169 , it may be changed in construction such that, for example, the coil spring  165  is disposed in a free state in which an initial load is not applied. Further, from the viewpoint of cushioning the reaction force received from the workpiece during hammering operation, the rubber ring  164  may be disposed between the coil spring  165  and the spring receiving ring  167 . 
     DESCRIPTION OF NUMERALS 
     
         
           101  hammer drill (impact power tool) 
           103  body (tool body) 
           105  motor housing 
           107  gear housing 
           109  handgrip 
           109   a  pivot 
           109   b  elastic spring 
           111  driving motor 
           113  motion converting mechanism (driving mechanism) 
           115  striking mechanism 
           117  power transmitting mechanics 
           119  hammer bit (hammer actuating member) 
           119   a  head edge portion 
           121  driving gear 
           123  driven gear 
           125  crank plate 
           126  eccentric shaft 
           127  crank arm 
           128  connecting shaft 
           129  piston 
           131  transmission gear 
           133  transmission shaft 
           134  small bevel gear 
           135  large bevel gear 
           137  tool holder 
           141  cylinder 
           141   a  air chamber 
           143  striker 
           145  impact bolt (hammer actuating member) 
           145   a  large-diameter portion 
           145   b  small-diameter portion 
           145   c  tapered portion 
           151  positioning member 
           153  rubber ring 
           155  front metal washer (intervening member) 
           157  metal washer 
           159  spacer 
           163  cylindrical weight (weight) 
           164  rubber ring (viscoelastic member) 
           165  coil spring (elastic element) 
           167  spring receiving ring 
           169  stopper