Patent Application
20250114923
This application claims priority to Japanese patent application no. 2023-172662 filed on Oct. 4, 2023, and to Japanese patent application no. 2024-104127 filed on Jun. 27, 2024, the contents of both of which are fully incorporated herein by reference.
The techniques disclosed in the present specification relate to an impact tool.
An impact rotary tool related to the present teachings is disclosed in U.S. Pat. No. 7,971,654.
To improve work efficiency when using an impact tool, there is demand for techniques that can curtail the occurrence of the cam-out phenomenon in which a tool accessory (undesirably) slips out of a cruciform (cross-shaped) groove, which is formed in a head portion of a screw (e.g., a Phillips-head screw), during a fastening or tightening operation.
It is one non-limiting object of the present teachings to disclose techniques that can curtail the occurrence of the cam-out phenomenon.
In one non-limiting aspect of the present teachings, the impact tool comprises: a motor supplied with a voltage of 18 V or more; a spindle rotated by the motor; a spindle groove formed in the spindle; a ball held in the spindle groove; a hammer supported on the spindle via the ball; at least one spring, which biases the hammer forward; and an anvil, which is impacted in a rotational direction by the hammer. An impact-start torque, which is the torque acting on the anvil when the hammer starts to impact the anvil, is 1,100 N·mm or less.
By utilizing techniques disclosed in the present specification, the occurrence of the cam-out phenomenon can be curtailed.
As was mentioned above, an impact tool according to one aspect of the present teachings preferably comprises: a motor supplied with a voltage of 18 V or more; a spindle rotated by the motor; a spindle groove formed in the spindle; a ball held in the spindle groove; a hammer supported on the spindle via the ball; a (at least one) spring, which biases the hammer forward; and an anvil, which is impacted in a rotational direction by the hammer. An impact-start torque, which is the torque acting on the anvil when the hammer starts to impact the anvil, is 1,100 N·mm or less.
During screw-tightening work, when a load torque from a screw greater than or equal to a prescribed value acts on the anvil, the rotation of the anvil and the hammer momentarily stops. Because the spindle continues to rotate owing to the motor, the spindle continues to rotate in the state in which the rotation of the hammer is momentarily stopped. As a result, the hammer moves (is caused to move) rearward relative to the spindle. After the hammer has moved rearward, the hammer moves forward while rotating owing to the biasing force of the spring. The anvil is impacted by the hammer in the rotational direction owing to the hammer moving forward while rotating. In the above-mentioned configuration, because the impact-start torque is relatively small, the timing of the start of impact of the hammer on the anvil is earlier. That is, an interval from a screw-tightening start time until an impact time, which is when the anvil is impacted by the hammer, becomes shorter.
During the screw-tightening work, the anvil is rotated while a tool accessory (e.g., a screwdriving bit) mounted in the anvil is inserted in a cruciform groove formed in a head portion of the screw (e.g., a Phillips-head screw). In an embodiment in which the timing of the start of impact of the hammer on the anvil is later (that is, if the interval from the screw-tightening start time until the impact time when the anvil is impacted by the hammer is relatively long), then the probability of the occurrence of the cam-out phenomenon, in which the tool accessory (undesirably) comes completely out of the cruciform groove, is relatively high. However, in the above-mentioned configuration, because the timing of the start of impact of the hammer on the anvil is earlier, the occurrence of the cam-out phenomenon can be reduced.
In one or more embodiments, the impact tool comprises a battery-mounting part, onto which a battery pack is mounted. The rated voltage of the battery pack is 18 V or more.
In the above-mentioned configuration, because the rated voltage of the battery pack is 18 V or more, a high output (torque) can be achieved for the impact tool.
In one or more embodiments, the depth of the spindle groove differs according to the location thereof in an axial direction of the spindle.
In the above-mentioned configuration, the occurrence of the cam-out phenomenon can be reduced.
In one or more embodiments, the depth of a rear portion of the spindle groove is deeper than the depth of a front portion of the spindle groove. For example, the depth of an end portion of the spindle groove is deeper than the depth of a center portion of the spindle groove.
In the above-mentioned configuration, because the center of the ball can sink (immerse deeper) into the inside of the spindle groove, among forces acting between the spindle and the ball, the force component in the rotational direction becomes larger. Consequently, the rotational speed of the hammer—when the hammer is moving forward while rotating—can be increased. Because the rotational speed of the hammer can be increased, the hammer can impact the anvil with a larger impact force.
In one or more embodiments, an end portion of the spindle groove is disposed more rearward than a center portion of the spindle groove. When the hammer moves rearward relative to the spindle, the ball moves to the (rearward) end portion of the spindle groove. At the (rearward) end portion of the spindle groove, the center of the ball is disposed on the inside of the spindle groove; i.e. the center of the ball is closer to the rotational axis of the spindle than the outer circumference (surface) of the spindle.
Thus, when the hammer moves rearward relative to the spindle, the ball moves from the center portion to the (rearward) end portion of the spindle groove. Thereafter, the hammer moves (is moved) forward relative to the spindle owing to the biasing force of the spring, whereby the ball moves from the (rearward) end portion back to the center portion of the spindle groove. In the above-mentioned configuration, because the center of the ball sinks into the inside of the spindle groove at the end portion of the spindle groove (i.e. as was noted above, the center of the ball is closer to the center of the spindle than the outer circumferential surface of the spindle), the rotational speed of the hammer—when the hammer is moving forward while rotating—can be increased. Because the rotational speed of the hammer can be increased, the hammer can impact the anvil with a larger impact force.
In one or more embodiments, the center of the ball is disposed outside of the spindle groove at the center portion of the spindle groove (i.e. the center of the ball is farther from the center of the spindle than the outer circumference (surface) of the spindle).
As was noted above, when the hammer impacts the anvil, the ball moves from the (rearward) end portion back to the center portion of the spindle groove. In the above-mentioned configuration, because the center of the ball that is at the center portion of the spindle groove is disposed outside of the spindle groove, the rotational speed of the hammer—at the moment the hammer impacts the anvil—can be increased. Because the rotational speed of the hammer can be increased, the hammer can impact the anvil with a larger impact force.
In one or more embodiments, the ball moves with circular motion so that the center of the ball moves along a virtual circle defined in the virtual plane. The center of the virtual circle is defined at a location that is offset from the rotational axis of the spindle.
In the above-mentioned configuration, the hammer can impact the anvil with a larger impact force.
In one or more embodiments, the impact tool may comprise: a motor supplied with a voltage of 18 V or more; a spindle rotated by the motor; a spindle groove formed in the spindle; a ball held in the spindle groove; a hammer supported on the spindle via the ball; a (at least one) spring, which biases the hammer forward; and an anvil, which is impacted in a rotational direction by the hammer. The maximum fastening torque of the impact tool may be higher than 220 N·m. The impact-start torque, which is the torque acting on the anvil when the hammer starts to impact the anvil, may be 1,300 N·mm or less. It is noted that, the maximum fastening torque is the torque when tightening an object (e.g., a screw) to be fastened, and generally refers to the torque measured by a supplemental fastening torque wrench, or the like, with respect to the object to be fastened after fastening. It is noted that, this is not a method in which a nut or a bolt is loosened and then measurement is performed. This maximum fastening torque is generally stated in each manufacturer's catalog.
In the above-mentioned configuration, the impact-start torque can be suppressed to 1,300 N·mm or less even though the maximum fastening torque is raised (increased) to higher than 220 N·m. Because the impact-start torque is relatively low, the timing of the start of impact of the hammer on the anvil is earlier. That is, the amount of screw rotation from the screw-tightening start time until the impact time when the anvil is impacted by the hammer becomes smaller. Because the timing of the start of impact of the hammer on the anvil is earlier, the occurrence of the cam-out phenomenon can be reduced.
In one or more embodiments, the impact tool may comprise a battery-mounting part, onto which a battery pack is mounted. The rated voltage of the battery pack may be 18 V or more.
Therefore, because the rated voltage of the battery pack is 18 V or more, a high output (torque) can be achieved for the impact tool.
In one or more embodiments, the depth the spindle groove may differ according to the location thereof in the axial direction of the spindle.
Therefore, the occurrence of the cam-out phenomenon can be reduced.
In one or more embodiments, the maximum fastening torque of the impact tool may be 230 N·m or more.
In the above-mentioned configuration, by reducing the impact-start torque to 1,300 N·mm or less even though the maximum fastening torque is made to be 230 N·m or more, both a higher maximum fastening torque and curtailment of the cam-out phenomenon can be achieved.
In one or more embodiments, the value calculated by dividing the maximum fastening torque (N·mm) of the impact tool by the impact-start torque (N·mm) may be greater than 165.
Generally speaking, the higher the maximum fastening torque is raised (becomes), the higher the impact-start torque tends to become. Furthermore, the higher the impact-start torque becomes and the later the timing of the start of impact becomes, the more easily (often) the cam-out phenomenon tends to occur. Because the value calculated by dividing the maximum fastening torque by the impact-start torque becomes larger the higher the maximum fastening torque and becomes larger the lower the impact-start torque, this value serves as an index that signifies (indicates) the lowness or smallness of the impact-start torque relative to the magnitude of the maximum fastening torque (i.e., the degree of curtailment of the cam-out phenomenon). In the above-mentioned configuration, because the value calculated by dividing the maximum fastening torque by the impact-start torque is greater than 165, it becomes possible to more effectively reduce the likelihood of the cam-out phenomenon occur than by simply changing (reducing) the magnitude of the maximum fastening torque.
In one or more embodiments, an impact tool may comprise: a motor; a spindle rotated by the motor; a cam element that includes a spindle groove, which is formed in the spindle, and a ball, which is held in the spindle groove; a hammer supported on the spindle via the ball; a (at least one) spring element which biases the hammer forward; and an anvil which is impacted in a rotational direction by the hammer. At the impact time, the hammer may resist the biasing force of the spring element owing to the rotational torque from the motor imparted thereto via the spindle and the cam element and, after having moved rearward while rotating, may advance while rotating owing to the elastic force of the spring element, which was compressed, and the cam element. The cam element and the spring element may each have a structure that nonlinearly increases a required torque, which is the torque needed to rotate the hammer, to suppress an increase in the required torque in a first region (hammer movement range) in which a rearward stroke of the hammer is shorter, and to boost the increase in the required torque in a second region (hammer movement range) in which the stroke of the hammer is longer.
In the above-mentioned configuration, the impact-start torque is reduced while raising (increasing) the maximum fastening torque, and thereby a high maximum fastening torque and curtailment of cam-out can both be achieved. That is, the output of the motor is raised to raise the rotational speed of the hammer in order to raise the maximum fastening torque. When the rotational speed of the hammer is higher, the time until the hammer arrives at an angle at which it impacts the anvil becomes shorter. The biasing force of the spring element is made to be larger and the hammer is caused to advance quickly in order to ensure that the advance of the hammer coincides with the timing at which the hammer impacts the anvil. However, when the biasing force of the spring element becomes larger, the impact-start torque also tends to become higher because the required torque, which is the torque needed to rotate and to move the hammer rearward, increases.
Thus, normally (i.e. in the past), when the maximum fastening torque is (was) raised (increased), the impact-start torque becomes (became) higher. Accordingly, as mentioned above, the cam element and the spring element of the present teachings can cause the change in the required torque, which is stroke dependent, to be nonlinear; therefore, by suppressing the increase in the required torque when the stroke is shorter, the rearward movement of the hammer can be made easier and the impact-start torque can be reduced. Furthermore, by boosting the increase in the required torque when the stroke is longer, the rotational speed of the hammer increases when the hammer is advancing, and thereby the maximum fastening torque can be raised (increased). As a result, the cam-out phenomenon can be curtailed while raising (increasing) the maximum fastening torque.
In one or more embodiments, when the first region is defined as extending from the location at which the rearward stroke of the hammer becomes its minimum up to the location at which the hammer and the anvil are no longer in contact and the second region is defined as a prescribed region that includes a location at which the stroke is longer than the stroke in the first region and the stroke becomes the maximum, a first degree of change in the required torque with respect to the stroke in the first region may be smaller than a second degree of change in the required torque with respect to the stroke in the second region.
In the above-mentioned configuration, because the degree of increase in the required torque, which is stroke dependent, is suppressed (reduced) in the first region, the hammer can be moved with a smaller required torque to a location at which the hammer and the anvil are no longer in contact. Accordingly, the impact-start torque can be lowered by reducing the required torque in the first region. However, if the degree of increase of the required torque, which is stroke dependent, were to (hypothetically) remain suppressed in the second region, then even when the hammer reaches its maximum stroke, the required torque would not rise sufficiently and the rotational speed of the hammer could not be raised. Consequently, a higher maximum fastening torque can be achieved by making the degree of increase in the required torque, which is stroke dependent, larger in the second region.
In one or more embodiments, the maximum value of the required torque in the first region may be 1,300 N·mm or less.
In the above-mentioned configuration, the impact-start torque can be effectively lowered.
In one or more embodiments, the maximum fastening torque of the impact tool may be 230 Nom or more and 350 N·m or less.
In the above-mentioned configuration, a higher maximum fastening torque can be achieved while suppressing (reducing) the impact-start torque.
In one or more embodiments, the depth of the rear portion of the spindle groove may be deeper than the depth of the front portion of the spindle groove.
In the above-mentioned configuration, the ball disposed in the spindle groove is most distant from the rotational axis at the front (forwardmost) portion of the spindle groove and closest to the rotational axis at the rear (rearmost) portion of the spindle groove. Consequently, at the front portion of the spindle groove, because the hammer can be caused to rotate with an even lower torque and to stroke (move) rearward, this contributes to reducing the impact-start torque. At the rear portion of the spindle groove, among the forces acting between the spindle and the ball, the force component in the rotational direction becomes relatively larger, and thus the rotational speed of the hammer when the hammer is moving forward while rotating can be increased. Because the rotational speed of the hammer can be increased, this contributes to increasing the maximum fastening torque.
In one or more embodiments, the hammer may have a (at least one) hammer groove in which at least a portion of the (at least one) ball is disposed. When a tangent vector is defined as having a direction parallel to a tangent of a circle about a rotational axis of the spindle and through a contact point between the ball and the hammer groove, and a load vector is defined as having a direction in which a load is transmitted from the ball to the contact point with the hammer, an angle formed between the tangent vector and the load vector at the rear portion of the spindle groove may be smaller than an angle formed between the tangent vector and the load vector at the front portion of the spindle groove.
The smaller that the angle formed between the tangent vector and the load vector becomes, the more that the force from the spindle acting on the ball acts exclusively in the rotational direction. In the above-mentioned configuration, the angle becomes relatively small at the front portion of the spindle groove, and among the forces from the spindle acting on the ball, the force component in the axial direction becomes relatively larger. Therefore, because the hammer can be caused to stroke (move) rearward with an even lower torque, this contributes to reducing the impact-start torque. At the rear portion of the spindle groove, among the forces from the spindle acting on the ball, the force component in the rotational direction becomes relatively larger. Therefore, because the rotational speed of the hammer—when the hammer moves forward while rotating—can be increased, this contributes to increasing the maximum fastening torque.
In one or more embodiments, the (at least one) spring element may include a first coil spring and a second coil spring. The first coil spring may continuously apply its biasing force to the hammer that urges the hammer to move forward. On the other hand, the second coil spring may apply its biasing force to the hammer only after the hammer has moved (or at least started to move) rearward.
In the above-mentioned configuration, only the first coil spring biases the hammer when the stroke of the hammer is shorter; therefore, the rearward movement of the hammer can be made easier and the impact-start torque can be reduced in the above-mentioned first region (range) of the hammer stroke. On the other hand, both the first coil spring and the second coil spring bias the hammer when the stroke of the hammer is longer (i.e. in the above-mentioned second region (range) of the hammer stroke), which makes it easy for the hammer to advance; thereby, the hammer can advance quickly and can impact the anvil appropriately even if the rotational speed of the hammer is raised (increased). Therefore, the maximum fastening torque can be raised (increased).
Embodiments are explained below, with reference to the drawings. In the embodiments, positional relationships among the parts are explained using the terms left, right, front, rear, up, and down. These terms indicate relative position or direction, wherein the center of an impact tool 1 is the reference. The impact tool 1 comprises a motor 6, which serves as the motive power source.
In the embodiments, the direction parallel to rotational axis AX of the motor 6 is called the axial direction where appropriate, the direction that goes around rotational axis AX is called the circumferential direction or the rotational direction where appropriate, and the radial direction of rotational axis AX is called the radial direction where appropriate.
Rotational axis AX extends in a front-rear direction. One side in the axial direction is forward, and the other side in the axial direction is rearward. In addition, in the radial direction, a location that is proximate to or a direction that approaches rotational axis AX is called radially inward where appropriate, and a location that is distant from or a direction that leads away from rotational axis AX is called radially outward where appropriate.
In the embodiments described below, the impact tool 1 is an impact driver, which is one type of screw-tightening tool according to the present teachings. The impact tool 1 comprises a housing 2, a rear cover 3, a hammer case 4, a hammer-case cover 5, the motor 6, a speed-reducing mechanism 7, a spindle 8, an impact mechanism 9, an anvil 10, a tool-holding mechanism 11, a fan 12, a battery-mounting part 13, a trigger lever 14, a forward/reverse-change lever 15, an operation-and-display part 16, a mode-change switch (quick mode-switching button) 17, and a light assembly 18.
The housing 2 is made of a synthetic resin (polymer). In the embodiment, the housing 2 is made of nylon (polyamide). The housing 2 comprises a left housing 2L and a right housing 2R, which is disposed rightward of the left housing 2L. The left housing 2L and the right housing 2R are fixed to each other by a plurality of screws 2S. Thus, the housing 2 is constituted from a pair of half housings.
The housing 2 comprises a motor-housing part 21, a grip part 22, and a battery-holding part 23.
The motor-housing part 21 has a tube shape. The motor-housing part 21 houses the motor 6. The motor-housing part 21 houses at least a portion of the hammer case 4.
The grip part 22 extends downward from the motor-housing part 21. The trigger lever 14 is provided at an upper portion of the grip part 22. The grip part 22 is gripped by the user.
The battery-holding part 23 is connected to a lower-end portion of the grip part 22. In both the front-rear direction and the left-right direction, the dimensions of the outer shape of the battery-holding part 23 are larger than the dimensions of the outer shape of the grip part 22.
The rear cover 3 is made of a synthetic resin (polymer), such as nylon. The rear cover 3 is disposed rearward of the motor-housing part 21. The rear cover 3 houses at least a portion of the fan 12. The fan 12 is disposed on the inner-circumference side of the rear cover 3. The rear cover 3 is disposed so as to cover an opening of a rear-end portion of the motor-housing part 21. The rear cover 3 is fixed to the rear-end portion of the motor-housing part 21 by two screws 3S.
The motor-housing part 21 has air-intake ports 19. The rear cover 3 has air-exhaust ports 20. Air outside of the housing 2 flows into the interior space of the housing 2 via the air-intake ports 19. Air in the interior space of the housing 2 flows out to the exterior of the housing 2 via the air-exhaust ports 20.
The hammer case 4 is made of a metal. In the embodiments herein, the hammer case 4 is made of aluminum. The hammer case 4 has a tube shape. The hammer case 4 is connected to a front portion of the motor-housing part 21. A bearing box 24 is fixed to a rear portion of the hammer case 4. A screw thread is formed on an outer circumferential portion of the bearing box 24. A thread groove is formed in an inner circumferential portion of the hammer case 4. The bearing box 24 and the hammer case 4 are fixed to each other by the coupling of the screw thread of the bearing box 24 and the thread groove of the hammer case 4 to each other. The hammer case 4 is sandwiched between the left housing 2L and the right housing 2R. At least a portion of the hammer case 4 is housed in the motor-housing part 21. The bearing box 24 is fixed to both the motor-housing part 21 and the hammer case 4.
The hammer case 4 houses the speed-reducing mechanism 7, the spindle 8, the impact mechanism 9, and at least a portion of the anvil 10. At least a portion of the speed-reducing mechanism 7 is disposed in the interior of the bearing box 24. The speed-reducing mechanism 7 comprises a plurality of gears.
The hammer case 4 has a first tube portion 401 and a second tube portion 402. The first tube portion 401 is disposed around the impact mechanism 9. The second tube portion 402 is disposed more forward than the first tube portion 401. The outer diameter of the second tube portion 402 is smaller than the outer diameter of the first tube portion 401.
The hammer-case cover 5 covers at least a portion of the surface of the hammer case 4. The hammer-case cover 5 protects the hammer case 4. The hammer-case cover 5 blocks contact between the hammer case 4 and objects around the hammer case 4.
The motor 6 is the motive power source of the impact tool 1. The motor 6 is an inner-rotor-type brushless motor. The motor 6 comprises a stator 26 and a rotor 27. The stator 26 is supported by the motor-housing part 21. At least a portion of the rotor 27 is disposed in the interior of the stator 26. The rotor 27 rotates relative to the stator 26. The rotor 27 rotates about rotational axis AX, which extends in the front-rear direction.
The stator 26 comprises a stator core 28, a front insulator 29, a rear insulator 30, and coils 31.
The stator core 28 is disposed more radially outward than the rotor 27. The stator core 28 comprises (is composed of) a plurality of laminated steel sheets. Each of the steel sheets is a sheet made of a metal in which iron is the main component. The stator core 28 has a tube shape. The stator core 28 comprises teeth that respectively support the coils 31.
The front insulator 29 is provided at a front portion of the stator core 28. The rear insulator 30 is provided at a rear portion of the stator core 28. The front insulator 29 and the rear insulator 30 are each an electrically insulating member that is made of a synthetic resin (polymer). The front insulator 29 is disposed so as to cover a forward portion of the surface of each of the teeth. The rear insulator 30 is disposed so as to cover a rearward portion of the surface of each of the teeth.
The coils 31 are mounted (wound) on the stator core 28 via (over) the front insulator 29 and the rear insulator 30. A plurality of the coils 31 is provided. The coils 31 are respectively disposed around the teeth of the stator core 28 via (over) the front insulator 29 and the rear insulator 30. The coils 31 and the stator core 28 are electrically insulated from each other by the front insulator 29 and the rear insulator 30. The coils 31 are electrically connected to a source of power (e.g., the battery pack) via fusing terminals 38.
The rotor 27 rotates about rotational axis AX. The rotor 27 has a rotor-core portion 32, a rotor-shaft portion 33, at least one rotor magnet 34, and at least one sensor magnet 35.
The rotor-core portion 32 and the rotor-shaft portion 33 are each made of steel. The rotor-shaft portion 33 protrudes in the front-rear direction from end surfaces of the rotor-core portion 32. The rotor-shaft portion 33 includes a front-side shaft portion 33F, which protrudes forward from a front-end surface of the rotor-core portion 32, and a rear-side shaft portion 33R, which protrudes rearward from a rear-end surface of the rotor-core portion 32.
The rotor magnet 34 is fixed to the rotor-core portion 32. The rotor magnet 34 has a circular-tube shape. The rotor magnet 34 is disposed around the rotor-core portion 32.
The sensor magnet 35 is fixed to the rotor-core portion 32. The sensor magnet 35 has a circular-ring shape. The sensor magnet 35 is disposed at the front-end surface of the rotor-core portion 32 and a front-end surface of the rotor magnet 34.
A sensor board 37 is mounted on the front insulator 29. The sensor board 37 is fixed to the front insulator 29 by a screw 29S. The sensor board 37 comprises a disk-shaped circuit board, in which a through hole is provided at the center, and at least one rotation-detection device, which is supported on the circuit board. At least a portion of the sensor board 37 opposes the sensor magnet 35. The rotation-detection device detects the location of the rotor 27 in the rotational direction by detecting the location of the sensor magnet 35 of the rotor 27.
The rotor-shaft portion 33 is supported by rotor bearings 39 in a rotatable manner. The rotor bearings 39 include a front-side rotor bearing 39F, which supports the front-side shaft portion 33F in a rotatable manner, and a rear-side rotor bearing 39R, which supports the rear-side shaft portion 33R in a rotatable manner.
The front-side rotor bearing 39F is held by the bearing box 24. The bearing box 24 has a recessed portion 24A, which is recessed forward from a rear surface of the bearing box 24. The front-side rotor bearing 39F is disposed in the recessed portion 24A. The rear-side rotor bearing 39R is held by the rear cover 3. A front-end portion of the front-side rotor-shaft portion 33F is disposed in the interior space of the hammer case 4 via an opening in the bearing box 24.
A pinion gear 41 is formed at a front-end portion of the rotor-shaft portion 33. The pinion gear 41 is coupled to an input portion of the speed-reducing mechanism 7. More specifically, the rotor-shaft portion 33 is coupled to the speed-reducing mechanism 7 via the pinion gear 41.
The speed-reducing mechanism 7 is disposed more forward than the motor 6. The speed-reducing mechanism 7 couples the rotor-shaft portion 33 and the spindle 8 to each other. The speed-reducing mechanism 7 transmits the rotation of the rotor 27 to the spindle 8. The speed-reducing mechanism 7 causes the spindle 8 to rotate at a rotational speed that is lower than the rotational speed of the rotor-shaft portion 33, but at an increased torque. The speed-reducing mechanism 7 comprises a planetary-gear mechanism.
The speed-reducing mechanism 7 comprises a plurality of gears. The gears of the speed-reducing mechanism 7 are driven by the rotor 27.
The planetary-gear mechanism of the speed-reducing mechanism 7 comprises a plurality of planet gears 42, which is disposed around the pinion gear 41, and an internal gear 43, which is disposed around the plurality of planet gears 42. The pinion gear 41, the planet gears 42, and the internal gear 43 are each housed in the hammer case 4. Each of the planet gears 42 meshes with the pinion gear 41. The planet gears 42 are respectively supported in a rotatable manner by the spindle 8 via (by) pins 42P. The spindle 8 is rotated by the planet gears 42. The internal gear 43 has inner teeth, which mesh with the planet gears 42. The internal gear 43 is fixed to the hammer case 4. The internal gear 43 is always non-rotatable relative to the hammer case 4.
When the rotor-shaft portion 33 rotates owing to the drive of the motor 6, the pinion gear 41 rotates, and the planet gears 42 revolve around the pinion gear 41. The planet gears 42 revolve while meshing with the inner teeth of the internal gear 43. Owing to the revolving of the planet gears 42, the spindle 8, which is connected to the planet gears 42 via the pins 42P, rotates at a rotational speed that is lower than the rotational speed of the rotor-shaft portion 33.
The spindle 8 is disposed more forward than at least a portion of the motor 6. The spindle 8 is disposed more forward than the stator 26. At least a portion of the spindle 8 is disposed more forward than the rotor 27. At least a portion of the spindle 8 is disposed forward of the speed-reducing mechanism 7. The spindle 8 is disposed rearward of the anvil 10. The spindle 8 is rotated by the rotor 27. The spindle 8 rotates owing to the rotational force of the rotor 27 transmitted thereto by the speed-reducing mechanism 7. The spindle 8 transmits the rotational force of the motor 6 to the anvil 10 via balls 48 and a hammer 47.
The spindle 8 comprises a flange portion 8A and a spindle-shaft portion 8B, which protrudes forward from the flange portion 8A. The planet gears 42 are supported on the flange portion 8A in a rotatable manner via the pins 42P. The rotational axis of the spindle 8 and rotational axis AX of the motor 6 coincide (are colinear) with each other. The spindle 8 rotates about rotational axis AX. The spindle 8 is supported by a spindle bearing 44 in a rotatable manner. Protruding portions 8C are provided on a rear side of the spindle 8. The protruding portions 8C protrude rearward from the flange portion 8A. The protruding portions 8C are disposed so as to surround the spindle bearing 44.
The bearing box 24 is disposed at least partly around the spindle 8. The spindle bearing 44 is held in the bearing box 24. The bearing box 24 has a protruding portion 24B, which protrudes forward from a front surface of the bearing box 24. The spindle bearing 44 is disposed around the protruding portion 24B.
The impact mechanism 9 is driven by the motor 6. The rotational force of the motor 6 is transmitted to the impact mechanism 9 via the speed-reducing mechanism 7 and the spindle 8. The impact mechanism 9 impacts the anvil 10 in the rotational direction using the rotational force of the spindle 8, which is rotated by the motor 6. The impact mechanism 9 comprises the hammer 47, the balls 48, a first coil spring 49, a second coil spring 50, a third coil spring 51, a first washer 52, and a second washer 53. The impact mechanism 9, which comprises the hammer 47, the balls 48, the first coil spring 49, the second coil spring 50, the third coil spring 51, the first washer 52, and the second washer 53, is housed in the first tube portion 401 of the hammer case 4.
The hammer 47 is disposed more forward than the speed-reducing mechanism 7. The hammer 47 is disposed around the spindle 8. The hammer 47 is held by the spindle 8. The balls 48 are disposed between the spindle 8 and the hammer 47. The hammer 47 comprises a tube-shaped hammer body 47D and hammer-projection portions 47E, which are provided at front portions of the hammer body 47D. A ring-shaped recessed portion (annular groove) 47C is provided in a rear surface of the hammer body 47D. The recessed portion 47C is recessed forward from the rear surface of the hammer body 47D.
The hammer 47 is disposed around the spindle-shaft portion 8B. The hammer 47 has a through hole 47A in which the spindle-shaft portion 8B is disposed.
The hammer 47 is rotated using the rotational energy output by the motor 6 as will be further explained below. The rotational force of the motor 6 is transmitted to the hammer 47 via the speed-reducing mechanism 7 and the spindle 8. The hammer 47 is rotatable, together with the spindle 8, using the rotational force of the spindle 8, which is rotated by the motor 6. The rotational axis of the hammer 47, the rotational axis of the spindle 8, and rotational axis AX of the motor 6 coincide with each other. The hammer 47 rotates about rotational axis AX.
The first washer 52 is disposed inside the recessed portion 47C. The first washer 52 is supported by the hammer 47 via a plurality of balls 54. The balls 54 are disposed more forward than the first washer 52.
The second washer 53 is disposed more rearward than the first washer 52 inside the recessed portion 47C. The outer diameter of the second washer 53 is smaller than the outer diameter of the first washer 52. The second washer 53 and the hammer 47 can move relative to each other in the front-rear direction.
The first coil spring 49 is disposed around the spindle-shaft portion 8B. A rear-end portion of the first coil spring 49 is supported by (contacts) the flange portion 8A. A front-end portion of the first coil spring 49 is disposed inside the recessed portion 47C and is supported by (contacts) the first washer 52. The first coil spring 49 continuously generates a biasing force (elastic force) that urges the hammer 47 to move forward.
The second coil spring 50 is disposed around the spindle-shaft portion 8B. The second coil spring 50 is disposed radially inward of the first coil spring 49. A rear-end portion of the second coil spring 50 is supported by (contacts) the flange portion 8A. A front-end portion of the second coil spring 50 is disposed inside the recessed portion 47C and is supported by (contacts) the second washer 53. The second coil spring 50 generates a biasing force (elastic force) that urges the hammer 47 to move forward, in particular, when (after) the hammer 47 has moved rearward.
The third coil spring 51 is disposed around the spindle-shaft portion 8B. The third coil spring 51 is disposed radially inward of the first coil spring 49. The third coil spring 51 is disposed inside the recessed portion 47C. A rear-end portion of the third coil spring 51 is supported by (contacts) the second washer 53. A front-end portion of the third coil spring 51 is supported by (contacts) the first washer 52. The third coil spring 51 generates a biasing force (elastic force) that urges the second coil spring 50 to move rearward. Owing to the biasing force of the third coil spring 51, the rear-end portion of the second coil spring 50 is pushed against the flange portion 8A. Thereby, the second coil spring 50 is suppressed from moving freely relative towards the flange portion 8A.
The balls 48 are made of a metal such as steel. The balls 48 are disposed between the spindle-shaft portion 8B and the hammer 47. The spindle 8 has spindle grooves 8D, in each of which at least a portion of the corresponding ball 48 is disposed. The spindle grooves 8D are provided in portions of an outer circumferential surface 8S of the spindle-shaft portion 8B. The hammer 47 has hammer grooves 47B, in each of which at least a portion of the corresponding ball 48 is disposed. The hammer grooves 47B are provided in portions of an inner surface of the hammer 47. The balls 48 are disposed between the spindle grooves 8D and the hammer grooves 47B, respectively. Each of the balls 48 can roll along the inner side of the corresponding spindle groove 8D and the inner side of the corresponding hammer groove 47B, respectively. The hammer 47 is movable along with the balls 48. The spindle 8 and the hammer 47 can move relative to each other in both the axial direction and the rotational direction within a movable range defined by the spindle grooves 8D and the hammer grooves 47B.
The anvil 10 is disposed more forward than the motor 6. The anvil 10 is an output part of the impact tool 1, which rotates based on the rotational force of the rotor 27. At least a portion of the anvil 10 is disposed more forward than the hammer 47. The anvil 10 has a tool hole 10A into which a tool accessory (e.g., a screwdriver bit) is inserted. The tool hole 10A is provided at (in) a front-end portion of the anvil 10. Thus, the tool accessory is mounted in (on) the anvil 10.
The anvil 10 has an anvil-protruding portion 10B. The anvil-protruding portion 10B is provided at a rear-end portion of the anvil 10. The anvil-protruding portion 10B protrudes rearward from the rear-end portion of the anvil 10. The spindle 8 is disposed rearward of the anvil 10. A spindle recessed portion 8E is provided at a front-end portion of the spindle-shaft portion 8B. The anvil-protruding portion 10B is disposed in the spindle recessed portion 8E.
The anvil 10 comprises a rod-shaped anvil-shaft portion 101 and anvil-projection portions 102. The tool hole 10A is provided at a front-end portion of the anvil-shaft portion 101. The tool accessory is mounted in the anvil-shaft portion 101. The anvil-projection portions 102 are provided at a rear side of the anvil 10. The anvil-projection portions 102 protrude radially outward from the anvil-shaft portion 101.
The anvil 10 is supported in a rotatable manner by bearings 46. The rotational axis of the anvil 10, the rotational axis of the hammer 47, the rotational axis of the spindle 8, and rotational axis AX of the motor 6 coincide (are colinear) with each other. The anvil 10 rotates about rotational axis AX. The bearings 46 are disposed around the anvil-shaft portion 101. The bearings 46 are disposed in the interior of the second tube portion 402 of the hammer case 4. The bearings 46 are held in the second tube portion 402 of the hammer case 4. The bearings 46 support a front portion of the anvil-shaft portion 101 in a rotatable manner.
O-rings 45 are disposed between the bearings 46 and the anvil-shaft portion 101. The O-rings 45 contact an outer circumferential portion of the anvil-shaft portion 101 and an inner circumferential portion of the bearings 46, respectively.
In the embodiment, two of the bearings 46 are disposed in the axial direction. The bearings 46 include a bearing 46A and a bearing 46B, which is disposed more rearward than the bearing 46A.
Additionally, in the embodiment, two of the O-rings 45 are disposed in the axial direction. The O-rings 45 include an O-ring 45A and an O-ring 45B, which is disposed more rearward than the O-ring 45A. The O-ring 45A is disposed between the bearing 46A and the anvil-shaft portion 101. The O-ring 45B is disposed between the bearing 46B and the anvil-shaft portion 101.
The bearings 46 are ball bearings. Each of the bearings 46 has an inner ring, balls, and an outer ring. The inner ring of the bearing 46A contacts the O-ring 45A. The inner ring of the bearing 46B contacts the O-ring 45B. The balls of the bearings 46 are disposed between the inner rings and the outer rings in the radial direction. The balls of the bearings 46 contact the inner rings and the outer rings, respectively. A plurality of the balls of the bearings 46 is disposed in the circumferential direction. The outer rings are disposed more radially outward than the inner rings and the balls. The outer ring of the bearing 46A contacts an inner circumferential surface of the second tube portion 402. The outer ring of the bearing 46B contacts an inner circumferential surface of the second tube portion 402.
At least a portion of the hammer 47 can contact the anvil-projection portions 102. The hammer-projection portions 47E are provided protruding forward from front portions of the hammer 47. The hammer-projection portions 47E and the anvil-projection portions 102 are capable of contacting each other. In the state in which the hammer 47 and the anvil-projection portions 102 are in contact with each other, the anvil 10 rotates together with the hammer 47 and the spindle 8 when the motor 6 is driven.
The anvil 10 can also be impacted in the rotational direction by the hammer 47. For example, when the load (torque) acting on the anvil 10 becomes high during screw-tightening work, the anvil 10 can no longer be caused to rotate by merely the load of the first coil spring 49. When the anvil 10 can no longer be caused to rotate by merely the load of the first coil spring 49, the rotation of the anvil 10 and the hammer 47 will momentarily stop. The spindle 8 and the hammer 47 can move relative to each other in the axial direction and the circumferential direction via the balls 48. When the rotation of the hammer 47 momentarily stops, the rotation of the spindle 8 continues owing to the motive power generated by the motor 6. Thus, when the rotation of the hammer 47 has momentarily stopped and the spindle 8 continue to rotate, the balls 48 move rearward while being guided by the spindle grooves 8D and the hammer grooves 47B. The hammer 47 receives a force from the balls 48 and moves rearward along with the balls 48. That is, while the rotation of the anvil 10 is momentarily stopped, the hammer 47 moves rearward owing to the rotation of the spindle 8. The contact between the hammer 47 and the anvil-projection portions 102 is released by the movement of the hammer 47 rearward.
As described above, the first coil spring 49 continuously generates a biasing force that urges the hammer 47 to move forward. The second coil spring 50 generates a biasing force that urges the hammer 47 to move forward after the hammer 47 has moved more rearward than a defined position. After having moved rearward, the hammer 47 moves forward owing to the biasing forces of the first coil spring 49 and the second coil spring 50. When the hammer 47 moves forward, it receives a force in the rotational direction from the balls 48. That is, the hammer 47 moves forward while rotating. When the hammer 47 moves forward while rotating, the hammer 47 contacts the anvil-projection portions 102 while rotating. Thereby, the anvil-projection portions 102 are impacted (hammered) in the rotational direction by the hammer-projection portions 47E of the hammer 47. The motive power of the motor 6 and the inertial force of the hammer 47 both act on the anvil 10. Accordingly, the anvil 10 can rotate about rotational axis AX with a higher torque.
The tool-holding mechanism 11 is disposed around a front portion of the anvil 10. The tool-holding mechanism 11 holds the tool accessory that is inserted in the tool hole 10A of the anvil 10. The tool accessory is detachable from the tool-holding mechanism 11.
The tool-holding mechanism 11 comprises balls 71, a leaf spring (flat spring, circular spring) 72, a sleeve 73, a coil spring 74, and a positioning member 75.
The anvil 10 has support recessed portions 76, which support the balls 71. The support recessed portions 76 are formed in the outer circumferential surface of the anvil-shaft portion 101. In the embodiment, two of the support recessed portions 76 are formed in the anvil-shaft portion 101.
The balls 71 are supported by the anvil 10 in a movable manner. The balls 71 are disposed in the support recessed portions 76. One of the balls 71 is disposed in each one of the support recessed portions 76.
Through holes, which connect the inner surfaces of the support recessed portions 76 and the inner surface of the tool hole 10A, are formed in the anvil-shaft portion 101. In the state in which the balls 71 are supported by the support recessed portions 76, at least a portion of each of the balls 71 is disposed in the interior of the tool hole 10A. The balls 71 can thereby fix the tool accessory that is inserted into the tool hole 10A. The balls 71 are capable of moving to an engaged position, at which the tool accessory is fixed, and a released position, at which the fixing of the tool accessory is released.
The leaf spring 72 generates an elastic force that urges the balls 71 to move toward the engaged position. The leaf spring 72 is disposed around the anvil-shaft portion 101. The leaf spring 72 generates an elastic force that urges the balls 71 to move radially inward.
The sleeve 73 is a circular-tube-shaped member. The sleeve 73 is disposed around the anvil-shaft portion 101. The sleeve 73 is capable of moving in (along) an axial direction relative to the anvil-shaft portion 101. The sleeve 73 can prevent the balls 71, which are disposed at the engaged position, from escaping from the engaged position. By moving in the axial direction, the sleeve 73 can transition (move) the balls 71 from the engaged position to the released position, that is, to the state in which the balls 71 are movable.
The sleeve 73 is capable of moving along the circumference of the anvil-shaft portion 101 to a blocked position, at which movement of the balls 71 radially outward is blocked, and to a permitted position, at which movement of the balls 71 radially outward is permitted.
By disposing the sleeve 73 at the blocked position, radially outward movement of the balls 71 disposed at the engaged position is blocked. That is, by disposing the sleeve 73 at the blocked position, the balls 71 disposed at the engaged position are blocked from escaping from the engaged position. By disposing the sleeve 73 at the blocked position, the state in which the tool accessory is fixed by the balls 71 is maintained.
By moving the sleeve 73 to the permitted position, the balls 71 disposed at the engaged position are permitted to move radially outward. By axially moving the sleeve 73 to the permitted position, the balls 71 are transitioned from the engaged position to the released position, that is, to the state in which the balls 71 are movable. That is, by disposing the sleeve 73 at the permitted position, the balls 71 disposed at the engaged position are permitted to escape from the engaged position. By disposing the sleeve 73 at the permitted position, the state in which the tool accessory is fixed by the balls 71 becomes releasable.
The coil spring 74 generates an elastic force to move the sleeve 73 toward the blocked position. The coil spring 74 is disposed around the anvil-shaft portion 101. The blocked position is defined more rearward than the permitted position. The coil spring 74 generates an elastic force that urges the sleeve 73 to move rearward.
The positioning member 75 is a ring-shaped member that is fixed to the outer circumferential surface of the anvil-shaft portion 101. The positioning member 75 is fixed at a position at which it is opposable to a rear-end portion of the sleeve 73. The positioning member 75 positions the sleeve 73 at the blocked position. The sleeve 73, to which the rearward-moving elastic force of the coil spring 74 is imparted, is positioned at the blocked position when it contacts the positioning member 75.
The fan 12 is disposed more rearward than the stator 26 of the motor 6. The fan 12 generates an airflow for cooling the motor 6. The fan 12 is fixed to at least a portion of the rotor 27. The fan 12 is fixed to a rear portion of the rear-side shaft portion 33R via a bushing 12A. The fan 12 is disposed between the rear-side rotor bearing 39R and the stator 26. The fan 12 rotates when the rotor 27 rotates. By rotating the rotor-shaft portion 33, the fan 12 rotates together with the rotor-shaft portion 33. By rotating the fan 12, air outside of the housing 2 flows into the interior space of the housing 2 via the air-intake ports 19. Air that has flowed into the interior space of the housing 2 flows through the interior space of the housing 2 and thereby cools the motor 6. By rotating the fan 12, air that has flowed through the interior space of the housing 2 flows out to the exterior of the housing 2 via the air-exhaust ports 20.
The battery-mounting part 13 is disposed on (at) a lower portion of the battery-holding part 23. The battery-mounting part 13 is connected to a battery pack 25. The battery pack 25 is mounted on the battery-mounting part 13. In the embodiment, there is one battery-mounting part 13. One battery pack 25 is mounted on the battery-mounting part 13. The battery pack 25 is detachable from the battery-mounting part 13. The battery pack 25 is mounted on the battery-mounting part 13 by being inserted into the battery-mounting part 13 from forward of the battery-holding part 23. The battery pack 25 is removed from the battery-mounting part 13 by being pulled forward from the battery-mounting part 13. The battery pack 25 comprises secondary batteries. In the embodiment, the battery pack 25 comprises rechargeable lithium-ion batteries. When mounted on the battery-mounting part 13, the battery pack 25 can supply electric power to the impact tool 1. The motor 6 is driven using the electric power supplied from the battery pack 25. The operation-and-display part 16 operates using electric power supplied from the battery pack 25.
The rated voltage of the battery pack 25 is 18 V or more. The rated voltage of the battery pack 25 may be 18 V, 36 V, or 72 V.
The trigger lever 14 is provided on the grip part 22. The trigger lever 14 is manipulated (pushed, slid) by the user to start the motor 6. The motor 6 switches between being driven and being stopped by the manipulation of the trigger lever 14.
The forward/reverse-change lever 15 is provided at an upper portion of the grip part 22. The forward/reverse-change lever 15 is manipulated by the user. By manipulating the forward/reverse-change lever 15, the rotational direction of the motor 6 switches from one of the forward-rotational direction and the reverse-rotational direction to the other. Switching the rotational direction of the motor 6 switches the rotational direction of the spindle 8.
The operation-and-display part 16 is provided on the battery-holding part 23. The operation-and-display part 16 is provided on an upper surface of the battery-holding part 23 more on the forward side than the grip part 22. The operation-and-display part 16 comprises a plurality of manipulatable buttons 16A. By manipulating (pressing) one of the manipulatable buttons 16A, the user can switch (change, cycle) the action mode of the motor 6.
The mode-change switch 17 is provided at an upper portion of the trigger lever 14. The user can also manipulate (press) the mode-change switch 17 to switch (change, cycle) the action mode of the motor 6.
The light assembly 18 emits illumination light. The light assembly 18 illuminates the anvil 10 and the periphery of the anvil 10 with the illumination light. The light assembly 18 illuminates forward of the anvil 10 with the illumination light. Additionally, the light assembly 18 illuminates the tool accessory that is mounted in the anvil 10 and the periphery of the tool accessory with the illumination light. In the embodiment, the light assembly 18 comprises a ring-shaped base member 18A and a plurality of light-emitting devices 18B that are held by the base member 18A. The base member 18A is disposed around the second tube portion 402 of the hammer case 4. Additionally, the light assembly 18 includes a ring member 18C that restrains (blocks) the base member 18A from coming off of the second tube portion 402 forward.
In the embodiment, a ring member 61 is disposed between the anvil-projection portions 102 and the bearing 46B, which is the rear-side bearing. The ring member 61 is a member having circular band (washer) shape. The ring member 61 is made of a metal. Iron is an illustrative example of a metal for forming the ring member 61. A front surface and a rear surface of the ring member 61 are each flat. At least a portion of the rear surface of the ring member 61 is disposed so as to oppose the front surfaces of the anvil-projection portions 102. At least a portion of the front surface of the ring member 61 contacts a rear-end surface of the bearing 46B, which is the rear-side bearing.
In the embodiment, a restraining member 62, which engages with the hammer case 4 and the ring member 61, is provided. The restraining member 62 restrains (blocks) the ring member 61 from coming off rearward. A snap ring and a C-ring are illustrative examples of the restraining member 62. The restraining member 62 is disposed so as to contact the ring member 61.
The hammer case 4 has a support surface 4A that opposes at least a portion of the front surface of the ring member 61. The support surface 4A of the hammer case 4 and the rear-end surface of the bearing 46, which is the rear-side bearing, are disposed substantially within the same plane. At least a portion of the ring member 61 is disposed between the front surfaces of the anvil-projection portions 102 and the support surface 4A of the hammer case 4.
The ring member 61 blocks contact between the hammer case 4 and the anvil-projection portions 102.
At least a portion of the front surface of the ring member 61 contacts the support surface 4A of the hammer case 4. Additionally, at least a portion of the front surface of the ring member 61 contacts the rear-end surface of the bearing 46B, which is the rear-side bearing.
A front-surface, outer-edge portion, which is an outer-edge portion of the front surface of the ring member 61, contacts the support surface 4A of the hammer case 4. A front-surface, inner-edge portion, which is an inner-edge portion of the front surface of the ring member 61, is capable of contacting a rear-end surface of the outer ring of the bearing 46B. The ring member 61 is disposed more radially outward than the inner ring of the bearing 46B. The ring member 61 and the inner ring of the bearing 46B do not make contact.
A rear-surface, outer-edge portion, which is an outer-edge portion of the rear surface of the ring member 61, contacts the restraining member 62.
A groove 4B, in which at least a portion of the restraining member 62 fits, is provided in an inner surface of the first tube portion 401 of the hammer case 4. Fluctuation in the relative positions of the hammer case 4 and the restraining member 62 is curtailed in at least the axial direction owing to the restraining member 62 being disposed in the groove 4B.
The bearings 46 are restrained (blocked) from coming off rearward (the other side in the axial direction) by the ring member 61 and the restraining member 62.
Front-surface, outer-edge portions, which indicate outer-edge portions of the front surfaces of the anvil-projection portions 102, tilt rearward as they go radially outward.
The restraining member 62 is disposed between the front-surface, outer-edge portions of the anvil-projection portions 102 and the rear-surface, outer-edge portion of the ring member 61 in the front-rear direction.
The spindle 8 is disposed more forward than the motor 6. The spindle 8 rotates about rotational axis AX in response to energization of the motor 6. The spindle grooves 8D are formed in the spindle 8. The spindle 8 comprises the flange portion 8A and the spindle-shaft portion 8B, which protrudes forward from the flange portion 8A. Two of the spindle grooves 8D are formed in the spindle-shaft portion 8B. The balls 48 are held in the spindle grooves 8D. One ball 48 is held in one corresponding spindle groove 8D.
Each of the spindle grooves 8D has a center portion 81 and a pair of end portions 82. The location of the center portion 81 and the location of the end portions 82 differ in the circumferential direction. One of the end portions 82 is disposed more toward one side than the center portion 81 in the circumferential direction, and the other end portion 82 is disposed more toward the other side than the center portion 81 in the circumferential direction. The end portions 82 of each of the spindle grooves 8D are disposed more rearward than the center portion 81 of the spindle groove 8D.
As shown in
As shown in
That is, the spindle grooves 8D and the hammer grooves 47B have arcuate shapes, respectively, in the virtual plane 91.
The spindle grooves 8D are formed by a cutting tool 92 such as an end mill. As shown in
The center 92C of the cutting tool 92 is disposed more radially inward than the outer circumferential surface 8S of the spindle-shaft portion 8B for at least a portion of the formation of the spindle grooves 8D. The center 92C of the cutting tool 92 is disposed more radially inward than the outer circumferential surface 8S of the spindle-shaft portion 8B when forming the end portions 82 of the spindle grooves 8D. The center 92C of the cutting tool 92 is disposed more radially outward than the outer circumferential surface 8S of the spindle-shaft portion 8B when forming the center portions 81 of the spindle grooves 8D.
The balls 48 move with circular motion so as to move along the virtual circle 90 defined by the virtual plane 91. As shown in
The depth of each of the spindle grooves 8D differs in accordance with its location in the axial direction (front-rear direction) of the spindle 8. In the embodiment, the depth at the rear portion of each of the spindle grooves 8D is deeper than the depth at the front portion of that spindle groove 8D. That is, the depth at the end portions 82 of each of the spindle grooves 8D is deeper than the depth at the center portion 81 of that spindle groove 8D. For example, and without limitation, the depth of the spindle grooves 8B at the end portions 82 is preferably at least 5% deeper than the depth of the spindle grooves 8B at the center portion 81, e.g., 5-15% deeper. In other words, a bottom or base of the spindle grooves 8B at the end portions 82 is preferably at least 5% greater from the outer circumferential surface 8S of the spindle-shaft portion 8B than the bottom or base of the spindle grooves 8B at the center portion 81, e.g., 5-15% greater.
The center 48C of each of the balls 48 may be disposed inside the corresponding spindle groove 8D at at least a portion of the corresponding spindle groove 8D. The center 48C of each of the balls 48 may be disposed inside the corresponding spindle groove 8D at the end portions 82 of the corresponding spindle groove 8D. The center 48C of each of the balls 48 may be disposed outside the corresponding spindle groove 8D at the center portion 81 of the corresponding spindle groove 8D. In other words, when each of the balls 48 is at one of the end portions 82, the center 48C of the balls 48 is disposed radially inward of the outer circumferential surface 8S of the spindle-shaft portion 8B, whereas when each of the balls 48 is at the center portions 81, the center 48C of the balls 48 is disposed radially outward of the outer circumferential surface 8S of the spindle-shaft portion 8B.
Next, the operation of the impact tool 1 will be explained.
When performing screw-tightening work on a work object (e.g., a Phillips-head screw), the tool accessory used for the screw-tightening work (a driver bit) is inserted in the tool hole 10A of the anvil 10. The tool accessory inserted into the tool hole 10A is held by the tool-holding mechanism 11. After mounting the tool accessory on (in) the anvil 10, the user grips the grip part 22 with, for example, their right hand and pulls the trigger lever 14. When the trigger lever 14 is pulled, electric power is supplied from the battery pack 25 to the motor 6, the motor 6 starts, and the light assembly 18 turns ON at the same time. Owing to the starting of the motor 6, the rotor-shaft portion 33 of the rotor 27 rotates. When the rotor-shaft portion 33 rotates, the rotational force of the rotor-shaft portion 33 is transmitted to the planet gears 42 via the pinion gear 41. Because the planet gears 42 are meshed with the inner teeth of the internal gear 43, the planet gears 42 revolve around the pinion gear 41 while rotating. The planet gears 42 are supported in a rotatable manner by the spindle 8 via the pins 42P. Owing to the revolving of the planet gears 42, the spindle 8 rotates at a rotational speed that is lower than the rotational speed of the rotor-shaft portion 33.
When the hammer 47 and the anvil-projection portions 102 are in continuous contact with each other and the spindle 8 rotates, the anvil 10 rotates together with the hammer 47 and the spindle 8. The screw-tightening work progresses owing to the rotation of the anvil 10. The rotational force of the spindle 8 is transmitted to the hammer 47 via the balls 48. The balls 48 are disposed at the center portions 81 of the spindle grooves 8D, respectively, while the spindle 8 and the hammer 47 are rotating together.
The rotation of the anvil 10 and the hammer 47 momentarily (intermittently) stops when, owing to the progress of the screw-tightening work, a load torque from the screw greater than or equal to a prescribed value acts on the anvil 10, e.g., when the screw has been driven deeply into a workpiece, such as a piece of wood. In the description below, the load torque when the rotation of the anvil 10 and the hammer 47 stops is called the stop-load torque where appropriate, and the point in time when the rotation of the anvil 10 and the hammer 47 has stopped is called the hammer-stop time where appropriate.
That is, when a torque equal to or greater than the stop-load torque from the screw acts on the anvil 10 during the screw-tightening work, the rotation of the anvil 10 and the hammer 47 momentarily stops. However, the spindle 8 continues to rotate owing to driving of the motor 6. Thus, as the spindle 8 continues to rotate in the state in which the rotation of the hammer 47 is momentarily stopped, the hammer 47 is caused to move rearward relative to the spindle 8.
Therefore, contact between the hammer 47 and the anvil-projection portions 102 is released when the hammer 47 has moved rearward relative to the spindle 8. In the description below, the point in time when the hammer 47 moves rearward relative to the spindle 8 and the contact between the hammer 47 and the anvil-projection portions 102 is released is called the hammer-release time where appropriate.
As shown in
When the contact between the hammer 47 and the anvil-projection portions 102 is released and the hammer 47 moves rearward relative to the spindle 8, the balls 48 move from the center portions 81 of the spindle grooves 8D to the end portions 82. At this time, the hammer 47 moves most rearward relative to the spindle 8 in the axial direction within the movable range of the hammer 47. In the description below, the point in time at which the hammer 47 moves most rearward relative to the spindle 8 is called the maximum-stroke time where appropriate.
As shown in
After having moved rearward, the hammer 47 moves forward while rotating owing to the forward biasing forces of the first coil spring 49 and the second coil spring 50. As the hammer 47 moves forward relative to the spindle 8, each of the balls 48 moves from one of the end portions 82 to the center portion 81 of the corresponding spindle groove 8D. In the description below, the point in time at which the hammer 47 moves forward while rotating and immediately before the hammer 47 impacts the anvil 10 is called the impact-start time where appropriate. As shown in
The anvil 10 is impacted by the hammer 47 in the rotational direction owing to the hammer 47 moving forward while rotating. Thereby, the anvil 10 is rotated about rotational axis AX with a higher torque. Consequently, the screw is tightened into the work pieces with a torque that is higher than the stop-load torque.
In the description below, the torque acting on the anvil 10 when the hammer 47 starts impacting the anvil 10 is called the impact-start torque where appropriate. The impact-start torque is equal to the stop-load torque described above. That is, the impact-start torque is equal to the load torque from the screw acting on the anvil 10 when the rotation of the anvil 10 and the hammer 47 momentarily stops during the screw-tightening work.
The effects of the impact tool 1 according to the embodiment are described below while making a comparison to a spindle 800 according to a conventional example. In the description below, the spindle 8 according to the embodiment described above is called the working example where appropriate, and the spindle 800 according to the conventional example is called the comparative example where appropriate.
As shown in
As shown in
As shown in
As shown in
As shown in
If the impact-start torque is relatively small, the occurrence of the cam-out phenomenon, in which the tool accessory (driver bit) adversely slips out of the cruciform groove of the head portion of the screw during the screw-tightening work, is less likely to occur. According to the findings of the present inventor, the timing of the start of impact on the anvil 10 is earlier in case the impact-start torque is small. That is, when the impact-start torque is smaller, the amount of screw rotation from the start of screw tightening to the impact-start time becomes smaller. The occurrence of the cam-out phenomenon is curtailed in case the screw-rotation amount from the screw-tightening start time until the impact-start time is smaller.
Using an impact tool according to the working example and an impact tool according to the comparative example, fastening work for a coarse thread of length 90 mm, a coarse thread of length 120 mm, and a metal screw of length 150 mm was performed.
As shown in
It is noted that angle α formed between tangent vector Vr and load vector Vp at the impact-start time in the working example is 43.1°. Angle α formed between tangent vector Vr and load vector Vp at the impact-start time in the comparative example is 57.3°.
It is noted that angle α formed between tangent vector Vr and load vector Vp at the impact time in the working example is 58.6°. Angle α formed between tangent vector Vr and load vector Vp at the impact time in the comparative example is 85.7°.
As described above, each of the end portions 82 of each of the spindle grooves 8D is disposed more rearward than the center portion 81 of the corresponding spindle groove 8D. When the hammer 47 moves rearward relative to the spindle 8, each of the balls 48 moves to one of the end portions 82 of the corresponding spindle groove 8D. In the embodiment, angle α formed between tangent vector Vr and load vector Vp changes based on the location of the balls 48 in the spindle grooves 8D (the hammer grooves 47B). In the embodiment, as the balls 48 move closer to the end portions of the spindle grooves 8D, angle α formed between tangent vector Vr and load vector Vp becomes smaller.
This means that the smaller angle α formed between tangent vector Vr and load vector Vp becomes, the more that the force from the spindle 8 acting on the balls 48 acts exclusively in the rotational direction. In the embodiment, among the forces from the spindle 8 acting on the balls 48, because the component in the axial direction is relatively small and the component in the rotational direction is relatively large, the hammer 47 can be moved forward while rotating at a higher rotational speed.
As explained above, in the embodiment, the impact tool 1 comprises: the motor 6; the spindle 8 disposed more forward than the motor 6 and rotated by the motor 6; the spindle grooves 8D formed in the spindle 8; the balls 48 held in the spindle grooves 8D; the hammer 47 supported by the spindle 8 via the balls 48; the first coil spring 49 and the second coil spring 50, which bias the hammer 47 forward; and the anvil 10, which is impacted by the hammer 47 in the rotational direction. The depth of each of the spindle grooves 8D differs in accordance with the location thereof in the axial direction of the spindle 8.
As described above, when the stop-load torque from the screw acts on the anvil 10 during the screw-tightening work, the rotation of the anvil 10 and the hammer 47 momentarily stops. However, the spindle 8 continues to rotate owing to the driving of the motor 6. Therefore, as the spindle 8 continues to rotate in the state in which the rotation of the hammer 47 is momentarily stopped, the hammer 47 moves (is moved) rearward relative to the spindle 8. After the hammer 47 has moved rearward, the hammer 47 is moved forward while rotating owing to the biasing forces of the first coil spring 49 and the second coil spring 50. The anvil 10 is impacted by the hammer 47 in the rotational direction owing to the hammer 47 moving forward while rotating.
In the above-mentioned configuration, the depth of each of the spindle grooves 8D differs in accordance with its location in the axial direction of the spindle 8; therefore, among the forces acting between the spindle 8 and the balls 48, the component in the rotational direction becomes larger. Consequently, the rotational speed of the hammer 47—when the hammer 47 is moving forward while rotating—can be increased. Because the rotational speed of the hammer 47 can be increased, the hammer 47 can impact the anvil 10 with a larger impact force.
Additionally, among the forces acting between the spindle 8 and the balls 48, the force component in the rotational direction becomes larger and the force component in the axial direction becomes smaller; therefore, even if the first coil spring 49 and the second coil spring 50 were to have small spring coefficients, the first coil spring 49 and the second coil spring 50 can still bias the hammer 47 forward. That is, even if the first coil spring 49 and the second coil spring 50 were to have small biasing forces, the hammer 47 can move forward while rotating owing to the biasing forces of the first coil spring 49 and the second coil spring 50. By using the first coil spring 49 and the second coil spring 50, which have smaller biasing forces, the timing of the start of impact of the hammer 47 on the anvil 10 can be made earlier. That is, the interval from the point in time when the rotation of the hammer 47 has stopped until the point in time when the anvil 10 is impacted by the hammer 47 becomes shorter.
During the screw-tightening work, the anvil 10 is rotated in the state in which the tool accessory mounted in the anvil 10 is inserted in a cruciform groove formed in the head portion of the screw. If the timing of the start of impact of the hammer 47 on the anvil 10 is later, that is, if the amount of screw rotation from the screw-tightening start time until the impact time when the anvil 10 is impacted by the hammer 47 is large, then the probability of the occurrence of the cam-out phenomenon, in which the tool accessory adversely comes completely out of the cruciform groove, becomes high. On the other hand, in the above-mentioned configuration, because the timing of the start of impact of the hammer 47 on the anvil 10 is earlier, the cam-out phenomenon is less likely to occur.
In the first embodiment, the end portions 82 of each of the spindle grooves 8D are disposed more rearward than the center portion 81 of that spindle groove 8D. When the hammer 47 moves rearward relative to the spindle 8, each of the balls 48 moves to one of the end portions 82 of the corresponding spindle groove 8D. The center 48C of each of the balls 48 is disposed inside the corresponding spindle groove 8D at one of the end portions 82 of the corresponding spindle groove 8D; i.e. the center 48C is below (instead of the outer diameter of) the outer circumferential surface 8S.
As the hammer 47 is moved rearward relative to the spindle 8, each of the balls 48 moves from the center portion 81 to one of the end portions 82 of the corresponding spindle groove 8D. As the hammer 47 is moved forward relative to the spindle 8 owing to the biasing forces of the first coil spring 49 and the second coil spring 50, each of the balls 48 moves from one of the end portions 82 to the center portion 81 of the corresponding spindle groove 8D. In the above-mentioned configuration, because the center 48C of each of the balls 48 sinks into the inside of the corresponding spindle groove 8D at one of the end portions 82 of the corresponding spindle groove 8D, the rotational speed of the hammer 47 when the hammer 47 moves forward while rotating can be increased. Because the rotational speed of the hammer 47 can be increased, the hammer 47 can impact the anvil 10 with a larger impact force.
In the first embodiment, the center 48C of each of the balls 48 is disposed outside the corresponding spindle groove 8D at the center portion 81 of the corresponding spindle groove 8D.
When the hammer 47 impacts the anvil 10, the balls 48 move from the end portions 82 to the center portions 81 of the spindle grooves 8D. In the above-mentioned configuration, because the center 48C of each of the balls 48 that is at the center portion 81 of the corresponding spindle groove 8D is disposed outside the corresponding spindle groove 8D (i.e. the center 48C is radially outward of the outer circumferential surface 8S), the rotational speed of the hammer 47—at the moment the hammer 47 impacts the anvil 10—can be increased. Because the rotational speed of the hammer 47 can be increased, the hammer 47 can impact the anvil 10 with a larger impact force.
In the first embodiment, the spindle 8 comprises the flange portion, which supports the rear-end portions of the first coil spring 49 and the second coil spring 50, and the spindle-shaft portion 8B, which protrudes forward from the flange portion. The spindle grooves 8D are formed in the spindle-shaft portion 8B. The trajectory of the center 48C of each of the balls 48 enters within the width (submerges below the outer diameter or outer circumferential surface 8S) of the spindle-shaft portion 8B.
In the above-mentioned configuration, because the balls 48 have a small range of motion, the rotational speed of the hammer 47 can be increased without enlarging the impact tool 1.
In the first embodiment, the trajectory of the balls 48 has an arcuate shape.
In the above-mentioned configuration, the balls 48 move with circular motion along a virtual circle having a small radius. Consequently, the hammer 47 can move in the axial direction while rotating smoothly. Because the sliding resistance of the hammer 47 becomes smaller, the rotational speed of the hammer 47 can be increased. Consequently, the hammer 47 can impact the anvil 10 with a larger impact force.
In the first embodiment, the balls 48 may move with circular motion within a virtual plane that is tilted by a prescribed angle from the rotational axis of the spindle 8.
In the above-mentioned configuration, because the sliding resistance of the hammer 47 becomes small, the rotational speed of the hammer 47 can be increased. Consequently, the hammer 47 can impact the anvil 10 with a larger impact force.
In the first embodiment, the balls 48 move with circular motion so that the center 48C of each of the balls 48 moves along a virtual circle defined in the virtual plane. The center 90C of the virtual circle is defined at a location that is offset from rotational axis AX of the spindle 8.
In the above-mentioned configuration, the hammer 47 can impact the anvil 10 with a larger impact force.
In the first embodiment, the spindle 8 comprises the flange portion, which supports the rear-end portions of the first coil spring 49 and the second coil spring 50, and the spindle-shaft portion 8B, which protrudes forward from the flange portion. The spindle grooves 8D are formed in the spindle-shaft portion 8B. The diameter of the virtual circle is smaller than the diameter of the spindle-shaft portion 8B.
In the above-mentioned configuration, because the balls 48 have a small range of motion, the rotational speed of the hammer 47 can be increased without enlarging the impact tool 1.
In the first embodiment, the impact-start torque, which is the torque acting on the anvil 10 when the hammer 47 starts impacting the anvil 10, is 1,100 N·mm or less.
In the above-mentioned configuration, because the impact-start torque is small, the timing of the start of impact of the hammer 47 on the anvil 10 will be earlier. That is, the amount of screw rotation from the screw-tightening start time until the impact time when the anvil 10 is impacted by the hammer 47 becomes smaller.
If the timing of the start of impact of the hammer 47 on the anvil 10 were to be later, that is, if the amount of screw rotation from the screw-tightening start time until the impact time when the anvil 10 is impacted by the hammer 47 is large, then the probability of the occurrence of the cam-out phenomenon, in which the tool accessory adversely comes completely out of the cruciform groove, becomes higher. On the other hand, in the above-mentioned configuration, the timing of the start of impact of the hammer 47 on the anvil 10 is earlier, and therefore the cam-out phenomenon is less likely to occur.
In the first embodiment, the impact tool 1 is provided with the battery-mounting part 13 to which the battery pack 25 is mounted. The rated voltage of the battery pack 25 is 18 V or more.
In the above-mentioned configuration, the rated voltage of the battery pack 25 is 18 V or more, and therefore a high output can be achieved for the impact tool 1.
In the first embodiment, the depth of each of the spindle grooves 8D differs in accordance with its location in the axial direction of the spindle 8. That is, the depth at the end portions 82 of the spindle grooves 8B is greater than at the center portion 81, preferably at least 5% deeper.
Consequently, the occurrence of the cam-out phenomenon is curtailed.
In other words, the depth at the rear portion of each of the spindle grooves 8D is deeper than the depth at the front portion of the corresponding spindle groove 8D. The depth at each of the end portions 82 of each of the spindle grooves 8D is deeper than the depth at the center portion 81 of the corresponding spindle groove 8D.
In the above-mentioned configuration, the center 48C of each of the balls 48 can sink into the inside of the corresponding spindle groove 8D at the end portions, and therefore, among the forces acting between the spindle 8 and the balls 48, the component in the rotational direction becomes large. Consequently, the rotational speed of the hammer 47—when the hammer 47 is moving forward while rotating—can be increased. Because the rotational speed of the hammer 47 can be increased, the hammer 47 can impact the anvil 10 with a larger impact force.
In the first embodiment described above, the widths of the end portions 82 of each of the spindle grooves 8D may be wider than the width of the center portion 81 of the corresponding spindle groove 8D. In the situation in which the widths of the end portions 82 are wider than the width of the center portion 81, among the forces acting between the spindle 8 and the balls 48, the component in the rotational direction becomes larger at the maximum-stroke time.
In the embodiment described above, the impact tool 1 is assumed to be an impact driver. The impact tool 1 may be an impact wrench. In the embodiment described above, the impact tool 1 may be an angle impact tool.
In the embodiment described above, the power supply of the impact tool 1 is not required to be the battery pack 25 and may be a commercial power supply (AC power supply). The commercial power supply supplies a voltage of 18 V or more to the motor 6.
In the embodiment described above, a configuration is described in which the first coil spring 49 and the second coil spring 50 are provided; however, the impact mechanism 9 of the impact tool 1 may be provided with a single spring. The single spring may be a linear spring, which has a substantially constant spring constant throughout the stroke range, or may be a nonlinear spring, which has a spring constant that changes in the course of the stroke.
A second embodiment of the present teachings will now be described. In the description below, structural elements that are identical or equivalent to those in the embodiment described above are assigned the same reference numeral, and descriptions of those structural elements are simplified or omitted.
The impact tool 1 according to the second embodiment comprises: a motor 6; a spindle 8 rotated by the motor 6; a cam element that includes spindle grooves 8D formed in the spindle 8, balls 48 held in the spindle grooves 8D, and hammer grooves 47B in each of which at least a portion of the corresponding ball 48 is disposed; a hammer 47 supported on the spindle 8 via the balls 48; spring elements (springs 49, 50), which bias the hammer 47 forward; and an anvil 10, which is impacted by the hammer 47 in the rotational direction, similarly to the above-described first embodiment (see
The motor 6 is supplied with a voltage of 18 V or more. The impact tool 1 according to the second embodiment is provided with a battery-mounting part 13, onto which the battery pack 25 is mounted, and the rated voltage of the battery pack 25 is 18 V or more.
The maximum fastening torque of the impact tool 1 according to the second embodiment is greater than 220 N·m. Specifically, the maximum fastening torque is 230 N·m or more and 350 N·m or less. In the impact tool 1 according to the second embodiment, the impact-start torque, which is the torque acting on the anvil 10 when the hammer 47 starts to impact the anvil 10, is 1,300 N·mm or less. Because the contact between the hammer 47 and the anvil-projection portions 102 is released, the torque acting on the anvil 10 instantaneously decreases when the impact from the hammer 47 starts. The torque at the point in time when the torque acting on the anvil 10 instantaneously decreases is the impact-start torque. In the impact tool 1 according to the second embodiment, the value calculated by dividing the maximum fastening torque (N·mm) by the impact-start torque (N·mm) is greater than 165.
Explanation of the Relationship between the Maximum Fastening Torque and the Impact-Start Torque
Next, the relationship between the maximum fastening torque and the impact-start torque will be described. Because the cam-out phenomenon tends to occur more often when the impact-start torque becomes higher, it would be desirable to lower the impact-start torque in order to curtail the occurrence of the cam-out phenomenon. On the other hand, normally the impact-start torque becomes higher when attempting to raise the maximum fastening torque. The reasons therefor are as follows.
First, the output of the motor 6 is increased to increase the rotational speed of the hammer 47 in order to increase the maximum fastening torque. Kinetic energy applied to the anvil 10 at the impact time increases owing to the rotational speed of the hammer 47 becoming higher, and thereby the maximum fastening torque becomes higher. However, when the rotational speed of the hammer 47 is relatively high, the time until the hammer 47 arrives at the angle at which it impacts the anvil 10 becomes shorter. When the rotational speed of the hammer 47 is higher, if the hammer 47, after having moved rearward once, rotates too much between advancing until impacting the anvil 10, the hammer 47 will, prior to impact, contact the rear surface of the anvil 10, resulting in a loss of kinetic energy at the impact time; thereby, a phenomenon occurs in which the fastening torque at impact time will not increase. To ensure that the advance of the hammer 47 coincides with the timing at which the hammer 47 impacts the anvil 10, the biasing force (the elastic force) of the spring element(s) is made to be larger and the hammer 47 is caused to advance more quickly.
However, when the biasing force (elastic force) of the spring element(s) becomes larger, although the hammer 47 may be advanced more quickly, a larger biasing force is applied to the hammer 47 also when the hammer 47 is being moved rearward. As a result, the torque required to rotate and to move the hammer 47 rearward (referred to as the “required torque” below) increases. The impact-start torque corresponds to the maximum value of the required torque when the hammer 47 is moved rearward to the location where the contact between the hammer 47 and the anvil-projection portions 102 is released. Accordingly, if the biasing force of the spring element(s) is made to be larger and the required torque becomes higher, then the impact-start torque becomes commensurately higher.
Thus, normally, when the maximum fastening torque is raised (increased), the impact-start torque becomes higher. In contrast, by making the impact-start torque 1,300 N·mm or less, even though the maximum fastening torque is higher than 220 N·m, the second embodiment realizes both a high maximum fastening torque and curtailment of cam-out (a low impact-start torque).
Next, a structure for lowering (reducing) the impact-start torque while raising (increasing) the maximum fastening torque will be described. In the second embodiment, the cam element and the spring elements each has a structure that causes the required torque, which is the torque needed to rotate the hammer 47, to nonlinearly increase in order to (a) suppress the increase in the required torque in a first region in which the rearward stroke of the hammer 47 is shorter, and (b) boost the increase in the required torque in a second region in which the stroke of the hammer 47 is longer.
At the impact time, after the hammer 47 has moved rearward against the biasing force of the spring elements 49, 50 while rotating owing to the rotational torque imparted from the motor 6 via the spindle 8 and the cam element, the hammer 47 advances while rotating owing to the elastic force of the compressed spring elements 49, 50 and the cam element. Referring to
First, the actions of the spring elements 49, 50, 51 will be described in greater detail with reference to
The spring elements include the first coil spring 49 and the second coil spring 50. In the embodiments, the spring elements further include the third coil spring 51.
As described above, the first coil spring 49 is supported by (contacts) the flange portion 8A and biases the hammer 47 via the first washer 52 and the balls 54. In
The second coil spring 50 is also supported by (also contacts) the flange portion 8A and biases the hammer 47 via the second washer 53, the third coil spring 51, the first washer 52, and the balls 54. The second coil spring 50 is configured (designed) to apply a biasing force to the hammer 47 only after the hammer has moved rearward away from location Q1; that is, at location Q1, the second coil spring 50 does not apply a biasing force to the hammer 47.
The third coil spring 51 is arranged between the second coil spring 50 and the hammer 47, and provides a biasing force (elastic force) that urges the second coil spring 50 to move rearward, thereby restricting (preventing) the second coil spring 50 from moving freely relative to the flange portion 8A. The third coil spring 51 is configured (designed) such that its spring constant is smaller than the spring constant of the first coil spring 49 and the spring constant of the second coil spring 50. The biasing force of the third coil spring 51 is preferably small enough to restrict (prevent) the second coil spring 50 from moving freely without significantly affecting the stroke (movement) of the hammer 47.
Therefore, in the state in which the stroke of the hammer 47 is 0 (i.e. when the hammer 47 is disposed at location Q1), the second coil spring 50 is in the state in which the second coil spring 50 is substantially at its natural length (i.e. its non-compressed and non-stretched state) and applies no biasing force to the hammer 47. The second coil spring 50 applies a biasing force, which is dependent on the stroke (axial position) of the hammer 47, to the hammer 47 when the hammer 47 has moved rearward to the location (Q4) at which the third coil spring 51 becomes its solid length. The spring constant of the second coil spring 50 is preferably larger than the spring constant of the first coil spring 49.
It is noted that, in the embodiments, the first coil spring 49, the second coil spring 50, and the third coil spring 51 are linear springs. For example, the spring constant of the first coil spring 49 may be 36.4 N/mm. The spring constant of the second coil spring 50 may be 150 N/mm. The spring constant of the third coil spring 51 may be 1.6 N/mm.
As shown in
In the embodiments, the degree of change in the required torque with respect to the stroke in first region R1 is smaller than the degree of change in the required torque with respect to the stroke in second region R2. That is, with regard to the spring elements, the degree of change in the biasing force (the elastic force) with respect to the stroke in first region R1 is smaller than the degree of change in the biasing force (the elastic force) with respect to the stroke in second region R2.
The spring elements bias the hammer 47 using the biasing force of substantially only the first coil spring 49 while the axial position of the hammer 47 is within first region R1. To be more precise, the reason is that while the biasing force of the third coil spring 51 is acting on the hammer 47, the biasing force of the third coil spring 51 is small enough that it can be ignored as compared to the biasing force of the first coil spring 49.
The spring elements bias the hammer 47 using the biasing force of both the first coil spring 49 and the second coil spring 50 while the axial position (stroke) of the hammer 47 is within second region R2. Therefore, the degree of change in the biasing force (the elastic force) with respect to the stroke becomes larger while the hammer 47 is in second region R2 as compared to while the hammer 47 is in first region R1.
Next, the cam element will be described.
As described above, the cam element includes the spindle grooves 8D, the balls 48, and the hammer grooves 47B. The depth of each of the spindle grooves 8D differs in accordance with the location thereof in the axial direction of the spindle 8. That is, as shown in
In first region R1, in which the stroke is shorter, each of the balls 48 is located at the front (forwardmost) portion of the corresponding spindle groove 8D. Because the front portions of the spindle grooves 8D are shallower than the rear portions of the spindle grooves 8D, the balls 48 are most distant from rotational axis AX. That is, because the contact points between the balls 48 and the inner surfaces of the hammer grooves 47B are relatively farther apart (away) from rotational axis AX in first region R1, in which the stroke is shorter, the required torque, which is the torque needed for rotating the hammer 47, is reduced. Therefore, because the required torque needed to rotate and to move the hammer 47 rearward is lower in first region R1, it is easier for the hammer 47 to reach the impact-start location.
In second region R2, in which the stroke is longer, each of the balls 48 is located at one of the rear portions of the corresponding spindle groove 8D. Because the rear portions of the spindle grooves 8D are deeper than the front portions of the spindle grooves 8D, the balls 48 will be closest to rotational axis AX. That is, because the contact points between the balls 48 and the inner surfaces of the hammer grooves 47B are relatively closer to rotational axis AX in second region R2, in which the stroke is longer, the required torque, which is the torque needed for rotating the hammer 47, is increased. Therefore, because the required torque can effectively be made larger, the advance of the hammer 47 can be started quickly in second region R2.
Angle α (see
According to such a configuration, with regard to the cam element, the degree of change in the required torque with respect to the stroke in first region R1 is made to be smaller than the degree of change in the required torque with respect to the stroke in second region R2.
In the example shown in
In the alternative example shown in
In this manner, the spring element and the cam element both have structures that nonlinearly increase the required torque to suppress (reduce) the increase in the required torque in first region R1 and boost the increase in the required torque in second region R2. In the second embodiment, because in first region R1 the hammer 47 can move rearward up to the impact-start location (location Q2) with an even lower torque owing to the synergistic effect of the spring element and the cam element, the impact-start torque decreases. In the second embodiment, the maximum value of the required torque (that is, the impact-start torque) in first region R1 is 1,300 N·mm or less. In the second embodiment, the hammer 47 is rotated at a higher speed with an even larger torque in second region R2 owing to the biasing force of the spring element and the shapes of the spindle grooves 8D. The impact tool 1 achieves a high maximum fastening torque by the hammer 47, which is being rotated at a higher speed, impacting the anvil 10. In the second embodiment, the maximum fastening torque of the impact tool 1 is 230 Nom or more and 350 N·m or less.
As explained above, in the second embodiment, the impact tool 1 comprises: the motor 6 supplied with a voltage of 18 V or more; the spindle 8 rotated by the motor 6; the spindle grooves 8D formed in the spindle 8; the balls 48 held in the spindle grooves 8D; the hammer 47 supported by the spindle 8 via the balls 48; the springs (the first coil spring 49 and the second coil spring 50), which bias the hammer 47 forward; and the anvil 10, which is impacted in the rotational direction by the hammer 47. The maximum fastening torque of the impact tool 1 is higher than 220 N·m. The impact-start torque, which is the torque acting on the anvil 10 when the hammer 47 starts to impact the anvil 10, is 1,300 N·mm or less.
In the above-mentioned configuration, the impact-start torque can be suppressed to 1,300 N·mm or less even though the maximum fastening torque is raised (increased) to higher than 220 N·m. Because the impact-start torque is lowered, the timing of the start of impact of the hammer 47 on the anvil 10 is earlier. That is, the amount of screw rotation from the screw-tightening start time until the impact time when the anvil 10 is impacted by the hammer 47 becomes smaller. Because the timing of the start of impact of the hammer 47 on the anvil 10 is earlier, the cam-out phenomenon is less likely to occur.
In the second embodiment, the impact tool 1 comprises the battery-mounting part 13, onto which the battery pack 25 is mounted. The rated voltage of the battery pack 25 is 18 V or more.
In the above-mentioned configuration, because the rated voltage of the battery pack 25 is 18 V or more, the impact tool 1 can have a high output (torque).
In the second embodiment, the depth of each of the spindle grooves 8D differs in accordance with the location thereof in the axial direction of the spindle 8.
Therefore, the occurrence of the cam-out phenomenon can be curtailed.
In the second embodiment, the maximum fastening torque of the impact tool 1 is 230 N·m or more.
In the above-mentioned configuration, by reducing the impact-start torque to 1,300 N·mm or less even though the maximum fastening torque is made to be 230 N·m or more, both a high maximum fastening torque and curtailment of the cam-out phenomenon can be achieved.
In the second embodiment, with regard to the impact tool 1, the value calculated by dividing the maximum fastening torque (N·mm) by the impact-start torque (N·mm) is greater than 165.
Generally speaking, the higher the maximum fastening torque is raised, the higher the impact-start torque tends to become. Furthermore, the higher the impact-start torque becomes and the later the timing of the start of impact becomes, the more likely the cam-out phenomenon tends to occur. Because the value calculated by dividing the maximum fastening torque by the impact-start torque (a) becomes larger the higher the maximum fastening torque and (b) becomes larger the lower the impact-start torque, this value serves an index that signifies (indicates) the smallness (size) of the impact-start torque relative to the maximum fastening torque (i.e., the degree of curtailment of the cam-out phenomenon). In the above-mentioned configuration, because the value calculated by dividing the maximum fastening torque by the impact-start torque is greater than 165, it becomes possible to more effectively curtail the cam-out phenomenon as compared to changing the magnitude of the maximum fastening torque.
In the second embodiment, the impact tool 1 comprises: the motor 6; the spindle 8 rotated by the motor 6; the cam element that includes the spindle grooves 8D formed in the spindle 8 and the balls 48 held in the spindle grooves 8D; the hammer 47 supported on the spindle 8 via the balls 48; spring elements, which bias the hammer 47 forward; and the anvil 10, which is impacted in the rotational direction by the hammer 47. At the impact time, after the hammer 47 moves rearward against the biasing force of the spring elements 49, 50 while rotating due to the rotational torque imparted from the motor 6 via the spindle 8 and the cam element, the hammer 47 advances while rotating owing to the elastic force of the compressed spring elements 49, 50 and the cam element. The cam element and the spring elements are preferably configured (designed) such that nonlinearly increases the required torque, which is the torque needed to rotate the hammer 47, nonlinearly increases to suppress (reduce) the increase in the required torque in first region R1 in which the rearward stroke of the hammer 47 is shorter, and to boost the increase in the required torque in second region R2 in which the stroke of the hammer 47 is longer.
In the above-mentioned configuration, the impact-start torque is reduced while raising the maximum fastening torque, and thereby a high maximum fastening torque and curtailment of cam-out can both be achieved. That is, the output of the motor 6 is raised to raise the rotational speed of the hammer 47 in order to raise the maximum fastening torque. If the rotational speed of the hammer 47 is higher, the time until the hammer 47 arrives at an angle at which it impacts the anvil 10 will become shorter. The biasing force of the spring elements is made to be larger and the hammer 47 is caused to advance quickly in order to ensure that the advance of the hammer 47 coincides with the timing at which the hammer 47 impacts the anvil 10. If the biasing force of the spring elements becomes larger, the impact-start torque will become higher because the required torque, which is the torque needed to rotate and to move the hammer 47 rearward, increases.
Thus, normally (i.e. in the past), when the maximum fastening torque is (was) raised, the impact-start torque becomes (became) higher. However, as explained above, the cam element and the spring elements of the present teachings cause the change in the required torque, which is stroke dependent, to be nonlinear; therefore, by suppressing (reducing) the increase in the required torque when the stroke is shorter, the rearward movement of the hammer 47 can be made easier and the impact-start torque can be reduced. Furthermore, by boosting the increase in the required torque when the stroke is longer, the hammer 47 advances more quickly and can impact the anvil 10 with greater force, even through the rotational speed of the hammer 47 is increased; thereby, the maximum fastening torque can be increased. As a result, the occurrence of the cam-out phenomenon can be curtailed even though the maximum fastening torque has been increased.
In the second embodiment, first region R1 is defined as extending from the location at which the rearward stroke of the hammer 47 becomes the minimum up to the location at which the hammer 47 and the anvil 10 are no longer in contact; and second region R2 is defined as the prescribed region, which includes the location at which the stroke is longer than the stroke in first region R1 and the stroke becomes the maximum. The degree of change in the required torque with respect to the stroke in first region R1 is smaller than the degree of change in the required torque with respect to the stroke in second region R2.
In the above-mentioned configuration, because the degree of increase in the required torque, which is stroke dependent, is suppressed (reduced) in first region R1, the hammer 47 can be moved with a smaller required torque to location (axial position) Q2 where the hammer 47 and the anvil 10 are no longer in contact. Because the hammer 47 becomes rotatable and impact begins owing to the hammer 47 and the anvil 10 no longer being in contact, the impact-start torque can be lowered by reducing the required torque in first region R1. However, if the degree of increase in the required torque, which is stroke dependent, were to (hypothetically) remain suppressed in second region R2, then even when the hammer 47 reaches its maximum stroke, the required torque would not rise sufficiently and the maximum fastening torque could not be raised. Therefore, according to the present teachings, a higher maximum fastening torque can be achieved by making the degree of increase in the required torque, which is stroke dependent, larger in second region R2.
In the second embodiment, the maximum value of the required torque in first region R1 is 1,300 N·mm or less.
Therefore, the impact-start torque can be effectively lowered.
Furthermore, in the second embodiment, the maximum fastening torque of the impact tool 1 is 230 Nom or more and 350 N·m or less.
As a result, a higher maximum fastening torque can be achieved while suppressing the impact-start torque.
In the second embodiment, the depth of the rear portion of each of the spindle grooves 8D is deeper than the depth of the front portion of the corresponding spindle groove 8D.
In the above-mentioned configuration, the rotational radius of each of the balls 48 disposed in the corresponding spindle groove 8D becomes relatively large at the front portion of the corresponding spindle groove 8D and becomes relatively small at the rear portions of the corresponding spindle groove 8D. Therefore, at the front portion of each of the spindle grooves 8D, among the forces acting between the spindle 8 and the balls 48, the force component in the rotational direction becomes relatively small, and the force component in the axial direction becomes relatively large. Therefore, because the hammer can be caused to stroke (move) rearward with an even lower torque, this contributes to reducing the impact-start torque. On the other hand, at the rear portions of each of the spindle grooves 8D, among the forces acting between the spindle 8 and the balls 48, the force component in the rotational direction becomes relatively large, and thus the rotational speed of the hammer can be increased when the hammer is moving forward while rotating. Because the rotational speed of the hammer can be increased, this contributes to increasing the maximum fastening torque in accordance with the present teachings.
In the second embodiment, the hammer 47 has the hammer grooves 47B, in each of which at least a portion of the corresponding ball 48 is disposed. Tangent vector Vr is defined as passing through the contact point 47F between the ball 48 and the hammer groove 47B in a direction parallel to a tangent of a circle about rotational axis AX of the spindle 8, and load vector Vp is defined as having the direction that a load is transmitted from the ball 48 to the contact point 47F with the hammer 47. Angle α formed between tangent vector Vr and load vector Vp at each of the rear portions of the spindle grooves 8D is smaller than angle α formed between tangent vector Vr and load vector Vp at each of the front portions of the spindle grooves 8D.
In the above-mentioned configuration, this means that the smaller that angle α formed between tangent vector Vr and load vector Vp becomes, the more that the force from the spindle 8 acting on the balls 48 acts exclusively in the rotational direction. Angle α becomes relatively small at the front portion of each of the spindle grooves 8D, and among the forces from the spindle 8 acting on the balls 48, the force component in the axial direction becomes relatively larger. Therefore, because the hammer can be caused to stroke (move) rearward with an even lower torque, this contributes to reducing the impact-start torque. However, at the rear portions of each of the spindle grooves 8D, among the forces from the spindle 8 acting on the balls 48, the force component in the rotational direction becomes relatively larger. Therefore, because the rotational speed of the hammer when the hammer moves forward while rotating can be increased, this contributes to increasing the maximum fastening torque in accordance with the present teachings.
In the second embodiment, the spring elements include the first coil spring 49 and the second coil spring 50. The first coil spring 49 continuously applies to the hammer 47 the biasing force that urges the hammer 47 to move forward. The second coil spring 50 applies the biasing force to the hammer 47 only after the hammer 47 has moved rearward from its forwardmost axial position (location Q1).
In the above-mentioned configuration, only the first coil spring 49 biases the hammer 47 when the stroke of the hammer 47 is shorter; therefore, the rearward movement of the hammer 47 can be made easier and the impact-start torque can be reduced. On the other hand, both the first coil spring 49 and the second coil spring 50 bias the hammer 47 when the stroke of the hammer 47 is longer which makes it easier for the hammer 47 to advance; thereby, the hammer 47 can advance more quickly and can impact the anvil appropriately, even if the rotational speed of the hammer 47 is raised (increased). Therefore, the maximum fastening torque can be raised.
Next, working examples will be described. Measurements of test data were performed with Working Example 1 and Working Example 2, as well as with Comparative Example 1 through Comparative Example 7, which are commercially-available impact tools; both the examples and specification values are compared. Working Example 1 corresponds to the impact tool according to the above-described first embodiment. Working Example 2 corresponds to the impact tool according to the above-described second embodiment. The maximum fastening torque of Working Example 2 is larger than that of Working Example 1. Consequently, Working Example 1 is compared with Comparative Example 1 through Comparative Example 3, which, among Comparative Example 1 through Comparative Example 7, is a group having relatively small maximum fastening torques. Working Example 2 is compared with Comparative Example 4 through Comparative Example 7, which, among Comparative Example 1 through Comparative Example 7, is a group having relatively large maximum fastening torques.
Table 1 shows Working Example 1 and Comparative Example 1 through Comparative Example 3.
The rated voltage of the battery of the impact tool according to Working Example 1 was 36 V. Therefore, the rated voltage of the battery pack 25 of the impact tool according to Working Example 1 was 18 V or more. That is, a voltage of 18 V or more was supplied to the motor 6 of the impact tool according to Working Example 1.
“No-Load Rotational Speed” means the number of rotations per minute (rpm) of the hammer 47 in a no-load state, in which no load is acting on the anvil 10. The no-load rotational speed of the impact tool according to Working Example 1 was 3,700 rpm.
“Impact-Start Torque” is the average value of the impact-start torque that was measured in an impact test performed a plurality of times. The impact start torque is expressed with three significant digits, with the last digit being rounded off.
The impact-start torque of the impact tool according to Working Example 1 was 1,000 N·mm (1.00 N·m). Therefore, with regard to the impact tool according to Working Example 1, the impact-start torque, which is the torque acting on the anvil 10 when the hammer 47 starts to impact the anvil 10, was 1,100 N·mm or less.
In contrast, the impact-start torque in every one of Comparative Example 1, Comparative Example 2, and Comparative Example 3 was higher than 1,100 N·mm.
The maximum fastening torque of the impact tool according to Working Example 1 was 220 N·m. Accordingly, the maximum fastening torque of the impact tool according to Working Example 1 was greater than 200 N·m. In contrast, the maximum fastening torque in Comparative Example 1 through Comparative Example 3 was 200 N·m or less.
Thus, even in Comparative Example 2, which had the closest maximum fastening torque to Working Example 1, the impact-start torque was 1,220 N·mm. In contrast, in Working Example 1, because the impact-start torque could be made to be 1,100 N·mm or less even though the maximum fastening torque was higher than 200 N·m, both a higher maximum fastening torque and curtailment of cam-out (a low impact-start torque) were achieved even within a product family having an equivalent torque level.
With regard to the impact tool according to Working Example 1, the value calculated by dividing the maximum fastening torque (N·mm) by the impact-start torque was 220. Accordingly, with regard to the impact tool according to Working Example 1, the value calculated by dividing the maximum fastening torque by the impact-start torque was greater than 165. In contrast, in every one of Comparative Example 1, Comparative Example 2, and Comparative Example 3, the value calculated by dividing the maximum fastening torque by the impact-start torque was less than 165. Therefore, the impact tool according to Working Example 1 achieved both a higher maximum fastening torque and a lower impact-start torque.
“Impacts Per Minute” means the number of impacts per minute (IPM) of the hammer 47. The impacts per minute of the impact tool according to Working Example 1 was 4,600. The spring constant in the impact tool according to Working Example 1 was 36.4 N/mm.
“Mounted Load” means the magnitude of the biasing force (the elastic force) that the spring elements (the springs) apply to the hammer 47 in the state in which the impact tool is not in operation, that is, the magnitude of the preload. The mounted load of the impact tool according to Working Example 1 was 120 (N). Thus, with regard to the impact tool according to Working Example 1, the mounted load of the spring elements, which bias the hammer, was 150 (N) or less.
In contrast, the mounted load of the spring element(s), which bias(es) the hammer in every one of Comparative Example 1, Comparative Example 2, and Comparative Example 3, was greater than 150 (N). Therefore, by having spring elements with a small mounted load in accordance with the present teachings, it was possible to effectively reduce the impact-start torque in the impact tool of Working Example 1.
“Hammer Inertia” means the magnitude of the moment of inertia of the hammer 47 that the impact tool has. The hammer inertia in the impact tool according to Working Example 1 was 32.0 (kg·mm2).
“Impact-Start Stroke” means the magnitude of the stroke of the hammer 47 up to the impact-start location, at which the hammer 47 and the anvil 10 are no longer in contact. Additionally, in the present specification, the stroke of the hammer 47 means the amount of displacement rearward from location Q1 (see
The product overall length of the impact tool according to Working Example 1 was 102.0 (mm). The impact-portion width of the impact tool according to Working Example 1 was 58.00 (mm). It is noted that “Impact-Portion Width” means the width dimension of the portion that is the impact mechanism 9.
The product mass of the impact tool according to Working Example 1 was 0.85 (kg). It is noted that “Product Mass” means the mass when the battery pack 25 is removed from the body of the impact tool. Because the size and the mass of the battery pack 25 can change depending on the battery capacity or the battery voltage, for the purpose of comparison between the working examples and the comparative examples, the product mass is assumed to be the mass of the product with the mass of the battery pack excluded.
“Impact-Time Rotational Amount” in Table 1 means the amount of rotation (rad) of the hammer per impact when the impact tool makes impact. The impact-time rotational amount was 1.57 for the impact tool according to Working Example 1, as well as in Comparative Example 1 and Comparative Example 3. It was 2.09 in Comparative Example 2.
Table 2 shows specifications for Working Example 2 and Comparative Example 4 through Comparative Example 7.
The rated voltage of the battery of the impact tool according to Working Example 2 was 36 V. Therefore, the rated voltage of the battery pack 25 of the impact tool according to Working Example 2 was 18 V or more. That is, a voltage of 18 V or more was supplied to the motor 6 of the impact tool according to Working Example 2.
The no-load rotational speed of the impact tool according to Working Example 2 was 3,900 rpm.
The impact-start torque of the impact tool according to Working Example 2 was 1,200 N·mm (1.20 N·m). Therefore, with regard to the impact tool according to Working Example 2, the impact-start torque, which is the torque acting on the anvil 10 when the hammer 47 starts to impact the anvil 10, was 1,300 N·mm or less.
The maximum fastening torque of the impact tool according to Working Example 2 was 230 N·m. Therefore, the maximum fastening torque of the impact tool according to Working Example 2 was higher than 220 N·m. The maximum fastening torque of the impact tool according to Working Example 2 was 230 Nom or more. The maximum fastening torque of the impact tool according to Working Example 2 was 230 N·m or more and 350 N·m or less.
Thus, with regard to the impact tool according to Working Example 2, the maximum fastening torque was higher than 220 Nm, and the impact-start torque was 1,300 N·mm or less. On the other hand, in Comparative Examples 4-7, the maximum fastening torque was 220 N·m or less, or the impact-start torque was higher than 1,300 N·mm.
In Comparative Example 7, which had the closest maximum fastening torque to Working Example 2, the impact-start torque became 1,730 N·mm. In contrast, according to Working Example 2, because the impact-start torque could be made to be 1,300 N·mm or less even though the maximum fastening torque was 230 N·m, which is higher than 220 N·m, both a higher maximum fastening torque and curtailment of cam-out (a low impact-start torque) were achieved even within a product family having an equivalent torque level.
With regard to the impact tool according to Working Example 2, the value calculated by dividing the maximum fastening torque (N·mm) by the impact-start torque (Nom) was 192. Accordingly, with regard to the impact tool according to Working Example 2, the value calculated by dividing the maximum fastening torque by the impact-start torque was greater than 180.
In contrast, in every one of Comparative Example 4 through Comparative Example 7, the value calculated by dividing the maximum fastening torque by the impact-start torque was less than 180. Therefore, the impact tool according to Working Example 2 achieved both a higher maximum fastening torque and a lower impact-start torque.
The impacts per minute of the impact tool according to Working Example 2 was 4,700. The spring constant in the impact tool according to Working Example 2 was 60.4 N/mm. The mounted load of the impact tool according to Working Example 2 was 120 (N). Thus, with regard to the impact tool according to Working Example 2, the mounted load of the spring elements, which bias the hammer, was 150 (N) or less.
The hammer inertia in the impact tool according to Working Example 2 was 33.9 (kg·mm2). The impact-start stroke in the impact tool according to Working Example 2 was 4.1 (mm).
The product overall length of the impact tool according to Working Example 1 was 102.0 (mm). The impact-portion width in the impact tool according to Working Example 2 was 58.00 (mm). The product mass of the impact tool according to Working Example 2 was 0.87 (kg). The impact-time rotational amount was 1.57 (rad) for the impact tool according to Working Example 2, and in Comparative Example 1 through Comparative Example 4.
Next, performance-evaluation indexes will be described with regard to the working examples and comparative examples described above. The performance-evaluation indexes are index values, which are computed on the basis of the specification values or the measurement values, for the purpose of evaluating specific performances of the impact tool. The present specification provides three index values from the viewpoint of the maximum fastening torque and the curtailment of cam-out.
First index value PIa is expressed by Equation (1).
Here, Nipm is the impacts per minute (IPM). I is the moment of inertia (kg·mm2) of the hammer (hammer inertia). θrot is the amount of rotation (rad) that the hammer rotates per impact. First index value PIa is an index value that expresses an effect on the maximum fastening torque, in which the larger first index value PIa, the more effectively the maximum fastening torque can be increased. The larger the impacts per minute (Nipm), the more easily the transmitted amount of the impact energy can be increased by increasing the number of impacts per unit time. In addition, the larger the hammer inertia, the less the rotation tends to fluctuate; therefore, the larger the moment of inertia I, the more easily the energy can be transmitted to the object that is contacted. The larger the amount of rotation θrot per impact, the larger the rotational energy of the hammer becomes, and thus the more easily the maximum fastening torque can be increased. Accordingly, first index value PIa, which is expressed in Equation (1) based on these parameters, is an index value that expresses the effect on the maximum fastening torque.
Second index value PIb is expressed by Equation (2).
Here, Nrpm is the no-load rotational speed (rpm). Fint is the mounted load (N) of the spring element. L is the impact-start stroke (mm). Second index value PIb is an index of the ease of making impact; the evaluation was that the larger the second index value PIb, the more easily the occurrence of cam-out can be curtailed. Second index value PIb has dimensionless coefficients, that is, moment of inertia I, mounting load Fint of the spring, and impact-start stroke L; these coefficients become indexes of the ease of making impact when the rotational speed increases. That is, the larger the moment of inertia I, the easier the rotational acceleration is converted into the stroke (hammer movement), and the less cam-out tends to occur. The smaller the mounting load Fint, the easier it is for the hammer to stroke (move), and the less cam-out tends to occur. The smaller the impact-start stroke L, the easier it is for the hammer to stroke (Move), and the less cam-out tends to occur. Additionally, the no-load rotational speed (Nrpm) expresses the increase in rotational speed per unit time and can be treated as the acceleration. The larger the no-load rotational speed (Nrpm), the easier it is for the hammer to stroke (move), and therefore the less cam-out tends to occur.
Third index value PIc is expressed by Equation (3).
Here, TM is the maximum fastening torque (N·mm) and TS is the impact-start torque (N·mm). Third index value PIc is the ratio of the maximum fastening torque to the impact-start torque and directly indicates how high the cam-out curtailment performance is (i.e., how low the impact-start torque is) in relation to the magnitude of the maximum fastening torque.
Table 3 shows the index values calculated with regard to the working examples and the comparative examples.
First index value PIa was 231.22 for the impact tool according to Working Example 1 and was 250.27 for the impact tool according to Working Example 2. With regard to the impact tools according to the embodiments, first index value PIa was 200 or more, preferably was 230 or more, and more preferably was 240 or more. As a result, the maximum fastening torque could be effectively increased.
Second index value PIb was 0.944 for the impact tool according to Working Example 1 and was 1.024 for the impact tool according to Working Example 2. With regard to the impact tools according to the embodiments, second index value PIb was 0.8 or more, preferably was 0.9 or more, and more preferably was 1.0 or more. As a result, the occurrence of cam-out could be effectively curtailed.
Third index value PIc is the same as the numerical value shown in Table 1 and Table 2 for “Maximum Fastening Torque/Impact-Start Torque”; third index value PIc was 220 for the impact tool according to Working Example 1 and was 192 for the impact tool according to Working Example 2. With regard to the impact tools according to the embodiments, third index value PIc was 165 or more, preferably was 180 or more, and more preferably was 200 or more. As a result, the cam-out curtailment performance could be raised in relation to the magnitude of the maximum fastening torque.
Representative, non-limiting examples of the present invention were described above in detail with reference to the attached drawings. This detailed description is merely intended to teach a person of skill in the art further details for practicing preferred aspects of the present teachings and is not intended to limit the scope of the invention. Furthermore, each of the additional features and teachings disclosed above may be utilized separately or in conjunction with other features and teachings to provide improved impact tools and similar power tools.
Moreover, combinations of features and steps disclosed in the above detailed description may not be necessary to practice the invention in the broadest sense, and are instead taught merely to particularly describe representative examples of the invention. Furthermore, various features of the above-described representative examples, as well as the various independent and dependent claims below, may be combined in ways that are not specifically and explicitly enumerated in order to provide additional useful embodiments of the present teachings.
All features disclosed in the description and/or the claims are intended to be disclosed separately and independently from each other for the purpose of original written disclosure, as well as for the purpose of restricting the claimed subject matter, independent of the compositions of the features in the embodiments and/or the claims. In addition, all value ranges or indications of groups of entities are intended to disclose every possible intermediate value or intermediate entity for the purpose of original written disclosure, as well as for the purpose of restricting the claimed subject matter.
| Number | Date | Country | Kind |
|---|---|---|---|
| 2023-172662 | Oct 2023 | JP | national |
| 2024-104127 | Jun 2024 | JP | national |
