BACKGROUND
The present disclosure relates to power tools, and more particularly to impacting ratchet tools.
SUMMARY
The present disclosure provides, in one aspect, an impact tool including a housing, a motor disposed within the housing, the motor including an output shaft rotatable about a first axis, and an impact mechanism. The impact mechanism includes a camshaft driven by the motor to rotate about the first axis, an anvil, and a hammer configured to reciprocate along the camshaft to impart rotational impacts to the anvil in response to rotation of the camshaft in a first direction about the first axis in a first operating mode of the impact tool and to directly drive the anvil in response to rotation of the camshaft in a second direction about the first axis opposite the first direction in a second operating mode of the impact tool. The impact tool also includes a crankshaft coupled to the anvil for co-rotation with the anvil, a yoke driven by the crankshaft to reciprocate about a second axis perpendicular to the first axis in response to rotation of the crankshaft, an output drive configured to receive a tool element, and a pawl configured to selectively couple the output drive to the yoke for co-rotation with the yoke in a first rotational locking direction about the second axis and to permit the yoke to rotate relative to the pawl in a second rotational locking direction about the second axis opposite the first rotational locking direction.
In another aspect, the disclosure provides an impact mechanism for an impact tool. The impact mechanism includes a camshaft, an anvil, a first hammer, and a second hammer. The camshaft is configured to rotate about an axis in response to receiving a motor output. The anvil includes anvil lugs. The first hammer is mounted to the camshaft and includes hammer lugs that are configured to rotationally impact the anvil lugs. The second hammer is sleeved onto the first hammer. One of the first hammer and the second hammer includes a plurality of key portions and the other of the first hammer and the second hammer includes a plurality of keyways, each of the key portions is configured to slide along a corresponding one of the keyways to allow for relative axial movement between the first hammer and the second hammer.
In another aspect, the disclosure provides an impact mechanism for an impact tool. The impact mechanism includes a camshaft configured to rotate about an axis in response to receiving a motor output. The anvil includes anvil lugs. The first hammer is mounted to the camshaft and includes hammer lugs that are configured to rotationally impact the anvil lugs. The second hammer is sleeved on the first hammer. One of the first hammer and the second hammer includes a protrusion that extends from the one of the first hammer and the second hammer toward the other of the first hammer and the second hammer to rotationally couple the first hammer and the second hammer together.
Other features and aspects of the disclosure will become apparent by consideration of the following detailed description and accompanying drawings.
BRIEF DESCRIPTION OF DRAWINGS
FIG. 1 is a perspective view of an impact tool according to an embodiment of the present disclosure.
FIG. 2 is a cross-sectional view of the impact tool of FIG. 1.
FIG. 3A is a perspective view of an impact mechanism and a ratchet mechanism of the impact tool of FIG. 1.
FIG. 3B is a perspective view of an anvil of the impact mechanism and a crankshaft of the ratchet mechanism of FIG. 3A.
FIG. 4 is a cross-sectional view of the impact mechanism and the ratchet mechanism of FIG. 3A.
FIG. 5 is a perspective view of the impact mechanism of FIG. 3A with portions removed.
FIG. 6 is a cross-sectional view of the impact mechanism of FIG. 5.
FIG. 7 is a front view of the impact mechanism of FIG. 5.
FIG. 8 is a front view of a camshaft of the impact mechanism of FIG. 3A.
FIG. 9 is a side view of the camshaft of FIG. 8 with portions removed.
FIG. 10 is a cross-sectional view of a ratchet mechanism of an impact tool according to another embodiment.
FIG. 11 is a cross-sectional view of a portion of an impact tool including an impact mechanism according to another embodiment.
FIG. 12 is a perspective view of a portion of the impact mechanism of FIG. 11.
FIG. 13 is a perspective view of an impact mechanism for an impact tool according to another embodiment.
FIG. 14A is a perspective view of a portion of the impact mechanism of FIG. 13 in a first orientation.
FIG. 14B is a perspective view of a portion of the impact mechanism of FIG. 13 in a second orientation.
FIG. 15A is a schematic illustration of an ellipse imposed over a conventional hammer for the impact mechanism of FIG. 13.
FIG. 15B is a front view of a double hammer assembly for the impact mechanism of FIG. 13.
FIG. 16 is a perspective view of an impact mechanism for an impact tool according to another embodiment.
FIG. 17 is a cross-sectional view of the impact mechanism of FIG. 16 taken along line 17-17.
FIG. 18 is a front view of the impact mechanism of FIG. 16.
FIG. 19 is a rear view of a portion of the impact mechanism of FIG. 16.
FIG. 20 is a cross-sectional view of the impact mechanism of FIG. 16 taken along line 20-20.
FIG. 21 is a cross-sectional view of the impact mechanism of FIG. 16 taken along line 21-21.
FIG. 22 is a zoomed-in view of a portion of the impact mechanism of FIG. 16.
Before any embodiments of the disclosure are explained in detail, it is to be understood that the disclosure is not limited in its application to the details of construction and the arrangement of components set forth in the following description or illustrated in the following drawings. The disclosure is capable of other embodiments and of being practiced or of being carried out in various ways. Also, it is to be understood that the phraseology and terminology used herein is for the purpose of description and should not be regarded as limiting.
DETAILED DESCRIPTION
With reference to FIGS. 1 and 2, an impact tool 10 in accordance with an embodiment of the disclosure includes a housing 14 and a secondary housing or yoke housing 18 coupled to the housing 14. The impact tool 10 further includes a motor 22 (FIG. 2) supported within the housing 14 and is preferably a brushless DC motor. The motor 22 has an output shaft 26 rotatable about a first axis 30 and configured to provide torque to an output drive 34 rotatably supported by the yoke housing 18 for rotation about a second axis 38 oriented perpendicular to the first axis 30. The output drive 34 includes a square drive interface 40, preferably with a nominal size (e.g., ⅜″, ½″, ¾″, 1″, etc.). The output drive 34 is configured to be coupled to a tool element (not shown) that performs work on a workpiece (e.g., a fastener, bit, or the like).
In the illustrated embodiment, the impact tool 10 includes a battery pack (not shown) received by a battery receptacle 46. The battery receptacle 46 electrically connects the battery pack to the motor 22 (via suitable electrical and electronic components, such as a PCBA containing MOSFETs, IGBTs, or the like). A trigger switch 50 is located on the housing 14, and actuation of the trigger switch 50 energizes the motor 22. The battery pack may be a 12-volt power tool battery pack that includes three lithium-ion battery cells. Alternatively, the battery pack may include fewer or more battery cells to yield any of a number of different output voltages (e.g., 14.4 volts, 18 volts, etc.). Additionally or alternatively, the battery cells may include chemistries other than lithium-ion such as, for example, nickel cadmium, nickel metal-hydride, or the like. Alternatively, the impact tool 10 may include an electrical cord for powering the motor 22 with a remote electrical source (e.g., a wall outlet), or the impact tool 10 may be pneumatic in yet other embodiments.
With reference to FIGS. 2-4, the illustrated impact tool 10 is an impacting ratchet tool that includes an impact mechanism 54 and a ratchet mechanism 58. The impact mechanism 54 includes a gear assembly in the form of a planetary gearset 60 operably coupled to the motor 22 to provide a speed reduction/torque increase from the motor 22. The planetary gearset 60 includes a pinion 61 coupled for co-rotation with the output shaft 26 of the motor 22, a plurality of planet gears 62 meshed with the pinion 61, and a ring gear 63 meshed with the planet gears 62. The ring gear 63 of the planetary gearset 60 is rotationally fixed within a gear case 64 fixed between the housing 14 and the yoke housing 18.
The impact mechanism 54 further includes a camshaft 66 driven by the output of the planetary gearset 60 (e.g., the camshaft 66 may define a planet carrier 70 of the planetary gearset 60), an inner hammer 74 coupled to the camshaft 66, an outer hammer 78 surrounding the inner hammer 74 and having a rear end rotationally supported by the camshaft 66, and an anvil 82 having opposed, radially outward extending anvil lugs 86, 90 (FIGS. 3A and 3B) that receive impacts from corresponding hammer lugs 94, 98 (FIG. 5) on the inner hammer 74. The illustrated impact mechanism 54 also includes a compression spring 80 (shown in FIG. 6) disposed within the outer hammer 78 and configured to bias the inner hammer 74 toward the anvil 82.
With continued reference to FIGS. 2-4, the ratchet mechanism 58 includes a crankshaft 106 rotatably coupled to the camshaft 66 and the anvil 82 of the impact mechanism 54. The crankshaft 106 has an eccentric member 110 arranged within a drive bushing 114. The ratchet mechanism 58 further includes a yoke 118 through which the output drive 34 extends. The yoke 118 has a recess 122 in which the drive bushing 114 is arranged. As explained in further detail below, when the crankshaft 106 is rotated, the drive bushing 114 pivots the yoke 118 in a reciprocating manner to drive the output drive 34.
The illustrated ratchet mechanism 58 includes a pawl 126 and a forward/reverse switch for the ratchet mechanism 58 in the form of a rotational member 130 (FIG. 1). The pawl 126 is provided within the yoke 118 and pivotably secured by a pin 134 (FIG. 4). The pin 134 is coupled to the rotational member 130, which has a gripping actuator 138 that is accessible through the yoke housing 18. The gripping actuator 138 can be used to rotate the rotational member 130, and thus, the pawl 126, between a first position corresponding to a first rotational locking direction 140a of the output drive 34 and a second position corresponding to a second rotational locking direction 140b of the output drive 34. The rotational member 130 is concentric with the output drive 34 in the illustrated embodiment; however, the rotational member 130 may be offset from the output drive 34 in other embodiments. In yet other embodiments, other types of forward/reverse switches may be used to move the pawl 126.
With continued reference to FIG. 4, the illustrated pawl 126 includes an angled first end 142 including teeth and an angled second end 146 including teeth. The first end 142 and the second end 146 of the pawl 126 are configured to engage inner teeth 150 of the yoke 118. When the rotational member 130 is rotated to arrange the pawl 126 in the first position, the first end 142 of the pawl 126 meshes with the inner teeth 150 of the yoke 118. The first position thereby prevents the output drive 34 from rotating relative to the yoke 118 in the first rotational locking direction 140a. In other words, the pawl 126 couples the output drive 34 for co-rotation with the yoke 118 in the first rotational locking direction 140a. The teeth on the first end 142 of the pawl 126 and/or the teeth 150 on the yoke 118 are angled to allow the teeth to slip past each other in the opposite direction 140b, thereby permitting the yoke 118 to “ratchet” and rotate relative to the output drive 34 in the second rotational locking direction 140b. When the rotational member 130 is rotated to arrange the pawl 126 in the second position (shown in FIG. 4), the second end 146 of the pawl 126 meshes with the inner teeth 150 of the yoke 118. The pawl 126 in the second position thereby prevents the output drive 34 from rotating relative to the yoke 118 in the second rotational locking direction 140b. In other words, the pawl 126 couples the output drive 34 for co-rotation with the yoke 118 in the second rotational locking direction 140b. The teeth on the second end 146 of the pawl 126 and/or the teeth 150 on the yoke 118 are angled to allow the teeth to slip past each other in the first rotational locking direction 140a, thereby permitting the yoke 118 to “ratchet” and rotate relative to the output drive 34 in the first rotational locking direction 140a.
With reference to FIGS. 5-7, a portion of the impact mechanism 54 is illustrated. The outer hammer 78 and inner hammer 74 form a two-piece or double hammer assembly able to provide increased rotating mass and, therefore, increased impact energy to the anvil 82 compared to the inner hammer 74 alone. In addition, the outer diameter of the hammer assembly may be less than a single hammer having an equivalent mass, permitting the impact tool 10 to be more compact.
In the illustrated embodiment, the outer hammer 78 is mounted to a portion of the camshaft 66 such that the outer hammer 78 is radially supported by the camshaft 66. Additionally, the outer hammer 78 is sleeved onto the inner hammer 74 (e.g., by cooperating projections, splines, pins, or any other suitable arrangement) such that the outer hammer 78 is coupled for co-rotation with the inner hammer 74, but the inner hammer 74 is able to move axially relative to the outer hammer 78. In particular, the illustrated inner hammer 74 is provided with keyways 158 configured to receive key portions 162 formed along the outer hammer 78. The inner hammer 74 functions similar to a hammer of a conventional impact mechanism, however the inner hammer 74 is provided with a reduced outer diameter to thereby reduce a size of the impact tool 10. As such, a mass of the inner hammer 74 is also reduced. Rotatably coupling the outer hammer 78 to the inner hammer 74 adds a substantial amount of mass to the inner hammer 74.
With reference to FIGS. 8 and 9, the camshaft 66 of the impact mechanism 54 is illustrated. The camshaft 66 includes a body 166 integrally formed with the planet carrier 70 of the planetary gearset 60. The camshaft 66 further includes a cam groove 170 defined along the body 166. Unlike conventional impact mechanism camshafts, which include a cam groove shaped has a double helix, thereby permitting operation of the impact mechanism in either rotational direction of the camshaft, the illustrated cam groove 170 forms a single helix about the body 166 of the camshaft 66. As described in greater detail below, the geometry of the cam groove 170 causes the impact mechanism 54 to produce rotational impacts in only one rotational direction of the camshaft 66. The impact mechanism 54 is disabled in the opposite rotational direction. Cam balls 174a, 174b (FIGS. 6-7) are received within the cam groove 170 and couple the inner hammer 74 to the camshaft 66.
With reference back to FIGS. 1 and 2, the impact tool 10 may be switched or toggled between a first mode (e.g., a non-impacting mode) and a second mode (e.g., an impacting mode) via a mode actuator 178. The mode actuator 178 is associated with a sensor (e.g., a microswitch, a Hall Effect sensor, or the like; not shown) configured to sense actuation of the mode actuator 178. The mode actuator 178 may be a sliding switch, a push-button, or the like. The sensor communicates with a controller (not shown) of the impact tool 10 to control a rotational direction of the motor 22 based on actuation (or the position) of the mode actuator 178.
For example, to operate the impact tool 10 in the impacting mode, a user may actuate the mode actuator 178 (e.g., to a first position A; FIG. 1). The motor 22 is configured to rotate the output shaft 26, and therefore, the camshaft 66, in a first rotation direction A about the first axis 30 (FIG. 9). As the camshaft 66 rotates in the first rotation direction A, the cam balls 174a, 174b are able to travel toward a first or rear end 170a of the cam groove 170. This permits the inner hammer 74 to translate rearward when the anvil 82 experiences a resistance torque greater than or equal to a trip torque of the impact mechanism 54 and thus remains stationary. The rear end 170a of the cam groove 170 accommodates enough rearward travel of the inner hammer 74 for the hammer lugs 94, 98 to clear the anvil lugs 86, 90, at which this point the inner hammer 74 rotates and is propelled forward along the camshaft 66 by the compression spring 80 to deliver a rotational impact to the anvil 82. The outer hammer 78 co-rotates with the inner hammer 74 providing an increased rotating mass, thereby delivering a greater amount of impact energy and torque to the anvil 82 compared to the inner hammer 74 alone. The crankshaft 106 is coupled to the anvil 82 to co-rotate therewith, such that the drive bushing 114 causes the yoke 118 to pivot in reciprocating manner relative to the yoke housing 18. The yoke 118 then transfers torque to the output drive 34 to either tighten or loosen a workpiece.
To operate the impact tool 10 in the non-impacting mode, the user may actuate the mode actuator 178 (e.g., to a second position B; FIG. 1). The motor 22 is configured to rotate the output shaft 26, and therefore, the camshaft 66, in a second rotation direction B about the first axis 30 (FIG. 9). As the camshaft 66 rotates in the second rotation direction B, the cam ball 174a engages a second or front end 170b of the cam groove 170. With the cam ball 174a at the front end 170b of the cam groove 170, the inner hammer 74 and the outer hammer 78 co-rotate with the camshaft 66. In addition, the inner hammer 74 is unable to translate rearward, such that the hammer lugs 94, 98 remain engaged with the anvil lugs 86, 90. Accordingly, the anvil 82 co-rotates with the camshaft 66 and the inner and outer hammers 74, 78, such that the impact mechanism 54 is disabled and instead operates as a direct-drive mechanism. The crankshaft 106 co-rotates with the anvil 82, such that the drive bushing 114 causes the yoke 118 to pivot in reciprocating manner relative to the yoke housing 18. The yoke 118 then transfers torque to the output drive 34 to either tighten or loosen a workpiece.
Whether the workpiece is tightened or loosened depends only on the position of the pawl 126, not the operating direction of the motor 22. Thus, the workpiece may be tightened or loosened, as desired, regardless of whether the impact tool 10 is operating in the impacting mode or the non-impacting mode.
Returning reference to FIGS. 2-3B, the assembling process of the anvil 82 and the crankshaft 106 in the illustrated embodiment advantageously allows the crankshaft 106 to have a more compact size relative to conventional crankshafts. Specifically, the anvil 82 and the crankshaft 106 may be assembled in a two-step process, rather than being formed integrally, which permits, or enables, the compact size of the crankshaft 106. In a first step, the crankshaft 106 may be slid through bearings 182 which are supported by a front portion 18a of the yoke housing 18. Specifically, the crankshaft 106 may be slid through the bearings 182 at an end of the crankshaft 106 where the anvil 82 would have been integrally formed. In some instances, the bearings 182 may first be slid onto crankshaft 106, and then the crankshaft 106 and the bearings 182 together may be slid through the front portion 18a of the yoke housing 18. In a second step, the anvil 82 may be pressed onto the crankshaft 106 after the crankshaft 106 is inserted through the bearings 182. Specifically, the anvil 82 includes a coupling aperture 82a, and the crankshaft 106 includes a coupling portion 106a that may be inserted through the coupling aperture 82a for securing the anvil 82 and the crankshaft 106 relative to one another. The coupling portion 106a of the crankshaft 106 is positioned at an end of the crankshaft 106 opposite from the eccentric member 110. The coupling aperture 82a and the coupling portion 106a are complimentarily shaped to one another. Specifically, in the illustrated embodiment, the coupling aperture 82a and the coupling portion 106a are oblong-shaped. In other embodiments, the coupling aperture 82a and the coupling portion 106a may have other shapes. By assembling the anvil 82 and the crankshaft 106 together with two steps, the crankshaft 106 (e.g., a main body of the crankshaft 106) is not required to have a diameter that is equal to the outer diameter of the lugs 86, 90 of the anvil 82 or the outer diameter of the eccentric member 110 in order for the crankshaft 106 to be supported by the bearings 182. That is, the bearings 182 are not required to have the same diameter as the outer diameter of the lugs 86, 90 of the anvil 82 or the eccentric member 110 because the crankshaft 106 is able to be inserted through the bearings 182 prior to coupling with the anvil 82. Accordingly, the two-step assembly process of the anvil 82 and the crankshaft 106 permits a more compact design of the crankshaft 106.
With reference to FIG. 10, an alternative ratchet mechanism 200 that may be incorporated into the yoke housing 18 is illustrated. The illustrated ratchet mechanism 200 includes an eccentric member 204 of a crankshaft arranged within a drive bushing 208 and a yoke 212 through which the output drive 34 extends. The yoke 212 has a recess 216 in which the drive bushing 208 is arranged. The ratchet mechanism 200 further includes a first pawl 220a and a second pawl 220b provided within the yoke 212. The first and second pawls 220a, 220b are pivotably secured by a respective pin 228 that is coupled to a rotational member (not shown). The rotational member is configured to be rotated to move the first and second pawls 220a, 220b between a first position corresponding to a first rotational locking direction 224a of the output drive 34 and a second position corresponding to a second rotational locking direction 224b of the output drive 34. Each pawl 220a, 220b includes an angled first end 232a, 232b having teeth and an angled second end 236a, 236b having teeth. The first end 232a, 232b and second end 236a, 236b of each pawl 220a, 220b are configured to engage inner teeth 240 of the yoke 212.
When the rotational member is rotated to arrange the pawls 220a, 220b in the first position, the first end 232a, 232b of each pawl 220a, 220b meshes with the inner teeth 240 of the yoke 212. The first position thereby prevents the output drive 34 from rotating in the first rotational locking direction 224a. Once the rotational member is rotated to arrange the pawls 220a, 220b in the second position, the second end 236a, 236b of each pawl 220a, 220b meshes with the inner teeth 240 of the yoke 212. The second position thereby prevents the output drive 34 from rotating in the second rotational locking direction 224b. Providing two pawls within the yoke 212 reduces the amount of stress that may be experienced by a single pawl and may permit a relatively higher torque output.
FIGS. 11 and 12 illustrate a portion of an impact tool 310 according to another embodiment of the disclosure. The impact tool 310 may be, for example, a rotary impact wrench, a high-torque impact wrench, or another similar wrench. The impact tool 310 includes a rear gear case 314, a front gear case 318, a motor 322, and an impact mechanism 326. The motor 322 includes an output shaft 330 that extends through the rear gear case 314 and is engaged with the impact mechanism 326 at an end of the output shaft 330 opposite from the motor 322. The motor 322 is configured to drive rotation of the output shaft 330 about an axis 334. As such, the motor 322 is configured to drive operation of the impact mechanism 326 via the output shaft 330. The impact mechanism 326 is supported within the rear gear case 314 and the front gear case 318.
The impact mechanism 326 includes a gear assembly in the form of a planetary gearset 338 operably coupled to the motor 322 to provide a speed reduction/torque increase from the motor 322. The planetary gearset 338 includes a pinion 342 coupled for co-rotation with the output shaft 330 of the motor 322, a plurality of planet gears 346 meshed with the pinion 342, and a ring gear 350 meshed with the planet gears 346. The ring gear 350 of the planetary gearset 338 is rotationally fixed within the rear gear case 314.
The impact mechanism 326 further includes a camshaft 354 driven by the output of the planetary gearset 338 (e.g., the camshaft 354 may define a planet carrier 358 of the planetary gearset 338), an inner hammer 362 coupled to the camshaft 354, an outer hammer 366 surrounding and radially supported by the inner hammer 362, a thrust bearing 368 that supports a rear end of outer hammer 366, and an anvil 370 having opposed, radially outward extending anvil lugs 374, 378 that receive impacts from corresponding hammer lugs 382, 386 on the inner hammer 362. The illustrated impact mechanism 326 also includes a compression spring 390 disposed within the outer hammer 366 and configured to bias the inner hammer 362 toward the anvil 370.
The outer hammer 366 and inner hammer 362 form a two-piece or double hammer assembly able to provide increased rotating mass and, therefore, increased impact energy to the anvil 370 compared to the inner hammer 362 alone. In addition, the outer diameter of the hammer assembly may be less than a single hammer having an equivalent mass, permitting the impact tool 310 to be more compact.
The outer hammer 366 is sleeved onto the inner hammer 362 (e.g., by cooperating projections, splines, pins, or any other suitable arrangement) such that the outer hammer 366 is coupled for co-rotation with the inner hammer 362, but the inner hammer 362 is able to move axially relative to the outer hammer 366. In particular, the illustrated outer hammer 366 is provided with keyways 394 configured to receive key portions 398 formed along the outer diameter of the inner hammer 362. In the illustrated embodiment, the key portions 398 extend outwardly from the outer surface of the inner hammer 362. In other embodiments, such as the embodiment illustrated in FIG. 7, the key portions 398 may extend inwardly from an inner surface of the outer hammer 366. The inner hammer 362 functions similar to a hammer of a conventional impact mechanism, however the inner hammer 362 is provided with a reduced outer diameter to thereby reduce a size of the impact tool 310. As such, a mass of the inner hammer 362 is also reduced. Rotatably coupling the outer hammer 366 to the inner hammer 362 adds a substantial amount of mass to the inner hammer 362.
The camshaft 354 includes a body 402 integrally formed with the planet carrier 358 of the planetary gearset 338. The camshaft 354 further includes a cam groove 406 defined along the body 402. Unlike conventional impact mechanism camshafts, which include a cam groove shaped has a double helix, thereby permitting operation of the impact mechanism in either rotational direction of the camshaft, the illustrated cam groove 406 forms a single helix about the body 402 of the camshaft 354. The geometry of the cam groove 406 causes the impact mechanism 326 to produce rotational impacts in only one rotational direction of the camshaft 354. The impact mechanism 326 is disabled in the opposite rotational direction. Cam balls 410 are received within the cam groove 406 and couple the inner hammer 362 to the camshaft 354.
FIG. 13 illustrate a portion of an impact tool according to another embodiment of the disclosure. The impact tool may be, for example, a rotary impact wrench, a high-torque impact wrench, or another similar wrench. The portion illustrated in FIG. 13 is an impact mechanism 510 for the impact tool.
The impact mechanism 510 includes a camshaft 514 that may support a planetary gearset configured to be engaged by the motor (e.g., the camshaft 514 may define a planet carrier 518 for the planetary gearset), an inner hammer 522 coupled to the camshaft 514, an outer hammer 526 surrounding and radially supported by the inner hammer 522, and an anvil 530 having opposed, radially outward extending anvil lugs 534 that receive impacts from corresponding hammer lugs 538 on the inner hammer 522. The illustrated impact mechanism 510 may also include a compression spring that is configured to bias the inner hammer 522 toward the anvil 530.
With reference to FIGS. 14A and 14B, the camshaft 514 includes a body integrally formed with the planet carrier 518 and a cam groove 542 defined along the body. In the illustrated embodiment, the cam groove 542 forms a double helix about the body of the camshaft 514 which allows the impact mechanism 510 to produce rotational impacts in multiple (e.g., two) rotational directions. Alternatively, the cam groove 542 may form a single helix about the body of the camshaft 514 which causes the impact mechanism 510 to produce rotational impacts in just a single rotational direction. Cam balls may be received within the cam groove 542 and couple the inner hammer 522 to the camshaft 514. The cam balls enable the inner hammer 522 to translate along camshaft 514 toward and away from the anvil 530. For example, the inner hammer 522 may move between a forward-most position, as illustrated in FIG. 14A, in which the inner hammer 522 is positioned to strike the anvil 530 (FIG. 13) and a rearward-most position, as illustrated in FIG. 14B, in which the inner hammer 522 is spaced away from the anvil 530 (FIG. 13).
The outer hammer 526 and inner hammer 522 form a two-piece or double hammer assembly able to provide increased rotating mass and, therefore, increased impact energy to the anvil 530 compared to the inner hammer 522 alone. In addition, the outer diameter of the hammer assembly may be less than a single hammer having an equivalent mass, permitting the impact tool to be more compact.
With reference to FIGS. 15A and 15B, the inner hammer 522 may be substantially similar to the inner hammer 362 of FIG. 12, except for the differences described herein. In the illustrated embodiment of FIGS. 14A-15B, the size of the inner hammer 522 is reduced relative to the inner hammer 362 of FIG. 12. FIG. 15A illustrates a conventional hammer 522′ with a schematic ellipse 546 imposed over the conventional hammer 522′. The inner hammer 522 of the illustrated embodiment, as illustrated in FIG. 15B, is formed based on what falls within the ellipse 546, as illustrated in FIG. 15A. Specifically, the inner hammer 522 is formed in the shape of the ellipse in which major radii 546a of the ellipse extend diametrically between the two hammer lugs 538 and minor radii 546b of the ellipse extend perpendicularly to the major radii 546a. The inner hammer 522 is formed without portions of a conventional hammer that would fall outside of the ellipse 546 (e.g., a portion of a rim 524′ for the conventional hammer 522′ that connects the lugs 538′ in the conventional hammer 522′). In the illustrated embodiment, the major radius 546a is 29 mm, and the minor radius 546b is 19 mm such that a ratio of the major radius 546a to the minor radius 546b is roughly 1.52. In some embodiments, the major radius 546a and the minor radius 546b may have different lengths while maintaining the ratio of 1.52. In other embodiments, the major radius 546a and the minor radius 546b may be variably different such that the ratio of 1.52 is not maintained. The elliptical shape of the inner hammer 522 reduces the overall mass and size of the translating inner hammer 522 compared to the inner hammer 362 of FIG. 12. By reducing the mass and size of the translating hammer 522, vibrations translated to the user as a result of operation of the impact tool are reduced.
With continued reference to FIGS. 15A and 15B, the outer hammer 526 is sleeved onto the inner hammer 522 such that the outer hammer 526 is coupled for co-rotation with the inner hammer 522, but the inner hammer 522 is able to move axially relative to the outer hammer 526. In particular, the illustrated outer hammer 526 is provided with a hammer shoulder 550 on each side of the inner hammer 522 that protrudes from an inner surface of the outer hammer 526 toward the inner hammer 522 to effectively rotationally couple the inner hammer 522 and the outer hammer 526 together. One hammer shoulder 550 is positioned on a first side the major radius 546a, and another hammer shoulder 550 is positioned on a second side of the major radius 546a that is opposite the first side. In the illustrated embodiment, with reference to FIGS. 13 and 15B, the hammer shoulders 550 are recessed from each end of the outer hammer 526. In the illustrated embodiment, the hammer shoulders 550 are positioned relatively closer to a forward end of outer hammer 526 than to a rearward end of the outer hammer 526 to ensure rotational coupling is secure at the point of impact between the hammer lugs 538 and the anvil lugs 534. In some embodiments, the shoulders 550 may be positioned substantially in the middle of the outer hammer 526.
Returning reference to FIGS. 15A and 15B, each of the hammer shoulders 550 is formed in the shape of the portions of the conventional hammer 522′ that fall outside of the ellipse 546. As such, in the illustrated embodiment, the hammer shoulders 550 are substantially crescent-shaped. In other embodiments, the hammer shoulders 550 may have different shapes. For example, the hammer shoulders 550 may have the shape of a semi-circle in which a flat side of the semi-circle extends tangentially to an outer edge of the ellipse 546 at the minor radius 546b. The shape of the hammer shoulders 550 permits translational movement of the inner hammer 522 relative to the outer hammer 526 (i.e., along the camshaft 514) but inhibits relative rotational movement of the hammers 522, 526. In other words, the hammer shoulders 550 effectively rotationally couple the inner hammer 522 and the outer hammer 526 together, thereby increasing the mass with which the hammer lugs 538 may strike the anvil lugs 534.
The impact mechanism of FIGS. 13-15B advantageously reduces the mass of the translating hammer (i.e., the inner hammer 522) while substantially maintaining the mass of the double hammer assembly. This enables the double hammer assembly to maintain the amount of torque that may be applied to the anvil 530 while reducing vibrations that may be transferred, or translated, to the user due to the reduced translational mass.
FIGS. 16-21 illustrate a portion of an impact tool according to another embodiment of the disclosure. The impact tool may be, for example, a rotary impact wrench, a high-torque impact wrench, or another similar wrench. The portion illustrated in FIGS. 16-18 is an impact mechanism 710 for the impact tool.
The impact mechanism 710 includes a camshaft 714 that may support a planetary gearset configured to be engaged by the motor (e.g., the camshaft 714 may define a planet carrier 718 for the planetary gearset), an inner hammer 722 coupled to the camshaft 714, an outer hammer 726 surrounding the inner hammer 722, and an anvil 730 having opposed, radially outward extending anvil lugs 734 that receive impacts from corresponding hammer lugs 738 on the inner hammer 722. The illustrated impact mechanism 710 also includes a compression spring 742 that is configured to bias the inner hammer 722 toward the anvil 730.
With reference to FIG. 17, the camshaft 714 includes a body integrally formed with the planet carrier 718 and a cam groove 746 defined along the body. In the illustrated embodiment, the cam groove 746 forms a double helix about the body of the camshaft 714 which allows the impact mechanism 710 to produce rotational impacts in multiple (e.g., two) rotational directions. Alternatively, the cam groove 746 may form a single helix about the body of the camshaft 714 which causes the impact mechanism 710 to produce rotational impacts in just a single rotational direction. Cam balls may be received within the cam groove 746 and couple the inner hammer 722 to the camshaft 714. The cam balls enable the inner hammer 722 to translate along the camshaft 714 toward and away from the anvil 730. For example, the inner hammer 722 may move between a forward-most position in which the inner hammer 722 is positioned to strike the anvil 730 and a rearward-most position in which the inner hammer 722 is spaced away from the anvil 730.
The outer hammer 726 and the inner hammer 722 form a two-piece or double hammer assembly able to provide increased rotating mass and, therefore, increased impact energy to the anvil 730 compared to the inner hammer 722 alone. In addition, the outer diameter of the hammer assembly may be less than a single hammer having an equivalent mass, permitting the impact tool to be more compact.
With reference to FIGS. 17-19, the inner hammer 722 may be substantially similar to the inner hammer 362 of FIG. 12 and/or the inner hammer 522 of FIG. 15A, except for the differences described herein. In the illustrated embodiment of FIGS. 16-19, the size of the inner hammer is reduced relative to the inner hammer 362 of FIG. 12 and is increased relative to the inner hammer 522 of FIG. 15A. The inner hammer 722 includes two flanges 750 and a circular portion 754 positioned between the two flanges 750. In the illustrated embodiment, the two flanges 750 are diametrically opposed from one another (i.e., diametrically across the circular portion 754), and the flanges 750 axially protrude from the circular portion 754 (i.e., along an extension direction of the camshaft 714). Each of the of the two flanges 750 includes one of the hammer lugs 738. The hammer lugs 738 may be substantially similar (e.g., have a similar shape) to the hammer lugs 382, 386 of FIGS. 11 and 12 and the hammer lugs 538 of FIGS. 14A and 14B. A first diameter D1 of the inner hammer 722 is defined by the outer edges 750a of the flanges 750. The circular portion 754 includes a central aperture for mounting the inner hammer 722 to the camshaft 714 and a plurality of key portions 758 that protrude from an outer perimeter of the circular portion 754. In the illustrated embodiment, the inner hammer 722 includes 10 key portions 758. The plurality of key portions 758 is provided along the perimeter of the circular portion 754 between the flanges 750. Specifically, 5 key portions 758 are provided between the flanges 750 on the circular portion 754 on a first side of the inner hammer 722 (e.g., a right side), and 5 key portions 758 are provided between the flanges 750 on the circular portion 754 on a second side of the inner hammer 722 (e.g., a left side). A second diameter D2 and a third diameter D3 are defined by the circular portion 754. Specifically, the second diameter D2 is defined by the outer perimeter of the circular portion 754 at a base 758a of the key portions 758, and the third diameter D3 is defined by an outer protruding edge 758b of the key portions 758. As such, the third diameter D3 is larger than the second diameter D2. Both the second diameter D2 and the third diameter D3 are smaller than the first diameter D1.
With reference to FIGS. 20 and 21, the outer hammer 726 is sleeved onto the inner hammer 722 such that the outer hammer 726 is coupled for co-rotation with the inner hammer 722, but the inner hammer 722 is able to move axially relative to the outer hammer 726. In particular, the illustrated outer hammer 726 is provided with keyways 762 configured to receive the key portions 758 from the circular portion 754. In other embodiments, the outer hammer 726 may have key portions that protrude inwardly from an inner circumference of the outer hammer 726 and are received in keyways formed in the inner hammer 722. The outer hammer 726 is also provided with flange receptacles 766 that are each configured to receive one of the flanges 750 from the inner hammer 722. The engagement between the key portions 758 and the keyways 762 and between the flanges 750 and the flange receptacles 766 inhibits relative rotation between the inner hammer 722 and the outer hammer 726 but permits the inner hammer 722 to slide relative to the outer hammer 726.
In the illustrated embodiment, with reference to FIGS. 18 and 22, the key portions 758 and the keyways 762 have complementary involute profiles. An involute profile is the locus of a point considered as the end of a taut string being unwound from a given curve in the plane of that curve. Accordingly, clearance may exist between the key portions 758 and the keyways 762 that allows for controlled-line action sliding between the inner hammer 722 and the outer hammer 726. In controlled-line action sliding, each of the key portions 758 may have a single point of contact 770 (in a given plane) with the outer hammer 726 in the keyways 762 one each side surface of the key portions 758. As such, the key portions 758 are configured to slide along a single line of engagement with the outer hammer 726 that extends along a center axis 774 (FIG. 17) of the impact mechanism 710 (e.g., the point of contact 770 extended along the center axis 774). The clearance between the key portions 758 and the keyways 762 may enable the key portions 758 to move within the keyways 762 more smoothly and with less friction, thereby inhibiting jamming between the inner hammer 722 and the outer hammer 726 during operation of the impact mechanism 710. Additionally, the point of contact 770 may change during operation such that the line of engagement between the inner hammer 722 and the outer hammer 726, as the inner hammer 722 slides, is curved. Conventional double hammer mechanisms without keys and keyways having involute profiles may undergo planar sliding and/or wedging between the inner hammer and the outer hammer in which planar surfaces of the keys interface with corresponding planar surfaces of the keyways. Due to the edges between adjacent planar surfaces, conventional double hammer mechanisms may be more susceptible to jamming. Therefore, the double hammer mechanism of the illustrated embodiment allows for more efficient and more predictable sliding between the inner hammer 722 and the outer hammer 726.
With reference to FIGS. 17 and 21, due to the clearance between the key portions 758 and the keyways 762, the inner hammer 722 does not provide constant radial support for the outer hammer 726. As such, the impact mechanism 710 further includes a needle bearing 778 that is mounted to a rear portion 726a of the outer hammer 726. The needle bearing 778 may be engaged with the housing of the power tool to radially support the outer hammer 726.
The impact mechanism 710 of FIGS. 16-21 advantageously reduces the mass of the translating hammer (i.e., the inner hammer 722) while maintaining the mass of the double hammer assembly and maintaining alignment between the inner hammer 722 and the outer hammer 726. This enables the double hammer assembly to maintain the amount of torque that may be applied to the anvil while reducing vibrations that may be transferred, or translated to the user due to the reduced translational mass and while maintaining alignment between the inner hammer 722 and the outer hammer 726 even under heavy loads and impacts.
Although the disclosure has been described in detail with reference to certain preferred embodiments, variations and modifications exist within the scope and spirit of one or more independent aspects of the disclosure as described.
Various features and aspects of the present disclosure are set forth in the following claims.
Patent Application
20250135611
