US2768546A - Torque control for impact wrenches - Google Patents

Description

Oct. 30, 1956 L". A. AMTsBERG 2,768,546

Y ToRQuE CONTROL FQR IMPACT wRENcHEs Filed April 26, 1954 3 Sheets-Sheer 1 mm .m mb .mm 1| mb mw Oct. 30, 1956 L. A. AMTsB-ERG 2,768,546

TORQUE CONTROL FOR IMPACT WRENCHES Filed April 26, 1954 u Y x V J4 MZ 5 Sheets-Shee1 2 Tui-7' za INVENTOR ATTORNEY Oct. 30, 1956 L A. AMTsBr-:RG 2,768,546

' TORQUE CONTROL FOR IMPACT WRENCHES Filed April 26, 1954 3 Sheets-Sheet 3 fg g '//5 www@ ATTORNEY United States Patent() TORQUE CONTROL FOR IMPACT WRENCHES Lester A. Amtsberg, Utica, N. Y., assignor to Chicago Pneumatic Tool Company, New York, N. Y., a corporation of New Jersey Application April 26, 1954, Serial No. 425,689

17 Claims. (Cl. 81-52.3)

This invention relates to impact wrenches and more particularly to a device for controlling the degree of tightness of a bolt, nut, screw or other object driven by the wrench.

In the use of impact wrenches, for example in the construction of buildings and bridges, it is very important to drive the nut to the proper degree of tightness within a short time. This is usually accomplished by a tool having sufcient power and speed to attain the correct tightness rather quickly, whereupon the power to the tool must be cut-olf immediately to prevent excessive tightness and possible damage to the threads. In most instances the 'cut-off is effected manually and since the tightness varies with the skill of the operator, the driven nut must be tested by a hand torque wrench to determine whether it is driven properly or is too loose or too tight.

The general object of this invention is the provision of a torque control for an impact wrench which automatically cuts off the power upon attainment of a predetermined tightness ofthe driven bolt. A more specific object is to obtain uniformity, within close limits, in the tightness of the driven bolt or nut, independently of variations in the air pressure, skill of the operator, the size of the driven bolt, and the dimensions and structural characteristics of the plates being bolted.

Another object is to avoid the necessity for testing the bolt after it has been driven.

A further object is to permit the motor to run at a suiiciently high speed to complete the impacting cycle quickly without danger of driving the bolt too tight.

A still further object is to minimize the work being done by the operator, by making the operation as automatic as possible.

The design of an automatic cut-off for an impact wrench presents a special problem because the instantaneous torque delivered by the hammer to the anvil is not necessarily an accurate indication of the degree of tightness of the driven nut. It often happens that the first blow in a series of impacts, which occurs under relatively slight resistance to rotation but at a relatively high speed of the motor, will be accompanied by an instantaneous torque of extremely short duration but disproportionately high value due to the inertia of the anvil upon acceleration thereof. Subsequent impacts are delivered at a slower speed of the hammer, but with greater resistance of the driven nut, as compared with the rst impact. As the impacting continues the torque delivered by the hammer to the anvil is greater than during the preceding impact, until the torque attains a value corresponding to the maximum tightness of the driven nut at which point the power is cut off.

It is accordingly a further object of this invention to provide an automatic cut-off which operates whenever a hammer blow of nite duration attains a predetermined value of torque, but which is unresponsive to a torsional impulse of infinitesimal duration, such as occurs on the delivery of the trst blow, even though the instantaneous 2,768,546 Patented Oct. 30, 1956 torque be greater than the amount at which the control device responds, to cut-olf the power.

Still another object is the provision of a torque control of simple construction that does not appreciably add to the size or Weight of the impact wrench.

Another object is a provision of a torque control of durable construction having long life and requiring minimum maintenance.

In accordance with this invention the torque control is driven by a motor shaft extending rearwardly or in the opposite direction from the shaft which drives the impact clutch. The rear shaft drives a ball cam which in turn drives a flywheel during the time that the motor is accelerating. When the hammer and driving motor are brought to a stop, upon delivery of a rotational impact to the anvil, the ywheel continues to rotate and delivers torque to the rear motor shaft by way of the ball cam connection. The torque thus delivered is proportional to the torque beingdelivered to the anvil during the impact blow, because it is proportional to the rate deceleration of the hammer. The axial component of this force exerted by the cam is likewise proportional. When this force exceeds the preload of a spring, the cam moves axially and trips an automatic control valve to the closed position. The valve remains closed until the hand operated throttle valve is closed.

Another object of this invention is to enable the impact wrench to be used for driving left hand as well as right hand bolts, without disturbing the setting of the control device, with the torque release in the left hand direction being the same as, or proportional to, the torque release for right hand rotation.

A feature of this invention resides in the use of ball bearings to minimize friction between the parts of the control device, and thereby insure accuracy. Another feature resides in an adjusting screw, conveniently accessible, for regulating the torque at which the control device operates.

Other objects and features will appear more fully from the description which follows.

In the accompanying drawings, Figs. 1-18 disclose one I form of invention; Figs. 19-25, 28 and 29 show a modification; and Figs. 26 and 27 show further modications. Fig. 1 is a longitudinal section of an impact wrench embodying one form of this invention.

Fig. 2 is a cross-section as indicated by the arrows 2 in Fig. l, but with the impact clutch hammer moved to a diiferent position.

Figs. 3 and 4 are cross-sections as indicated by the arrows 3 and 4 respectively in Fig. l.

Fig. 5 is a fragmentary cross-section, in the same plane as Fig. 2, showing the hammer element in a different position.

Fig. 6 is a fragmentary view in longitudinal section similar to Fig. 1 but on a larger scale and with the parts in initial running condition.

Fig. 7 is a view similar to Fig. 6 but with the torque control apparatus in the process of operation, and with the automatic control valve moving toward closed position.

Fig. 8 is a cross-section along the irregular line 8 8 in Fig. l.

Fig. 9 is a fragmentary view, in elevation, of the rear end of the rotor shaft.

Fig. 10 is an end view of the same.

Fig. 11 is alongitudinal section of the ily-wheel or inertia sleeve as indicated by the arrows 11 in Fig. l2.

Fig. 12 is an elevational view of the rear side of the 1ly-wheel.

Fig. 13 is a longitudinal section of the ily-wheel as indicated by the arrows 13 in Fig. 12.

Fig. 14 is a front elevational view of the cam which drives the ywheel.

Fig. 15 is a side elevational view of the cam.

Fig. 16 is a rear elevational view of the cam.

Fig. 17 is a development showing the relation between the cam, ball and fly-wheel when the parts of the tool are in their normal position.

Fig. 18 is a View similar to Fig. 17 but with the ball and cam shifted to the operated position.

Fig. 19 is a fragmentary view in longitudinal section, of a modified torque control apparatus.

Fig. 20 is a view similar to Fig. 19 but with the trigger operated and the control valve moving to closed position.

Fig. 21 is a view partly in side elevation and partly in lection of the modied cam forming part of the device of Fig. 22 is a rear view in elevation of the cam shown in Fig. 21.

Fig. 23 is a front view of the modified ily-wheel.

Fig. 24 is a longitudinal section through said fly-wheel.

Fig. 25 is a development of the cam groove or pocket in said flywheel.

Figs. 26 and 27 are developments similar to Fig. 25 but showing modified cam grooves.

Fig. 28 is a development of the ball cam connection in the Fig. 19 device.

Fig. 29 is a view similar to Fig. 28 but with the parts in the operated position.

Fig. 6-16 and 19-24 are drawn to a larger scale than Figs. l-5. The development views in Figs. 17, 18 and 25-29 are on a still larger scale.

In the embodiment of invention illustrated in Figs. l- 18, the impact wrench, which is of the pneumatic type, comprises a clutch housing 18, a motor housing 19, and `a control housing 2G, secured in xed relation by any suitable means, such as the usual arrangement of bolts and flanges (not shown). The control housing extends rearwardly to form a grip handle 21. by means of which the tool may be manually held. An auxiliary handle (not shown) extends radially from clutch housing 22.

A reversible air motor 23 within the motor housing includes a cylinder or cylinder liner 24, the ends of which abut against end plates 25. The rear end plate has a flange 26 fitting a recess in the control housing 20 and a peripheral portion fitting the motor housing 19. The ilange 26 surrounds and supports a ball bearing 27.

Abutting the forward side of the front end plate 25 is a ball bearing 28 supported within motor housing 19. Ball bearings 27 and 2S `respectively support rear and front shafts 29 and 30 integral with and projecting from a rotor 31. As shown in Figs. 1 and 4, the rotor 31 is of cylindrical shape and is arranged coaxially with its shafts and with the clutch housing 1S, but eccentric with the cylinder 24 to provide a crescent shaped chamber 32 between the rotor and cylinder. The rotor has a pluraliy of radial slots 33 in which blades 34 are mounted for movement with their outer edges in scraping contact with the cylinder 24 to divide the crescent shaped chamber into a series of pockets between the inlet and exhaust ends. A reverse valve 35, operated by a lever 36, controls the direction of ow of the live air and hence the direction of `rotation of the motor 23.

Impact clutch Positioned centrally of the clutch housing is a rotatable tool head 38 having a elongated shank 39 and having an anvil portion comprising jaws or shoulders 41 adapted to receive rotational impacts as hereinafter described. Under the usual operating conditions, these impacts tend to misalign the tool axis. To resist such tendency the tool head is supported at its forward portion in a steel bushing 42 afiixed to the clutch housing and at its rear end by the rotor shaft 3G which projects into a recess 43 formed in the tool head. Also supporting `the rear end of the tool head is a clutch driving cam 44 whose front portion surrounds the recessed end of the tool head and whose rear portion is supported in a ball bearing 4S mounted in the front wall of the motor housing 19. The driving cam 44 has a splined connection 46 with the front rotor shaft 30 and rotates in unison therewith. The driving cam rotates at times on the tool head 58 but docs not move axially thereon, being confined between the inner race of ball bearing 28 and a shoulder 47 on the tool head. Shoulder 47 of course also prevents rearward axial movement of the tool head. Forward movement of the tool head is prevented by the engagement of the front end of the anvil jaws 41 with a sleeve 48 which iits over the shank 38 of the tool head and which abuts against a front washer 49 seated against bushing 42. The front end of the tool head 38 has a driving connection with a wrench socket 5t) adapted to receive the head of a bolt or other element (not shown) which is to be driven.

The hammer assembly surrounds the tool head 38 and comprises a cage 51 having front and rear plates 52 and 53 integrally connected by an arcuate web 54. The cage is confined against axial movement and at its opposite ends is arranged to abut against the front washer 49 and a rear washer 5S which is disposed in front of the inner race of ball bearing 45. The hammer cage 51 is arranged for coaxial rotative movement relative to the tool head 38 and accordingly the front plate 52 is bored to iit the sleeve 48 on the shank 39. rl'he rear plate 53 is bored to t over a cylindrical portion of the driving cam 44 and is permitted to oscillate or turn relative to the cam 44 and motor shaft 3i). Extending between the front and rear plates 52, 53, on the side of the cage opposite the connecting portion 44, is a pin 56. The pin extends parallel to the axis of the tool head 38 and provides a pivotal support for a hammer' element 57. As shown in Figs. l and 3, the hammer element comprises a central pivotal portion 57P apertured to t over the pin 56 and loosely abutting the front and rear plates 52 and 53 on the hammer cage 51. The hammer element 57 also comprises a pair of dogs 57R and 57L symmetrically arranged for right and left hand drive respectively and extending forwardly and inwardly from the pivotal portion 57i. At its front and inner extremity each dog has a striking surface or impact shoulder 57S engageable with the anvil jaws or shoulders 41 as hereinafter more fully described.

The hammer element 57 is provided with a cam surface 57C which extends between the striking surfaces 57S. The cam surface has the general shape of a sector of a cylinder whose center lies on the near side of the axis of revolution of the clutch `and extends parallel to said axis.

The impact clutch when driven in a clockwise direction (looking forward) operates in the following manner: The driving cam 44, having a direct connection with the rotor shaft 30, delivers force to the hammer element 57, along a line which lies intermediate the axis of revolution of the hammer cage assembly and the individual axis 56 about which the hammer element 57 may oscillate. The force of the driving cam is resolved into two components, one of which imparts a motion of revolution to the hammer element about the axis of tool head, and another of which ends to rock the hammer dog 57K out of driving engagement with the associated anvil jaw 41. The cage, which includes the plates 52 and 53, web 54 and pin 56, is carried along with the hammer element 57 `as it revolves. The dog 57K is guided for rocking movement about the pivot pin 56 due to camming engagement of the concave inner face 57C of the hammer element with the tool head surface adjacent the anvil jaw 41. Upon completion of the rocking movement the shoulder 57S on the dog is meshed with the anvil jaw 11. lf the resistance to rotation of the driven bolt, nut, or screw, is relatively slight, the clutch parts may remain for a considerable period of time in meshed relation due to ric tion between the hammer element 57 and the pivot pin 56, and also frictional engagement between the driving and driven shoulders'57S and 41 on the hammer dog and anvil jaw respectively.

`When the tool head 38 encounters substantial resistl'ance to rotation, the forces holding the clutch in mesh are overcome by the declutching force set up by the driving cam 44 against the hammer element 57 and the dog 57R is thereby locked in a releasing direction, which is counterclockwise, looking forward as in Fig. 3. As soon as the dog 57R is de-clutched, the motor 23 is relieved of its load and accelerates with the cage and hammer element to accumulate kinetic energy during one revolution of the motor, after which the motor, cage and hammer element are stopped upon delivery of an impact through the shoulders 57S and 41. The rotational impacts are repeated as long as the operator holds the wrench socket 50 in engagement with the torque resisting bolt and continues the supply of air to the motor, unless sooner terminated by the automatic control of the present invention.

Rotary air motor Live air is supplied to the tool from an air hose connection 60 on the grip handle 21, through a suitable filter 61, and thence through a throttle valve 62 to a handle passage 63, the throttle valve being controlled by the usual manipulative lever 64 operating against the holding force of a spring 65. After leaving the handle passage the air passes around an automatic valve 66 (the purpose of which will be described later) into a valve bushing 67, through a series of radial ports 68 in the bushing and thence into an air supply chamber 69 formed in the control housing 20. Referring to Figs. 1 and 4, live air flows from the supply chamber through a longitudinal passage 71 and a port leading to the interior of reverse valve bushing 72. From there the live air is directed, by means of reverse valve 35, into a housing recess 73R or 73L, depending on the direction of rotation, which is selected by manipulation of the reverse valve lever 36. From the housing recess the live pressure fluid passes through a longitudinal bore 74R (or 74L), a radial port 75R (or 75L) and another longitudinal port 76K y(or 76L) leading to a cylinder recess 77R (or 77L) at 'one end of the crescent shaped chamber in the motor. 'The rotor 31 turns, in a manner well known in the art, )by the action of live air passing through the crescent :shaped chamber 32 and expanding in the pockets or :spaces between the blades 34. The exhaust air flows from the cylinder recess 77L (or 77K) through longitudinal bore 76L (or 76R), radial port 75L (or 75R), longitudinal bore 74L (or 74K), housing recess 73L (or 73K), through an opening into reverse valve bushing 72 and thence through exhaust port 79 to atmosphere.

The impact clutch 51, 57, 38, the air motor 23 for driving the clutch and the reverse valve 35, 36 for controlling the direction of rotation, all of which have been described in detail, are conventional structures by themselves. The present invention relates to the combination of such a device with the automatic valve 66 which is arranged to cut olf the supply of live air and therefore stop the rotation of the clutch automatically upon attainment of a predetermined maximum load on the tool head 38. The novel mechanism for automatically releasing the automatic valve will now be described.

Automatic cut-O valveV Referring to Figs. l, 6 and 7, the valve 66 is surrounded by a coiled compression spring 81, interposed between the head 82 of the valve and the bottom of the recess in the valve bushing 67, whereby the spring at all times urges the valve outward or toward the open position. When the throttle valve 62 is closed, as shown in Fig. 1, the pressure of the spring 81 is unopposed and acts to seat the outer extremity of the automatic control valve 66 against a plug 83 secured to the control housing 20. rlfhe automatic valve 66 has a stem 84 slidably tting the bushing 67 and terminating in an inward extension or projection 85 separated from the main stem by a shoulder 86. In the idle condition of the parts, as shown in Fig. 1, the

projection acts as a stop for a double arm lever or trigger 87 pivoted on a pin 88 supported transversely in the control housing 20. The trigger is held in contact with the valve extension 85 by yieldable means such as compression spring 89. Upon depression of thethrottle lever 64, the operator admits live air through the handle passage 63 to initiate operation of the tool as previously described. The uid pressure in passage 63 acts against the outer end of the automatic valve and, being unbalanced over the area of the valve stern 84, overcomes the relatively light spring 81 and shifts the control valve inward until the valve shoulder 86 seats against the trigger 87 as shown in Fig. 6. As long as the trigger remains in the Fig. 6 position it holds the control valve 66 open to drive the motor 23.

Control for automatic valve The apparatus for automatically operating the trigger 87, to release it from the path of the automatic valve, will now be described. Referring to Figs. 6, 9 and 10, the rear shaft 29 on the rotor 31, which projects into the control housing 20, has three longitudinally extending grooves 91 adapted for the reception of balls 92. A cam sleeve 93, shown best in Figs. 14, l5 and 16, is provided with internal grooves 94 cooperating with the balls to form a splined connection allowing limited relative axial (but not rotative) movement between shaft 29 and the cam sleeve. The peripheral surface of the cam sle-eve is generally cylindrical and tits a counterbore within a ilywheel or inertia sleeve 95, shown best in Figs. l1 and 12. In the idle condition of the parts, and also during the first part of the operation of the tool, the cam sleeve 93 occupies a forward position in the ily-wheel as shown in Fig. 6.v The cam sleeve drives the ily-wheel through a ball cam arrangement adapted to transmit to the ily-wheel a rotational component of force and in addition an axial component tending to shift the ily-wheel forward or the cam sleeve rearward. The ily-wheel 95 cannot move forward because its front face seats against the inner race of ball bearing 27. Therefore the camming action between the cam sleeve 93 and the ily-wheel tends to move the cam sleeve rearward toward the position illustrated in Fig. 7. In order to effect the desired camming action the cam sleeve 93 is provided on its outside surface with three equally spaced grooves or pockets 97, all of the same size and shape, each having sides 97K and 97L, tapering from the front edge of the cam sleeve and joined by a rounded vertex. As shown in Fig. 14, each groove 97 is rounded in cross section to t a ball 98 along either side of the groove. The ball is adapted in its normal position to seat in the rounded vertex between sides 97R and 97L' and to roll forward along one of the sides toward the open end of cam groove 97 upon relative movement between cam sleeve 93 and the ily-wheel.

Fly-wheel 95 has three tapered grooves or pockets 99 formed in its counterbore and open at the rear side of the latter. Each groove has sides 99K and 99L joined by a rounded vertex at the front end of the groove. As shown in Fig. 12 each groove 99 is rounded in cross section to t the ball 98 along either side of the groove. The ball is adapted in its normal position (Fig. 6) to seat in the rounded vertex between sides 99R and 99L and to roll rearward along side 99R or 99L toward the open end of groove 99 upon relative movement between cam sleeve 93 and fly-wheel 95. The shape and dimensions of groove 99 conform as nearly as possible with groove 97 and the sides 99R and 99L are spirally inclined at the same angle, say, an angle of 15 Assuming that the tool is running in a right hand direction (clockwise, looking forward), the ball 98 is arranged for rolling engagement along the sides 97R and 99K as the cam sleeve 93 is displaced relative to the iiy-Wheel from the Fig. 17 to the Fig. 18 position and back again. Similar rolling engagement takes place between the sides 97L and 99L when the tool is operated to deliver impacts in a left hand direction, as will be explained more fully hereinafter.

Yieldable means is provided to constantly urge the cam sleeve forward and to tend to hold it in the position of Figs. 6 and 17 with the balls 93 seated in the vertices of cam grooves 97 and 99. Such means is located within the control housing which also encloses the cam sleeve 93, fly-wheel 95 and trigger 87. The yieldable means comprises a leaf spring 101 (Figs. 6, 7 and 8) having a fulcrum portion partly encircling a transverse pin 102, a short arm extending rearward therefrom, and a long arm extending inward from the fulcrum portion. At or near its free end the long arm is interposed between the inner end of trigger 87 and a button 103. The buttonis mounted in a ball bearing 104 which in turn is supported in a cam sleeve extension 105 seated in the rear end the cam sleeve 93. A set screw 106 threaded in the wall of control housing 20, has an inner end abutting against the short arm of leaf spring 101 near the free end thereof. The purpose of the set screw is to provide an adjustable tension on spring 101 which acts through the button 103, ball bearing 104, cam sleeve extension 105, and cam sleeve 93. The set screw 106 has a kerf or slot 107 adapted to receive a screw driver by means of which the screw may be adjusted from outside the housing 20.

Assume that the operator desires to tighten a bolt, having right hand threads, to the required degree of tightness as predetermined by adjustment of the set screw 106. He checks, or re-sets, the position of the reverse valve 35, moves the entire tool until the wrench socket 50 rests on the head of the bolt, depresses the throttle lever 64 and simply holds the lever down with the tool in position. The action from that time on is automatic and independent of the skill of the operator. Live air flows past the throttle valve 62, through handle passage 63, past the automatic control valve 66 (in the Fig. 6 position) into air supply chamber 69, through passage 71, past reverse valve 35, into motor housing recess 73K (Fig. 4) and through passages 74R, 75l?` and 76K to cylinder recess 77R. The live air drives the motor 23 in a clockwise direction (looking forward) by acting against the slidable blades 34 in passing through the crescent shaped chamber 32. The exhaust air flows from the cylinder recess 77L, through passages 75L, 74L, 73L and past the reverse valve to atmospheric exhaust port '79.

As the motor 23 turns, its front rotor shaft 30 carries with it the clutch driving cam 44 (Fig. 3) which imparts clockwise revolution to the hammer assembly including hammer element 57 and supporting cage 51. The dog 57R on the hammer element engages the anvil jaw 41 to drive the tool head 38, wrench socket and driven bolt (not shown) all at the same speed as the motor. As the head of the bolt comes into contact with the work, there is a sudden increase in the resistance to rotation of the bolt, wrench socket S0, tool head 3S and anvil jaw 41. The reaction is transmitted back to hammer element 57, hammer cage 51, clutch driving cam 44 and motor 23, causing the motor to decelerate from a free running speed of, say, 3000 R. P. M. to a somewhat lower speed. During this deceleration, the impact clutch is released because the hammer element 57 is driven with a declutching component of force tending to rock the dog 57K counterclockwise about the pivot pin 56. Following release, the motor is permitted to accelerate during one complete revolution as the hammer element 57 is carried around the anvil jaws 41 and moves through the positions shown in Figs. 2 and 5. When the hammer element 57 reengages, as in Fig. 3, it delivers an impact over the shoulders 57S and 41, and the frictional engagement therebetween locks `the hammer element against declutehing movement until the impact has terminated. During the delivery of the impact the clutch driving cam and the motor are alsoloeked against anymovement relativeto 8 the anvil jaws 41. Upon termination of the impact the clutch driving Vcam 44 rocks the hammer element 57 out of driving position and the rotating parts accelerate for another 360 degrees until the hammer dog 57R again strikes the anvil. The rotational impacts are delivered in rapid succession, each blow being effective to turn the tool head 38 by a few degrees and thus to increase the tightness of the driven bolt or threaded fastener, in a manner well known in the art.

Generally speaking, each succeeding blow of a series occurs over a shorter period of time, with a shorter degree of turning movement of the anvil, with a greater force of blow and with a greater amount of deceleration of the motor, as the bolt is driven tighter and meets with increasing resistance. It should be understood, however, that an instantaneous deceleration of the motor which occurs only during an infinitesimal period of time is not a reliable indication of the tightness of the driven nut. The reason is that the force of the blow is delivered in two stages, the first of which is instantaneous and variable with factors unrelated to the tightness of the bolt, and the second stage of which is of finite duration and proportional to the resistance or tightness of the driven element. When the striking face 57S of the hammer dog 57K hits the anvil jaw 41, the `first effect is to start the tool head 38 and wrench socket 50 from rest and accelerate them to the speed at which they turn the driven bolt. The tool head accelerating stage occurs only for an infinitesimal time which is followed by the werking stage of finite duration in which the tool head does the work of turning the bolt against its resistance. As the bolt becomes tighter, the distance traveled during each working stage of impact becomes less and the time required to arrest rotation of the hammer assembly 51, 57 is correspondingly shortened with a corresponding increase in the force of blow and in the deceleration of the hammer assembly. In the case of the first stage of the impact, however, the amount of force required to accelerate the tool head 38 and wrench socket 50 is not proportionate to the degree of tightness of the driven bolt, but depends chiey on the speed of the motor just before impact. In fact, the maximum instantaneous impact force probably occurs on the first blow because of the high motor speed, whereas on subsequent impacts the motor is limited to the speed that it can attain after starting from rest and turning only 360 degrees, or perhaps a few additional degrees in the case of rebound.

In accordance with the present invention, the torque control device is arranged to be responsive to a predetermined amount of motor deceleration of finite duration, but unresponsive to motor deceleration of the same or even higher value which occurs instantaneously or for an infinitesimal period of time. Referring to Figs 6, ll, l5 and 17, when the motor 23 starts from rest, either upon initial operation of the tool or at the end of one impact in a series, the rear shaft 29 on the rotor drives the inertia sleeve through a connection which includes cam edge 971., ball 98 and cam edge 99L. This connection is arranged to resolve the driving force into two components, one rotational and the other in an axial direction tending to displace the cam 93 rearwardly. During the time that the motor is increasing its speed the driving force through the ball cam connection is not sufficient to overcome the holding pressure of leaf spring 101 and the ball cam 98 therefore remains in its normal or Fig. 17 position up to the time of delivery of an impact. The cam 93 and fly-wheel 95 rotate in unison in the direction indicated by the arrows in Fig. l7. When the hammer dog 57R strikes the anvil jaw 4l, the motor 23 stops suddenly or encounters an abrupt reduction in speed. Upon such deceleration of the motor with its rear shaft 29, the ball cam arrangement reverses its action and the fly-wheel 95 drives the motor shaft .through the edge 99K, ball 98 and edge 97K. The inclination of these sides-or edgestendsto shift the cam 93 rearward. Upon delivery of the lirst impact there is a sharp instantaneous deceleration of the motor, as the hammer element 57 accelerates the tool head 38 and wrench socket 50. During this brief instant of maximum deceleration the driving force of the flywheel 95 rises to a value suicient to dislodge the balls 98 from the bottoms of the pockets 99 and 97 and to cause the cam sleeve 93 to shift rearward. Before the sleeve has moved substantially, the force of the blow is attenuated and `during the working portion of the rst impact the cam sleeve is restored to the position of Figs. 6 and 17. During subsequent impacts the ball 98 either remains in the Fig. 17 position or moves only slightly therefrom, along the edges 99R and 97K. When the bolt is driven to its maximum predetermined degree of tightness, the force of impact and the rate of deceleration of motor 23 cause the cam 93 to be moved rearward as above explained.

Torque delivered by the fly-Wheel through the ball cam connection is proportional to the impact torque delivered to the anvil 38 because it is proportionate to the rate of deceleration of the hammer assembly 51, 57. The axial component of force on the cam 93 is likewise proportional. Since the deceleration is sustained during the working portion of the impact, the cam continues to move rearward, against the pressure of spring 101, until it reaches the position shown in Figs. 7 and 18. During such movement the cam acts against the cam extension 105, ball bearing 104, button 103 and spring 101 to rock the trigger 87 against the pressure of spring 89. As soon as the trigger is rocked to the Fig. 7 position it frees the shoulder 86 of automatic control valve 66 to permit the valve to move inward. The valve is moved inward by the pressure of live air action over the stem portion 84 of the valve until the head 82 seats against the valve bushing 67 to cut olf the supply of air to the motor 23.

The control valve remains closed and the motor remains at rest as long as the operator holds the throttle level 64 down. In order to condition the tool for starting a new cycle of operation, the operator releases lever 64 to close throttle valve 62. Thereupon the pressure acting on control valve 66 is relieved to permit the light spring 81 to open the valve with the result that the trigger 87 moves back to its Fig. l position, seated against the reduced extension 85 on the valve. In order to overcome delay in restoration of the valve which might be caused by live air trapped in handle passage 63 the valve 66 is provided with a central opening 109. When the valve is in its innermost position the opening 109 bleeds air from handle passage 63 into the control housing 20 which is vented by any suitable means, such as a hole 110, or a passage into the motor exhaust chamber.

The operator moves the impact wrench from one bolt to another, merely applying the wrench socket`50 to the bolt head, depressing the throttle lever 64, waiting until the tightening operation is completed releasing the lever and pressing again for the next bolt. To establish the desired adjustment a driven bolt is tested with a hand torque wrench. If greater (or less) tightness is desired, the set screw 106 is turned to increase (or reduce) the pre-load on spring 101.

The forces and design factors of the torque control for the impact wrench illustrated in Figs.' l-l8 are related as follows:

24= T zl LXI F=pre1oad of spring 101 (in pounds) T=delivered torque at cut-olf (foot pounds) =moment of inertia of fly-wheel 95 I=mon1ent of inertia of hammer assembly 57, 51,

clutch cam 44 and rotor 31 f L=lead of ball cam edges 97R and 99R (inches per revolution) Preferably the combined inertia of the tool head or anvil 38 should not exceed 5 percent of the hammer and rotor assembly 57, 51, 44, 31 to provide reasonable torque accuracy and prevent premature cut-off. The amount of displacement of cam 93 required to release the control valve 66 should be small enough to provide reasonabletorque accuracy but large enough to prevent prematurel cut-off due to inertia of tool head 38. In a physical. embodiment of the invention which has been tested suc-- cessfully, the cam 93 is displaced about 40 or 50 thou-v sandths of an inch against a spring force which may be: adjusted within the range between l5 and 30 pounds.. The torque delivered by the tool is adjustable between 300 and 600 foot pounds. Torque variation is less than,I plus or minus l0 percent. The accuracy of the device is due in part to the use of balls which minimize friction. Torque delivered at any setting of the screw 106 is virtually unaffected by live air pressure variations between 50 and 100 pounds per square inch providing the pressure is high enough to develop the required torque. In short the performance of the control ydevice is not seriously affected by normal variations in motor performance. Torque delivered to a 3M inch bolt is the same as that delivered to a one (l) inch bolt. Torque delivered at any setting is virtually unaffected by the addition of long extension Shanks, universal joints, slip chucks, etc. providing the addition of such accessories does not reduce the ultimate torque to a value lower than that to which the control device is adjusted. It is alsosubstantially independent of the Work rate. That is, a gathering job, in which the bolt approaches its maximum re sistance gradually will receive the same nal torque as a sudden stop job.

The embodiment of invention illustrated in Figs. l-l8 is symmetrical, and upon selected setting of the reverse lever 36 will drive a bolt with left hand threads to the same degree of tightness as a right hand bolt.

Modified )ty-wheel and cam In a modied form of this invention, illustrated in Figs. 19-24, and diagrammatically in Figs. 28 and 29, the fly-wheel 112 moves axially when cam action occurs. As shown in Figs. 24, the fly-wheel has a bore 113 and a counterbore 114 extending forward therefrom to the front face of the ily-wheel. An inner cam member 115 (Figs. 2l and 22) has a cylindrical portion 116 tting the counterbore 114 and a rear extension 117 fitting the bore 113.` The cylindrical surface of the counterbore is interrupted by three spaced grooves or pockets 118 open at lthe front end. Each groove has sides 118R and 118L tapering rearwardly and joined by a rounded vertex at the rear end. Each pocket is rounded in cross section to receive a ball 120. The inner cam 115 is provided with three spaced pockets 121 formed on the cylindrical portion 116 and open at the rear end thereof. Each pocket has sides 121K and 121L tapering forwardly and joined by a rounded vertex at the front end. The shape, dimensions and functions of the pockets 118 and 121 correspond to those of pockets 99 and 97 respectively in the embodiment first described, except that they are oppositely directed. That is to say, the cam drives the fly-wheel, and the balls normally rest in the vertices of the pockets in the case of the Fig. l embodiment, but upon relative rotary move ment, the fly-wheel 112 rather than the inner cam member 115 is displaced rearward.

In front of the cylindrical portion 116, the inner cam 115 has a iiange 122 which seats against the inner raceway of ball bearing 27. The front face of flange 122 is slightly spaced from the rear extremity of a shaft 123 which extends rearward from rotor 31. The rotor shaft has a recess receiving the front end of cam 115 with a splined connection 125, so that the cam member rotates in unison with the rotor 31 at all times.

The rear extremity of ily-wheel 112 abuts against the inner race of ball bearing 126 which is mounted within a `1 1 button 127. The rear or closed side of the button abuts against the spring 101 which provides a forward tension on the button, adjustable by means of set screw 106 as previously described.

In the operation of the embodiment of invention shown in Figs. 19-25, the rotor 31 starts from rest and as its front shaft drives the impact clutch, its rear shaft 123 drives the fly-wheel 112, acting through the splined connection 125, cam 115, and ball cam connection 121L, 120, 118L. As the fly-wheel is thus accelerated the inclination of pocket sides 121L and 118L has a tendency to shift the ily-wheel rearward but this tendency is overcome by the pressure of leaf spring 101. Accordingly, the ily-wheel remains in its normal relation to the driving cam 115, as shown in Figs. l9and 28, with the balls 12@ seated in the vertices of cam pockets 121 and 11S. When motion (in a right hand direction) of the rotor 31 is arrested bythe delivery of an impact, the iiy-wheel 112 continues under its inertia and drives the rotor shaft 123 through the ball cam connection HSR, 120, 121R. At the time of each impact the ball rolls along the pocket sides 118K and 121R, increasing its degree of movement as blows of increased force are delivered to the head 38. The balls roll back to the normal position (Figs. 19 and 28) after each impact. Upon delivery o-f a sustained blow, of nite duration and of predetermined magnitude, the balls 120 and y-wheel 112 are displaced to the position shown in Figs. 20 and 29. The ily-wheel moves button 127, which acts through spring 101 to press the trigger 87 to release position. At this stage the control valve 66 is released for closing movemen-t as indicated in Fig. 20.

In the form of invention shown in Figs. 19-24 the rotor 31 is relieved of end thrust which otherwise might cause it to rub against the forward end plate 25.

Modified cam grooves Fig. shows, in development, one of the cam pockets 118 in the y-wheel 112. The sides 121L and 121K are symmetrical. That is to say the angle of inclination L of the former is equal to the angle of inclination R of the latter. As a result the control arrangement of Figs. 19-25 will operate the same for both directions of rotation and will release at the same degree of tightness whether the wrench is driving a right hand or a left hand bolt. It is sometimes desired to release at a different degree of tightness for diterent directions without requiring the re-adjustment of set screw 106. Fig. 26 shows a pocket 118 unsymmetrically shaped in order to accomplish that end. In this modification the angle L is less than the angle R. The result is that upon delivery of rotational impacts in a left hand direction the balls 120 transmit a force which has a lesser component in an axial direction that would be the case if the angle L were equal to the angle R. Therefore the wrench continues to operate on a left hand bolt until it is driven tighter than would be the case with a bolt having right hand threads. Also, by making the side 121K steeper than the side 121L it is possible to remove a conventional right hand bolt that has been tightened without changing the setting of screw 106 even though the bolt has been somewhat frozen.

ln Fig. 27 the angle L is reduced to zero which means that the ball cam connection transmits only rotary force, with no axial component, to the fly-wheel 112 upon delivery of left hand or counter-clockwise impacts. A series of pockets so shaped may be used to provide an automatic control for tightening the bolts and a manual control for loosening the same bolts.

lf the fly-wheel pockets 118 are modified as shown in Fig. 26 or 27, corresponding changes should be made in the angles of the cam pockets 121. It willbe understood further that the variations in the pocket shapes as shown in Figs. 26 and 27 are equally applicableto the fly-wheel -12 pockets 99 and cam pockets 97 in the form of invention shown in Figs. 1-18.

While the invention has been described with reference to a clutch of the type which is driven directly by the motor, it is also applicable to impact wrenches of the accumulator type in which the motor runs continuously. In adapting the present invention to an accumulator type impact wrench, the ball cam drive for the y-Wheel would be operated not by the motor but by a spindle or other element that starts and stops along with the hammer assembly. in the case of an electric wrench the control device of the present invention may be arranged to open a switch instead of closing a valve.

If desired, the screw 106 may be provided with a calibrated torque setting dial. This would be useful in yautomotive repair shops where it is desired to adapt the tool quickly for driving different kinds of nuts having known torque requirements.

What is claimed is:

l. An impact wrench comprising a rotatable anvil, a rotatable hammer assembly arranged to drive said anvil and to deliver a series of rotational impacts thereto, a rotatable motor having a rotor for driving the hammer assembly, a rotating shaft connected with the hammer assembly and arranged to start and stop in substantial unison with the hammer assembly and rotor whereby the combinedv kinetic energy of the rotor, hammer assembly and shaft is delivered to the anvil upon impact, and means responsive to a predetermined deceleration of the shaft prior to the stopping thereof for cutting off the supply of power to the motor.

2. An impact wrench comprising a tool head adapted to drive a threaded fastener, a rotatable hammer arranged to deliver a series of rotational impacts to the tool head, a rotating shaft connected to the hammer and arranged to decelerate together with the hammer, a trigger device, inertia means for operating the trigger device in response to a predetermined deceleration of the rotating shaft, means for automatically terminating rotation of the hammer upon operation of the trigger device, and means for delaying operation of the trigger device, to prevent operation of the terminating means in response to hammer deceleration of extremely `short duration upon acceleration of the tool head during delivery of the first impact of a series.

3. An impact wrench according to claim 2 in which the inertia means comprises an inertia sleeve and a driving connection between the rotating shaft and the inertia sleeve.

4. An impact wrench according to claim 3 in which the driving connection between the rotating shaft and the inertia sleeve comprises mutually cooperating cam elements carired by the shaft and inertia sleeve respectively, said cam elements having relative axial movement upon development of relative rotative movement between the shaft and inertia sleeve.

5. An impact wrench according to claim 4 in which the trigger is operated in response to such relative axial movement.

6. An impact wrench comprising a rotatable anvil, a rotatable hammer element arranged to drive said anvil and deliver a series of rotational impacts thereto, a rotary air motor having a rotor for driving said hammer ele' ment, said rotor and hammer element being arranged to start and stop in substantial unison, a valve for supplying live air to the motor for driving the same, and means for automatically closing the valve to stop the motor upon development of a predetermined deceleration of the hammer element, said means being operable during the delivery of an impact but being ineffective upon delivery of an impact of short duration occurring as the first impact of a series.

7. An impact wrench comprising a rotatable anvil, a rotatable hammer arranged to drive said anvil and to dev liver a series of rotational impacts thereto, a rotary air motor for driving said hammer, a rotating shaft connected with the hammer and arranged to decelerate along with the hammer upon delivery of a rotational impact, a valve for supplying live air to the motor for driving the same, and inertia means for automatically closing the valve to stop the motor upon development of a predetermined deceleration of the rotating shaft, said inertia means comprising a ily-wheel, and a driving cam interposed between the shaft and fly-wheel to drive the latter, said ily-wheel being rotatable relative to the shaft upon deceleration thereof, and said driving cam being axially movable relative to the shaft and fly-wheel upon rotation of the fly-wheel relative to the shaft.

8. An impact wrench comprising a rotatable anvil, a rotatable hammer arranged to drive said anvil and deliver a series of rotational impacts thereto, means for driving the hammer, means including a trigger for stopping the rotation of the means for driving the hammer, automatic inertia means responsive to a predetermined deceleration of the hammer for operating the trigger, and means for delaying operation of the automatic means to prevent actuation of the trigger in response to deceleration of extremely short duration.

9. An impact tool comprising an anvil, a rotatable hammer arranged to deliver a series of impacts to said anvil, a rotatable shaft connected to the hammer and having intermittent rotary movement in substantial unison with the hammer, a rotatable driving means for the hammer, a fly-wheel driven by the shaft through a flexible connection, said ily-wheel being movable at times in unison with the shaft but being capable of rotating ahead of the shaft upon sudden deceleration of the latter, automatic means responsive to relative movement between the fly-wheel and shaft for discontinuing rotation of the means for driving the hammer, and means for inhibiting operation of said automatic means in response to an impact of short dura-tion occurring as the first impact of a series.

10. An impact tool according to claim 9, in which the exible connection comprises one or more pairs of cooperating inclined grooves, one groove carried by the shaft and the other by the ily-wheel, said grooves being arranged to cause relative axial movement between the fly-wheel and shaft upon relative rotative movement therebetween.

11. An impact tool according to claim 10 in which a spring is arranged to resist such axial movement.

l2. An impact tool according to claim 11 in which manipulative means is provided to adjust the tension of the spring and hence the rate of deceleration of the shaft required to discontinue rotation of the means for driving the hammer.

13. An impact tool according to claim 11 in which one 0r more balls are inserted between the cooperating grooves to minimize friction.

14. An impact tool according to claim 13 in which the groove carried by the shaft is in the form of a V-shaped pocket, the sides of which function in different directions, the inertia member being provided with a complementary V-shaped pocket.

15. An impact tool according to claim 14 in which the sides of the pocket are symmetrical whereby the means for discontinuing the driving of the hammer responds to fthe same deceleration thereof in either direction of rotation.

16. An impact tool according to claim 14 in which the sides of the pocket are asymmetrical whereby the means for discontinuing the driving of the hammer responds to a different deceleration of the hammer when driven in one direction than when driven in the other direction.

17. An impact tool according to claim 14 in which one side of the pocket is inclined and the other parallel to the axis of rotation whereby the means for discontinuing rotation of the hammer is elfective only for one direction of rotation.

References Cited in the file of this patent UNITED STATES PATENTS 2,261,204 Amtsberg Nov. 4, 1941 2,326,347 Forss Aug. 10, 1943 2,425,793 Fosnot Aug. 19, 1947 2,543,979 Maurer Mar. 6, 1951 FOREIGN PATENTS v 1,050,484 France Sept. 2, 1953

Images