Abstract:
A clutching arrangement of a positive feed drill comprises a first shaft and a second shaft. The first shaft has a first longitudinal axis and the second shaft has a second longitudinal axis. The first longitudinal axis and the second longitudinal axis being generally aligned. A first clutch gear assembly being rotatable about the first axis. A second clutch gear assembly being rotatable about the second axis. The second clutch gear assembly and the second shaft being axially moveable toward and away from the first clutch gear assembly and the first shaft.

Description:
CROSS-REFERENCE TO RELATED APPLICATIONS 
     The present application claims the priority benefit under 35 U.S.C. §119(e) of U.S. Provisional Application No. 61/155,412, filed on Feb. 25, 2009, which is hereby incorporated by reference in its entirety. 
    
    
     BACKGROUND OF THE INVENTION 
     1. Field of the Invention 
     The present invention generally relates to clutching arrangements for positive feed drills and, in particular, to clutching arrangements for positive feed drills that feature two clutching gears mounted to separate shafts. 
     2. Description of the Related Art 
     Positive feed drills are used to produce accurately placed and accurately dimensioned holes in workpieces. One application of this type of drill is in the aircraft industry where the holes are formed in materials that can be very difficult to cut. In some environments, the drills may periodically bind during operation. The binding can result from overloading caused by drilling deep holes or by heavy reaming, which results in metal chips being packed heavily in between the cutter flutes. The torque and thrust loads can be very high under overload conditions, which may resist the automatic withdrawal of the cutter from the workpiece. With reference to  FIG. 1 , a prior construction of a positive feed drill will be discussed to illustrate some of the causes of the failures being experienced. 
     SUMMARY OF THE INVENTION 
     With reference to  FIG. 1 , a clutching arrangement  18  of a prior positive feed drill will be described. As shown, the prior clutching arrangement  18  was positioned within a housing  19 . A motor (not shown) had a drive shaft that carried a first bevel gear (not shown). The first bevel gear drove a second bevel gear  20 . The second bevel gear  20  was connected to an input gear  22 . The input gear  22  meshed with a lower clutch gear  24 . The lower clutch gear  24  meshed with a drive gear  26 . Typically, the lower clutch gear  24  and the drive gear  26  had a 1:1 ratio so that the drive gear  26  rotated at the same speed as the clutch gear  24 . Thus, a 1:1 driving connection was established between the motor and the drive gear  26 . The drive gear  26  powered a cutter spindle (not shown) such that the cutter spindle rotated a cutting tool (not shown). 
     An upper clutch gear  30  meshed with a feed gear  32 . As will be described below, the feed gear  32  had a threaded relationship with the cutter spindle such that, if the feed gear  32  was stationary relative to the cutter spindle, the cutter spindle would move in one axial direction and, if the feed gear  32  was moving faster than the cutter spindle, the cutter spindle would move in an opposite axial direction. 
     To control the movement of the feed gear and the direction of axial movement, the upper clutch gear  30  moved in an axial direction. The upper clutch gear  30  included axial teeth  34  and the lower clutch gear  24  included axial teeth  36 . The two sets of axial teeth  34 ,  36  would engage each other when the upper clutch gear  30  moved axially downward into engagement with the lower clutch gear  24 . When the teeth  34 ,  36  engaged, the upper and lower clutch gears  30 ,  24  rotated together, which caused rotation of the feed gear  32  and the drive gear  26  respectively. When the teeth  34 ,  36  disengaged, the lower clutch gear  24  continued to rotate while the upper clutch gear  30  stopped rotating, which caused the drive gear  26  to continue rotating while the feed gear  32  stopped rotating. The upper clutch gear  30  also comprised upwardly extending teeth  38  that meshed with stationary teeth  40  such that the upper clutch gear  30  could be secured against rotation. 
     Thus, the prior clutching arrangement was designed to control the rotation of the feed gear  32 , which in turn caused the cutter spindle (not shown) and the cutting tool (not shown) to advance into or retract from a work piece. In the prior configuration, when the rate of rotation of the feed gear  32  was greater than the rate of rotation of the drive gear  26 , the cutting tool (not shown) advanced toward a work piece and, when the rate of rotation of the feed gear  32  was less than the rate of rotation of the drive gear  26 , the cutting tool (not shown) retracted from the work piece. 
     As shown in  FIG. 1 , the lower clutch gear  24  and the upper clutch gear  30  were mounted along a single solid shaft  42 . The solid shaft  42  extended through a spring  46  that was positioned between the lower clutch gear  24  and the upper clutch gear  30 . The axial position of the lower clutch gear  24  was generally fixed by the housing and a bushing member but the upper clutch gear  30  was designed to travel axially along with the axial movement of the solid shaft  42 . A snap ring  44  secured the upper clutch gear  30  in axial location along the solid shaft  42  from the top while the spring  46  generally urged the upper clutch gear  30  against the snap ring  44  to hold the clutch gear  30  in axial location along the solid shaft  42  from the bottom. 
     The solid shaft  42  moved axially up and down. Because the upper clutch gear  30  was generally secured in location along the solid shaft  42 , the upper clutch gear  30  followed the axial movement of the solid shaft  42 . In other words, when the solid shaft  42  moved downward along the axis of rotation R of the upper clutch gear  30 , the upper clutch gear  30  made the same axial translation. 
     As shown, a piston  50  divided an air cylinder  52  into an upper chamber  54  and a lower chamber  56 . The piston  50  was secured to the upper end of the solid shaft  42  by a bolt  58 . When the pressure in the upper chamber  54  was sufficiently higher than the pressure in the lower chamber  56  combined with the force from the spring  46 , the solid shaft  42  moved downward, thereby moving the upper clutch gear  30  toward the lower clutch gear  24 . When the pressure in the lower chamber  56  combined with the force from the spring  46  was sufficiently higher than the pressure in the upper chamber  54 , the solid shaft  42  moved upward, thereby moving the upper clutch gear  30  away from the lower clutch gear  24 . 
     Under certain conditions, such as when drilling a deep hole, performing heavy reaming, or when metal chips are tightly packed within the cutting tool&#39;s flutes, the torque and thrust load on the drill were very high. These loads often prevented automatic withdrawal of the cutter from the work piece because the high thrust load coupled with high torque transferred through the lower clutch gear  24  and the upper clutch gear  30  to the solid shaft  42 . The loads would cause a misalignment of the centerlines of the clutch gears  24 ,  30  and the solid shaft  42 . It was found, however, that even a minute misalignment resulted in increased friction, binding, and wear on the clutch gears  24 ,  30  and the solid shaft  42 . In addition, it was discovered that the slight misalignment reduced the likelihood of the upper clutch gear  30  properly engaging and disengaging the lower clutch gear  24 . 
     In addition, the high-load dynamic conditions sometimes caused the snap ring  44  to snap out of its groove along the solid shaft  42 . To replace the snap ring  44  within the groove, the drill had to be turned off and disassembled so that the snap ring  44  could be replaced. This maintenance downtime was extensive because of the retaining ring&#39;s central positioning in the clutching arrangement. 
     Accordingly, a split shaft clutching assembly has been developed to relieve the loading issues while an integrated rib has been developed to address the snap ring failures. Thus, in accordance with one embodiment that is arranged and configured in accordance with certain features, aspects and advantages of the present invention, a clutching arrangement comprises a first shaft secured to a housing. The first shaft defines a first axis of rotation. A second shaft is mounted in the housing and defines a second axis of rotation. The second axis of rotation generally is aligned with the first axis of alignment. A first clutch gear assembly is rotatably mounted on the first shaft. The first clutch gear assembly comprises a first tooth. A second clutch gear assembly is rotatably mounted on the second shaft. The second clutch gear assembly comprises a second tooth. The first tooth and the second tooth are positioned between the first clutch gear assembly and the second clutch gear assembly. The second clutch gear assembly and the second shaft are adapted to move axially along the second axis of rotation toward the first clutch gear assembly and the first shaft such that the first tooth and the second tooth can selectively engage each other. 
     In another configuration, a clutching arrangement is provided for a positive feed drill. The positive feed drill comprises a housing that defines a clutching arrangement accommodating chamber. The clutching arrangement comprises a lower clutch gear that is drivingly engaged with a drive gear. A stationary stub shaft is mounted to the housing. The stationary stub shaft comprises a first shaft longitudinal axis. A first bearing is positioned on the stationary stub shaft. The first bearing is mounted within the lower clutch gear. A first retainer ring secures the first bearing within the lower clutch gear. The first bearing is interposed between at least a portion of the lower clutch gear and the stub shaft. The lower clutch gear is rotatable around the first shaft longitudinal axis relative to the stationary stub shaft. A clutch ring is mounted to the housing. A floating shaft comprises a second shaft longitudinal axis and a flange. The floating shaft extends through the clutch ring. The floating shaft is capable of axial movement along the second shaft longitudinal axis relative to the clutch ring. A second bearing is supported by the clutch ring. The second bearing is interposed between the clutch ring and a portion of the floating shaft. An upper clutch gear is drivingly engaged with a feed gear. A third bearing is positioned within the upper clutch gear. The third bearing is interposed between a portion of the upper clutch gear and the floating shaft. A second retainer ring secures the third bearing within the upper clutch gear. The upper clutch gear is rotatable around the second shaft longitudinal axis relative to the floating shaft. The upper clutch gear is capable of movement along the second shaft longitudinal axis relative to the housing. A spring is positioned between the lower clutch gear and the upper clutch gear. The spring surrounds an upper portion of said stationary stub shaft and a lower portion of said floating shaft. The spring applies an upward spring force to the upper clutch gear. The upward spring force biases the upper clutch gear away from the lower clutch gear. A lower clutch gear tooth is positioned on an upward-facing surface of the lower clutch gear. An upper clutch gear tooth is positioned on a downward-facing surface of the upper clutch gear. The upper clutch gear is in an advance position when the lower clutch gear tooth engages the upper clutch gear tooth, whereby the upper clutch gear drives the feed gear. A first clutch ring tooth is positioned on a downward-facing surface of the clutch ring. A second clutch ring tooth positioned on an upward-facing surface of the upper clutch gear. The upper clutch gear is in a retract position when the first clutch ring tooth engages the second clutch ring tooth, whereby the upper clutch gear stops the feed gear. 
     In yet another configuration, a clutching arrangement for a positive feed drill comprises a first clutch gear that comprises a first longitudinal axis. The first clutch gear is rotatable about a fixed stub shaft that defines the first longitudinal axis. A floating shaft comprises a second longitudinal axis. The floating shaft is axially moveable along the second longitudinal axis. The first longitudinal axis and the second longitudinal axis are generally aligned. A second clutch gear is rotatable relative to the floating shaft around the second longitudinal axis. The second clutch gear translates along the second longitudinal axis along with the floating shaft. A first clutch gear tooth is positioned on the first clutch gear. A second clutch gear tooth is positioned on the second clutch gear. The second clutch tooth engages the first clutch gear tooth when the second clutch gear approaches the first clutch gear. 
    
    
     
       BRIEF DESCRIPTION OF THE DRAWINGS 
       These and other features, aspects and advantages of certain embodiments of the present invention will be described with references to the accompanying drawings. 
         FIG. 1  is a sectioned side elevation view of a prior positive feed drill drive featuring a clutching arrangement. 
         FIG. 2  is a perspective view of a positive feed drill having a clutching arrangement that is arranged and configured according to certain features, aspects, and advantages of an embodiment of the present invention. 
         FIG. 3  is an exploded perspective view of the positive feed drill depicted in  FIG. 2 , illustrating an air motor housing, a planetary gear reducer, and an angle head module. 
         FIG. 4  is an exploded perspective view of major subassemblies of the angle head module of the positive feed drill depicted in  FIG. 2 . 
         FIG. 5  is a sectioned side elevation view of the angle head module of the positive feed drill depicted in  FIG. 2  showing the clutching arrangement in a position wherein a cutting tool would be retracted. 
         FIG. 6  is a sectioned side elevation view of the angle head module of the positive feed drill depicted in  FIG. 2  showing the clutching arrangement in a position wherein the cutting tool would be advanced. 
         FIG. 7  is an enlarged sectioned elevation view of the clutching arrangement of the positive feed drill depicted in  FIG. 2  wherein the clutching arrangement is in a retract position. 
         FIG. 8  is an enlarged sectioned elevation view of the clutching arrangement of the positive feed drill depicted in  FIG. 2  wherein the clutching arrangement is in an advance position. 
         FIG. 9  is a perspective view of the lower clutch gear of the clutching arrangement depicted in  FIG. 7 . 
         FIG. 10  is a perspective view of the lower side of the upper clutch gear of the clutching arrangement depicted in  FIG. 7 . 
         FIG. 11  is a perspective view of the upper side of the upper clutch gear depicted in  FIG. 9 . 
         FIG. 12  is a perspective view of a clutch ring of the clutching arrangement depicted in  FIG. 7 . 
         FIG. 13  is an enlarged sectioned view of an upper clutch gear tooth of the upper clutch gear depicted in  FIG. 9  interlocking with a lower clutch gear tooth of the lower clutch gear depicted in  FIG. 12 . 
         FIG. 14  is an exploded perspective view of the drive gear and cutter spindle of the positive feed drill depicted in  FIG. 2 . 
     
    
    
     DETAILED DESCRIPTION OF THE PREFERRED EMBODIMENT 
       FIGS. 2 and 3  illustrate a positive feed drill  100  that is arranged and configured in accordance with certain features, aspects and advantages of an embodiment of the present invention. As illustrated in  FIGS. 2 and 3 , the drill  100  comprises a housing assembly  102 . The illustrated housing assembly  102  comprises a number of components that are secured together to define the outer shell of the drill  100 . 
     An end housing  104  is connected to an air motor housing  106  and the air motor housing  106  is connected to an angle head module  110 . The illustrated angle head module  110  comprises an upper housing subassembly  112 , a middle housing subassembly  114  and a lower housing subassembly  116 . In the illustrated embodiment, the subassemblies  112 ,  114 ,  116  are secured together with threaded fasteners  120 . Other configurations are possible. 
     The housing assembly  102  houses a drive train of the drill  100 . In the illustrated configuration, the drill  100  is pneumatically driven and, as such, the drive train can be considered to start with an air inlet fitting  124 . The air inlet fitting  124  is mounted to the end housing  104  and supplies air to an air motor  126  through a system of air channels contained within the housing assembly  102 . 
     A portion of the air motor  126  is shown in  FIG. 3 . The air motor  126  is positioned within the air motor housing  106  and comprises an output shaft  130 . In the illustrated embodiment, the output shaft  130  comprises splines. Preferably, the air motor  126  is powered by 90-120 psi air pressure, which causes the output shaft to rotate at high rotational speeds with low torque output. 
     The output shaft  130  couples to a planetary gear reducer  132 . The planetary gear reducer  132  decreases the rotational speed from the output shaft and increases the torque output. Different planetary gear reducers can be provided to provide different speed outputs and different torque outputs. Thus, the planetary gear reducers can be interchanged as desired. As shown in  FIG. 3 , the illustrated planetary gear reducer  132  comprises a threaded surface  134 . The threaded surface  134  threads into a bore (not shown) formed in an end of the air motor housing  106 . Thus, the air motor  126  and the planetary gear reducer  132  can be coupled to the air motor housing  106 . 
     The planetary gear reducer  132  comprises an output shaft. The output shaft preferably comprises a splined portion  136 . The planetary gear reducer  132  preferably is received within a recess  140  formed within the angle head module  110 . With reference now to  FIG. 5 , the splined portion  136  of the output shaft is received within a splined portion  142  of a first bevel gear  144 . 
     The first bevel gear  144  drives a second bevel gear  146 . A spur gear  150  is mounted to the second bevel gear  146  such that the spur gear  150  rotates with the second bevel gear  146 . In the illustrated embodiment, a shaft  152  supports the second bevel gear  146  and the spur gear  150 . 
     The spur gear  150  drives a lower clutch gear  154 . As shown in  FIG. 5 , the lower clutch gear  154  is supported by a first stub shaft  156 . In the illustrated embodiment, the first stub shaft  156  supports a bearing  160  and the bearing  160  supports the lower clutch gear  154 . The stub shaft  156  comprises a threaded portion  158 , which is secured in position within the middle housing subassembly  114 . Preferably, the threaded portion  158  threads into position from an outside of the middle housing subassembly  114  such that the enlarged threaded portion  158  is sandwiched between the middle housing subassembly  114  and the lower housing subassembly  116 . Such a configuration secures the shaft  156  and reduces the likelihood of the shaft loosening over time. Other arrangements also can be used to secure the stub shaft  156  in position. The fixed stub shaft  156  provides a strong and sturdy support for the lower clutch gear  154 , which rotates about the fixed stub shaft  156 . Accordingly, such a construction can withstand high side loads. 
     In the illustrated configuration, a retainer ring  162  secures the bearing  160  in place within the lower clutch gear  154 . In other words, the bearing  160  is secured between the retainer ring  162  and a shoulder  163  formed in the lower clutch gear  154 . The shoulder  163  can be integrally formed or can be a separate component, such as a retainer ring, for example but without limitation. The shoulder  163 , however, provides improved reliability due to the decreased number of components. 
     The lower clutch gear  154  drives a drive gear  164 . The drive gear  164  is joined for rotation to a cutter spindle  166  such that rotation of the drive gear  164  results in rotation of the cutter spindle  166 . In one configuration, such as that shown in  FIG. 14 , the drive gear  164  comprises internal keys  170  that match keyways  172  formed in the cutter spindle  166 . Other arrangements also can be used to couple the drive gear  164  for rotation with the cutter spindle  166 . In view of the rotational coupling of the drive gear  164  and the cutter spindle  166 , and in view of the engaged gear coupling of the lower clutch gear  154  and the drive gear  164 , rotation of the lower clutch gear causes rotation of the cutter spindle  166 . 
     With reference now to  FIG. 9 , the lower clutch gear  154  comprises an upwardly-facing surface  180 . The upwardly facing surface  180  comprises a plurality of clutching teeth  182 . While the illustrated configuration comprises four clutching teeth  182  that are generally equally spaced around the upwardly-facing surface  180 , other configurations are possible. Each tooth  182  preferably comprises a sloping trailing surface  184  as well as a sloping leading surface  186 . Preferably, the leading surface  186  slopes at about 7 degrees from perpendicular while the trailing surface  184  slopes at about 45 degrees from perpendicular. The leading surface  186  defines a contact surface in the illustrated configuration. 
     As shown in  FIG. 13 , the lower clutch gear  154  selectively engages with an upper clutch gear  190 .  FIGS. 10 and 11  provide two views of the upper clutch gear  190  in the illustrated embodiment. As illustrated, the upper clutch gear  190  comprises a downwardly-facing surface  192 . The downwardly facing surface  192  comprises a plurality of clutching teeth  194 . In the illustrated configuration, the upper clutch gear  190  comprises four clutching teeth  194  that are generally equally spaced about the downwardly facing surface  192  but other configurations can be used. In addition, while the number of clutching teeth  182  on the lower clutch gear  154  matches the number of clutching teeth  194  on the upper clutch gear  190 , other configurations also are possible. 
     As shown in  FIG. 10 , the clutching teeth  194  of the upper clutch gear  190  comprise a sloping trailing surface  196  and a sloping leading surface  198 . Preferably, the leading surface  198  slopes at an angle of about 7 degrees from perpendicular while the trailing surface  196  slopes at an angle of about 45 degrees. With reference again to  FIG. 13 , the leading surface  198  of the upper clutch gear  190  preferably defines a contract surface that can engage with the leading surface  186  of the lower clutch gear  154 . 
     With reference now to  FIGS. 7 and 8 , the upper clutch gear  190  is carried on a floating shaft  200 . As shown, the floating shaft  200  comprises an upper end  202  and a lower end  204 . Intermediate of the upper end  202  and the lower end  204 , the illustrated floating shaft  200  comprises a rib  206 . While the illustrated rib  206  is integrated into the floating shaft  200 , other arrangements may utilize a separable component. The integrated rib  206 , however, improves reliability and decreases manufacturing costs. 
     A bearing  210  is positioned within the upper clutch gear  190 . The bearing  210  is mounted over the lower end  204  of the floating shaft  200 . As with the lower clutch gear  154 , the upper clutch gear  190  comprises a shoulder  212  and a retainer ring  214  secures the bearing  210  within the upper clutch gear  190  between the shoulder  212  and the retainer ring  214 . 
     As shown in  FIG. 7 , the lower end  204  of the floating shaft  200  protrudes below the upper clutch gear  190 . With reference now to  FIG. 8 , the floating shaft  200  comprises an axis A 2  that generally aligns within an axis A 1  of the fixed stub shaft  156 . In the illustrated construction, absolutely perfect alignment of the two axes A 1 , A 2  is unnecessary and slight misalignments can be accommodated. In addition, because the floating shaft  200  is separate of the stub shaft  156 , side loads encountered by the stub shaft  156  are not transferred from the stub shaft  156  to the floating shaft  200 , which allows significantly more freedom in movement for the floating shaft  200  relative to the single shaft construction described above. 
     The floating shaft  200  moves vertically generally along the axis A 2 . In its lowermost position, the lower end  204  of the floating shaft  200  abuts against an upper end  216  of the lower clutch gear  154 . With reference to  FIGS. 8 and 13 , the length of the shafts  156 ,  200  between the rib  206  of the floating shaft  200  and a shoulder  218  of the stub shaft  156  is slightly longer than a sum of (1) the distance between the downwardly facing surface  192  and the shoulder  212  of the upper clutch gear  190 , (2) the distance between the upwardly facing surface  180  and the shoulder  163  of the lower clutch gear  154  and (3) the largest height of the teeth  182 ,  194 . In other words, when the floating shaft  202  moves into contact with the stub shaft  156 , a gap  220  is defined between the teeth  182  and the downwardly facing surface  192  and a gap  222  is defined between the teeth  194  and the upwardly facing surface  180 . In some configurations, the gap ranges from about 0.010 inch to about 0.020 inch. 
     Due to the gaps  220 ,  222  Thus, the contact surface area between the lower clutch gear  154  and the upper clutch gear  190  is greatly reduced compared to a construction in which the gear teeth engage along the full length. In other words, only portions of the contact surfaces  186 ,  198  are in contact when the clutch gears  154 ,  190  are engaged. Accordingly, the illustrated configuration results in less friction, which reduces binding and encourages smooth engagement and disengagement of the clutch gears  154 ,  190 . Moreover, the illustrated configuration provides for smoother transmission of torque and rotation from the lower clutch gear  154  to the upper clutch gear  190 . Furthermore, the angled contact surfaces (e.g., the contact surfaces are angled at about 7 degrees, as discussed above) create an upward force (F 1 ) as shown in  FIG. 13 , which upward force F 1  creates a natural tendency for the clutch gears  154 ,  190  to separate. Thus, the angled contact surfaces also encourage improved disengagement. 
     With reference again to  FIGS. 7 and 8 , a biasing member  230  can be interposed between the lower clutch gear  154  and the upper clutch gear  190 . Any suitable biasing member  230  can be used. In some embodiments, the biasing member  230  is a compression spring. The compression spring  230  in the illustrated configuration extends between the inner race of the lower bearing  160  and the inner race of the upper bearing  210  and the compression spring  230  generally encircles a portion of the stub shaft  156  and a portion of the floating shaft  200 . Such a configuration allows use of a smaller diameter compression spring  230 . Other configurations also can be used, such as larger diameter compression springs, which may help to maintain better alignment of the axes A 1 , A 2 . 
     A bearing  232  supports the upper end  202  of the floating shaft  200 . The bearing  232  can be secured in position with a retaining ring  233 . In the illustrated configuration, the bearing  232  comprises a needle bearing. The bearing  232  allows relatively free axial and rotational movement of the floating shaft  200 . A clutch ring  234  supports the bearing  232 , which supports the floating shaft  200 . The clutch ring  234  can be secured in position with one or more mechanical fasteners (e.g., screws). 
     As shown in  FIGS. 7 and 8 , the clutch ring  234  generally encloses a region defined by a recess  236  within the upper housing subassembly  112 . The enclosed recess  236  contains a piston  240  that divides the recess  236  into an upper chamber  242  and a lower chamber  244 . The piston  240  is secured to or supported by the upper end  202  of the floating shaft  200 . In some embodiments, the piston  240  merely bears against the extreme upper end  202  of the floating shaft  200 . 
       FIG. 8  shows a port  246  that can be used to alter a pressure within the upper chamber. When air is supplied to the upper chamber  242 , the piston  240  moves downward within the recess  236 , which moves the floating shaft  200  downward. The rib  206  on the floating shaft  200  bears against the bearing  210 , which causes the bearing  210  and associated upper clutch gear  190  to move downward and compress the biasing member  230 . Thus, the upper clutch gear  190  moves toward, and eventually into engagement with, the lower clutch gear  154 . As the pressure in the upper chamber  242  decreases, the biasing force of the biasing member  230  overcomes the pressure in the upper chamber  242 , which allows the piston  240  to travel upward within the recess  236  and which allows the upper clutch gear  190  to move away from the lower clutch gear  154 . 
     With reference now to  FIGS. 11 and 12 , the upper clutch gear  190  comprises an upwardly facing surface  250 . The upwardly facing surface  250  comprises a plurality of upwardly extending teeth  252 . Similarly, the clutch ring  234  comprises a downwardly facing surface  254  and the downwardly facing surface  254  comprises a plurality of teeth  256 . The two sets of teeth  252 ,  256  are arranged and configured to stop rotation of the upper clutch gear  190  when the teeth  252 ,  256  are engaged. While the illustrated configuration comprises sets of eight, generally symmetrically disposed teeth, other rotation-limiting configurations also can be used. 
     With reference again to  FIGS. 5 and 6 , the upper clutch gear  190  meshes with a feed gear  260 . The upper clutch gear  190  translates axially along the axis A 2  relative to the feed gear  260  but remains engaged with the feed gear  260  during this axial translation. The feed gear  260  has a threaded inner surface that receives a threaded portion  262  of the cutter spindle  166  (see  FIG. 14 ). 
     In operation, the drive gear  164  is constantly driven by the lower clutch gear  154 . The cutter spindle  166  is fixed for rotation with the drive gear  164 . Thus, the cutter spindle  166  constantly rotates so long as the lower clutch gear  154  is rotating. When the upper clutch gear  190  is secured against rotation by the clutch ring  234 , the upper clutch gear  190  and the feed gear  260  are secured against rotation. Because the cutter spindle  166  has a threaded connection with the feed gear, rotation of the cutter spindle  166  relative to the feed gear  260  causes the cutter spindle to move upward relative to the feed gear  260 . On the other hand, when the upper clutch gear  190  engages with the lower clutch gear  154 , the upper clutch gear  190  rotates at the same speed as the lower clutch gear  154 . The gear ratio between the upper clutch gear  190  and the feed gear  260  is such that the feed gear  260  turns faster than the drive gear  164  and the cutter spindle  166 . Because the feed gear  260  is rotating faster than the cutter spindle  166 , the cutter spindle  166  moves down relative to the feed gear  260 . The gear ratios can be varied to obtain the desired movements of the cutter spindle  166  relative to the feed gear  260 . 
     In other words, the downward movement of the cutter spindle  166  results because of a clockwise differential rotation between the drive gear  164  and the feed gear  260  and the left-handed threaded engagement between the feed gear  260  and the cutter spindle  166 . The downward movement will occur at a constant feed rate but altering the gear ratios between the upper clutch gear  190  and the feed gear  260 , for example, can be used to adjust the fed rate. The feed rate (F) in inches per revolution (in./In.) is determined by the following equation: F=[(R 1 /R 2 )−1]/T, where R 1 =N 1 /N 2  and R 2 =N 3 /N 4 , where N 1 =number of upper clutch gear teeth, N 2 =number of feed gear teeth, N 3 =number of lower clutch gear teeth, N 4 =number of drive gear teeth, and T=number of threads per inch of the cutter spindle. In the illustrated configuration, because the number of teeth of the lower clutch gear equals the number of teeth of the drive gear, then R 2 =1 and the equation reduces to F=[R 1 −1]/T. In one preferred configuration, the threaded portion  262  of the cutter spindle  166  comprises about 40 threads per inch. Such a configuration provides a wide range of feed rates that can be achieved simply by changing the feed gear  260  and/or the upper clutch gear  190 . 
     Although the present invention has been described in terms of a certain embodiment, other embodiments apparent to those of ordinary skill in the art also are within the scope of this invention. Thus, various changes and modifications may be made without departing from the spirit and scope of the invention. For instance, various components may be repositioned as desired. Moreover, not all of the features, aspects and advantages are necessarily required to practice the present invention. Accordingly, the scope of the present invention is intended to be defined only by the claims that follow.