Patent Publication Number: US-2021162574-A1

Title: Impact tool

Description:
CROSS-REFERENCE TO RELATED APPLICATIONS 
     This application claims priority to co-pending U.S. Provisional Patent Application No. 62/831,779 filed on Apr. 10, 2019, the entire content of which is incorporated herein by reference. 
    
    
     FIELD OF THE INVENTION 
     The present invention relates to power tools, and more particularly to impact tools. 
     BACKGROUND OF THE INVENTION 
     Impact tools use an impact mechanism, such as a rotary impact mechanism, to impart repeated rotational impacts to a workpiece to perform work on the workpiece. 
     SUMMARY OF THE INVENTION 
     The present invention provides, in one aspect, an impact tool comprising a housing having a handle portion defining a first axis, a motor supported by the housing and defining a motor axis, and a rotary impact mechanism arranged on a second axis that is perpendicular to the first axis. The rotary impact mechanism is configured to convert a continuous rotational input from the motor to consecutive rotational impacts upon a workpiece. The rotary impact mechanism includes a chamber containing a hydraulic fluid, an anvil positioned at least partially within the chamber, and a hammer for imparting the consecutive rotational impacts upon the anvil. The hydraulic fluid is configured to attenuate a noise of the rotary impact mechanism that is created by the hammer impacting the anvil. The impact tool further comprises a gear train that receives torque from the motor and includes a rotational input that transfers torque to the rotary impact mechanism. 
     The present invention provides, in another aspect, an impact tool comprising a housing having a handle portion defining a first axis, a motor supported by the housing, and a rotary impact mechanism arranged on the first axis and configured to receive torque from the motor. The rotary impact mechanism is configured to convert a continuous rotational input from the motor to consecutive rotational impacts upon a workpiece. The rotary impact mechanism includes a chamber containing a hydraulic fluid, an anvil positioned at least partially within the chamber, and a hammer for imparting the consecutive rotational impacts upon the anvil. The hydraulic fluid is configured to attenuate a noise of the rotary impact mechanism that is created by the hammer impacting the anvil. The impact tool further comprises an output member for receiving torque from the rotary impact mechanism. The output member is arranged on a second axis that is perpendicular to the first axis. 
     Other features and aspects of the invention will become apparent by consideration of the following detailed description and accompanying drawings. 
    
    
     
       BRIEF DESCRIPTION OF THE DRAWINGS 
         FIG. 1  is a schematic view of an impact tool in accordance with an embodiment of the invention. 
         FIG. 2  is a schematic view of the impact tool of  FIG. 1  in accordance with an embodiment of the invention. 
         FIG. 3  is a schematic view of the impact tool of  FIG. 1  in accordance with another embodiment of the invention. 
         FIG. 4A  is an assembled, cross-sectional view of a first impact mechanism of the impact tool of  FIG. 1  in accordance with an embodiment of the invention. 
         FIG. 4B  is an exploded perspective view of a first impact mechanism of  FIG. 4A . 
         FIG. 5  is a cross-sectional view of an output shaft of the impact mechanism shown in  FIG. 4A . 
         FIG. 6  is an assembled, cross-sectional view of a portion of the impact mechanism of  FIG. 4A . 
         FIG. 7  is a perspective view of a second impact mechanism in accordance with another embodiment of the invention. 
         FIG. 8  is an exploded view of the impact mechanism of  FIG. 7 . 
         FIG. 9  is a cross-sectional view of the impact mechanism of  FIG. 7 , taken along section  4 - 4  in  FIG. 7 . 
         FIG. 10  is a cross-sectional view of the impact mechanism of  FIG. 7 , illustrating an overview of a retraction phase. 
         FIGS. 11A-11C  are cross-sectional views of the impact mechanism of  FIG. 7 , illustrating operation of the retraction phase. 
         FIGS. 12A-12C  are cross-sectional views of the impact mechanism of  FIG. 7 , illustrating operation of a return phase. 
         FIG. 13  is a schematic view of a first gear train arrangement of the impact tool embodiment of  FIG. 3 . 
         FIG. 14  is a schematic view of a second gear train arrangement of the impact tool embodiment of  FIG. 3 . 
         FIG. 15  is a schematic view of a third gear train arrangement of the impact tool embodiment of  FIG. 3 . 
         FIG. 16  is a schematic view of a fourth gear train arrangement of the impact tool embodiment of  FIG. 3 . 
         FIG. 17  is a schematic view of a fifth gear train arrangement of the impact tool embodiment of  FIG. 3 . 
         FIG. 18  is a schematic view of a sixth gear train arrangement of the impact tool embodiment of  FIG. 3 . 
         FIG. 19  is a schematic view of a seventh gear train arrangement of the impact tool embodiment of  FIG. 3 . 
         FIG. 20  is a schematic view of an eighth gear train arrangement of the impact tool embodiment of  FIG. 3 . 
         FIG. 21  is a schematic view of a ninth gear train arrangement of the impact tool embodiment of  FIG. 3 . 
         FIG. 22  is a schematic view of a tenth gear train arrangement of the impact tool embodiment of  FIG. 3 . 
         FIG. 23  is a schematic view of an eleventh gear train arrangement of the impact tool embodiment of  FIG. 3 . 
         FIG. 24  is a schematic view of a twelfth gear train arrangement of the impact tool embodiment of  FIG. 3 . 
         FIG. 25  is a schematic view of a thiteenth gear train arrangement of the impact tool embodiment of  FIG. 3 . 
         FIG. 26  is a schematic view of a fourteenth gear train arrangement of the impact tool embodiment of  FIG. 3 . 
         FIG. 27  is a schematic view of a fifteenth gear train arrangement of the impact tool embodiment of  FIG. 3 . 
         FIG. 28  is a schematic view of a sixteenth gear train arrangement of the impact tool embodiment of  FIG. 3 . 
         FIG. 29  is a schematic view of an seventeenth gear train arrangement of the impact tool embodiment of  FIG. 3 . 
         FIG. 30  is a schematic view of a eighteenth gear train arrangement of the impact tool embodiment of  FIG. 3 . 
         FIG. 31  is a schematic view of a nineteenth gear train arrangement of the impact tool embodiment of  FIG. 3 . 
         FIG. 32  is a schematic view of the impact tool of  FIG. 1  in accordance with an embodiment of the invention. 
     
    
    
     Before any embodiments of the invention are explained in detail, it is to be understood that the invention 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 invention 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 
       FIG. 1  illustrates a right angle impact tool  10  having a housing  14  with a handle portion  18  defining a first, handle axis  22  that is perpendicular to a second, output axis  26  of the tool  10 .  FIG. 2  schematically illustrates a first embodiment of the impact tool  10 . Specifically, in the embodiment of  FIG. 2 , the right angle impact tool  10  includes a motor  30 , a gear train  34  receiving torque from the motor  14 , a rotary impact mechanism  38  that receives torque from the gear train  34 , and a rotational input  42  that receives torque from the impact mechanism  38  and drives an output member  43  defining the output axis  26 . The output member  43  has a hexagonal receptacle therein for receipt of a tool bit. In some embodiments, instead of a hexagonal receptacle, the output member  43  includes a square drive, a hex drive, or a spline drive. In some embodiments, instead of a rotary impact mechanism  38 , a mechanical impact mechanism may be used. 
     As shown schematically in  FIG. 32 , in a specific implementation of the embodiment of  FIG. 2 , the rotary impact mechanism  38  is arranged on the first axis  22 , the gear train  34  is a first gear train, and the right angle impact tool  10  includes a second gear train  45  configured to transfer torque from the rotary impact mechanism  38  to the output member  43 . The second gear train  45  includes a pinion  47  that is coupled for co-rotation about the first axis  22  with an output  48  of the rotary impact mechanism  38 . The second gear train  45  also includes a ring gear  49  that is engaged with the pinion  47 , coupled to the output member  43 , and rotatable about the output axis  26 , such that the ring gear  49  functions as the rotational input  42  to transfer torque to the output member  43 . As also shown in  FIG. 2 , the rotary impact mechanism  38  is coaxial with an output  52  of the first gear train  34  and is also coaxial with the pinion  47 , which is the input of the second gear train  45 . 
       FIG. 3  schematically illustrates a second embodiment of the impact tool  10 . In the embodiment of  FIG. 3 , the right angle impact tool  10  includes the motor  30  and the gear train  34 , but instead of receiving torque from the impact mechanism  38 , the rotational input  42  is the final element of the gear train  34 . Further, unlike the embodiment of  FIG. 2 , in the embodiment of  FIG. 3 , the impact mechanism  38  is downstream of the rotational input  42  and coaxial with the output axis  26 , such that the rotational input  42  at the end of the gear train  34  provides rotational input to the impact mechanism  38 .  FIGS. 1-3  also illustrate a trigger  44  to actuate the motor  30 . 
       FIGS. 4-6  illustrate a first embodiment 1000 of the rotary impact mechanism  38  and  FIGS. 7-12C  illustrate a second embodiment 2000 of the rotary impact mechanism  38 . Specifically, with reference to  FIGS. 4A and 4B , the rotary impact mechanism  1000  includes a hammer or cylinder  1026  coupled for co-rotation with an output of the gear train  34  ( FIG. 2 ) or rotational input  42  ( FIG. 3 ). The rotary impact mechanism  1000  also includes a camshaft  1038 , the purpose of which is explained in detail below, attached to the cylinder  1026  for co-rotation therewith about a longitudinal axis  1034 . Although the camshaft  1038  is shown as a separate component from the cylinder  1026 , the camshaft  1038  may alternatively be integrally formed as a single piece with the cylinder  1026 . 
     With reference to  FIG. 6 , the cylinder  1026  includes a cylindrical interior surface  1042 , which partly defines a cavity  1046 , and a pair of radially inward-extending protrusions  1050  extending from the interior surface  1042  on opposite sides of the longitudinal axis  1034 . In other words, the protrusions  1060  are spaced from each other by 180 degrees. The rotary impact mechanism  1000  further includes an anvil or output shaft  1054  ( FIGS. 4-5 ), a rear portion  1058  of which is disposed within the cavity  1046  and a front portion  1062 . In the embodiment of  FIG. 3 , the front portion  1062  extends from the housing  14  and includes a hexagonal receptacle  1066  ( FIG. 5 ) therein for receipt of a tool bit. 
     The rotary impact mechanism  1000  also includes a pair of pulse blades  1070  ( FIGS. 4 and 6 ) protruding from the output shaft  1054  to abut the interior surface  1042  of the cylinder  1026  and a pair of ball bearings  1074  are positioned between the camshaft  1038  and the respective pulse blades  1070 . The output shaft  1054  has dual inlet orifices  1078  ( FIG. 5 ), each of which extends between and selectively fluidly communicates the cavity  1046  and a separate high pressure cavity  1082  within the output shaft  1054 . The output shaft  1054  also includes dual outlet orifices  1086  ( FIG. 5 ) that are variably obstructed by an orifice screw  1090  ( FIGS. 4A and 4B ), thereby limiting the volumetric flow rate of hydraulic fluid that may be discharged from the output shaft cavity  1082 , through the orifices  1086 , and to the cylinder cavity  1046 . The camshaft  1038  is disposed within the output shaft cavity  1082  and is configured to selectively seal the inlet orifices  1078 . 
     With reference to  FIGS. 4A and 4B , the cavity  1046  is in communication with a bladder cavity  1094 , defined by an end cap  1098  attached for co-rotation with the cylinder  1026  (collectively referred to as a “cylinder assembly”), located adjacent the cavity  1046  and separated by a plate  1102  having apertures  1108  for communicating hydraulic fluid between the cavities  1046 ,  1094 . A collapsible bladder  1104  having an interior volume  1142  filled with a gas, such as air at atmospheric temperature and pressure, is positioned within the bladder cavity  1094 . The bladder  1104  is configured to be collapsible to compensate for thermal expansion of the hydraulic fluid during operation of the rotary impact mechanism  1000 , which can negatively impact performance characteristics. 
     As shown in  FIGS. 4A and 4B , prior to the end cap  1098  being threaded into the cylinder  1026 , the collapsible bladder  1104  is bent into an annular shape and set into the bladder cavity  1094 , which is also annular. Alternatively, the collapsible bladder  1104  can take any shape that permits the bladder to be set by fitment with the cavity  1094  and still effectively compensate for thermal expansion of the hydraulic fluid in the cavities  1046 ,  1094 . After the end cap  1098  is threaded to the cylinder  1026 , the collapsible bladder  1104  is trapped via fitment within the cavity  1094 , having its annular shape maintained by the shape of the cavity  1094  itself. 
     In operation, upon activation of the motor  30  (e.g., by depressing a trigger  44 ), torque from the motor  30  is transferred to the cylinder  1026  via the gear train  34  ( FIG. 2 ) or rotational input  42  ( FIG. 3 ), causing the cylinder  1026  and camshaft  1038  to rotate in unison relative to the output shaft  1054  until the protrusions  1050  on the cylinder  1026  impact the respective pulse blades  1070  to deliver a first rotational impact to the output shaft  1054 . Just prior to the first rotational impact, the inlet orifices  1078  are blocked by the camshaft  1038 , thus sealing the hydraulic fluid in the output shaft cavity  1082  at a relatively high pressure, which biases the ball bearings  1074  and the pulse blades  1070  radially outward to maintain the pulse blades  1070  in contact with the interior surface  1042  of the cylinder. For a short period of time following the initial impact between the protrusions  1050  and the pulse blades  1070  (e.g., 1 ms), the cylinder  1026  and the output shaft  1054  rotate in unison. 
     Also at this time, hydraulic fluid is discharged through the outlet orifices  1086  at a relatively slow rate determined by the position of the orifice screw  1090 , thereby damping the radial inward movement of the pulse blades  1070 . Once the ball bearings  1074  have displaced inward by a distance corresponding to the size of the protrusions  1050 , the pulse blades  1070  move over the protrusions  1050  and torque is no longer transferred to the output shaft  1054 . The camshaft  1038  rotates independently of the output shaft  1054  again after this point, and moves into a position where it no longer seals the inlet orifices  1078  thereby causing fluid to be drawn into the output shaft cavity  1082  and allowing the ball bearings  1074  and pulse blades  1070  to displace radially outward once again. The cycle is then repeated as the cylinder  1026  continues to rotate, with torque transfer occurring twice during each 360 degree revolution of the cylinder. In this manner, the output shaft  1054  receives discrete pulses of torque from the cylinder  1026 . 
     As noted above,  FIGS. 7-12C  illustrate a second embodiment 2000 of the rotary impact mechanism  38 . Specifically, with reference to  FIGS. 7-9 , the rotary impact mechanism  2000  includes an anvil  2026 , a hammer  2030 , and a cylinder  2034 . A driven end  2038  of the cylinder  2034  is coupled to the electric motor  2022  to receive torque therefrom, causing the cylinder  2034  to rotate. The cylinder  2034  at least partially defines a chamber  2042  ( FIG. 9 ) that contains an incompressible fluid (e.g., hydraulic fluid, oil, etc.). The chamber  2042  is sealed and is also partially defined by an end cap  2046  secured to the cylinder  2034 . The hydraulic fluid in the chamber  2042  reduces the wear and the noise of the rotary impact mechanism  2000  that is created by impacting the hammer  2030  and the anvil  2026 . 
     With continued reference to  FIGS. 7-9 , the anvil  2026  is positioned at least partially within the chamber  2042  and includes an output shaft  2050 . In the embodiment of  FIG. 3 , the output shaft  2050  includes a hexagonal receptacle  2054  therein for receipt of a tool bit. In some embodiments, instead of a hexagonal receptacle, the output shaft  2050  includes a square drive, a hex drive, or a spline drive. The output shaft  2050  extends from the chamber  2042  and through the end cap  2046 . The anvil  2026  rotates about a rotational axis  2058  defined by the output shaft  2050 . 
     With continued reference to  FIGS. 7-9 , the hammer  2030  is positioned at least partially within the chamber  2042 . The hammer  2030  includes a first side  2062  facing the anvil  2026  and a second side  2066  opposite the first side  2062 . The hammer  2030  further includes hammer lugs  2070  and a central aperture  2074  extending between the sides  2062 ,  2066 . As discussed in greater detail below, the central aperture  2074  permits the hydraulic fluid in the chamber  2042  to pass through the hammer  2030 . The hammer lugs  2070  correspond to lugs  2078  formed on the anvil  2026 . The rotational rotary impact mechanism  2000  further includes hammer alignment pins  2082  and a hammer spring  2086  (i.e., a first biasing member) positioned within the chamber  2042 . The hammer alignment pins  2082  are coupled to the cylinder  2034  and are received within corresponding grooves  2090  formed on an outer circumferential surface  2094  of the hammer  2030  to rotationally unitize the hammer  2030  to the cylinder  2034  such that the hammer  2030  co-rotates with the cylinder  2034 . The pins  2082  also permit the hammer  2030  to axially slide within the cylinder  2034  along the rotational axis  2058 . In other words, the hammer alignment pins  2082  slide within the grooves  2090  such that the hammer  2030  is able to translate along the axis  2058  relative to the cylinder  2034 . The hammer spring  2086  biases the hammer  2030  toward the anvil  2026 . 
     The impact mechanism  2000  further defines a trip torque, which determines the reactionary torque threshold required on the anvil  2026  before an impact cycle begins. In one embodiment, the trip torque is equal to the sum of the torque due to seal drag, the torque due to the spring  2086 , and the torque due to the difference in rotational speed of the hammer  2030  and the anvil  2026 . In particular, the seal drag torque is the static friction between the O-ring and the anvil  2026 . The spring torque contribution to the total trip torque is based on, among other things, the spring rate of the spring  2086 , the height of the lugs  2070 , and the coefficient of friction between the anvil lugs  2078  and the hammer lugs  2070 . The torque from the difference in rotational speed of the anvil  2026  and the hammer  2030  is included in the torque calculation during impaction only, and has little to no effect on determining the trip torque threshold (i.e., is the damping force of the fluid rapidly moving through the orifice  2122 ). In some embodiments, the trip torque is within a range between approximately 10 in-lbf and approximately 30 in-lbf. In other embodiments, the trip torque is greater than 20 in-lbf. Increasing the trip torque increases the amount of time the hammer  2030  and the anvil  2026  are co-rotating (i.e., in a continuous drive). 
     With reference to  FIGS. 8 and 9 , the rotary impact mechanism  2000  further includes a valve assembly  2098  positioned within the chamber  2042  that allows for various fluid flow rates through the valve assembly  2098 . As described in greater detail below, the valve assembly  2098  adjusts the flow of the hydraulic fluid in the chamber  2042  to decrease the amount of time it takes the hammer  2030  to return to the anvil  2026 . In other words, the valve assembly  2098  reduces the time it takes to complete a single impact cycle. In particular, the flow rate through the valve assembly  2098  varies as the hammer  2030  translates within the cylinder  2034  along the axis  2058 . The valve assembly  2098  includes a valve housing  2102  (e.g., a cupped washer), a valve (e.g., an annular disc  2106 ), and a spring  2110  (i.e., a second biasing member) positioned between the valve housing  2102  and the disc  2106 . The valve housing  2102  includes a rear aperture  2108  and defines a cavity  2114  in which the disc  2106  and the spring  2110  are positioned. The spring  2110  biases the disc  2106  toward the hammer  2030 , and the hammer spring  2086  biases the valve housing  2102  toward the hammer  2030 . In particular, the valve housing  2102  includes a circumferential flange  2118  against which the spring  2086  is seated to bias the valve housing  2102  toward the hammer  2030 . In other words, the valve housing  2102  is at least partially positioned between the spring  2086  and the hammer  2030 . With reference to  FIG. 9 , the hammer  2030  defines a recess  2120  and the valve assembly  2102  is at least partially received with the recess  2120 . 
     With reference to  FIG. 8 , the disc  2106  includes a central aperture  2122  and at least one auxiliary opening  2126 . The aperture  2122  of the disc  2106  is in fluid communication with the aperture  2074  formed in the hammer  2030  ( FIG. 9 ). In the illustrated embodiment, the auxiliary openings  2126  are positioned circumferentially around the aperture  2122  and are formed as grooves in the outer periphery of the disc  2106 . In other embodiments, the auxiliary openings may be apertures formed in any location on the disc  2106 . In further alternative embodiments, the auxiliary opening may be formed as part of the central aperture  2122  to form one single aperture with less than the entire aperture in fluid communication with the aperture  2074  during at least a portion of operation. In other words, the auxiliary openings may be formed as cutouts or scallops contiguous with the central aperture  2122  that are sometimes blocked and sometimes opened by the hammer  2066  during operation of the impact mechanism  2000 . 
     With continued reference to  FIG. 9 , the central aperture  2122  defines an orifice diameter  2123  and the hammer  2030  defines a hammer diameter  2031 . A ratio R of the hammer diameter  2031  to the orifice diameter  2123  is large and beneficially allows less reliance on tolerances and removes a feature that requires calibration. Additionally, the large ratio R makes leak paths less significant relative to fluid moved by the hammer  2030 . Furthermore, the impact tool  2010  has a greater total amount of fluid contained within the rotary impact mechanism  2000 . As such, a greater volume of fluid is moved with each stroke of the hammer  2030 . In one embodiment, the total fluid in the rotary impact mechanism  2000  is greater than approximately 18,000 cubic mm (18 mL). In another embodiment, the total fluid in the rotary impact mechanism  2000  is greater than approximately 20,000 cubic mm (20 mL). In another embodiment, the total fluid in the rotary impact mechanism  2000  is greater than approximately 22,000 cubic mm (22 mL). Likewise, the amount of fluid moved with each stroke of the hammer  2030  in one embodiment is greater than approximately 1000 cubic mm (1 mL). In another embodiment, the fluid moved with each stroke of the hammer  2030  is greater than approximately 1250 cubic mm (1.25 mL). In another embodiment, the fluid moved with each stroke of the hammer  2030  is approximately 1500 cubic mm (1.5 mL). A greater amount of fluid moved with each stroke of the hammer  2030  results in fluid leak paths having a proportionally smaller effect on the performance of the tool  2010 . Additionally, by moving a greater area of fluid, the rotary impact mechanism  2000  experiences less pressure for the same amount of torque. 
     The disc  2106  is moveable between a first position ( FIG. 9 ) that permits a first hydraulic fluid flow rate in the chamber  2042  from the second side  2066  to the first side  2062  of the hammer  2030 , and a second position ( FIG. 12B ) that permits a second hydraulic fluid flow rate in the chamber  2042  from the first side  2062  to the second side  2066  of the hammer  2030 . In the illustrated embodiment, the second fluid flow rate is greater than the first fluid flow rate, and the disc  2106  is in the second position ( FIG. 12B ) when the hammer  2030  moves along the axis  2058  toward the anvil  2026 . In particular, the hammer  2030  defines a rear surface  2130  on the second side  2066  and the disc  2106  engages the rear surface  2130  when the disc  2106  is in the first position ( FIG. 9 ). In contrast, the disc  2106  is spaced from the rear surface  2130  when the disc  2106  is in the second position ( FIG. 12B ). 
     With reference to  FIGS. 8 and 9 , when the disc  2106  is in the first position, the hydraulic fluid flows through the central aperture  2122  but does not flow through the auxiliary openings  2126 . In other words, when the valve assembly  2098  is in a closed state ( FIG. 9 ), the spring  2110  biases the disc  2106  against the hammer  2030 , blocking the auxiliary openings  2126  with the rear surface  2130  while the central opening  2122  remains in fluid communication with the aperture  2074  formed in the hammer  2030  ( FIG. 9 ). When the disc  2106  is in the second position, the hydraulic fluid flows through the central aperture  2122  and the auxiliary openings  2126 . In other words, when the valve assembly  2098  is in an open state ( FIG. 12B ), the disc  2106  separates from the hammer  2030 , which unblocks the auxiliary openings  2126  and places the auxiliary openings  2126  in fluid communication with the central aperture  2074  of the hammer  2030 . As a result, the valve assembly  2098  provides an increased hydraulic fluid flow rate in one direction, which allows faster fluid pressure equalization when the hammer  2030  is translating along the axis  2058  toward the anvil  2026 . 
     With continued reference to  FIGS. 8 and 9 , the impact tool  2010  further includes an expansion chamber  2134  defined in the cylinder  2034 . The expansion chamber  2134  contains the hydraulic fluid and is in fluid communication with the chamber  2042  by a passageway  2138  (e.g., a pin hole) formed within the cylinder  2034 . A plug  2142  is positioned within the expansion chamber  2134  and is configured to translate within the expansion chamber  2134  to vary a volume of the expansion chamber  2134 . In other words, the plug  2142  moves with respect to the cylinder  2134  to vary the volume of the expansion chamber  2134 . The size of the passageway  2138  is minimized to restrict flow between the expansion chamber  2134  and the chamber  2142  and to negate the risk of large pressure developments over a short period of time, which may otherwise cause significant fluid flow into the expansion chamber  2134 . In some embodiments, the diameter of the passageway  2138  is within a range between approximately 0.4 mm and approximately 0.6 mm. In further embodiments, the diameter of the passageway  2138  is approximately 0.5 mm. In the illustrated embodiment, the plug  2142  includes an annular groove  2146  and an O-ring  2150  positioned within the annular groove  2146 . The O-ring  2150  seals the sliding interface between the plug  2142  and the expansion chamber  2134 . As such, the plug  2142  moves axially within the expansion chamber  2134  to accommodate changes in temperature and/or pressure resulting in the expansion or contraction of the fluid within the sealed rotational rotary impact mechanism  2000 . As such, a bladder or the like compressible member is not required in the cylinder  2034  to accommodate pressure changes. 
     Over extended periods of use, the output torque of the rotary impact mechanism  2000  may degrade because the fluid within the sealed rotational rotary impact mechanism  2000  generates heat and as the temperature increases, the fluid viscosity changes. A fluid with a higher viscosity index (VI) is utilized to reduce the change in viscosity due to changes in temperature, thereby providing more consistent performance. In one embodiment, the fluid viscosity index is greater than approximately 2035. In another embodiment, the fluid viscosity index is greater than approximately 2080. In another embodiment, the fluid viscosity index is within a range between approximately 2080 and approximately 2110. In the embodiment of the impact mechanism  2000 , the impact tool  10  includes a temperature sensor that senses the temperature of the fluid within the rotary impact mechanism  2000  and communicates the fluid temperature to a controller. The controller is configured to then electrically compensate for changing fluid temperature in order to output consistent torque at different temperatures. 
     During operation of the impact mechanism  2000 , the hammer  2030  and the cylinder  2034  rotate together and the hammer lugs  2070  rotationally impact the corresponding anvil lugs  2078  to impart consecutive rotational impacts to the anvil  2026  and the output shaft  2050 . When the anvil  2026  stalls, the hammer lugs  2070  ramp over and past the anvil lugs  2078 , causing the hammer  2030  to translate away from the anvil  2026  against the bias of the hammer spring  2086 .  FIG. 10  illustrates an overview of a hammer retraction phase, and  FIGS. 11A-11C  illustrate step-wise operation of the retraction phase.  FIG. 11A  illustrates the rotary impact mechanism  2000  when the hammer lugs  2070  first contact the anvil lugs  2078 .  FIG. 11B  illustrates the rotary impact mechanism  2000  when the hammer  2030  begins to translate away from the anvil  2026 . As the hammer  2030  moves away from the anvil  2026 , the hydraulic fluid in the chamber  2042  on the first side  2062  of the hammer  2030  is at a low pressure while the hydraulic fluid in the chamber  2042  on the second side  2066  of the hammer  2030  is at a high pressure ( FIG. 10 ). In addition, the valve assembly  2098  translates with the hammer  2030 , away from the anvil  2026 . The hydraulic fluid flows from the second side  2066  to the first side  2062  by traveling through the central aperture  2122  of the disc  2106  and the hammer aperture  2074 . At the end of the retraction phase ( FIG. 11C ), the hammer spring  2086  is compressed and the hammer lugs  2070  have almost rotationally cleared the anvil lugs  2078 . 
     Once the hammer lugs  2070  rotationally clear the anvil lugs  2078 , the spring  2086  biases the hammer  2030  back towards the anvil  2026  in a hammer return phase ( FIG. 12A-12C ).  FIG. 12A  illustrates the rotary impact mechanism  2000  when the hammer  2030  begins to translate toward the anvil  2026 . As the hammer  2030  moves toward the anvil  2026 , the hydraulic fluid in the chamber  2042  on the first side of the hammer  2030  is at a nominal pressure while the hydraulic fluid in the chamber  2042  on the second side  2066  of the hammer  2030  is at a low pressure ( FIG. 12A ).  FIG. 12B  illustrates the rotary impact mechanism  2000  with the valve assembly  2098  in the open state as the hammer  2030  translates toward the anvil  2026 . The hammer spring  2086  keeps the flange  2118  of the valve housing  2102  in contact with the rear surface  2130  of the hammer  2030  as the disc  2106  separates from the rear surface  2130  due to the pressure differential between the two sides  2062 ,  2066  of the hammer  2030 . 
     With the valve disc  2106  unseated from the hammer  2030 , the auxiliary openings  2126  are placed in fluid communication with the hammer aperture  2074 , thereby providing for additional fluid flow through the valve assembly  2098 . In other words, the disc  2106  deflects away from the hammer  2030  as the hammer  2030  is returning toward the anvil  2026 , which creates additional fluid flow through the valve assembly  2098 . Once the hammer  2030  has axially returned to the anvil  2026 , the valve assembly  2098  returns to the closed state ( FIG. 12C ), and the impact assembly is ready to begin another impact and hammer retraction phase. In other words, when the hammer  2030  has returned, the pressure on both sides  2062 ,  2066  of the hammer  2030  has equalized and the disc  2106  is re-seated against the rear surface  2130  of the hammer  2030  by the bias of the valve spring  2110 . As such, the valve assembly  2098  provides for additional fluid flow through the valve assembly  2098  when the hammer  2030  is returning toward the anvil  2026  in order to more quickly reset the hammer  2030  for the next impact cycle. In other words, the valve assembly  2098  reduces the amount of time it takes to complete an impact cycle. 
       FIGS. 13-31  schematically illustrate different arrangements of the second embodiment of the impact tool  10  shown in  FIG. 3 . 
       FIG. 13  illustrates an embodiment in which the motor  30  includes a motor pinion  46  coupled to a first intermediate shaft  50  having a first bevel gear  54 . In the embodiment of  FIG. 13 , the motor  30  has a motor axis  56  that is parallel to or coaxial with the handle axis  22  and perpendicular to the output axis  26 . The first bevel gear  54  is engaged with a second bevel gear  58  on the end of a second intermediate shaft  62 . The second intermediate shaft  62  also includes a first spur gear  66  that is engaged with a second spur gear  70 , which functions as the rotational input  42  in the embodiment of  FIG. 13 . The second spur gear  70  drives the rotary impact mechanism  38 . 
       FIG. 14  illustrates an embodiment that is similar to the embodiment of  FIG. 13 , except that a third intermediate shaft  74  with third and fourth spur gears  78 ,  82  is interposed between the first spur gear  66  and second spur gear  70 , with the third spur gear  78  in meshing engagement with the first spur gear  66  and the fourth spur gear  82  in meshing engagement with the second spur gear  70 . 
       FIG. 15  illustrates an embodiment that is similar to the embodiment of  FIG. 13 , except that instead of the motor pinion  46  directly driving the first intermediate shaft  50 , the motor pinion  46  drives a fifth spur gear  86  that is in meshing engagement with a sixth spur gear  90  on the end of the first intermediate shaft  50 . 
       FIG. 16  illustrates an embodiment that is similar to the embodiment of  FIG. 15 , except that a first face gear  94  is interposed between the fifth and sixth spur gears  86 ,  90  to transfer torque therebetween. The first face gear  94  rotates about a third axis  96  that is parallel to the second axis  26  when transferring torque from the fifth spur gear  86  to the sixth spur gear  90 . 
       FIG. 17  illustrates an embodiment in which the motor pinion  46  functions as a sun gear  102  of a first planetary gear stage  98  in the gear train  34 . The first planetary gear stage  98  also includes a plurality of planet gears  106  encircling the sun gear  102  and rotatable about the sun gear  102  within a rotationally fixed ring gear  110 . A planet carrier  114  is coupled to the planet gears  106 , such that rotation of the planet gears  106  about the sun gear  102  causes rotation of the planet carrier  114 . A fourth intermediate shaft  118  extends from the planet carrier  114  and includes a third bevel gear  122  in meshing engagement with a fourth bevel gear  126  that in the embodiment of  FIG. 17  functions as the rotational input  42 . 
       FIG. 18  illustrates an embodiment that is similar to the embodiment of  FIG. 13 , except that instead of the motor pinion  46  directly driving the first intermediate shaft  50 , the first intermediate shaft  50  is driven by the planetary stage  98  of the embodiment of  FIG. 17 . 
       FIG. 19  illustrates an embodiment that is similar to the embodiment of  FIG. 17 , except that instead of the third bevel gear  122 , the fourth intermediate shaft  118  includes a worm gear  130 , and instead of the fourth bevel gear  126 , the rotational input  42  is an eighth spur gear  134  that is driven by the worm gear  130 . 
       FIG. 20  illustrates an embodiment that is similar to the embodiment of  FIG. 19 , except that a fifth intermediate shaft  138  with ninth and tenth spur gears  142 ,  146  is interposed between the worm gear  130  and the eighth spur gear  134 , with the ninth spur gear  142  engaged with the worm gear  130  and the tenth spur gear  146  engaged with the eight spur gear  134 . 
       FIG. 21  illustrates an embodiment that is similar to the embodiment of  FIG. 17 , except that instead of the third bevel gear  122  and the fourth bevel gear  126 , the fourth intermediate shaft  118  includes a pinion  150  that interfaces with a second face gear  154  that functions as the rotational input  42 . 
       FIG. 22  illustrates an embodiment that is similar to the embodiment of  FIG. 18 , except that instead of the first bevel gear  54  and second bevel gear  58 , the first intermediate shaft  50  includes the pinion  150  that interfaces with a third face gear  158  on the second intermediate shaft  62 . 
       FIG. 23  is similar to the embodiment of  FIG. 13 , but the gear train  34  also includes a planetary gear stage  162  between the second spur gear  70  and the rotary impact mechanism  38 , with a pinion  164  of the second spur gear  70  functioning as a sun gear of the planetary gear stage  162 . The planetary gear stage  162  also includes planet gears  166 , a fixed gear ring  168 , and a planet carrier  169  coupled to the planet gears  166 , such that rotation of the planet gears  166  about the pinion  164  causes rotation of the planet carrier  169 , which in turn drives the impact mechanism  38 . Thus, in the embodiment of  FIG. 23 , the planet carrier  169  functions as the rotational input  42  instead of the second spur gear  70 . 
       FIG. 24  is similar to the embodiment of  FIG. 17 , with the following differences. First, instead of the fourth intermediate shaft  118 , a second planetary gear stage  170  is driven by a sun gear  172  of the planet carrier  114  of the first planetary gear stage  98 . The second planetary gear stage  170  includes a plurality of planet gears  174 , a fixed ring gear  176 , and a planet carrier  180  coupled to the planet gears  174 , such that rotation of the planet gears  174  about the sun gear  172  causes rotation of the planet carrier  180 . Additionally, a sixth intermediate shaft  182  with an eleventh spur gear  184  is driven by the planet carrier  180  of the second planetary gear stage  170  and the fourth intermediate shaft  118  includes a twelfth spur gear  186  that is engaged with the eleventh spur gear  184 . 
       FIG. 25  illustrates an embodiment that is similar to the embodiment of  FIG. 17 , with the following differences. In the embodiment of  FIG. 25 , the motor axis  56  is perpendicular to the handle axis  22  and parallel to the output axis  26 . Additionally, the planet carrier  114  of the first planetary gear stage  98  drives a thirteenth spur gear  188  that engages a fourteenth spur gear  190  that functions as the rotational input  42 . 
       FIG. 26  illustrates an embodiment that is similar to the embodiment of  FIG. 25 , except that the third intermediate shaft  74  of the embodiment of  FIG. 14  is interposed between the thirteenth spur gear  186  and the fourteenth spur gear  190 , with the third spur gear  78  in meshing engagement with the thirteenth spur gear  188  and the fourth spur gear  82  in meshing engagement with the fourteenth spur gear  190 . 
       FIG. 27  illustrates an embodiment that is similar to the embodiment of  FIG. 13 , except that a drive wheel  192  replaces the first spur gear  66 , a driven wheel  194  replaces the second spur gear  70  to function as the rotational input  42 , and an endless drive member  198  is interposed between the drive wheel  192  and the driven wheel  194  to transfer torque therebetween. In some embodiments, the drive and driven wheels  192 ,  194  are pulleys and the endless drive member  198  is a belt. In other embodiments, the drive and driven wheels  192 ,  194  are sprockets and the endless driven member  198  is a chain. 
       FIG. 28  illustrates an embodiment that is similar to the embodiment of  FIG. 14 , except that the drive wheel  192  replaces the first spur gear  66 , the driven wheel  194  replaces the fourth spur gear  92 , the endless drive member  198  is interposed between the drive wheel  192  and the driven wheel  194  to transfer torque therebetween, and the third spur gear  78  is in meshing engagement with the second spur gear  70 . In some embodiments, the drive and driven wheels  192 ,  194  are pulleys and the endless drive member  198  is a belt. In other embodiments, the drive and driven wheels  192 ,  194  are sprockets and the endless driven member  198  is a chain. 
       FIG. 29  illustrates an embodiment that is similar to the embodiment of  FIG. 17 , except that the motor pinion  46  and the planetary gear stage  98  is omitted and the motor  30  directly drives the fourth intermediate shaft  118 . 
       FIG. 30  illustrates an embodiment that is similar to the embodiment of  FIG. 15 , except that the second intermediate shaft  62  is omitted and the second spur gear  70  is replaced with a fifth bevel gear  202  that functions as the rotational input  42  and is in meshing engagement with the first bevel gear  54 . Also, the first intermediate shaft  50  defines a fourth axis  204  that is parallel to the motor axis  56 , and the first intermediate shaft  50  is arranged such that the fifth bevel gear  202  is arranged between the motor axis  56  and the fourth axis  204 . 
       FIG. 31  illustrates an embodiment that is similar to the embodiment of  FIG. 13 , except that a seventh intermediate shaft  206  is driven by the motor pinion  46  instead of the first intermediate shaft  50 , and the seventh intermediate shaft  206  includes a fifteenth spur gear  210  that is in meshing engagement with the sixth spur gear  90 , such that the fifteenth spur gear  210  drives the sixth spur gear  90 . 
     In each of the embodiments of  FIGS. 13-31  that include spur gears, other types of parallel axis gears (e.g., helical gears) may be used. Although  FIGS. 13-31  schematically illustrate different arrangements of the second embodiment of the impact tool  10  shown in  FIG. 3 , each of the embodiments of  FIGS. 13-31  could be modified to arrange the rotary impact mechanism  38  upstream of the rotational input  42 , with the output of the rotational input  42  driving the output member  43 , as shown in  FIG. 2 . 
     Various features of the invention are set forth in the following claims.