Patent Document

FIELD OF THE INVENTION 
   This invention relates generally to electrically or pneumatically powered hand tools and, more specifically, to dynamically compensated electrically or pneumatically powered hand tools. 
   BACKGROUND OF THE INVENTION 
   Sanders are generally described by the characteristic motion by which drive their abrasive; sanders may be orbital, in-line, disk, or belt sanders. In-line, disk, and belt sanders gouge distinct abrading marks on the surface of the workpiece, by the cumulative effects of the abrading medium as it travels in the same direction. To produce a suitable finish, another tool, such as an orbital sander later must remove the resultant abrasion marks. Orbital sanders produce a more random abrading pattern, therefore, a more uniform and desirable surface finish. In general, using belt, inline, and disk sanders is limited to aggressive surface abrading of the workpiece surface. 
   Orbital sanders drive a sanding pad in an eccentric orbit around the motor shaft centerline. Operators prefer orbital sanders because of their controllability. When abrading a surface, an operator has excellent control of sander position, which is important because it allows the operator to abrade a precisely defined area, such as abrading next to masking tape or to a perpendicular surface. In contrast, belt, in-line, or disk, apply a reactionary force to the operator, opposite the direction of sanding medium motion. To keep such a sander in one location, the operator must always provide an equal reactionary force. As a result, belt, in-line, and disk sanders are more difficult to control. 
   Orbital sanders, however, generate relatively high vibration levels, up to 30 m/s 2 . With long exposures, these levels are often injurious to the operator, resulting in serious long-term nerve, vascular, or musculoskeletal damage of an upper extremity. The vibration is the result of imbalanced rotational forces along the shaft-assembly. These forces are dependent on operator pushing force as well as variations in counterweight mass, sanding pad mass, and sanding medium mass. 
   Orbital sanders have been limited in use to less aggressive abrading tool because of their vibration levels. A more aggressive orbital sander is one that swings its sanding pad at larger orbits that is with greater eccentricity rather than by increasing rotational speed. As a result, the sander drives the pad to abrade more area per orbit. The most aggressive orbital-sanders typically have ⅜-inch diameter orbits with rotational speeds between 10,000 and 12,000 orbits per minute. 
   Orbital sander manufacturers have not been able to design the vibration out of orbital sanders. The vibration results from imbalance, and in the design of orbital sanders, imbalance, in large part, stems from the displacement of a center of gravity from a center of rotation. Given the variety of weights of sandpapers, any replacement of sandpaper can offset the center of gravity from the center of rotation. Due to the varying weight of sandpaper, a single offset design is not possible. 
   The disadvantages associated with current orbital sanders have made it apparent that a new orbital sander that generates less vibration and is more aggressive is needed. 
   SUMMARY OF THE INVENTION 
   A system for active dynamic balancing of a rotating power tool driven by a motor having a shaft supported by a first and second bearing on opposing ends of the motor includes an acceleration sensing assembly configured to sense radial accelerations on the shaft producing an acceleration signal indicative of the radial accelerations. A correcting mass assembly is configured to rotate with the shaft and to move at least one mass radially to the shaft responsive to a correcting signal. A controller is configured to receive the acceleration signal generating a correcting signal by means of a closed-loop iterative algorithm. 
   An active dynamic rotational balancing system corrects for both the radial imbalance forces and the operator pushing force generated by orbital sander operation. When these corrections are made, all rotational force interactions with the handgrip are greatly reduced; this results in lower handgrip vibration levels. 
   A system uses a programmable microcontroller to implement the feedback control algorithm and to operate two miniature stepper motors that reposition correction masses. Two accelerometers integrated into the bearing mounts provide feedback information. The programmable microcontroller compensates for a phase shift difference with reference to an optical sensor. Each stepper motor operates a lead screw to move correction masses radially in the two planes of imbalance to correct for both the radial imbalance forces and for the operator pushing force generated by orbital sander operation. When the stepper motor compensates for them, all rotational force interactions vibrating the handgrips are greatly reduced. 
   A force biasing mechanism is incorporated into the system to provide four times the compensating force of a system without the mechanism. Using acceleration data from feedback sensors imbedded into the handgrip as well as a shaft position sensor and microcontroller, the mechanism is directed to correctly redistribute correction mass in two planes, which are perpendicular to the rotating shaft, to dynamically balance the entire rotational system. An active rotational balancing system corrects for variations in the rotational system, to produce a balanced force system. 
   As will be readily appreciated from the foregoing summary, the invention provides an active dynamic rotation balancing system for a rotating tool. 

   
     BRIEF DESCRIPTION OF THE DRAWINGS 
     The preferred and alternative embodiments of the present invention are described in detail below with reference to the following drawings. 
       FIG. 1   a  is a force analysis diagram for an active dynamic rotation balancing system for a rotating tool; 
       FIG. 1   b  is a force analysis diagram for the active dynamic rotation balancing system for a rotating tool showing a phase-angle shift; 
       FIG. 2  is a block diagram of an electronic control assembly for the active dynamic rotation balancing system for a rotating tool; 
       FIG. 3  is a flow chart of an algorithm for controlling the active dynamic rotation balancing system for a rotating tool; 
       FIG. 4  is a cross-sectional view of an orbital sander having the active dynamic rotation balancing system; 
       FIG. 5  is a perspective view of a balancing mass assembly for the active dynamic rotation balancing system; and 
       FIG. 6  is an exploded diagram of the orbital sander having the active dynamic rotation balancing system. 
   

   DETAILED DESCRIPTION OF THE PREFERRED EMBODIMENT 
   Referring to  FIG. 1   a , a force diagram  20  aids in the description and analysis of the forces causing vibration in an orbital sander. In this diagram the forces are all coplanar. A motor shaft  21  spinning about that axis at a rotation speed of ω provides a frame of reference. The motor shaft  21  is the principal moving part of the orbital sander and drives the operational components including the sanding pad with attached sandpaper. For purposes of analysis the sanding pad assembly may be fairly represented by a point mass located at a center of the constituent mass. The motor shaft  21 , can be assumed symmetrical around the motor shaft axis and with homogeneous density, thereby not contributing to any imbalance in the system. 
   The mass of the sanding pad assembly with sandpaper can be represented by mass  24  at distance a from a motor shaft axis  21  of rotation. At rotational speed ω, the mass  24  imparts a rotational force  27  on the motor shaft axis  21 , that is the product of radius a times the magnitude of the mass  24 , and the square of the rotational velocity, i.e. ω 2  (F=mrω 2  ). Vibration results as time-varying reactionary forces, and in the case of the orbital sander transmits through the bearings and contributes to the horizontal top bearing force  45  and the bottom bearing force  51 . 
   Along with forces imparted simply by the rotation of the shaft, vibration stems from time-varying reactionary forces fed to the orbital sander motor shaft  21  by action of the operator. The operator pushing the sander across the surface of the workpiece and pressing the sander to the workpiece with a vertical pushing force  42  that together with the gravitational force impart a vertical pushing force  48  through the workpiece acting on the shaft, thus forming a force couple. Vibration results as time-varying reactionary forces, and in the case of the orbital sander transmits through the bearings and contributes to the horizontal top bearing force  45  and the bottom bearing force  51 . 
   To counteract the reactionary forces, i.e. the horizontal top bearing force  45  and the bottom bearing force  51 , a top correction-mass  30  and a bottom correction-mass  36  spin with motor shaft  21 , at radii c and e respectively, and produce forces respectively. As set forth above, the resulting forces, forces  33  and  39  are proportional to the rotational velocity squared ω 2 , and respective radii c and e. The force diagram demonstrates that by suitably selecting the radii c and e respectively, the reactionary forces are effectively counterbalanced eliminating the reactionary forces, i.e. the horizontal top bearing force  45  and the bottom bearing force  51 . Suitably varying the radii c and e is a dynamic process as the pushing force  42  varies 
   Referring to  FIG. 1   b  (the elements present remain as set forth as in  FIG. 1   a  discussed above), as the rotational velocity ω, a phase-shift phenomenon exists, resulting from the time difference between when the rotational system produces a maximum force and when the corresponding forces are measured. In other words, the force measured is not necessarily coplanar to the correcting forces  33  and  39 . The measured force must be adjusted by the phase-angle φ to obtain bearing forces that are coplanar to the correction-forces  33  and  39 . If the phase-angle φ equals zero, then the measured bearing forces  45  and  51  are coplanar to the correction-forces  33  and  39 . 
   The phase-angle φ is measured using an optical sensor  75  in a presently preferred embodiment though as will readily be perceived by those skilled in the art, any suitable motor shaft  21  indexing device will serve to measure the phase-angle φ. The purpose of the indexing device such as the optical sensor  75  is to inform the controller of the phase-angle of the motor shaft  21  as it rotates, whereas the accelerometers  54 ,  57  indicate the magnitudes of the top bearing force  45 , and the bottom bearing force  51 . 
   Referring to  FIGS. 1   a ,  1   b , and  2 , a controller  63  comprises two processing channels, a top channel  66   a  and a bottom channel  66   b . The top channel  66   a  is configured to minimize the top bearing force  45  and the bottom channel  66   b  is configured to minimize the bottom bearing force  51 . The controller  63  controls the radial positions, at radii c and e respectively, of the top correction-mass  30  and the bottom correction-mass  36 , to produce forces  33  and  39 . As forces  33  and  39  are optimized, the top bearing force  45  and the bottom bearing force  51  are minimized. Characteristic of a closed-loop program, the outputs are measured with accelerometer  54  and accelerometer  57 , then fedback and compared to the desired input. If they are not the same the controller  63  makes adjustments to drive them to be the same. 
   The controller receives inputs from mixers  69   a  and  69   b  by the top channel  66   a  and the bottom channel  66   b  of the controller  63  respectively. The mixers receive a signal as a negative input from the accelerometers  54  and  57  for the top bearing acceleration, which is represented by the top bearing force  54 , and the bottom bearing acceleration, which is represented by the bottom bearing force  57  respectively. Since accelerometers measure acceleration, the controller  60 , works in acceleration instead of working in force values. Force and acceleration are proportional. The mixers  69   a  and  69   b  receive inputs representative of a zero acceleration input as a positive input for comparison with the output of the top and bottom bearing accelerometers  54  and  57  respectively. These inputs are corrected for phase angle information received by the optical sensor  75  to determine an appropriate signal for determining a position for varying the positions of the top correction-mass  30  and the bottom correction-mass  36  by varying radii c and e respectively. 
   A second mixer  71   a  modifies the output of the top channel  66   a  as a second mixer  71   b  modifies the output of the bottom mixer according to the input of a force disturbance that could be from several different sources, such as the operator pushing on the sander and/or a change in sandpaper mass from either installing a new piece of sandpaper, loading the current sandpaper with work-piece particles, or degrading the current sandpaper by loosing abrasive particle media. 
   Referring to  FIG. 3 , in a presently preferred embodiment an effective two-channel control algorithm  100  begins at a block  102  and operates continuously while the sander drives the sandpaper and only ends when the orbital sander is turned off. The purpose of the control algorithm  100  is to move correction-masses  30  and  36  until both phase-corrected bearing acceleration  45  and bearing acceleration  51  are near or equal to zero. Another objective of algorithm  100  is to get both phase-corrected bearing acceleration  45  and phase-corrected bearing acceleration  51  to zero or near zero in a short amount of time. Although there could be other more efficient algorithms, algorithm  100  has proven to function effectively. 
   The algorithm  100  has the feature to change from coarse to fine resolution by assuming a large step size (displacement) of correction-mass position. In the current algorithm, the low resolution displacement value is ten times longer than the high resolution displacement value. 
   In algorithm  100 , after both bearing accelerations have been corrected for the phase-angle φ offset they are combined for comparison purposes. This combined value has the advantage of having only one acceleration level to compare instead of two. Determining the absolute value of each top bearing acceleration  45  and bottom bearing acceleration  51  calculate this comparison level. The higher of the two absolute acceleration levels is the comparison value. When the comparison value is near or equal to zero, then the bearing accelerations  45  and  51  are also near or equal to zero and the sander housing will transmit minimal vibration to the operator&#39;s hand. 
   At a block  105 , the controller receives signals from the optical sensor  75  and the accelerometers  54  and  57  to derive the phase corrected top and bottom bearing accelerations. At a block  105 , the highest absolute acceleration level (the comparison value) is calculated, as described above. This initial highest absolute acceleration is the baseline level. 
   At a block  108 , the mass displacement, or movement step size is set to the low-resolution value. 
   At a block  111 , the controller moves the bottom correcting mass  36  by decreasing the radius e. 
   Again, at a block  114 , the highest absolute acceleration level is calculated, as described above, as in the block  105 . 
   At a decision block  117 , the algorithm compares the new highest absolute acceleration level to the baseline level in order to determine if the movement of the mass at block  111  has reduced the acceleration. 
   If the new highest absolute level is lower than the baseline level, then the new highest absolute level is made equal to the baseline level, and the old baseline level is erased. At a decision block  120 , the algorithm determines whether to use the low resolution mass displacement value or the high-resolution displacement value. In either case, again at a block  111 , the controller moves the bottom correcting mass  36  by decreasing the radius e. Again the steps in block  114  and decision block  117  are repeated. While the new highest absolute level is lower than the baseline, steps in block  120 , block  111 , block  114  and decision block  117  are repeated again and again until the new highest absolute level is higher than the baseline. 
   When at decision block  117 , the new highest acceleration level is higher than the baseline level, the step in block  126  is initiated. At a block  126 , the controller moves the top correcting mass  30  by decreasing the radius c. At a block  129 , and as in a block  105 , the highest absolute acceleration level is calculated. At a decision block  132 , the algorithm compares the new highest absolute acceleration level to the baseline level in order to determine if the movement of the mass at block  126  has reduced the acceleration. 
   If the new highest absolute level is lower than the baseline level, then the new highest absolute level is made equal to the baseline level, and the old baseline level is erased. Again at a block  126 , the controller moves the top correcting mass  30  by decreasing the radius c. Again the steps in block  129  and decision block  132  are repeated. While the new highest absolute level is lower than the baseline, steps in block  126 , block  129  and decision block  132  are repeated again and again until the new highest absolute level is higher than the baseline. 
   When at decision block  132 , the new highest acceleration level is higher than the baseline level, the step at block  135  is initiated. At a block  135 , the controller moves the top correcting mass  30  by increasing the radius c. At a block  138 , and as in a block  105 , the highest absolute acceleration level is calculated. At a decision block  141 , the algorithm compares the new highest absolute acceleration level to the baseline level in order to determine if the movement of the mass at block  135  has reduced the acceleration. 
   If the new highest absolute level is lower than the baseline level, then the new highest absolute level is made equal to the baseline level, and the old baseline level is erased. Again at a block  135 , the controller moves the top correcting mass  30  by increasing the radius c. Again the steps in block  138  and decision block  141  are repeated. While the new highest absolute level is lower than the baseline, steps in block  135 , block  138  and decision block  141  are repeated again and again until the new highest absolute level is higher than the baseline. 
   When at a decision block  141 , the new highest acceleration level is higher than the baseline level, the next step is initiated. At a block  144 , the controller moves the bottom correcting mass  30  by increasing the radius c. At a block  147 , and as in a block  105 , the highest absolute acceleration level is calculated. At a decision block  150 , the algorithm compares the new highest absolute acceleration level to the baseline level in order to determine if the movement of the mass at block  144  has reduced the acceleration. 
   If the new highest absolute level is lower than the baseline level, then the new highest absolute level is made equal to the baseline level, and the old baseline level is erased. Again at a block  144 , the controller moves the bottom correcting mass  36  by increasing the radius e. Again the steps in block  147  and decision block  150  are repeated. While the new highest absolute level is lower than the baseline, steps in block  144 , block  147  and decision block  150  are repeated again and again until the new highest absolute level is higher than the baseline. 
   When at a decision block  150 , the new highest acceleration level is higher than the baseline level, the next step at the decision block  120  is initiated. At a decision block  120 , the algorithm determines whether to use the low resolution mass displacement value or the high-resolution displacement value. In the current algorithm, the low-resolution mass displacement value is used to implement a minimum of two mass displacement cycles, defined as performing the steps listed from block  111  to the decision block  150 . After two mass displacement cycles, in the decision block  120 , the baseline acceleration level from using the prior mass displacement value is compared to the new baseline acceleration level using the current mass displacement value. While the new baseline acceleration is lower than the prior baseline acceleration level, the low-resolution mass displacement value is used and the system continues implementing additional mass displacement cycles. When no change in two consecutive baseline accelerations occurs, the algorithm changes to using the high resolution mass displacement value. 
   Referring to  FIG. 4 , a cross-section view of a presently preferred embodiment of the inventive orbital sander  20   c  reveals a compact and functional sanding machine. A sander housing  22  is configured to enclose the workings of the sander and also to serve as an advantageous shaped handgrip. The sander housing  22  encloses a drive train with elements found in non-inventive orbital sander systems: a motor  25  (either electric or pneumatic), a motor shaft  21   a , a top bearing  44  and top bearing mount  43 , a bottom bearing  51  and a bottom bearing mount  52 , an orbital bearing assembly  97 , and a sanding pad  99 . Collectively these elements form a drive train similar to that found in a conventional sander. 
   Inventive elements of a dynamic balancing system include a controller  60 , slip ring brushes  79  along with a slip brush plate  77  to convey signals to a top stepper motor  83   a  and a bottom stepper motor  83   b  mounted respectively on an top motor plate  87   a  and a bottom motor plate  87   b . In the top correcting assembly  78   a , a top stepper motor  83   a  drives a top biased correction-mass assembly  85   a  and in a bottom correcting assembly  78   b , the bottom stepper motor  83   b  drives a bottom biased correction-mass assembly  86   b . A top thrust transfer pad  84   a  supports a top thrust bearing  89   a  as the top stepper motor  83   a  drives the top biased correction-mass assembly  85   a . Similarly, the bottom thrust transfer pad  84   b  supports the bottom thrust bearing  89   b  as the bottom stepper motor drives the bottom correction-mass assembly  85   b . These elements affect the placement of corrective masses in the respective correction-mass assemblies  85   a  and  85   b  at the direction of controller  60 . The controller  60  receives input from the advantageously placed top bearing accelerometer  54 , the bottom bearing accelerometer  57  and the optical sensor  75 . 
   Referring to  FIG. 5 , an exemplary correcting assembly  78  represents both the top correcting assembly  78   a  ( FIG. 4 ) and bottom correcting assembly  78   b  ( FIG. 4 ). Each correcting assembly  78  is configured to nest with a second correcting assembly  78  that is rotated 180 degrees around a minor (vertical) axis and flipped across a horizontal plane. In this manner, opposed masses are oriented for parallel radial movement with respect to the shaft  21   a  ( FIG. 4 ) while each are axially offset from the motor  25  ( FIG. 4 ) distinct distances. So configured, the masses of the rotating stepper motors  83  are at equal radial distances in a horizontal plane, thereby neutralizing their masses in the horizontal plane in the rotating system, but they are vertically offset to create the vertical distance between correction-mass  36  and correction-mass  30 . Similarly, placement of the motor plate  87 , the thrust transfer pad  84 , and the thrust bearing  89 , are placed to compensate for each other in the horizontal plane in the rotating system. Although the motor plate  87 , the thrust transfer pad  84 , and the thrust bearing  89  are vertically offset from each corresponding other, the active dynamic rotational balancing system correctly compensates for this offset. Stepper motor mount  80 , is held in place by motor plate  87  and contains the stepper motor  83 , thrust transfer pad  84 , and the thrust bearing  89 . 
   Built on the motor plate  87  to give rigidity and exact placement of remaining elements, the correction-mass assembly  78  includes the stepper motor mount  80 , thrust transfer pad  80 , the thrust bearing  89 , the stepper motor  83 , a configured correction-mass  85  and a matched pair of biasing springs  82 . A stepper motor armature  88  rotates 1/20th of a revolution for each step with a pitch advantageously selected to allow fine resolution movement of the correction-mass  85 , a 0.25 mm screw pitch is selected in the presently preferred embodiment so the correction-mass  85  is moved 0.0125 mm for each step. 
   In operation, during high-speed rotation of the correction-mass assembly  78 , a rotational acceleration acts on the armature  88  of the stepper motor  83 . The rotational acceleration applies a force to the armature  88  causing misalignment. The thrust transfer pad  84  supporting a thrust bearing  89  is advantageously included to support the armature  88  from misalignment, assuring optimal operation of the stepper motor  83 . 
   The inventive configuration of the correction-mass assembly  78  amplifies the force used to move the correction-masses often against rotational acceleration. In the presently preferred embodiment, the stepper motor  83  can only provide 3 lbs of thrust (radial force) to accomplish the movement of correction-masses. To achieve more than 11 lbs of balancing force, two springs  82  supply a biasing force to counteract the rotational acceleration on the correction-masses  85 . In the presently preferred embodiment, when correction-masses  85  at an extreme range of the designed travel, a rotational force of 11 lbs is exerted on the correction-mass. Advantageously in this position the springs  82  supply a total of 9 lbs biasing in opposition to the rotational force. Thus, at even the extreme end of the range there are only 2 lbs. of thrust that the stepper motor  83  must supply to move the correction-masses  85  inward. 
   Referring to  FIG. 6 , an exploded view of the inventive sander  20   c  sets forth the several components of the presently preferred embodiment. Though illustrated with an electric motor  25 , the presently preferred embodiment may be driven by any suitable motive means including a pneumatic motor as will readily be appreciated by one skilled in the arts. 
   The housing  22  is, advantageously, formed to enclose the driving means and to conform to an operator&#39;s hand. Two bearings, a top bearing  44  in the top bearing mount  43  and a bottom bearing  53  in its bottom bearing mount  52  hold the motor shaft  21   a  in fixed relationship to the housing  22 . Additionally, the top bearing mount  43  provides a suitable mount for the top bearing accelerometer  54  ( FIG. 4 ) and the optical sensor  75  ( FIG. 4 ), both advantageously placed to note movement of the motor shaft  21   a . Similarly, the bottom bearing mount  52  provides a suitable mount for the bottom bearing accelerometer  57 . As discussed above the accelerometers  54 , and  57  along with the optical sensor  75  or other suitable indexing device such as a Hall effect sensor, allow for measurement and determination of the phase-corrected accelerations on the motor shaft  21   a . With the determinations of the phase-corrected accelerations on the shaft, the controller  63  can suitably move the correction-masses  85   a ,  85   b  into optimal position to minimize the phase-corrected accelerations. 
   The motor shaft  21   a  drives the sanding pad  99  and the orbital bearing assembly  97 . The orbital bearing assembly  97  contains an offset axis and produces an orbital motion in any designated one of known modes such as random orbital, dual-action, or jitterbug. The motor shaft  21   a  drives the sanding pad  99  in an eccentric orbit around the motor shaft axis  21  ( FIGS. 1   a ,  1   b ). For a random orbital sander, the circular sanding pad  99  is mounted to a bearing on its axis; during operation sanding pad  99  is allowed to slip on a sanding pad axis. In a dual-action, the operator can select one of two modes of operation, one being the random orbital operation, the other being a locked pad mode. In the locked pad mode, the pad does not slip on its axis. 
   In most orbital sanders, the sanding pad  99  is suitably configured to accept round pads with either pressure sensitive adhesive or a hook and pile system. In a jitterbug orbital sander, the sanding pad is square or rectangular and contains two clips to attach the sanding medium. The advantage of a square pad is that the square pad will accept standard sheet sanding medium, and the sheet sanding medium can be cut to the correct size. 
   The controller  63  ( FIG. 2 ) controls the stepper motors  83  by means, in the presently preferred embodiment, of four voltage sources for each of two stepper motors thus by means of eight voltage signals. Therefore, an eight channel slip-ring system  92  includes a eight channel slip-ring  81  with contact rings in each of the defined channels. Eight contact brushes  79  each contact one of the individual contact rings. Suitable wiring (not shown) allows the voltage signals sent by the controller  63 , at the contact rings to reach the two stepper motors  83   a ,  83   b.    
   To place the signal on the contact rings, brush springs  94  suitably bias the contact brushes  79  against the contact rings while conducting signals to the brushes by biased contact. A non-conductive slip brush plate  77  holds the slip brushes  79  in orthogonal relation to the contact rings while allowing axial movement of the slip brushes  79 . A keeper  96  and an insulated pin  93  fix the biasing slip brush springs  94  in relationship to the slip brushes  79  to suitably apply the biasing force. Both the keeper  96  and the pins are of a nonconductive material to prevent cross-talk between distinct voltage channels. 
   While the preferred embodiment of the invention has been illustrated and described, as noted above, many changes can be made without departing from the spirit and scope of the invention. For example, an additional adjusting mechanism that allows the operator to increase the orbital eccentricity might be inserted to allow for more aggressive sanding. Accordingly, the scope of the invention is not limited by the disclosure of the preferred embodiment. Instead, the invention should be determined entirely by reference to the claims that follow.

Technology Category: 7