Abstract:
A gear bearing having a first gear and a second gear, each having a plurality of teeth. Each gear operates on two non-parallel surfaces of the opposing gear teeth to perform both gear and bearing functions simultaneously. The gears are moving at substantially the same speed at their contact points. The gears may be roller gear bearings or phase-shifted gear bearings, and may be arranged in a planet/sun system or used as a transmission.

Description:
ORIGIN OF THE INVENTION  
         [0001]    The invention described herein was made by an employee of the United States Government, and may be manufactured and used by or for the Government for governmental purposes without the payment of any royalties thereon or therefor.  
         BACKGROUND OF THE INVENTION  
         [0002]    1 . Field of the Invention  
           [0003]    The present invention relates to a gear bearing having components that perform gear and bearing functions in a single component using two non-parallel surfaces simultaneously.  
           [0004]    2 . Description of the Related Art  
           [0005]    A first known planetary gear system uses helical planetary gears with two sets of ball bearings and drives its output off a carrier. In driving off the carrier, mechanical advantage and efficiency are sacrificed. The two sets of ball bearings locate and stabilize the operation of the gears. This arrangement takes up space and the interfaces make for a weaker system. By using two, or even one separate bearing, separate interfaces require a separate attachment/detachment means which tends to rattle and is so flexible that the structure is weakened.  
           [0006]    A second known design involves harmonic drives. Harmonic drives come in two main types, a pancake type (short axial length) and a cup type (with a larger axial length). Harmonic drives operate by means of a wave generator which rotates and, in so doing, periodically pushes a flexible spline (with teeth) radially outward in two diametrically opposite places. As the spline deflects outwards, its teeth push against the sides of the teeth of the output ring, causing the output ring to move to one side. As the wave generator turns, the points of the flexible spline turn with it, and the output ring moves with it also. There is generally one less tooth in the output ring than in the flex spline, so as the flex spline makes a complete cycle, the output ring rotates a total width of two teeth. The pancake version is not as easy to lubricate and is not as efficient as the cup type, though they are more compact. The cup type is more efficient but is not as compact. Both types are expensive and structurally weak, and the flexible splines tend to fail due to stripping.  
           [0007]    A third known design is disclosed in U.S. Pat. No. 5,409,431 , involving Carrier-Less, Anti-Backlash Planetary Drive System (Apr. 25, 1995 , by John M. Vranish). This system requires a spring loading which complicates the design and makes the system too large. Furthermore, anti-backlash is not frequently needed and is very expensive.  
         SUMMARY OF THE INVENTION  
         [0008]    Accordingly, it is an object of the present invention to provide a gear bearing system that overcomes the above disadvantages of the known designs.  
           [0009]    It is a further object of the present invention to provide a gear bearing system that results in a high mechanical advantage.  
           [0010]    It is yet another object of the present invention to provide a gear bearing system that has improved efficiency, strength and structural rigidity.  
           [0011]    It is still another object of the present invention to provide a gear bearing system that has a simpler construction and lower parts count.  
           [0012]    It is another object of the present invention to provide a gear bearing system that is superior in bearing strength.  
           [0013]    It is another object of the present invention to provide a gear bearing system in which it is easy to assemble, locate and stabilize gear systems.  
           [0014]    It is still another object of the present invention to provide a gear bearing system common to both the electric motor drive in order to allow the armature and stator to function properly, and to provide a high mechanical advantage output.  
           [0015]    A fourth embodiment of the present invention includes a system including a component that performs both gear and bearing functions in a single component using two or more orthogonal surfaces simultaneously. In a first embodiment, this system includes roller gear bearings, and in a second embodiment, this system includes phase-shifted gear bearings. In a third embodiment, the bearings are helical.  
           [0016]    The foregoing objects of the invention are further achieved by a fixed ratio transmission based on gear bearings.  
           [0017]    The foregoing objects of the invention may also be achieved by a gear bearing system comprising a first gear having a plurality of first gear teeth comprising a first plurality of contact points, each of said first plurality of contact points moving at a first speed, and a second gear having a second plurality of gear teeth comprising a second plurality of contact points, each of said second plurality of contact points moving at a second speed. The first plurality of contact points and second plurality of contact points are in contact with each other and the first speed is substantially the same as the second speed.  
           [0018]    The foregoing objects of the invention are still further achieved by a gear bearing system comprising a first gear having a plurality first gear teeth, each of the first gear teeth having a first surface and a second surface, and a second gear operating on the first surface and second surface of the first gear teeth.  
           [0019]    Additional objects and advantages of the invention will be set forth in part in the description which follows, and, in part, will be obvious from the description, or may be learned from practice of the invention. 
       
    
    
     BRIEF DESCRIPTION OF THE DRAWINGS  
       [0020]    These and other objects and advantages of the invention will become apparent and more readily appreciated from the following description of the preferred embodiments, taken in conjunction with the accompanying drawings of which:  
         [0021]    [0021]FIG. 1 is a perspective view of a spur gear as used in a first embodiment of the present invention.  
         [0022]    [0022]FIG. 2A is an overhead view of a planetary system using the spur gears of the first embodiment of the present invention. FIG. 2B is a cross-sectional view taken through line A-A of FIG. 2A.  
         [0023]    [0023]FIG. 3A is a partial top view of the spur gears of FIGS. 1 and 2 interacting.  
         [0024]    [0024]FIG. 3B is a partial top view of a spur gear interacting with a ring gear.  
         [0025]    [0025]FIG. 3C is a sectional view taken through line A-A of FIG. 3B.  
         [0026]    [0026]FIG. 4 is a perspective view of a phase-shifted gear bearing as used in a second embodiment of the present invention.  
         [0027]    [0027]FIG. 5A is a top view of a planetary system utilizing the phase-shifted gear bearing of FIG. 4.  
         [0028]    [0028]FIG. 5B is a cross-sectional view taken through line A-A of FIG. 5A.  
         [0029]    [0029]FIG. 6A is a partial top view illustrating the interaction of the phase-shifted gear bearings of FIG. 4.  
         [0030]    [0030]FIG. 6B is a sectional view taken through line A-A in FIG. 6A.  
         [0031]    [0031]FIG. 6C is an edge view of FIG. 6A. FIG. 6D is a sectional view taken through line B-B of FIG. 6A.  
         [0032]    [0032]FIG. 7A is a partial top view illustrating the interaction of the phase-shifted spur gears and the ring gear of FIG. 4.  
         [0033]    [0033]FIG. 7B is a sectional view taken through line A-A in FIG. 7A.  
         [0034]    [0034]FIG. 7C is an edge view of FIG. 7A.  
         [0035]    [0035]FIG. 7D is a sectional view taken through line B-B of FIG. 7A.  
         [0036]    [0036]FIG. 8A is a perspective view of a helical phase-shifted spur gear as used in a third embodiment of the present invention.  
         [0037]    [0037]FIG. 8B illustrates a peeled open edge view of the teeth of the phase-shifted helical gear bearing of FIG. 8A.  
         [0038]    [0038]FIG. 9 is a sectional view of a planetary transmission featuring roller gear bearings according to the present invention.  
         [0039]    [0039]FIG. 10 is a sectional view of a planetary transmission using phase-shifted gear bearings according to the present invention.  
         [0040]    [0040]FIG. 11 is a sectional view of an existing electric motor.  
         [0041]    [0041]FIG. 12 is a sectional view of an electric motor including the present gear bearings. 
     
    
     DESCRIPTION OF THE PREFERRED EMBODIMENTS  
       [0042]    Reference will now be made in detail to the present preferred embodiments of the present invention, examples of which are illustrated in the accompanying drawings, wherein like reference numerals refer to like elements throughout.  
         [0043]    As illustrated in FIGS. 1, 2 and  3 , a first embodiment of the present invention relates to spur roller gear bearings. Referring to FIG. 1, spur roller gear bearings consist of a spur gear  10  which has a roller  12  coaxially mounted on its top. Spur gear teeth  14  extend radially from spur gear  10 , and have a pitch radius R 1 . The radius of the roller R 2  is equal to the pitch radius R 1  of the spur gear teeth. The tops of the spur gear teeth  14  form a crown  16 . The radius to the crown top R 3  is equal to the pitch radius and the roller radius. Since R 1 , R 2  and R 3  are equal, the points at these radii move at the same speed.  
         [0044]    Referring now to FIG. 2A, we see that spur gears  10  can be configured with a ring gear  18  formed of ring gear teeth  24  to form planetary system  20 . Planet spur gears  10 B revolve around sun spur gear  10 A. Since spur gears  10 A, B are identical in size, the speed at R 1 , R 2 , and R 3  is identical for each spur gear  10 A, B. The crown  16  of each planet spur gear  10 B interfaces with the roller  12  of a sun spur gear  10 A, and vice versa. The teeth  14 B of each planet gear  10 B also interface with the teeth  14 A of the sun gear  10 A. Specifically, the teeth  14 A, B contact each other at pitch radius R 1 .  
         [0045]    Referring to FIG. 2B, which is a cross-section of the planetary system  20  of FIG. 2A, we see that a ring roller  22  is coaxially mounted on the top of the ring gear  18  such that the diameter of the ring roller  18  is set equal to the pitch diameter of the ring gear teeth  24 . The tops of the ring gear teeth  24  are crowned at (not shown) the point where they interface with the spur gear teeth  14 B.  
         [0046]    This planetary system  20  is held together without further structures. As can be seen from FIGS. 2A and 2B, if a planet spur gear  10 B is pushed down, its teeth  14 B will slide with respect to ring gear  18  and the sun spur gear teeth  14 A, but the planet spur gear roller  12 B will be blocked by the upper surface of the ring gear teeth  24 . If planet spur gear  10 B is pushed upwards, the ring roller  22  will block the upper surface of the planet spur gear teeth  14 B. If the sun spur gear  10 A is pushed down, sun spur gear roller  12 A will be blocked by the upper surface of each of the three planet spur gears  10 B, so that the planet spur gears  10 B will likewise be pushed down. However, planet spur gears  10 B will each, in turn, be blocked by the ring gear  18  so, ultimately, the sun spur gear  10 A cannot be pushed down. Likewise, the sun spur gear  10 A cannot be pulled up.  
         [0047]    [0047]FIGS. 3A, 3B and  3 C further illustrate how gears  10 A,  10 B,  18  interact with each other. FIG. 3A shows spur gears  10  interacting with other spur gears, illustrating the case where the sun spur gears  10 A are interacting with the planet spur gears  10 B. The teeth  14 A of sun spur gear  10 A contact the teeth  14 B of planet spur gear  10 B at point C 1 . The teeth  14 B of planet spur gear  10 B contact the roller  12 A of sun spur gear  10 A at point C 2 . FIG. 3B shows planet spur gears  10 B interacting with ring gear  18 . The ring gear teeth  24  contact the spur gear teeth  14 B at point C 3 . FIGS. 3A and 3B show that the spur gear on spur gear case is essentially the same as the spur gear on ring gear case in terms of matching speeds for both the roll and gear surfaces. Spur roller  12 B contacts ring roller  22  at point C 4 , and spur gear teeth  14 B contact ring roller surface  22  at point C 5 . FIG. 3C shows that by crowning the tops of the spur gear  10  with the apogee of the crown  16  at the same radial distance as the roller and tooth pitch radii, (R 1 , R 2 ) that thrust bearing contact must occur at the apogee point and so speed matching can be achieved for simultaneous and/or individual contacts between interfacing rollers, gear teeth and thrust bearing tooth tops/roller bottoms. This means, a planetary roller gear system will perform with great efficiency and strength. Furthermore, the addition of the rollers must, inevitably, greatly improve the accuracy with which the gears mesh. The rollers precisely set gear locations with respect to each other. On the other hand, the gears act as a highly efficient and precise caging/carrier mechanism for the rollers. The cumulative result is a superior system that is also very simple and low cost.  
         [0048]    Turning now to FIGS. 4, 5 and  6 , we discuss the second embodiment of the present invention, which involves phase-shifted gear bearings. FIG. 4 illustrates a phase-shifted spur gear  26  for use in phase-shifted gear bearings. Phase-shifted spur gear  26  includes an upper gear half  28  comprising upper gear teeth  30 , and a lower gear half  32  comprising lower gear teeth  34 . Upper gaps  36  and lower gaps  38  are formed between the gear teeth  32 ,  34 . Upper gear half  28  is rotated with respect to lower gear half  32  so that the two halves are exactly out of phase with respect to each other. That is, upper gear teeth  30  are positioned above lower gaps  38 , and lower gear teeth  34  are positioned below upper gaps  36 . Thus, phase-shifted spur gear  26  could mesh with a phase-shifted gear just like it. As one gear turned and drove the other, both halves would be continuously contacting each other but, in different phases of contact. In FIG. 4, the lower gear teeth  34  are bevelled and extended slightly between the upper gear teeth  30 . The upper gear teeth  30  are bevelled and slightly extended between the lower gear teeth  34  for both phase-shifted spur gears. Thus, the bevelled tooth surfaces contact each other much in the same manner as a four-way thrust bearing, and gear teeth  30 ,  34  contact each other and engage in conventional spur gear motion. The two motions can be timed so as to maximize efficiency, strength and smoothness.  
         [0049]    Referring now to FIGS. 5A and 5B, we see that phase-shifted spur gears  26  can be configured with a ring gear  40 , having upper ring gear teeth  44  and lower ring gear teeth  46 , to form a planetary system  42 , much like the system shown in FIGS. 2A and 2B. The planetary system  42  stays together in a similar manner to planetary system  20  of FIGS. 2A, 2B.  
         [0050]    [0050]FIGS. 6A and 6B further illustrate how phase-shifted spur gears  26  interact with each other. FIG. 6A particularly illustrates upper gear teeth  30  of one phase-shifted spur gear  26  contacting lower gear teeth  34  of a second phase-shifted spur gear  26  at contact point C 6 . FIGS. 6B, 6C and  6 D further illustrate contact points C 7 , where upper gear teeth  30  of the phase-shifted spur gears  26  contact; and contact points C 8 , where lower gear teeth  34  contact. FIG. 6C is an edge view of FIG. 6A.  
         [0051]    [0051]FIGS. 7A and 7B illustrate how phase-shifted spur gears  26  interact with ring gear  40 . FIG. 7A particularly illustrates upper ring gear teeth  44  contacting lower gear teeth  34  of phase-shifted spur gear  26  at point C 9 . FIGS. 7B, 7C and  7 D further illustrate contact points C 10 , where the upper gear teeth  30  of the spur gear contact the upper ring gear teeth  44 ; and contact points C 11 , where lower gear teeth  34  contact the lower ring gear teeth. (not shown)  
         [0052]    [0052]FIGS. 8A and 8B illustrate the third embodiment of the present invention, namely, helical gear bearings, in which spur gear  26  is replaced by a helical (or herring bone) gear  48 . The same timing issues and geometries that worked for the phase-shifted spur gear  26  apply in this embodiment. Although, FIGS. 8A and 8B show the case of phase-shifted helical gear bearings, a conventional roller gear bearing with helical teeth is also possible. FIG. 8B illustrates a peeled open edge view of upper helical teeth  50  and lower helical teeth  52 .  
         [0053]    The number of variations on the gear bearing arrangement of the present invention are endless, but only two will be discussed here. FIGS. 9 and 10 illustrate planetary transmissions using roller gear bearing and phase-shifted gear bearings, respectively. These planetary transmissions are fixed mechanical advantage transmissions which show great promise in being strong, compact, very efficient, carrierless, simple and capable of great speed reduction. The two concepts are functionally very similar, thus the explanation for roller gear bearings can easily be extended to the phase-shifted case.  
         [0054]    The roller gear bearing planetary transmission generally operates as follows. The transmission  54  comprises an input system  56  and an output system  58 . Input system  56  comprises input sun roller gear  60 , input roller gear planets  62  and ground ring roller gear  64 . Output system  58  comprises output roller gear planets  66 , output roller gear sun  68  and output ring roller gear  70 . The planets  62 ,  66  of both systems  56 ,  58  are axially joined together and thus, have the same angular velocity and must orbit about the center of the transmission  54  at the same angular velocity. The input sun roller gear  60  drives the input roller gear planets  62  which, in turn, react against the ground ring roller gear  64  by rotating at some angular velocity and orbiting about the center of transmission  54  at some orbital angular velocity. Thus, the orbital angular velocity and the rotational angular velocity for the planets  62 ,  66  are set. However, the output roller gear planets  66  have a different tooth pitch diameter than the input roller gear planets  62 . Thus, the output ring roller gear  70  has a different speed than the ground ring roller gear  64  and the transmission  54  exhibits speed reduction. The output roller gear sun  68  is in place primarily to provide strength and rigidity to transmission  54 , keeping the output system  58  together with strength and precision just as the output sun roller gear sun  68  does for input system  56 . Thus, the two systems  56 ,  58  are independantly strong and rigid and the combined system is even stronger.  
         [0055]    We will now derive the transfer function for the transmission  54  and, in so doing obtain further understanding of how it works.  
         ω OR   R   0 −ωp R   PO =ω O   R   O   (1)  
         ω OR   R   0 +ωp R   PO =ω S   R   S   (2)  
         [0056]    Where:  
         [0057]    ω OR =planet angular orbital velocity.  
         [0058]    R 0 =transmission output radius.  
         [0059]    ω P =planet angular velocity.  
         [0060]    R PO =planet output radius.  
         [0061]    ω O =output orbital angular velocity.  
         [0062]    R PI =planet input radius.  
         [0063]    ω S =sun angular velocity.  
         [0064]    R S =sun radius.  
         ( R   S +2 R   PI )θ OR   =R   PI θ PI   (3)  
         [0065]    Where:  
         [0066]    θ PI =some arbitrary angle a planet rotates.  
         [0067]    θ OR =the corresponding angle the planet orbits.  
         [0068]    Taking the time derivative of both sides of eq. (3) we get:  
                 (       R   S     +     2        P   PI         )          δθ     -     OR   δτ           =       R   PI          δθ       PI   _     δτ                 (   4   )                               
 
               δ                   θ     OR   δτ         =     ω   OR             (   5   )                               
 
               δ                   θ     PI   δτ         =         ω   PI     -     ω   PO       =     ω   P               (   6   )                               
 
         [0069]    Eqs. (4), (5) and (6) come from the basic definition of angular velocity and from the fact that a planet must have a single angular velocity for both the input and output interfaces and establish the relationship between ω P  and ω OR .  
                     ω   P          R   PI          R   O           (   R          S   +     2        R   PI         )       -       ω   P          R   PO         =       ω   O          R   O               (   7   )                       ω   P          R   PI          R   S           (   R          S   +     2        R   PI         )       +       ω   P          R   PI         =       ω   S          R   S               (   8   )                               
 
         [0070]    Esq. (7) and (8) come from substituting for ω OR .  
                   ω   S          R   S           ω   O          R   O         =           R   PI          R   O       +       R   PI          (       R   S     +     2        R   PI         )               R   PI          R   O       -       R   PO          (       R   S     +     2        R   PI         )                   (   9   )                               
 
                 ω     S   -         ω   O       =         (     R   O     )     [       2          (     R   PI     )     2       +       R   PI          (       R   O     +     R   S       )               R   S          [         R   PI          R   O       -       R   PO          (       R   S     +     2        R   PI         )         ]                 (   10   )                   ω   S       ω   O       =         (     R   O     )     [       2          (     R   PI     )     2       +       R   PI          (       R   O     +     R   S       )               R   S          [         R   PI          (       R   O     -     2        R   PO         )       -       R   PO          R   S         ]                 (   11   )                               
 
         [0071]    Eq. (11) comes from rearranging terms in eqs. (9), (10).  
           ω   S       ω   O       =     -   117                 T   O       T   I       =       -   99.5                     (     SAY              -   100     )                             
 
         [0072]    Eq. (12) states conservation of energy. Let:  
         [0073]    ∝=85%  
         [0074]    R PI =R S =0.25 in.  
         [0075]    R PO −R PI =0.020 in. Since:  
         [0076]    R O =R S +R PI +R PO    
         [0077]    R O =0.770 in.  
         [0078]    So:  
           ω   S       ω   O       =     -   117                 T   O       T   I       =       -   99.5                     (     SAY   -   100     )                             
 
         [0079]    An estimate that the transmission can withstand 60 ft-lb output torque is derived as follows:  
         [0080]    The largest stress will be on the planet teeth that push off against the ground ring roller gear  64 . This is because the lower planet radius is slightly smaller than the upper planet radius and because it will take slightly more load. Assuming  20  teeth in the planet.  
           1.024945      E3                 lbs                   (     .770                   in   .       )         12                     in   .     /     ft   .           =     65                     ft   .     /     lbs   .                               
 
         [0081]    E 3(.6) (.25 in.)=max allowable shear load per tooth =342 lbs.  
         [0082]    Where:  
         [0083]    L=length of teeth that can resist shear=0.25 in.  
         [0084]    58E3=yield strength of material with 2:1 safety factor.  
         [0085]    0.6=shear factor.  
         [0086]    40=20 teeth+20 spaces between teeth.  
         [0087]    With 3 planets we get 1.024945 E3 lbs.=F. And  
                   (     π2                   R   PI        L     )        58                 E                 3                   (   .6   )                     (     .25                   in   .       )       40     =     max                 allowable                 shear                 load                 per                 tooth                 =     342                   lbs   .                                   
 
         [0088]    Assembly of transmission  54  will now be discussed. To assemble transmission  54 , the roller portions of the planets  62 ,  66  are positioned in output ring roller gear  70 . The roller of input sun roller gear  60  is then positioned in the arrangement. The other planet gear teeth cylinders  72  are then tightly fit over each of the bottom roller portions of the planets  62 ,  66  and, at the same time, meshed with the teeth of the output ring roller gear  70 . The input sun roller gear teeth cylinder  73  is then tightly fit over the roller of the input sun roller gear  60 , meshed with the teeth of the planets  62 ,  66  and fastened in place with an assembly screw. The ground ring roller gear  64  is then slipped in place, its teeth meshing with the teeth of the planets  62 ,  66  as it goes. Next, input sun roller gear  60  is slipped into place, its teeth meshing with those of the planets  62 ,  66  as it goes. The three bottom portions of the planet rollers are each then fit tightly into their respective planet gear teeth cylinders and splined into the roller portion of the planet already in place. Then, each of the planets  62 ,  66  is finalized in its assembly with a fastening screw. The entire transmission  54  is now assembled, aligned and ready to function.  
         [0089]    Disassembly is accomplished by reversing the steps. It should be noted that if the output roller gear sun  68  can be manufactured in a single piece, and the assembly/disassembly process can proceed, essentially unchanged.  
         [0090]    Referring to FIG. 10, we see a sectional view of a planetary transmission  82  using phase-shifted gear bearings. The phase shifted gear bearing transmission  82  has a similar structure to roller gear bearing transmission  54 . However, the corresponding input and output sun and planet gears, as well as the ring gear, comprise phase-shifted gear bearings as opposed to roller gear bearings. The assembly/disassembly process for the phase-shifted gear bearing transmission of FIG. 10 is essentially identical to that described with respect to the transmission  54  of FIG. 9.  
         [0091]    The present gear bearing can also be used to improve electric motors. FIG. 11 is a sectional view of an existing electric motor  90 , requiring two sets of ball bearings  92 , which separate armature  94  from stator  96 . The armature  94  includes permanent magnet  95 , and the stator  96  includes coils  97 . The ball bearings  92  also allow the armature  94  to rotate with respect to the stator  96 , typically by using the weak forces of electric motors. The motor  90  further includes a motor mount screw  98 , and an output drive  99 .  
         [0092]    [0092]FIG. 12 is a sectional view of an electric motor  100  using the gear bearings of the present invention. Motor  100  is similar to the existing motor design in that it includes armature  104 , including permanent magnets  105 , stator  106 , including coils  107 , and motor mount screw  108 . These elements form a housing  110 . Motor  100  also comprises an output  112 , including an output screw  109 .  
         [0093]    Instead of using ball bearings, motor  100  has sun gear bearing  114 , a gear bearing transmission  120 , comprising a sun gear bearing which drives plant gear bearings  116 , which in turn drive the output  112 . An idler  118  acts as a stiffener and is placed between planet gear bearings  116 .  
         [0094]    The gear bearing transmission  120  results in a smaller, simpler design, that is easier to assemble as compared to the existing ball bearing design.  
         [0095]    Although a few preferred embodiments of the present invention have been shown and described, it would be appreciated by those skilled in the art that changes may be made in these embodiments without departing from the principles and spirit of the invention, the scope of which is defined in the claims and their equivalents.