Abstract:
The present invention is a gear system generally having a driven gear, rotatably mounted on an axis, having teeth defined by unique geometrical profiles; a driving means contacting the teeth of the driven gear; and round pins or spheres for engaging the teeth of the driven gear rotatably mounted on the driving means, the round pins or spheres for engaging the teeth of the driven gear roll along a profile of the tooth of the driven gear; and the gear system has a variable pressure angle and a variable contact ratio.

Description:
[0001]    Priority for this non-provisional patent application is claimed under 35 U.S.C. § 119, pursuant to Applicant&#39;s provisional patent application, application No. 60/291,981, filed on May 21, 2001. 
     
    
     
       FIELD OF INVENTION  
         [0002]    The present invention relates to an improved gear system and specifically to a gear system wherein at least one circular gear in the system includes a plurality of rotating pins or spheres instead of involute teeth. In this manner, the at least one circular gear including a plurality of rotating pins or spheres, can provide a variable pressure angle and variable contact ratio allowing force to be delivered to a driven gear as maximum torque for a correspondingly longer period of time.  
         BACKGROUND OF THE INVENTION  
         [0003]    Pressure angle is one of the primary defining characteristics for gear pairs. It is a measure of how effective the gear pairs are at transmitting torque from a driving gear to a driven gear. In a frictionless system, the driven gear with no frictional losses due to the sliding of the gear teeth on each other, receives the maximum amount of torque from the driving gear.  
           [0004]    However, in a conventional gearing system friction is present in all gears. Gearing systems that have a larger pressure angle may be able to accept a greater load or force on the gears as compared to gearing systems with smaller pressure angles. However, larger forces applied to gearing systems, which result from a larger pressure angle, leads to more friction losses and greater gear wear. In conventional gearing systems, a large pressure angle generally indicates a correspondingly larger radial component force. Pressure angle and radial component force are proportional to each other.  
           [0005]    The present invention relates to gearing systems having at least one circular gear, such as gear reducers, as well as rack and pinion gearing systems.  
         SUMMARY OF THE INVENTION  
         [0006]    In general, the present invention is a gearing system having:  
           [0007]    [a] a driven gear, rotatably mounted on an axis, where the gear has unique tooth geometries compatible with the driver gear means consistent with the law of gearing;  
           [0008]    [b] a driving gear means contacting the teeth of the driven gear means;  
           [0009]    [c] a means for engaging the teeth of the driven gear means rotatably mounted on the driving gear means, the means for engaging the teeth of the driven means having the ability to roll along a profile of the tooth of the driven means; and the gear system including a variable pressure angle and variable contact ratio.  
           [0010]    The means for engaging the teeth of the driven means is typically a plurality of pins that engage the teeth and roll on their profile. The contact of the pins on the teeth results in gearing geometry such that a gear pair can have a variable pressure angle. This has significant advantages on gear load and speed. The gear system of the present invention may be in a circular gear system or rack and pinion gear arrangement.  
         OBJECTS OF THE INVENTION  
         [0011]    It is the primary object of the invention to provide an improved gear system which has a variable pressure angle.  
           [0012]    It is another object of the present invention to provide an improved gear system in which the pressure angle can be reduced so that the load on the gear system may be increased.  
           [0013]    It is a further object of the present invention to provide an improved gear system which can operate at high velocities and heavy loads.  
           [0014]    It is still a further object of the present invention to provide an improved gear system which may be retrofitted into an existing gear system.  
           [0015]    It is yet another object of the present invention to provide an improved gear system in which the contact ratio between the gears may be varied.  
           [0016]    It is also an object of the invention to provide an improved method of rolling contact between the driven teeth and the driver.  
           [0017]    Another object of the invention is to reduce or eliminate the amount of lubrication needed for power or motion transmission between the gears.  
           [0018]    Yet another object of the present invention is to provide an improved gear system which requires less energy to operate as compared to conventional gearing systems.  
           [0019]    Other objects, features and advantages of the present invention will become apparent from the following detailed description taken in conjunction with the accompanying drawings.  
       
    
    
     BRIEF DESCRIPTION OF THE DRAWINGS  
       [0020]    The accompanying drawings which are incorporated into and constitute a part of this specification, illustrate a preferred embodiment of the invention and together with a general description of the invention given above and the detailed description of the preferred embodiment given below serve to explain the principals of the invention.  
         [0021]    [0021]FIG. 1 is a schematic cross-sectional view of two gears of the prior art.  
         [0022]    [0022]FIG. 2 is a schematic cross-sectional view of a rack and pinion gear system of the prior art.  
         [0023]    [0023]FIG. 3 is a schematic cross-sectional view of a circular gear system of the present invention.  
         [0024]    [0024]FIG. 4 is a schematic cross-sectional enlarged view of part of the gear system of FIG. 3, showing a pin and tooth relationship.  
         [0025]    [0025]FIG. 5 is a graph of pin radius versus pressure angle relating to the present invention wherein the gears of the present invention are circular.  
         [0026]    [0026]FIG. 6 is a schematic cross-sectional view of a rack and pinion gear system of the present invention.  
         [0027]    [0027]FIG. 7 is a graph of pin radius versus pressure angle relating to the present invention wherein the present invention employs a rack and pinion gear system.  
         [0028]    [0028]FIG. 8 is a schematic cross-sectional view of the split tooth and split pin embodiment of the present invention.  
         [0029]    [0029]FIG. 9 is a schematic cross-sectional view of the stationary pin and sleeve embodiment of the present invention. 
     
    
     DETAILED DESCRIPTION OF THE INVENTION  
       [0030]    Nearly all gears used in standard gearing operations today have teeth based upon an involute profile. Referring now to FIG. 1, two prior art gears  2 ,  10  are illustrated. A first gear  2  is rotatable on an axis  4  and has a base circle  6  defined by radius r bc . Given that base circle  6 , having radius r bc , the shape of the working part of each of the teeth  8  is defined as shown in FIG. 1. Teeth  8  circumscribe gear  2 , but are illustrated only in one arced portion for simplicity. While other parameters of gear  2  such as undercut, the thickness of teeth  8 , and the spacing of individual teeth  8  are required to completely define gear geometry, the base circle  6  alone determines the fundamental contact geometry.  
         [0031]    In operative engagement with the first gear  2 , in FIG. 1, is a second gear  10  rotatable on an axis  12 . Gear  10  has a plurality of circumferentially positioned teeth  14  which contact teeth  8  of gear  2  at a mesh area  16 . Teeth  14  circumscribe gear  10 , but are illustrated only in one arced portion for simplicity. It is in mesh area  16  where a driving gear  10  exerts a force, typically a torque, on a driven gear  2 . A driving gear  10  refers to a gear which is rotated by a power source and a driven gear  2  refers to a gear which is rotated by another gear, such as the driving gear  10 . In FIG. 1 it is arbitrary which gear is the driving gear  10  and which is the driven gear  2 . Thus for example only, the gear  10  will be the driving gear  10  and the gear  2  will be the driven gear  2 .  
         [0032]    The axis  4  of gear  2  and the axis  12  of gear  10  are aligned along a center line  18 . The shortest distance between the axis  4  of gear  2  and the axis  12  of gear  10  is called the center distance C, denoted by the letter C in FIG. 1. In FIG. 1, the distance on the center line  18  between the axis  4  and the axis  12  is the center distance C for gear  2  and gear  10 .  
         [0033]    There exists along center line  18  a point of contact between gear  2  and gear  10  called a pitch point  20 . Pitch point  20  divides the center distance C in the same proportions as the gear ratio for gear  2  and gear  10 . The gear ratio for two gears is the ratio of the number of teeth  8 ,  14  in two engaging gears  2 ,  10 , or the ratio of the gear spends or the ratio of the gear diameters.  
         [0034]    Pitch point  20  defines the location of pure rolling contact between gear  2  and gear  10 . It is at pitch point  20  where gear  2  and gear  10  contact in a rolling motion without slipping. For purposes of defining gear geometry, each gear, gear  2  and gear  10 , has a boundary called a pitch circle which intersects pitch point  20  and circumscribes each gear. For example, gear  2  has a pitch circle  22  shown by the dashed circle in FIG. 1 defined by radius r pc . Pitch circle  22  intersects pitch point  20 . Likewise, gear  10  has a pitch circle  24 , shown by a dashed circle, which intersects pitch point  20 . Pitch point  22  and pitch circle  24  are tangent to each other.  
         [0035]    Also intersecting pitch point  20  is a pitch line  26  which passes through the pitch point  20  and is perpendicular to the center line  18 . Pitch line  26  is also tangent to pitch circle  22  and pitch circle  24 .  
         [0036]    Another line intersecting pitch point  20  is a line called the line of action  28 . For involute gears the line of action  28  passes through pitch point  20  and is tangent to base circle  6  of gear  2 . The line of action  28  is the line and direction along which the pressure of individual teeth  8  acts. In other words, the line of action  28  shows the direction along which maximum force or torque is transferred from gear  10 , the driving gear, to gear  2 , the driven gear, with a minimum of energy losses. The maximum torque transferred from gear  10 , the driving gear, to gear  2 , the driven gear is at the point of action  20  from teeth  14  to teeth  8 .  
         [0037]    Defined by the pitch line  26  and the line of action  28  is a pressure angle  30 , also denoted in FIG. 1 by the angle Ø. Pressure angle  30  defines the angle which maximum torque may be transferred from gear  10 , the driving gear, to gear  2 , the driven gear. In the prior art gear system as shown in FIG. 1, the pressure angle  30  can be defined by the following equation: Ø=Sin[r b ÷r].  
         [0038]    For involute spur gear systems, such as that in FIG. 1, a gear alone does not have a fixed pressure angle  30 . The base circle  20  and defined operating or radius r pc  of pitch circle  22  determine the pressure angle  30 . For any gear pair the, pressure angle  30  is determined by the base circle  6  and the center distance C for two gears. For the involute spur gear system of FIG. 1, the pressure angle  30  is constant throughout mesh area  16  and the line of action  28  always intersects center line  18  at the pitch point  20 . This relationship satisfies the fundamental law of gearing and results in a constant angular velocity ratio between gear  2  and gear  10 .  
         [0039]    Referring now to FIG. 2, illustrating a prior art rack and pinion gear system, a pinion gear  32  rotatable on an axis  34  has a base circle  36  defined by radius r bc . Given base circle  36 , having radius r bc , the shape of the working part of each of the teeth  38  is defined as shown in FIG. 2. Teeth  38  circumscribe pinion gear  32 , but are illustrated only in one arced portion for simplicity.  
         [0040]    In operative engagement with, and contacting the pinion gear  32 , in FIG. 2, is a rack  40 . Rack  40  has teeth  42  which contact teeth  38  of pinion gear  32  at a mesh area  44 . It is in mesh area  44  where rack  40  exerts a force on pinion gear  32 .  
         [0041]    In FIG. 2, there exists along a vertical line  46  a point of contact between pinion gear  32  and rack  40  called a pitch point  48 . Pitch point  48  defines the location of pure rolling contact between pinion gear  32  and rack  40 . It is at pitch point  48  where pinion gear  32  and rack  40  contact in a rolling motion without slipping. For purposes of defining gear geometry, pinion gear  32  has a boundary called a pitch circle  50  which intersects pitch point  48  and circumscribes pinion gear  32 . For example, pinion gear  32  includes pitch circle  50  shown in FIG. 2 as the dashed circle and defined by radius r pc . Pitch circle  50  intersects pitch point  48 .  
         [0042]    Also intersecting pitch point  48  is a pitch line  52  which passes through the pitch point  48  and is perpendicular to the vertical line  46 . Pitch line  52  is also tangent to pitch circle  50 .  
         [0043]    Another line intersecting pitch point  48  is a line called the line of action  54 . In a rack and pinion gear system, the line of action  54  is normal to the profile  56  of the individual teeth  42  of rack  40 . Again, the line of action  54  is the line and direction along which the pressure of individual teeth  42  acts.  
         [0044]    Defined by the pitch line  52  and the line of action  54  is a pressure angle  58 , also depicted in FIG. 2 by the angle Ø. Pressure angle  58  defines the angle at which maximum torque may be transferred from rack  40  to pinion gear  32 . In the prior art rack  40  and pinion gear  32  system shown in FIG. 2, the pressure angle  58  can also be defined by Ø=Sin[r b ÷r].  
         [0045]    For the rack and pinion gear system of FIG. 2, the pressure angle  58  is determined by the angle of individual teeth  42  because the line of action  54  is normal to the profile  56  of individual teeth  42 . Moving the rack  40  further from the axis  34  of pinion gear  32  has no effect on pressure angle  58 .  
         [0046]    The gear system of the present invention  100 , as illustrated in the embodiment in FIG. 3, is constructed similar to a spur gear system, except that substituted in place of the involute teeth of one of the gears, there are rotating pins  60  which may be cylindrical or spherical in shape. The rotating pins  60  are preferably made of a material that has a strong wear resistance, for example, i.e. metal. The rotating pins  60  and gear teeth  8  in which they engage may preferably be coated with a friction reducing material. The gearing system of the present invention  100 , having rotating pins  60  that interact with gear teeth  8 , has advantages not found in common involute gearing. Instead advantages are created in the present invention that never before have been realized.  
         [0047]    In FIG. 3 , the reference numbers that are the same as the reference numbers of FIG. 1 correspond to like parts. Gear  10  has a plurality of pins  60  around the entire gear, instead of involute teeth. Pins  60  are a means for engaging teeth  8  of gear  2 . Again it is arbitrary which gear is the driving gear and which is the driven gear, but for the embodiment shown in FIG. 3, gear  10  will be the driving gear and gear  2  will be the driven gear.  
         [0048]    In the gear system of the present invention  100 , the teeth  8  and the pins  60  of gear  10  produce a variable pressure angle  30 . The reason for the variable pressure angle  30  is explained below.  
         [0049]    In the gear system of the present invention  100 , the pressure angle  30  is variable. This is because as each of the pins  60  rotate on the individual teeth  8  in the mesh area  16  resulting in constant direct physical contact between pins  60  and each of the teeth  8 , as illustrated in FIG. 4. FIG. 4 is an enlarged view of one of pins  60 , having a radius r p , in contact with the profile  61  of one of teeth  8 . The solid circle  63  represents a single pin  60  in a first position A and the dashed circle  65  indicates the single pin  60  in a second position B.  
         [0050]    [0050]FIG. 4 illustrates that the point of contact  20  moves along the profile  61  of each individual teeth  8 , from position A to position B, as each of the individual pins  60  rotates and rolls up the profile  61 . As the contact point  20  moves along the edge of profile  61  from position A to position B, each of the pins  60  is in substantially perfect rolling contact with one of the teeth  8 . It is along this profile  61  that maximum torque can be delivered from the driving gear to the driven gear with a minimum of loss.  
         [0051]    The moving contact point  20  between the pin  60  and tooth  8  results in a line of action  28  which rotates and must intersect the pitch point  20  at all times to satisfy the fundamental law of gearing. The line of action  28  moves from position A to position B and changes the pressure angle  30  defined between the line of action  28  and the pitch line  26 . Therefore, the pressure angle  30  is not fixed during the time of contact of each of individual pins  60  on each of individual teeth  8 .  
         [0052]    The moving contact point  20  allows force to be delivered along the line of action  28  for a longer period of time, allowing force to be delivered as maximum torque for a correspondingly longer period of time. This therefore allows greater torque to be delivered to gear  2  for a longer period of time. It is in this way that pressure angle  30  is variable.  
         [0053]    As illustrated in FIG. 3, pressure angle  30  is found by solving the following equation: C=[r b ÷cos Ø]+r p  Sin Ø, where C is the center distance between gear  2  and gear  10 . r b  is the base circle radius and r p  is the radius of the pin  60 . For the present invention  100  illustrated in FIG. 3, the pressure angle  30  will vary about 2 to 7 degrees.  
         [0054]    As an example of the variation in pressure angle  30  in the system shown in FIG. 3, FIG. 5 shows a graph of pin radius versus pressure angle at a center distance of C=5.684. The radius of the base circle  6  of gear  2  is equal to 5.446. As shown in FIG. 5, the pressure angle  30  decreases as pin radius of each of individual pins  60  on gear  10  increases.  
         [0055]    The reason for the decrease in pressure angle  30  with an increase in the radius of each of individual pins  60  is that the area of contact between a larger pin, and an involute tooth is larger and therefore the pitch point does not change or move as significantly with the moving of the larger pin. In other words, the angle of separation between the larger pin on the involute tooth is smaller and therefore the pitch point does not move as much with a larger pin as with a smaller pin. If the pitch point does not change as much then the line of action will not change as much and the pressure angle will be smaller.  
         [0056]    The gear system of the present invention  100 , as illustrated in another embodiment in FIG. 6, is constructed similar to a rack  40  and pinion gear  32  system, except in place of the involute teeth of the rack are rotating pins  62  which may be cylindrical or spherical in shape. The rotating pins  62  are preferably made of a material that has a strong wear resistance, like metal. The gearing system of this embodiment of the present invention  100  with rotating pins  62  that interact with involute gear teeth  38  does not follow the conventional and fundamental laws of gearing.  
         [0057]    In FIG. 6, the reference numbers that are the same as the reference numbers in FIG. 2 correspond to like parts. Rack  40  has a plurality of pins  62 . Pins  62  are a means for engaging teeth  38  of pinion gear  32 .  
         [0058]    In the gear system of FIG. 6, the involute teeth  32  of pinion gear  38  and the pins  62  of rack  40  produce neither a constant angular velocity or a constant pressure angle  58  for the same reasons as explained for the embodiment of FIG. 3 and shown in FIG. 4. The pressure angle can be found by solving the following non-linear equation, where R 2 =r p   2 +[C−(r b ÷cos Ø)]−{2r p [C−(r b ÷cos Ø)]Sin Ø} and where C is the center distance, r b  is the base circle radius and r p  is the radius of the pin. For the present invention illustrated in FIG. 6, the pressure angle  58  will vary about 2 to 7 degrees.  
         [0059]    As an example of the variation in pressure angle  58  in the system shown in FIG. 6, FIG. 7 shows a graph of pin radius versus pressure angle at a center distance of C=8.7. The radius of base circle  36  of gear  32  is equal to 5. As shown in FIG. 7, pressure angle  58  decreases as pin radius of each of individual pins  62  on gear  32  increases.  
         [0060]    For the present invention in FIGS.  3 - 7  the pressure angle can be minimized if the radius of the pin is maximized. For the gear system of the present invention, in which the radius of the means for engaging the involute teeth of the driven means is defined at a maximum as effectively less than the distance between midpoints of two involute teeth and for a constant distance between the driven means and the driving means, if the radius of the means for engaging the involute teeth of the driven means is maximized then the pressure angle of the gear system is minimized. Minimizing the pressure angle decreases the radial forces on the gear. However the advantage of the present invention is that even though the pressure angle is minimized, maximum torque can be delivered for the duration of the contact time between each individual gear teeth and each of the plurality of pins.  
         [0061]    In both the embodiments shown in FIG. 3 and  6 , the contact ratio varies as a function of pin diameter. The contact ratio is a measure of the number of gear teeth in full contact with the pins. A higher contact ratio provides increased load and torque capacity. In the present invention contact ratio varies from about 1 to 2.5 for a pin diameter of 20 to 90 mm.  
         [0062]    Another embodiment of the present invention is a split tooth and split pin arrangement, as illustrated in FIG. 8. FIG. 8 shows a split involute tooth  64  on a gear  66  engaging with a split involute pin  68 . Split involute pin  68  could be mounted on either a gear or a rack of a rack and pinion gear system. The difference between the split tooth and split pin arrangement is that the contact areas between the pin and gear tooth are separated by a gap in which there is no contact between the pin and the tooth.  
         [0063]    A further embodiment is the sleeve arrangement illustrated in FIG. 9. This embodiment uses a plurality of stationary pins  70  on which are mounted sleeves  72  which can rotate around pins  70 . This embodiment has several advantages, namely, ease of assembly and ease of retrofitting. Further the sleeve  72  replaces the need to make the pins  70  rotatable on both ends. The pin diameter can also be reduced.  
         [0064]    While there has been illustrated and described several embodiments of the present invention, it will be apparent that various changes and modifications thereof will occur to those skilled in the art. It is intended in the appended claims to cover all such changes and modifications that fall within the true spirit and scope of the present invention.