Variable speed transmission

A variable speed transmission transmits power between an input and output in which a pinion gear, a rotor and a cam are mounted for relative rotation therebetween. First and second gear rack pairs have opposite facing toothed gear racks which are spaced apart sufficiently to receive the pinion gear and cam therebetween when the pinion gear engages at least one of the gear racks, and the outer surface of the cam is adjacent at least one of the gear racks. First and second rack supports support the rack pairs against lateral movement to ensure full engagement between the pinion and sequentially engaged racks. The rack support slidingly engage the rotor to transmit power therebetween in such a manner that a resultant force vector from contact between the pinion and gear rack passes closely adjacent a point of contact between a rack support and a rotor. Teeth of the gears have inclined tips to facilitate meshing.

BACKGROUND OF THE INVENTION 
The invention relates to a fully variable power transmission apparatus in 
which rotary power applied to an input member is transferred to an output 
member, with means to permit a continuous or "stepless" change in speed 
and torque between the input and output members. 
Many different types of variable speed power transmissions are available, 
some types using trains of journalled toothed gears, or chains and 
sprockets, or pulleys and belts. Many types of transmissions provide a 
relatively limited speed range or torque ratio, with a number of 
pre-selected ratios provided between minimum and maximum gear ratios. Such 
ratios are in so-called "steps" or are discontinuous, i.e. to progress 
through the full range of gears a series of incremental steps must be 
taken, intermediate positions between the steps being unattainable. That 
is, this type of transmission does not provide a continuously variable 
speed or torque throughout the entire range between minimum and maximum 
gear ratios. 
While attempts have been made to produce "stepless" or continuously 
variable transmissions, many additional complications are usually 
introduced often with limitations on maximum power that can be 
transmitted. A stepless transmission has the advantage of providing 
continuous variations in torque and speed between input and output 
members, but usually the range of gear ratios is relatively limited 
compared with conventional "stepped" transmissions as previously 
described. 
Examples of stepless gear transmissions are shown in U.S. Pat. No. 
4,411,165 issued to Evans in 1983 and in U.S. Pat. No. 4,800,768, issued 
to Kazuta in 1989. While both of these patents disclose stepless 
transmissions which provide a relatively wide gear ratio range, in the 
applicant's opinion the designs appear to be limited to relatively low 
power applications, and it is anticipated that difficulty can be 
experienced in adapting the patented transmissions to heavy duty 
application. Both patents disclose an input pinion surrounded by four 
toothed racks which are connected together as two pairs of oppositely 
disposed racks, and which transfer power to an output rotor. Spacing 
between axes of the input pinion and output rotor is variable, and the 
transmission can be likened to an input pinion driving an output annular 
gear of variable diameter. The pinion is engaged by the racks in sequence 
which rotate thereabouts so that there is intermittent meshing between the 
pinion and the racks in sequence. Both patents disclose resilient means to 
permit the racks to move radially relative to the pinion. 
One problem that appears to be common to both patents is the difficulty of 
ensuring smooth transfer of torque from the pinion gear to each rack in 
sequence. As the pinion rotates, each rack is engaged by the pinion in 
sequence, and thus there is constant engagement and re-engagement of racks 
by the pinion. Even for a constant speed ratio between the input and 
output members, difficulty can be experienced when the pinion gradually 
disengages from one rack, and re-engages with the next rack in sequence. 
When the gear ratio is actually changing, by changing spacing between axes 
of the input and output members, the difficulties of ensuring a smooth 
transfer of torque between the pinion and the racks is increased. 
Sometimes a tooth of the pinion can come into direct tip-to-tip contact 
with an opposing tooth of a rack about to be engaged. If this occurs, the 
resilience provided in the mounting of the rack permits the rack to move 
generally radially outwardly to some extent, but this movement is often 
insufficient to prevent high forces from being generated between the 
pinion and the rack, which can cause damage to the rack. In any event, 
when such tip-to-tip contact occurs, an impulsive load is inadvertently 
applied to the rotor, causing intermittent output of power therefrom. 
Furthermore, in each of the two said patents, relatively high friction 
forces can be generated between sliding members, with a corresponding loss 
of power. Furthermore, the direction of transfer of forces between some 
adjacent components is less than optimum, significantly increasing 
mechanical losses in the transmission. 
SUMMARY OF THE INVENTION 
The invention reduces many of the difficulties and disadvantages of the 
prior art by providing a "stepless transmission" of a type similar to the 
two patents above with improved sequential meshing of the racks with the 
pinion by a specific structure for mounting the racks with respect to the 
pinion. Furthermore, sliding frictional losses are reduced considerably by 
improving cooperation between members transferring power from the racks to 
the rotor. Also force from the pinion transferred to the racks is then 
directed to the rotor in a more efficient manner than in the patented 
devices. 
A power transmission according to the invention has a body, a pinion gear 
mounted for rotation about a pinion axis relative to the body, and a cam 
means having curved inner and outer surfaces. The cam means and the pinion 
gear are mounted for relative rotation therebetween. The transmission 
further includes first and second gear rack pairs. Each gear rack pair has 
oppositely facing toothed gear racks which are spaced apart sufficiently 
to receive the pinion gear and cam means therebetween when the pinion gear 
fully engages at least one of the toothed gear racks, and the outer 
surface of the cam means is adjacent at least one of the oppositely facing 
toothed gear racks. The transmission also includes first and second rack 
support means and a rotor. The first and second rack support means support 
the first and second gear rack pairs respectively against lateral movement 
of the rack pair relative to the respective support means. This ensures 
full engagement between the pinion and the respective racks sequentially 
as required, thus reducing problems associated with sequential engagement 
between racks of the prior art. The rotor is mounted for rotation about a 
rotor axis and has rotor engaging means for engaging the first and second 
rack support means so as to transmit power therebetween. 
Also, in a transmission apparatus generally as described above, preferably, 
each rack support means has oppositely disposed driving sides disposed 
generally normally to the rack pairs mounted thereon. Each driving side 
has a driving side inner portion and two driving side outer portions 
spaced on opposite sides of the respective driving side inner portion. The 
driving side outer portions project outwardly further from the respective 
driving side inner portion relative to the pinion. The rotor engaging 
means include four engaging surfaces disposed as a square about the rotor 
axis. The two driving side outer portions of each rack support means 
engage respective engaging surfaces of the rotor to permit relative 
movement along the engaging surfaces. Preferably, a resultant force vector 
from contact between the pinion and the respective gear rack passes 
closely adjacent a point of contact between an adjacent driving side outer 
portion and the respective engaging surface of the rotor. 
Also, in a transmission apparatus generally as described above, the pinion 
and gear racks have teeth, some of which preferably have obliquely 
inclined tooth tips. In this way, the teeth generate oblique forces on 
each other if opposing tooth tips contact each other momentarily prior to 
full engagement with each other. The oblique forces between the teeth 
shift the rack teeth generally tangentially relative to the pinion teeth 
immediately prior to engagement so as to facilitate smooth meshing of the 
teeth. 
A detailed disclosure following, related to drawings, describes a preferred 
embodiment of the invention which is capable of expression in structure 
other than that particularly described and illustrated.

DETAILED DISCLOSURE 
FIGS. 1 through 3 
A variable speed transmission apparatus 10 according to the invention has 
an input or pinion shaft 12 and an output rotor 14, and thus is described 
initially by showing direction of power transmitted through the apparatus 
from the input shaft 12 to the output rotor 14. Furthermore, for 
convenience of description, direction of rotation of the input shaft or 
member 12 is shown as an arrow 16, which is also the direction of rotation 
of the output rotor or member 14. However, the direction of rotation of 
the input and output members can be reversed, and direction of power 
transmission through the apparatus can be reversed i.e., the input and 
output as disclosed can be interchanged. Also, for convenience of 
description, some components are shown to be stationary, while others 
rotate relative thereto. Clearly, in certain applications some of these 
configurations can be interchanged. 
Many components of the apparatus 10 and their function resemble closely 
equivalent components shown in U.S. Pat. No. 4,800,768, issued to Kazuta 
as previously described. Consequently, the disclosure of this patent is 
incorporated herein by reference. However, to clarify understanding of the 
present invention, even those components similar to the structure shown in 
Kazuta will now be briefly described. The apparatus includes a pinion gear 
19 secured to the shaft 12 for rotation therewith about a pinion axis 21. 
The shaft 12 is mounted for rotation in a stationary shaft journal 23, and 
receives power from an external source, not shown. 
Referring to FIG. 1, the apparatus 10 also includes a crescent-shaped cam 
means 26 and a generally rectangular cam mounting means 28 which mounts 
the cam and has a bore 29 to journal the cam mounting means on the shaft 
12. Thus the cam means and the pinion gear are mounted for relative 
rotation therebetween. The cam means 26 has curved inner and outer cam 
surfaces 31 and 32, the inner surface 31 being concentric with the pinion 
axis 21 and being spaced from teeth 33 of the pinion to permit relative 
rotation therebetween. 
The rotor 14 is an annulus and is mounted for rotation about a rotor axis 
35 on a disc-like rotor mounting means 37. The rotor axis 35 is spaced 
from the pinion axis 21 by an axis spacing 36 which, in FIG. 1, is a 
minimum axis spacing which represents the narrowest gear ratio attainable 
by the apparatus. As will be described, the spacing 36 between the axes 21 
and 35 is variable, and variation of the axis spacing varies gear ratio 
between the pinion and rotor. The rotor mounting means 37 has a circular 
rotor bearing surface 41 which is concentric with the axis 35 and 
complementary to an annular bearing surface 43 of the rotor. Preferably, a 
plurality of bearing elements such as balls or rollers 45 are provided to 
reduce friction of the rotor rotating relative to the rotor mounting 
means. A cam axis 38 coincides with a diameter of the pinion and rotor and 
thus intersects the pinion axis 21 and the rotor axis 35 and passes 
through a generally mid position of the cam means 26. The axis 38 clearly 
serves as a general axis of symmetry of the apparatus, with the exception 
of the cam means as will be described. 
As best seen in FIG. 2, the apparatus has a body 50 which includes the 
rotor mounting means 37 and also stationary circular first and second body 
side members 54 and 55 respectively, which have undesignated aligned 
central clearance openings therein to receive the pinion shaft 12. An 
outer annular plate 56 is secured to the rotor 14 to rotate therewith, and 
has an inner clearance opening 57 which is sealed by the first body side 
member 54, to provide an adequate overlap therewith, similarly to the 
Kazuta patent. A bolt 59 secures the second body side member 55 to the cam 
mounting means 28 and also to the stationary shaft journal 23. An inner 
annular plate 58 is secured to the rotor mounting means 37 so as to move 
therewith, has a clearance opening 60 which is sealed by the second body 
side member 55 in a manner similar to that described with reference to the 
Kazuta patent. Thus, the side members 54 and 55, the cam mounting means 28 
and cam means are secured relative to the stationary shaft journal. As 
will be described, the rotor mounting means 37 is non-rotatably movable 
relative to the body and journal 23 to permit variation in the axis 
spacing 36. 
As best seen in FIG. 1, the apparatus includes first and second gear rack 
pairs 61 and 62 respectively, the pairs being disposed perpendicularly to 
each other and surrounding the pinion gear 19. The first gear rack pair 
has oppositely facing toothed gear racks 65 and 66 and the second gear 
rack pair has similar toothed gear racks 67 and 68 respectively. The gear 
racks and pinion gear have similar meshing teeth, in which the racks are 
engaged sequentially by the pinion gear as will be described. Inter-rack 
spacing 63 between opposite gear racks of a pair is equal and critical, 
and the present invention provides a means of ensuring that there is 
negligible variation in the said spacing. The gear racks of a particular 
pair are spaced apart at the spacing 63 which is sufficient to receive the 
pinion gear 19 and the cam means 26 therebetween when the pinion gear 
fully engages at least one of the toothed gear racks, for example the gear 
rack 65 as shown, and the outer surface 32 of the cam means is in contact 
with at least one of the oppositely facing toothed gear racks, in this 
instance the gear rack 66. Thus, teeth 69 of the rack 65 are in engagement 
with the teeth 33 of the pinion, and at least one tooth of teeth 70 of the 
rack 66 is in contact with the outer surface 32 the cam means. 
Again referring to FIG. 1, the apparatus 10 includes first and second rack 
support means 71 and 72 respectively for supporting the first and second 
gear rack pairs 61 and 62 respectively against lateral movement of the 
rack pair relative to the respective support means. This is to maintain 
the spacing 63 constant, which ensures full engagement between the pinion 
and respective rack, the engagement occurring sequentially as will be 
described. The first rack support means 71 has a pair of parallel recesses 
75 and 76 respectively disposed on opposite sides of the rack support 
means and spaced apart on opposite sides of the pinion. The recesses 75 
and 76 retain the respective toothed racks 65 and 66 and have respective 
outer bearing surfaces 77 and 78 which are accurately linear and smooth to 
reduce sliding friction thereon. The racks 65 and 66 have similar linear 
outer bearing surfaces 81 and 82 respectively disposed along outer edges 
of the respective racks opposite to the teeth 69 and 70 of the racks. The 
outer bearing surfaces 81 and 82 of the gear racks 65 and 66 are in 
engagement with the outer bearing surfaces 77 and 78 of the respective 
recesses of the rack support means 71. 
The rack 65 has oppositely disposed rack end portions 87 and 88, space 
between the end portions defining length 90 of the rack. The recess 75 has 
a pair of oppositely spaced recess end portions 92 and 93 for receiving 
the rack end portions 87 and 88 respectively. Space between the recess end 
portions 92 and 93 defines length 95 of the recess, which is greater than 
the length 90 of the rack retained therein, so as to permit limited 
longitudinal movement of the rack relative to the rack recess per arrow 
84. The ends 87 and 88 of the rack 65 have axially aligned openings which 
receive similar elongated resilient members 98 therein. Thus a resilient 
means, which can be made from an elastomeric or rubber compound, is 
mounted at each end of each rack, each resilient member cooperating with 
an adjacent end portion 92 and 93 respectively of the rack recess so as to 
apply an inwards axial force to the rack. The inwards force applied at one 
end of the rack is opposed by an opposite inwards force at an opposite end 
of the rack, which tends to center the rack within its recess. Preferably, 
the length of the recess is slightly more than one tooth width longer than 
the rack. Thus, when the rack is centered in the recess, it can move in 
either direction along the recess approximately one half of the tooth 
width. 
The end portions 92 and 93 of the recess 75 have inner edge portions 101 
and 102 disposed generally oppositely to the outer bearing surface 77 of 
the recess. Space between an inner edge portion of the recess and an 
oppositely facing adjacent end portion of the outer bearing surface of the 
recess is essentially equal to transverse width of a respective end 
portion of the rack. There is clearance sufficient only for axial movement 
of the rack with negligible lateral movement of the rack relative to the 
recess, thus maintaining the spacing 63 constant for reasons to be 
described. 
The rack 66 is similarly resiliently mounted in the respective recess 76 by 
a similar structure and is not described. The racks 67 and 68 are 
similarly resiliently mounted in respective recesses within the second 
rack support means 72 and are also not described. 
The first rack support means 71 is generally rectangular, and has 
oppositely disposed, slightly inwardly curved driving sides 104 and 105 
disposed generally normally to the racks 65 and 66 mounted thereon. The 
support means 71 also has a pair of oppositely disposed straight, 
interconnecting sides 108 and 109 which are parallel to the racks 65 and 
66 and interconnect the driving sides. The driving side 104 has a driving 
side inner portion 111, and two driving side outer portions 113 and 114 
spaced on opposite sides of the inner portion 111. The driving side outer 
portions project further outwardly from the inner side portion 111 
relative to the pinion and preferably have bearing means to reduce 
friction and wear as will be described. It can be seen that the driving 
side 105 is generally similar to the driving side 104 and thus both 
driving sides are generally concave so as to extend smoothly inwardly to 
the respective driving side inner portion from the two respective driving 
side outer portions thereof. 
The second rack support means 72 is generally similar to the first rack 
support means 71 and thus has similar curved driving sides 117 and 118 
disposed generally normally to the rack pair 62 mounted thereon, and 
interconnected by straight interconnecting sides 119 and 120 which are 
parallel to the rack pair 62. While the driving sides are shown to be 
concave, any shape which has outer portions projecting further outwardly 
than the inner portions thereof will suffice. 
The rotor 14 has four engaging surfaces 131 through 134 disposed as a 
square symmetrically about the rotor axis 35. The two driving side outer 
portions 113 and 114 of the driving side 104 are in contact with the 
engaging surface 131 of the rotor. Similarly, outer portions 125 and 126 
of the driving side 105 contact the engaging surface 133 of the rotor. 
Corresponding outer portions of driving sides 117 and 118 of the second 
rack mounting means 72 contact the engaging surfaces 132 and 134 of the 
rotor as shown partially in FIG. 1. Space between the outer portions of 
each rack support means, or, if provided, the bearing means thereof, is 
closely matched to space between oppositely facing engaging surfaces of 
the rotor to essentially eliminate any relative rotation therebetween, to 
reduce the chances of "binding" of the rack support means and to limiting 
the movement thereof to smooth longitudinal movement along the engaging 
surfaces. 
As seen in FIG. 3, the rack 65 has a thickness 136 defined by space between 
outer and inner side faces 137 and 138 thereof. Also the rack support 
means 71 has a thickness 139 defined by space between opposite side faces 
of the rack support means, namely space between an outer side face 141, 
and an inner side face 142. The recess 75 has a depth 145 less than the 
thickness 139 of the rack support means. The recess 75 has a recess rear 
face portion 147 which engages the inner side face 138 of the rack 65. The 
rear face portion 147 has a thickness 151 which is no greater than 
difference between the thickness 139 of the rack support means and the 
thickness 136 of the rack. Thus, when the rack 65 is received in the rack 
recess 75, the outer side face 137 of the rack does not project beyond, 
and preferably is slightly recessed with respect to, the adjacent inner 
side face 142 of the rack support means. 
The above relative dimensions provide a dimensional relationship for a rack 
and its respective rack support means, and a similar dimensional 
relationship relates to the rack 66 in its recess 76, and the racks 67 and 
68 in respective recesses 153 and 154 in the second rack support means 72. 
As seen in FIG. 3, the recesses of second rack support means 72 face in an 
opposite direction than the recesses of the first rack support means. In 
this way, outer faces of the racks 67 and 68 of the second rack support 
means are closely adjacent outer faces of the racks 65 and 66 of the first 
support means. This results in the recess rear face portions of each rack 
support means being disposed oppositely from each other to be exposed on 
the outside of the pair of rack support means which constitute a 
self-contained "package" of components for assembly, which are quite well 
protected. As seen also in FIGS. 2 and 3, the engaging surfaces 132 and 
134 of the rotor 14 have a width 158 at least equal to sum of the 
thicknesses of the first and second rack support means so as to provide 
adequate engagement therewith. 
From the above, it can be seen that the rack pairs of the present invention 
are mounted in respective rack support means to cooperate with the rotor 
and the pinion in a manner quite different from that shown in the patents 
to Evans and Kazuta. In both said patents, the rack means are resiliently 
mounted to permit some relative lateral movement of the rack, that is 
movement of the rack that has a generally radial component relative to the 
pinion gear during engagement. Such generally radial movement relative to 
the pinion gear has been shown to increase the risk of head-to-head 
contact between teeth of the pinion and a rack means about to be engaged, 
as will be shortly described. 
As best seen in FIGS. 1 and 3, the driving side 117 of the second rack 
support means 72 has an outer portion 155 provided with a roller 156 to 
serve as the bearing means as previously described. The roller 156 is 
journalled on an axle 157 secured in a recess of the outer portion 155, 
and an axle stop 159 retains the roller on the axle. A portion of the 
roller projects beyond the outer portion 155 to contact the engaging 
surface 132 of the rotor. The three remaining outer portions of the rack 
support means 72 are similarly provided with rollers to serve as bearing 
means, which are shown in broken outline. Similarly the outer portions 
113, 114, 125 and 126 of the rack support means 71 are also supplied with 
similar rollers or equivalent means to reduce sliding friction and wear of 
the rack support means as it moves along the engaging surface as will be 
described with reference to FIGS. 6 and 7. In most of the following 
description, for simplicity the rollers adjacent the outer portions of the 
rack support means are not referred to specifically. 
Similarly to the said Kazuta patent, as best seen in FIG. 1 the cam 
mounting means 28 of the present invention has spaced parallel mounting 
sides 161 and 162 which are disposed parallel to the cam axis 38 which 
passes through the pinion axis 21. The rotor mounting means 37 has an 
elongated rectangular recess 166 having a pair of spaced parallel recess 
side walls 168 and 169 parallel to the axis 38, and spaced parallel recess 
end walls 170 and 171 perpendicular to the axis 38, the walls being shown 
in broken outline. The side walls 168 and 169 engage the mounting side 
walls 161 and 162 of the cam mounting means 28 to permit relative axial 
sliding movement between the rotor mounting means 37 and the cam mounting 
means 28 along the cam axis 38. Because the rotor axis is fixed relative 
to the rotor mounting means, this axial movement of the rotor mounting 
means varies the spacing 36 between the pinion axis 21 and the rotor axis 
35. A gear shift connector 164 is a short strip aligned with the cam axis 
38 and connected to the rotor mounting means 37 by a bolt 165. Axial 
movement of the strip 164 per a double-headed arrow 163 similarly shifts 
the rotor mounting means with respect to the cam mounting means 28 so as 
to vary spacing 36, as will be described with reference to FIG. 4. Thus, 
the rotor mounting means 37 is non-rotatably mounted relative to the cam 
means 26 but can move axially relative thereto. The mounting sides 161 and 
162, and the recess side walls 168 and 169 have complementary guide means 
to maintain the axial movement within a main plane of the rotor 14. Thus 
the rotor mounting means 37 is mounted for guided movement along the cam 
axis 38 with negligible lateral deviation therefrom. It is seen that the 
rotor itself is movable along a radius passing through the pinion axis. 
As seen in FIG. 1, the curved outer surface 32 of the cam means 26 is 
asymmetrical with respect to the cam axis 38. This contrasts with the cam 
means shown in the two prior art patents which are symmetrical with 
respect to the corresponding axis. The surface 32 is asymmetrical because 
the gear racks have teeth which are asymmetrical and have obliquely 
inclined tooth tips as will be described with the reference to FIGS. 5 and 
6. 
FIGS. 1 and 4 
The teeth of the gear racks 65 through 68, and the teeth 33 of the pinion 
have generally standard involute tooth form for driven or driving faces of 
the teeth, with modified tooth tips as will be described with reference to 
FIG. 7. Face profiles of the teeth function in a normal manner as below. 
Referring particularly to FIG. 4, a known resultant force vector 175 is 
generated at a point of contact 176 of a side 173 of a particular driving 
tooth 33 of the pinion gear 19 with a side 174 of an engaged driven tooth 
69 of the respective rack 65. As seen in FIG. 1, the vector 175 passes 
closely adjacent a point of contact 177 between the adjacent driving side 
outer portion 125 and the respective engaging surface 133 of the rotor. As 
is well known, the angle of the resultant force vector is dependent on the 
"pressure angle" of the involute tooth form, which in this instance is 20 
degrees. Consequently, the resultant force vector 175 is inclined at an 
angle 179 of 20 degrees to a tangent 178 passing through the point of 
contact 176 of the teeth, i.e. it is inclined at 20 degrees to a line 
normal to the cam axis 38 or is inclined at 20 degrees to a longitudinal 
axis of the rack. Clearly, there will be some variation in the actual 
direction of the resultant force vector and the point of contact 177 is 
selected to be generally in line with the resultant force. This enables 
the force from the pinion, applied to the rack and thence to the rack 
support means to be applied at an essentially maximum radius from the 
pinion axis, thus ensuring essentially optimum transfer of force in the 
pinion to the rotor. This is a considerable improvement of driving force 
application and direction when compared with the said prior art. 
As will be described with reference to FIGS. 5 and 6, the resultant force 
vector 175 can be resolved into two components of force, namely a major 
component disposed at 90 degrees to the engaging surface 133 which applies 
a driving force to the rotor to generate torque, and a minor component 
parallel to the engaging surface 133 which applies a force to the rack 
support means 71 in the direction of an arrow 172. When the pinion first 
engages the rack 65, the first rack support means initially moves relative 
to the engaging surface 133 in direction of the arrow 172, and thus is 
assisted by the minor component of force. 
FIG. 5 
The apparatus 10 is shown in a widest gear ratio configuration between the 
rotor and pinion in which the pinion and rotor axes 19 and 35 are spaced 
apart by a maximum spacing 180. Similarly to the apparatus described in 
the Kazuta patent aforesaid, the gear ratio of the present invention can 
be changed by actuating the gear ratio change means or shift connector 164 
which causes the rotor to shift relative to the pinion by relative sliding 
between the sides 161 and 162 of the cam mounting means, and the side 
walls 168 and 169 of the rectangular recess in the rotor mounting means. 
As previously described, the rotor can shift diametrically relative to the 
rotor mounting means and attain any position intermediate the minimum and 
maximum spacing between the axes 21 and 35, to attain any intermediate 
gear ratio as in a "stepless transmission". 
The resultant force vector 175, generated by contact between the pinion 
teeth 33 and the rack tooth 69 again passes closely adjacent the outer 
portion 125 of the driving side 105 where it contacts the engaging surface 
133 of the rotor. Clearly, as the rack support means 71 moves relative to 
the rotor 14, the point of contact 177 similarly moves, and the force 
vector 175 moves with it. Thus, for the two extreme positions of the rack 
support means with respect to the rotor, as shown in FIGS. 1 and 5, the 
point of contact 177 is intersected by, or is closely adjacent, the force 
vector 175, thus ensuring efficient transfer of force to the rotor. 
As briefly described previously, the resultant force vector 175 has a major 
force component 191 inclined normally to the engaging surface 133 of the 
rotor, and a minor force component 192 inclined parallel to the surface 
133. As the pinion rotates and first drives the rack 65, the minor force 
component 192 applies a force to the rack support means 71, which 
initially moves in direction of the arrow 172 until the interconnecting 
side 108 contacts the engaging surface 132 of the rotor as shown in FIG. 
5. This contact occurs when the pinion is about half-way along the rack 65 
and represents the outer limit of travel of the rack support means 71 with 
respect to the rotor. The side 108 remains in contact with the surface 132 
until the transfer of force between the rack 65, which is presently 
engaged by the pinion, to a following rack 67, which is about to be 
engaged by the pinion, as will be described with reference to FIG. 6. It 
is added that the interconnecting sides of the rack support means only 
contact the engaging surfaces of the rotor when in the widest gear ratio 
setting. For a narrower ratio (eg. as shown in FIG. 1) the displacement of 
the rack support means relative to the rotor is considerably reduced and 
the rack support means reciprocate between close outer limits without the 
interconnecting sides contacting the engaging surfaces. 
The rotor has a rotor index mark 181, and the pinion has a pinion index 
mark 182, shown in a datum position in which the two marks are coincident 
with the cam axis 38, and are both positioned in a "12 o'clock" position. 
This establishes a datum from which the gear ratio is to be determined, 
with reference to FIG. 6. 
FIGS. 5 and 6 
In FIG. 5, the index marks 181 and 182 are shown in initial positions on 
the axis 38, i.e. both marks are shown at 12 o'clock. In FIG. 6, the 
pinion gear 19 has rotated through an angle 183 so that the pinion mark 
182 has assumed a new position 182.1 at approximately 4 o'clock. In 
consequence, the rotor 14 has rotated through an angle 184 and the index 
mark 181 has assumed a new position 181.1 at approximately 2 o'clock. The 
gear ratio between the two marks is thus approximately 3:1, which 
represents the widest gear ratio for this specific example as described, 
in which the pinion axis 21 is spaced from the rotor axis 35 by the 
maximum axis spacing 180. Clearly, increasing diameter of the rotor and 
lengths of the engaging surfaces 131-134 of the rotor would increase this 
ratio. 
In FIG. 6, the pinion gear 19 is shown in transition and is essentially 
simultaneously disengaging from the "leading" rack 65, and engaging the 
"following" rack 67, which is the next rack in sequence as the rotor 
rotates per the arrow 16. A very short period of time before the 
configuration shown in FIG. 6, the rack 65 was under load, the 
interconnecting side 108 contacted the surface 132 and the resultant force 
vector 175.1, shown in broken outline, passed closely adjacent the point 
of contact 177 of the outer portion 125 with the engaging surface 133 of 
the rotor recess. When the rack 65 is unloaded from force from the pinion 
gear, the rack moves in direction of an arrow 185 to be re-centered within 
the respective recess 75 due to the resilient means acting at opposite 
ends thereof. In contrast, after the rack 67 has been synchronized for 
smooth meshing with the pinion and is subjected to force the rack 67, the 
rack 67 moves in a direction of arrow 186 to be moved to a right hand end 
of its respective recess 153, as viewed in FIG. 6. A second resultant 
force vector 187 from contact between the pinion gear 19 and the rack 67 
passes closely adjacent a point of contact 189 between an outer portion 
188 of the driving side 117 of the second rack support means 72 and the 
engaging surface 132 of the rotor 14. 
Referring again to FIG. 6, just prior to the relative positions as shown, 
the resultant force vector 175.1 from contact between the pinion and the 
rack 65 has a minor component of force 192.1 acting in direction as shown. 
This minor component acts on the rack support means 71 to hold it against 
the surface 132. As the pinion disengages the rack 65 and starts to engage 
the rack 67, the minor component is reduced to zero and the support means 
71 reverses direction and moves away from the engaging surface 133 as 
shown. When force from the pinion is fully transferred to the rack 67, the 
resultant force vector 187 has a minor force component 193 which acts on 
the second rack support means 72, moving it in direction of the arrow 193, 
until it contacts the engaging surface 131 of the rotor. 
Thus, in all positions of the rack support means, the resultant force 
vector generated from contact between the pinion and rack teeth acts to 
transfer force from the pinion and rack to the rotor at a generally 
enhanced leverage position from that shown in the previous patents. Also, 
in all gear ratio settings, a minor component of force of the resultant 
force vector initially acts on the particular rack support means so as to 
assist in moving the rack support means in the direction of the minor 
component until an outer limit of travel of the rack support means is 
reached. This initial direction is in the same direction of movement of 
the rack support means along the engaging surface due to operation of the 
apparatus. Thus, the minor force component, which is inevitable in any 
gear transmission of this general type, for a major portion of the 
operation acts on the rack support means in a direction so as to assist in 
movement of the rack support means. 
In FIG. 6, it can be seen that the pinion 19 is meshed instantaneously with 
teeth of the adjacent racks 65 and 67. If the apparatus were stopped in 
this configuration, because the pinion engages both racks, relative 
lateral motion between the pinion and racks is prevented and thus the 
rotor cannot be shifted with respect to the cam mounting means. Thus, when 
the apparatus is stationary in this "double meshed" configuration, the 
gear ratio cannot be changed. In practice, even though double meshing 
occurs four times per revolution of the rotor, it does not appear to be a 
problem as the apparatus quickly attains a configuration in which the gear 
ratio can be changed. 
FIGS. 6 and 7 
FIG. 6 shows a driving configuration an instant of time after transfer of 
force between adjacent racks 65 and 67, in which teeth of the pinion 19 
have fully engaged the following rack 67. In an ideal situation with 
accurately manufactured components, smooth transfer of drive between the 
leading rack and the following rack can occur with negligible interference 
between the teeth. However, in some circumstances tip-to-tip contact 
between teeth can occur momentarily prior to full engagement of the teeth. 
The present invention provides tips of the teeth with a modified form or 
geometry which is designed to essentially eliminate the possibility of 
damage resulting from tip-to-tip contact between teeth. With the present 
invention, if tip-to-tip contact occurs, it occurs for a fraction of a 
second only under negligible load, and results in relative movement 
between the rack teeth and pinion teeth, so that synchronized meshing can 
occur before a full load is applied to the following rack which is just 
being engaged. 
In contrast, in prior art structures if tip-to-tip contact between teeth 
occurs for a sustained length of time, full load was applied to the 
following rack with a considerable magnification of force from the pinion 
tooth to the rack tooth, commonly resulting in damage to the rack, and 
possibly the pinion. Thus sustained tip-to-tip contact of prior art 
devices must be eliminated for efficient use of the present mechanism, and 
the inclined tooth tips of the present invention as will be described 
alleviate this problem. 
FIG. 7 shows instantaneous tip-to-tip contact between a typical pinion 
tooth 33 and a typical rack tooth 195 of the rack 67. The teeth 33 and 195 
have respective main longitudinal axes 196 and 197 respectively, the axis 
196 being a radial axis passing through the pinion axis 21 and the axis 
197 being disposed normally to a longitudinal axis at the rack 67, not 
shown. The pinion tooth 33 has a tooth tip having a tip end face 199 which 
is inclined at an angle 200 to the axis 196. The rack tooth 195 has a tip 
having a tip end face 202 which is inclined at an angle 203 to the axis 
197. It can be seen that the tip end faces of teeth are obliquely inclined 
and generally straight, and radiused at opposite ends thereof to blend 
smoothly with adjacent faces of the teeth. 
As shown in FIG. 6, as the pinion transfers force from the rack 65 to the 
rack 67, in some instances initial contact between the pinion teeth and 
the teeth 195 of the following rack to be engaged can be generally of the 
type shown in FIG. 7. The obliquely inclined tooth tips are shaped 
relative to each other so as to generate oblique forces on each other if 
opposing tooth tips contact each other momentarily prior to full 
engagement with each other. As can be seen, oblique forces are generated 
on the tooth tip end face 202 which shift the rack 67 generally 
tangentially per an arrow 205 relative to the pinion teeth. This shift is 
generally "in reverse" to normal rack movement and occurs immediately 
prior to full engagement of the teeth so as to facilitate the meshing of 
the teeth. The rack 67 shifts initially rearwardly because the pinion gear 
19 is the input and is rotating at a constant velocity. Because the rack 
67 is resiliently mounted, initially the rack is free to shift to 
accommodate misalignment of the teeth by shifting axially relative to the 
respective rack support means. The resilient means, which otherwise tend 
to center the rack within its recess, thus yield to accommodate the forces 
and provide limited rearward axial shifting of the rack in response to the 
oblique forces generated on the tooth tips. In most situations of even 
slight tip-to-tip contact, the rack would tend to shift in direction of 
the arrow 205 to permit the pinion tooth 33 to continue its rotation 
whereupon it assumes a position forward of the said tooth 69, to attain a 
relative position as shown in FIG. 4. The resilient means cooperating with 
the rack would thus be required to allow the rack to shift approximately 
one half tooth width from center, to permit such engagement, after which 
the rack end portion contacts and is held against the recess end portion. 
After the rearward shifting of the rack, a driving force from the pinion is 
applied to a side 209 of an adjacent tooth 210 which shifts the rack 67 in 
direction of an arrow 207 which is opposite to the arrow 205. Thus, from 
the first instance of tip-to-tip contact between the teeth, the force 
acting on the rack 67 shifts from a rearward direction per arrow 205 to a 
forward direction per arrow 207. As full engagement of the teeth occurs, 
the rack 67 shifts to compress the resilient means at a forward end of the 
rack, after which full force is transferred to the rack support means, and 
thence to the rotor. Clearly, the above description occurs in a fraction 
of a second and results in negligible fluctuation of driving force applied 
to the rotor. 
The pinion and rack teeth as shown both have oblique tooth tip end faces 
199 and 202, inclined at similar angles 200 and 203 respectively which are 
approximately 70 degrees, but can be between 85 degrees and 45 degrees to 
longitudinal axes of the teeth. For manufacturing simplicity, it may be 
preferable to use conventional involute tooth forms for either the pinion 
or the racks, and thus if preferred, only one set of teeth need be 
modified, that is either the pinion teeth or the rack teeth. If the pinion 
teeth are modified, geometry of the pinion teeth is as follows. The pinion 
tooth has a leading end portion 212 and a trailing end portion 213 with 
respect to a direction of rotation of the pinion gear, shown per arrow 16. 
The leading end portion is spaced at a radial distance 215 from the pinion 
axis 21, which is greater than spacing 216 of the trailing end portion 213 
from the pinion axis. Thus the tip end face 199 of the pinion effectively 
has a clearance angle similar to a clearance angle of a rotating single 
point cutting tool. Similarly, the rack tooth 195 has a leading end 
portion 217 and a trailing end portion 218 with respect to direction of 
movement of the rack under forward driving load. i.e. per the arrow 207. 
The leading end portion 217 is spaced at a distance from the rack axis, 
not shown, which is less than a similar spacing of the trailing end 
portion from the rack axis. 
Because the tip end faces of the rack teeth are asymmetrical, the profile 
of the outer cam surface 32 is similarly asymmetrical as seen in FIG. 6. 
The outer cam surface 32 has a geometrical profile which reflects a locus 
of contact between tooth tips of a non-engaged rack sweeping past and 
contacting the outer surface, while a diametrically opposed rack is 
engaged with the pinion means. The process of generation of the 
geometrical profile of the surface 32 follows standard tooth profile 
generation practice and is not discussed further. It is added that tightly 
controlled clearances are necessary to ensure smooth meshing. 
It appears that, because the inwardly projecting trailing end portions of 
the rack teeth sweeping the surface 32 are displaced towards the right 
hand side of respective tooth axes as seen in FIG. 6, the outer surface 32 
of the cam means is enlarged towards the right hand end of the cam means 
when compared with the left hand side. Thus, with respect to rotation of 
the pinion 19, the cam profile resembles to some extent a highly cambered 
aerofoil section in which a leading end thereof, designated 220, has a 
greater cross-sectional area than a trailing end thereof, designated 221. 
Clearly, in this definition the leading end is approached and passed by a 
particular pinion tooth prior to the trailing end. Because the inner cam 
surface 31 is concentric with the axis 21, the outer cam surface 32 is 
disposed at a greater radial distance 223 from the pinion axis 21 adjacent 
the leading end 220 thereof than an equivalent radial distance 224 
adjacent the trailing end 221 thereof. 
OPERATION 
The apparatus 10 operates generally similarly to the apparatus described in 
the aforesaid Kazuta patent, with important exceptions as previously 
referred to, and as briefly described as follows. Referring to FIG. 1, the 
input or pinion shaft 12 is rotated per arrow 16 and is engaged with the 
rack 65 which compresses the resilient member 98 so that the rack 65 moves 
into the recess end portion 93 to contact the end face thereof. Force from 
the pinion is transferred as the resultant force vector 175 through the 
point of contact 177, which results in rotation of the rotor 14. If an 
outwards force is applied to the gear shift connector 164, the rotor 
mounting means 37, and with it the rotor, move outwardly radially along 
the axis 38 so as to increase the spacing between the axes 19 and 35, to 
eventually attain the widest axial spacing 180 as shown in FIG. 5. This 
represents the widest gear ratio. In both FIGS. 1 and 5, the rack support 
means 71 is shown stationary at a limit of its travel from the pinion axis 
19. 
Referring to FIG. 6, after rotation of the pinion axis through the angle 
183, the racks have swung through an angle of approximately 45 degrees as 
shown, and force from the pinion to the rack 65 is rapidly decreasing to 
zero, while force on the rack 67 is rapidly increasing to generate the 
resultant force 187 which passes closely adjacent the point of contact 189 
of the driving side outer portion 188 with the engaging surface 132. It 
can be seen that, irrespective of whether the racks are in a wide ratio 
position, or a narrow ratio position, fully engaged by the pinion, or 
partially engaged by the pinion, the point of contact of the rack support 
carrying that rack is essentially intersected by a resultant force vector 
generated by force between the pinion teeth and a particular rack tooth in 
contact. 
It is to be understood that, if the direction of rotation of the input 
shaft 12 were reversed, so that power was applied to the shaft 12 in a 
direction opposite to the arrow 16, the angle of the tips of the teeth of 
the pinion and/or the racks would similarly be reversed. Also, if the 
input power were applied to the rotor 14, corresponding changes in the 
angle of the teeth would be required, depending on the direction of 
rotation. 
It can be seen that the four racks function unitarily as an internal ring 
gear of varying diameter to vary gear ratio between the input pinion and 
the output rotor.