U.S. patent application number 17/351674 was filed with the patent office on 2021-10-07 for speed change device.
The applicant listed for this patent is Genesis Advanced Technology Inc.. Invention is credited to James B. Klassen.
Application Number | 20210310546 17/351674 |
Document ID | / |
Family ID | 1000005666431 |
Filed Date | 2021-10-07 |
United States Patent
Application |
20210310546 |
Kind Code |
A1 |
Klassen; James B. |
October 7, 2021 |
SPEED CHANGE DEVICE
Abstract
A speed change device comprising an inner race having an outer
surface, an outer race having an inner surface, and set of orbital
rollers including inner rollers in rolling contact with the outer
surface of the inner race and outer rollers in rolling contact with
the inner surface of the outer race.
Inventors: |
Klassen; James B.; (Langley,
CA) |
|
Applicant: |
Name |
City |
State |
Country |
Type |
Genesis Advanced Technology Inc. |
Langley |
|
CA |
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|
Family ID: |
1000005666431 |
Appl. No.: |
17/351674 |
Filed: |
June 18, 2021 |
Related U.S. Patent Documents
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Application
Number |
Filing Date |
Patent Number |
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16150022 |
Oct 2, 2018 |
11067153 |
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17351674 |
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14403942 |
Nov 25, 2014 |
10132392 |
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PCT/CA2013/050400 |
May 24, 2013 |
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16150022 |
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61652148 |
May 25, 2012 |
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Current U.S.
Class: |
1/1 |
Current CPC
Class: |
F16C 33/526 20130101;
F16C 33/36 20130101; F16H 1/28 20130101; F16C 33/523 20130101; F16C
33/585 20130101; Y02T 10/86 20130101; F16H 13/08 20130101; F16C
19/505 20130101; F16C 19/38 20130101; F16C 33/363 20130101 |
International
Class: |
F16H 13/08 20060101
F16H013/08; F16C 19/38 20060101 F16C019/38; F16C 33/36 20060101
F16C033/36; F16C 33/58 20060101 F16C033/58; F16C 19/50 20060101
F16C019/50; F16C 33/52 20060101 F16C033/52; F16H 1/28 20060101
F16H001/28 |
Claims
1. A speed change device comprising: an inner race having an outer
surface; an outer race having an inner surface; a set of orbital
rollers including inner rollers in rolling contact with the outer
surface of the inner race and outer rollers in rolling contact with
the inner surface of the outer race, wherein at least one of: (a)
each outer roller is formed of a pair of axially aligned roller
sections, or (b) each inner roller is formed of a pair of axially
aligned roller sections; each and every inner roller being in
rolling contact with two outer rollers and wherein the inner
rollers are longer than the outer rollers and the inner rollers are
connected to be driven by an input ring coaxial with the inner race
and the outer race; and each and every outer roller being in
rolling contact with two inner rollers.
2. The speed change device of claim 1, in which the input ring has
an inward facing surface and the inward facing surface is in geared
contact with the inner roller.
3. The speed change device of claim 1, in which each of the inner
race and outer race are centered on an axis, and, for each pair of
a first inner roller that contacts a first outer roller, where a
traction angle o is defined as the angle between a first line
extending outward from the axis through a center of the first inner
roller and a second line extending from the contact point of the
first outer roller with the outer race and a contact point of the
first inner roller with the inner race, and the first inner roller
contacts the inner race with a first coefficient of friction cf1and
the first outer roller contacts the outer race with a second
coefficient of friction cf2, cf1>tan(o) and cf2>tan(o).
4. The speed change device of claim 1, in which each inner roller
is formed of two or more axially aligned roller sections and each
outer roller is in contact with each roller section of each inner
roller that the outer roller contacts.
5. The speed change device of claim 1, integrated with a
preliminary speed change stage, and wherein the preliminary speed
change device has an integrated preliminary speed change stage
using the self-energizing present speed change system such that
camming of the rollers in the first stage increases the contact
force of these rollers on the outer race of the first stage such
that the outer race of the first stage can expand to transmit this
increased radial force to the input traction surfaces on the
rollers of the second stage.
6. The speed change device of claim 1, further comprising a ring
concentric with the inner race and outer race and the ring being
connected to the orbital rollers to drive or be driven by the
orbital rollers.
7. The speed change device of claim 6, in which the ring is
connected to the inner rollers to drive the inner rollers.
8. The speed change device of claim 6, in which the ring is
connected to the outer rollers to drive the outer rollers.
9. The speed change device of claim 1, in which the outer rollers
have a different diameter contact with the inner surface of the
outer race, than the diameter of contact of the outer rollers with
the inner rollers.
10. The speed change device of claim 1, in which the outer rollers
have a larger diameter primary torque transmitting contact with the
inner surface of the outer race, than the diameter in primary
torque transmitting contact with the inner rollers.
11. The speed change device of claim 1, in which the rollers of the
orbital rollers have conical faces that contact conical faces of
other rollers of the orbital rollers when the rollers are not
axially centered.
12. The speed change device of claim 1, further comprising
integrated rotational sensors.
13. A speed change device comprising: an inner race having an outer
surface; an outer race having an inner surface; a set of orbital
rollers including inner rollers in rolling contact with the outer
surface of the inner race and outer rollers in rolling contact with
the inner surface of the outer race, wherein at least one of: (a)
each outer roller is formed of a pair of axially aligned roller
sections, or (b) each inner roller is formed of a pair of axially
aligned roller sections; each and every outer roller being in
rolling contact with two inner rollers and wherein the outer
rollers are longer than the inner rollers and the outer rollers are
connected to be driven by an input ring coaxial with the inner race
and the outer race; and each and every inner roller being in
rolling contact with two outer rollers.
14. The speed change device of claim 13, in which the input ring
has an outward facing surface and the outward facing surface is in
geared contact with the outer rollers.
15. The speed change device of claim 13, in which each of the inner
race and outer race are centered on an axis, and, for each pair of
a first inner roller that contacts a first outer roller, where a
traction angle o is defined as the angle between a first line
extending outward from the axis through a center of the first inner
roller and a second line extending from the contact point of the
first outer roller with the outer race and a contact point of the
first inner roller with the inner race, and the first inner roller
contacts the inner race with a first coefficient of friction cf1
and the first outer roller contacts the outer race with a second
coefficient of friction cf2, cf1>tan(o) and cf2>tan(o).
16. The speed change device of claim 13, further comprising
integrated rotational sensors.
17. The speed change device of claim 13, in which the rollers of
the orbital rollers have conical faces that contact conical faces
of other rollers of the orbital rollers when the rollers are not
axially centered.
18. The speed change device of claim 13, in which the rollers of
the orbital rollers have conical faces that contact conical faces
of other rollers of the orbital rollers when the rollers are not
axially centered.
19. The speed change device of claim 13, integrated with a
preliminary speed change stage, and wherein the preliminary speed
change device has an integrated preliminary speed change stage
using the self-energizing present speed change system such that
camming of the rollers in the first stage increases the contact
force of these rollers on the outer race of the first stage such
that the outer race of the first stage can expand to transmit this
increased radial force to the input traction surfaces on the
rollers of the second stage.
20. The speed change device of claim 13, in which the outer rollers
have a different diameter contact with the inner surface of the
outer race, than the diameter of contact of the outer rollers with
the inner rollers.
Description
CROSS-REFERENCE TO RELATED APPLICATIONS
[0001] This patent application is a continuation of U.S. patent
application Ser. No. 16/150,022, filed Oct. 2, 2018, which is a
continuation of U.S. patent application Ser. No. 14/403,942, filed
Nov. 25, 2014, now issued as U.S. Pat. No. 10,132,392, which is a
national phase entry of PCT Appl. No. PCT/CA2013/050400, filed May
24, 2013, which claims the benefit under 35 USC 119(e) of U.S.
Patent Provisional Application No. 61/652,148, filed May 25, 2012,
all of which are incorporated herein by reference in their
entirety.
TECHNICAL FIELD
[0002] Speed change devices.
BACKGROUND
[0003] Gear transmission speed change devices are capable of high
torque but are complex to manufacture, exhibit backlash, gear
noise, and typically require multiple stages to achieve high
reduction ratios.
[0004] Traction drive speed change systems offer certain advantages
but typically produce less torque then a geared speed change device
of the same size.
[0005] Generally speaking, with traction drive systems it is
desirable, for many applications, to provide pure rolling contact
between torque transmitting members for low friction, high
efficiency, long service life, and increased traction. Furthermore,
it is desirable for the contact forces between the torque
transmitting members to increase automatically as torque increases,
so high loads are only generated between traction components when
necessary to allow increased torque output. This would provide the
benefit of increased service life and efficiency by reducing wear
surface loading when the actuator is not transmitting high
torque.
[0006] Other desirable characteristics of a rotary actuator are
zero backlash, backdriveability, low vibration, non-cogging output,
high rigidity, and quiet operation. High torque capacity for size
and weight are also desirable, as are a wide range of speed change
ratio possibilities (including very high speed change ratios) high
input speed capability (to allow the use of low torque input drive
systems) low input inertia, and a relatively large center thru-hole
for internal wiring (or possibly to allow an integrated electric or
other type of rotary drive motor within (or partially within) the
inner diameter (ID) of the actuator.
SUMMARY
[0007] There is disclosed a speed change device comprising an inner
race having an outer surface, an outer race having an inner
surface, and set of orbital rollers including inner rollers in
rolling contact with the outer surface of the inner race and outer
rollers in rolling contact with the inner surface of the outer
race.
[0008] In an embodiment, each and every inner roller is in rolling
contact with two outer rollers and each and every outer roller
being in rolling contact with two inner rollers.
[0009] In an embodiment, a set of orbital rollers includes 19 or
more inner rollers.
[0010] In an embodiment, each of the inner race and outer race are
centered on an axis, and, for each pair of a first inner roller
that contacts a first outer roller, where a traction angle o is
defined as the angle between a first line extending outward from
the axis through a center of the first inner roller and a second
line extending from the contact point of the first outer roller
with the outer race and a contact point of the first inner roller
with the inner race, and the first inner roller contacts the inner
race with a first coefficient of friction cf1 and the first outer
roller contacts the outer race with a second coefficient of
friction cf2, cf1>tan(o) and cf2>tan(o).
[0011] In an embodiment, either A or B or both is present each
inner roller is formed of two or more axially aligned roller
sections, B each outer roller is formed of two or more axially
aligned roller sections.
[0012] In an embodiment, at least each of the inner rollers or each
of the outer rollers have two different diameters, and, depending
on the ratio of the diameter or diameters of the inner rollers to
the diameter or diameters of the outer rollers, the orbital rollers
orbit in a direction the same as or opposite to rotation of the
inner race.
[0013] In an embodiment, the orbital rollers extend between axial
ends and the orbital rollers have decreasing diameter towards the
axial ends.
[0014] In an embodiment, the rollers of the orbital rollers have
conical faces that contact conical faces of other rollers of the
orbital rollers when the rollers are not axially centered.
[0015] In an embodiment, a multiple stage speed change device is
disclosed in which at least one of the stages comprises a speed
change device according to one of the disclosed embodiments.
[0016] In an embodiment, at least some of the orbital rollers are
spaced by one or more of a geared ring, a cage attached to at least
some of the orbital rollers and cooperating circular and boss
members on spaced apart orbital rollers.
[0017] Various of the embodiments accomplish one or more of the
desirable characteristics of a speed change device.
[0018] In various embodiments, there may be included any one or
more of the following features: one or more of the orbital rollers
are hollow, inner or outer rollers or both are formed of axially
aligned roller sections, the races are split, outer rollers or
inner rollers are longer and the longer rollers are driven by a
ring that is coaxial with the races, a drive or output ring as
geared contact with either inner rollers or outer rollers, traction
angles and/or friction coefficients have particular limits on their
values, the rollers are pre-loaded, the inner race consists of a
single cylindrical surface, the inner race consists of two or more
cylindrical surfaces separated by an area large enough to allow
room for other components, the outer race consists of a single
cylindrical surface, the outer race comprises two or more
cylindrical surfaces separated by an area large enough to allow
room for other components, the inner and or outer races consists of
two cylindrical surfaces on either side of a plane perpendicular to
the axis of the races on or near a center position, measured
axially from the outer edges of the cylindrical contact faces of
the inner and or outer races, a ring concentric with the inner race
and outer race and the ring is connected to the orbital rollers to
drive or be driven by the orbital rollers, the ring is connected to
the inner rollers to drive the inner rollers, the ring has an inner
surface and the inner surface is in contact with one or more inner
rollers, the inner surface of the ring is in geared contact with
one or more inner rollers, the inner surface of the ring is in
traction contact with one or more inner rollers, the ring is
connected to the outer rollers to drive the outer rollers, the ring
has an outer surface and the outer surface of the ring is in
contact with one or more outer rollers, which the outer surface of
the ring is in geared contact with one or more outer rollers, the
outer surface of the ring is in traction contact with one or more
outer rollers, all rolling contacts are geared contacts, all
rolling contacts are traction contacts, contacts of each inner
roller with the inner race and with the outer rollers that transmit
torque have a torque transmitting diameter and all the torque
transmitting diameters are equal, contacts of each outer roller
with the outer race and with the inner rollers that transmit torque
have a torque transmitting diameter and all the torque transmitting
diameters are equal, the outer rollers have a different diameter
contact with the inner surface of the outer race, than the diameter
of contact of the outer rollers with the inner rollers, the outer
rollers have a larger diameter primary torque transmitting contact
with the inner surface of the outer race, than the diameter in
primary torque transmitting contact with the inner rollers, the
outer rollers have a larger diameter primary torque transmitting
contact with the primary torque transmitting contact inner surface
of the outer race, than the primary torque transmitting contact
diameter in contact with the inner rollers, and the inner rollers
have the same primary torque transmitting contact diameter with the
outer rollers as with the primary torque transmitting contact
diameter of the inner race, the inner rollers have larger diameter
primary torque transmitting contact with the outer rollers and
smaller primary torque transmitting contact with the outer surface
of the inner race, the outer rollers have two diameters and inner
rollers have one diameter such that rotation of rollers causes
orbiting of the rollers in one direction resulting in rotation of
inner race in the same direction as roller orbit direction when the
outer race is fixed, the outer rollers have two diameters and the
inner rollers have one diameter such that rotation of rollers
causes orbiting of rollers in one direction resulting in rotation
of the inner race in the opposite direction as the roller orbit
direction when the outer race is fixed, the outer rollers have two
diameters and inner rollers have two different diameters such that
rotation of rollers causes orbiting of rollers in one direction
resulting in rotation of inner race in the same direction as roller
orbit direction when outer race is fixed, the outer rollers with
two diameters and inner rollers with two different diameters such
that rotation of the rollers causes orbiting of the rollers in one
direction resulting in rotation of the inner race in the opposite
direction as the roller orbit direction when outer race is fixed,
the orbital rollers extend between axial ends and at least some of
the orbital rollers have decreasing diameter towards at least one
of the axial ends, the rollers of the orbital rollers have conical
faces that contact conical faces of other rollers of the orbital
rollers when the rollers are not axially centered, structural
members are connected to one or both of the inner race and the
outer race, the orbital rollers are configured to bear radial load,
the inner race has a center through hole, a motor is integrated
with the speed change device to provide an input drive, and other
features listed in the claims or disclosure.
BRIEF DESCRIPTION OF THE FIGURES
[0019] Embodiments will now be described with reference to the
figures, in which like reference characters denote like elements,
by way of example, and in which:
[0020] FIG. 1 shows a principle of operation of two equal sized
contacting rollers positioned vertically with parallel axes between
two fixed horizontal contact surfaces.
[0021] FIG. 2 shows an example where the contact rollers of FIG. 1
are of different sizes.
[0022] FIG. 3 shows three equal sized contact rollers positioned in
a triangular stack between two fixed horizontal contact
surfaces.
[0023] FIG. 4 shows two races spaced apart by two large cylindrical
roller bearings between two fixed horizontal contact surfaces.
[0024] FIG. 5 shows two circular races and their direction of
rotation about their individual axes while rotating in tandem
between two fixed curved contact surfaces.
[0025] FIG. 6 charts the effect of the number of contacting rollers
on theoretical max torque.
[0026] FIG. 7 shows the angle between a radial line from the center
axis of the races through a contact point of a roller on a race,
called the traction angle line.
[0027] FIG. 8 shows a variety of traction angle lines and the
required coefficient of friction to produce traction.
[0028] FIG. 9 shows an example of a set of rollers on a race using
34 roller sets, where all of the inner rollers are contacting two
outer rollers, and all of the outer rollers are contacting two
inner rollers.
[0029] FIG. 10 shows an example of a set of rollers on a race as in
FIG. 9 where the outer rollers are larger than the inner
rollers.
[0030] FIG. 11 shows the minimum coefficient of friction required
to achieve the desire effect in the embodiment of FIG. 10.
[0031] FIG. 12 shows an example of a set of rollers on a race using
100 roller sets.
[0032] FIG. 13 shows an example of a set of rollers on race using
313 roller sets.
[0033] FIG. 14 shows geometry of traction angles.
[0034] FIG. 15 shows a simplified partial section of an embodiment
of a speed change device.
[0035] FIG. 16 shows an example of the conical angle of the
contacting alignment faces shown in FIG. 15.
[0036] FIG. 17 shows a simplified partial section of the present
device with the cover of the inner race removed.
[0037] FIG. 18 show a partial view of an embodiment of a speed
change device showing a preferable elliptical shape of the
bosses.
[0038] FIG. 19 shows a sectional view of bosses of FIG. 18.
[0039] FIG. 20 shows a schematic view of a rolling contact roller
spacing system.
[0040] FIG. 21 shows a simplified partial section a multiple stage
multi-element self-energizing speed change actuator.
[0041] FIG. 22 shows the embodiment of FIG. 21 where the inner race
is removed.
[0042] FIG. 23 shows an example of a radially expanding input ring
removed from the actuator.
[0043] FIG. 24 shows an alternative embodiment of the radially
expanding input ring of FIG. 23.
[0044] FIG. 25 shows a preferred embodiment of a geared input drive
of an embodiment of a speed change device.
[0045] FIG. 26 shows an alternative view of the geared input drive
of FIG. 25 with a pinion.
[0046] FIG. 27 shows an embodiment of a speed change device with
motor.
[0047] FIG. 28 is a cut-away view of the embodiment of FIG. 27.
[0048] FIG. 29 is a first detail of gearing of the embodiment of
FIG. 27.
[0049] FIG. 30 is a second detail of gearing of the embodiment of
FIG. 27.
[0050] FIG. 31 shows an embodiment of a speed change device with
fixed arm and output arm.
[0051] FIG. 32 shows a detail of the embodiment of FIG. 31.
[0052] FIG. 33 shows principles of operation of a gearing
system.
[0053] FIGS. 34-43 show various views of geared rollers and
races.
[0054] FIGS. 44-50 show an example of a thru-hole, high torque
actuator using an embodiment of a speed change device.
[0055] FIGS. 51-54 show various roller arrangements for embodiments
of a speed change device.
[0056] FIGS. 55-59 show examples of five configurations of an
embodiment of a speed change device as a single stage
reduction.
[0057] FIG. 60 shows an embodiment of a roller that tapers
axially.
[0058] FIG. 61 shows a cross-section of an embodiment of a speed
change device with different torque transmitting diameters.
[0059] FIG. 62 shows a view, partly in section, of a further
embodiment of a speed change device with different torque
transmitting diameters
DETAILED DESCRIPTION
[0060] The preferred embodiment of the present speed change device
is believed to be capable of providing numerous of the desired
benefits as well as others described in this disclosure. The
Multi-Element Self Energizing Speed Change Device is, in many of
its preferred embodiments, a torque reactive traction drive speed
change device with a number of unique features and benefits.
[0061] With a high enough reduction ratio, low enough input
friction, and a high enough input speed capability, it is believed
to be possible to use a low torque, high speed drive input such as,
but not limited to, a boundary layer air turbine or low torque
electric motor such a compact inductance or variable reluctance
electric motor.
[0062] Fluid (smooth, non-erratic) motion combined with precise
control is a performance goal that is very challenging with many
conventional actuators. The use of an air turbine input with the
present device is believed to allow for the precise control of
actuated systems such as, but not limited to robotic arm movements,
with a very high level of precision and fluidity. A proportional
air flow (or even a pulse width modulated flow) valve controls the
air flow to the turbine in either direction to produce rotatory
motion.
[0063] By taking advantage of the potentially low friction (low
resistance input rotation torque) and high speed input drive
characteristic of the present device, it is believed by the
inventor that the beginning and end of each actuator movement can
be controlled precisely, while at the same time allowing the
natural acceleration and deceleration characteristics of the
actuator to define, to a certain extent, what the acceleration and
deceleration rates will be.
[0064] The traction version of the present device works on a dual
idler roller torque transfer mechanism. The following description
is a step by step description of the principle starting with a few
prerequisite concepts.
[0065] Referring to FIG. 1, if two equal sized contacting rollers
12 are positioned vertically with parallel axes between two fixed
horizontal contact surfaces 14, they will both roll at the same
speed and maintain their vertical alignment (assuming no slippage
occurs). If one roller rotates, so must the other. If one roller is
prevented from rotating, so will the other.
[0066] Referring to FIG. 2, according to this same principle, if
one of the rollers 16 is smaller than the other roller 18, it will
spin faster than the larger roller but will still maintain the
vertical alignment provided no slippage occurs.
[0067] Referring to FIG. 3, if three rollers 12 are used in a
triangular stack, the same principle holds true, but it becomes
possible to transmit force through the rollers from one horizontal
contact surface to the other. Rollers 12 can roll freely but races
cannot move relative to each other without slipping. Force A may be
applied to horizontally movable race 20. The force is transmitted
through roller contacts to upper fixed or horizontally movable race
22.
[0068] In this case, the rollers can move relative to the
horizontal surfaces, but (assuming a fixed distance between the
horizontal surfaces) the horizontal surfaces are unable to move
relative to each other without sliding on the rollers. Therefore,
by applying a horizontal force to one horizontal surface (referred
to here as a "race") which is perpendicular to the roller axes,
that force will be transmitted through the roller/race contacts
with one race and roller/roller contacts to the other race through
the other roller/race contact. Only two rollers are required to
transmit force in one direction, as long as the angle of the two
rollers is low enough, combined with a high enough coefficient of
friction between the two rollers, and between the rollers and the
races, to allow the camming action of the rollers to maintain a
non-sliding traction contact at the surface contacts. If the
coefficient of friction is too low, or if the angle of the rollers
is too high, then force will only be transferred from one race to
the other through the traction contact if the rollers are
preloaded. In this diagram, preload would be accomplished by
exerting a force on the bottom two rollers together, which will, in
turn, push the single top roller upward and the bottom two rollers
downward.
[0069] The fact that the rollers can roll freely while transmitting
force from one race to the other, but at the same time, the races
are not able to move relative to each other, may be
counterintuitive for some observers because we are accustomed to
roller bearings allowing race contacts to roll freely against one
another but while the surfaces of a roller bearing are allowed to
roll freely in a roller bearing with a single row of rollers, the
rollers themselves roll at a fixed ratio compared to the races
(unless slippage occurs) and are therefore not free to move
independently of the races.
[0070] By comparison, the dual rollers of the present device which
are transmitting force from one horizontal surface to the other in
FIG. 3 are able to roll freely at the same time as they are able to
prevent movement of one race relative to the other.
[0071] Referring to FIG. 4, another principle at work here is the
self-energizing characteristic that results from the angled rollers
"camming" against each other when force is applied thru them from
one race to the other. This figure shows two races spaced apart by
two large conventional cylindrical roller bearings 26. Only two
force transmitting rollers 12 of the present device are used to
illustrate force transfer in a single direction. In this case the
large cylindrical rollers 26 set the distance between the races and
a horizontal force B applied to the bottom race 24 is transmitted
through the two smaller rollers 12 to the top race 32. The most
significant characteristic demonstrated in this figure is how a
horizontal force applied to the rollers will create a proportional
vertical force on the races due to the traction angle 34 (from the
broken vertical line) between contact 28 and contact 30.
[0072] This vertical "camming" force increases the contact force of
the rollers on the races (and between the rollers) to prevent
sliding at the traction contacts as the horizontal force between
the races increases. This is true, even if there is very little
vertical preload on the traction contacts rollers before the
horizontal force is applied.
[0073] With coefficient of friction of 0.37 or greater, the
traction angle of 20.degree. shown here would allow force transfer
from race to race in one direction without the preloading help of a
third roller (as shown in FIG. 3). A metallic material with a
coefficient of friction (CF) of up to 0.6 such as a boron infused
steel such as boronizing surface treatment available from Richter
Precision Inc. in East Petersburg, Pa., U.S.A. has a coefficient of
friction of 0.4''
[0074] A typical beryllium copper will have a CF of up to 0.8
unlubricated against steel. An example of a plastic material with a
CF of 0.4-0.5 is PC/PET. Certain grades of Torlon have a
coefficient of friction of 0.35 and could be used with a traction
angle of approximately 19.degree. or less
[0075] In FIG. 3, a third force transfer roller is used to allow
horizontal force transmission between the races in both directions.
With a small amount of preload pressing the inner rollers and outer
rollers together, this force transmission mechanism will allow the
direction of force transmission to be reversed with zero backlash.
Unlike gear teeth where preload between the gears can lead to high
levels of friction, wear, noise and even seizing, this preload is
not detrimental to the present system because the rollers operate
with pure rolling contact so there is very little efficiency loss
that results from this preload.
[0076] Rigidity in a rotary actuator application of the speed
change device, as one of many examples of how the present device
could be used, is often a very desirable characteristic as it
provides precise control and predictability of an actuated system.
The rotational rigidity of the force transfer effect of the present
device is potentially very high with the compression and partial
shear of the preloaded (and the potentially very rigid) rollers
being the primary area of deformation when loaded. Configurations
like those used in the present device also allow the use of a high
number of rollers (such as 19 or more and up to 50 or 100 or more)
for increased traction and rigidity.
[0077] Note that although the traction embodiment of this
disclosure has many advantages, gear tooth interfaces can be used
on some or all if the contact areas and are also disclosed
here.
[0078] The force transmission device described in FIGS. 1 through 4
demonstrate some of the working principles of the present device
but does not provide a speed change function between the two races.
It is, in effect, an infinite speed ratio device where the ratio of
motion from the direct movement if the rollers: to the output
(horizontal race relative movement) is 1:0 because the surface
speed of the roller to race contacts is the same against both
races.
[0079] This does not provide useful function as a speed change
device but can be used for the support and isolation of components
within an assembly.
[0080] To use this principle to transfer torque from a fixed or
rotatable race to another fixed or rotatable race, one or more sets
of orbital rollers must be positioned between the circular outer
diameter (OD or outer surface) of an inner race and the circular
inner diameter (ID or inner surface) of an outer race. As the
rollers are caused to rotate, they roll along the races due to
traction (or geared) contact. The surface speed of the
roller-to-race contacts is similar (as in the above examples with
non-curved races) but the circumference of the contact surface of
the inner race is less than the circumference of the contact
surface of the outer race, so as the rollers rotate with the same
or similar surface speed on each race and (assuming the rollers
remain in contact with each other) they will cause the two circular
races to rotate relative to each other as shown in FIG. 5. In FIG.
5, the outer race 36 fixed in and the inner race 38 rotates
clockwise, and input to rollers 12 causes them to orbit
counterclockwise.
[0081] The smaller the rollers, the closer the race diameters will
be to each other (for a given traction angle) and the higher the
speed change ratio between the roller orbiting speed and the speed
of the output race relative to the reference (or fixed) race.
(Note: For the purpose of this disclosure, one (or more in some
configurations) of the races will be referred to as an output race,
and the other/s will be referred to as the fixed or reference
race/s and will be assumed to be fixed in space unless otherwise
noted).
[0082] With less than ten roller sets (referring here to one "inner
roller"--i.e. The roller in the set of two contacting rollers that
is in contact with the OD of the inner race--and one "outer roller"
(that is, the roller in the set of two contacting rollers that is
in contact with the outer race) and assuming a full complement of
roller sets (such that all the outer rollers are in contact with
the inner roller in that set, and also in contact with the inner
roller of the adjacent set) the speed change ratio is typically
less than is desirable for many high torque applications like
robotics. A further disadvantage, if less than ten roller sets is
used in a full complement arrangement (requiring relatively large
rollers) is the uneven force distribution of the rollers on the
races which requires thicker and heavier races to prevent unwanted
race deformation. This race deformation is important to minimize in
robotics and other applications because deformation of the races
will be transferred into the structure which it is actuating
causing unwanted movement and vibration as the rolling or geared
components orbit around the actuator center axis.
[0083] A further non-obvious benefit of using more than ten roller
sets is the increase of total traction force that results from a
greater number of traction contacts.
[0084] Increasing the number of rollers, without increasing the ID
of the outer race, requires smaller rollers. Smaller rollers have
the disadvantage if increased Hertzian stress for the same load per
roller, but the attached investigation reveals that for a set outer
race ID, the benefit of additional rollers increases dramatically
from ten to fifteen roller sets even though the reduced Hertzian
stress limit of the smaller rollers requires a reduction of the
maximum load On each roller. Above fifteen rollers, the torque
capacity benefit of additional roller sets is significantly less.
The advantages of higher speed change ratios, and more consistent
force distribution do continue to increase as the number of rollers
is increased however. Actuators with 10 or more, 19 or more, 20 or
more, 30 or more, 40 or more, 50 or more, 60 or more, 70 or more,
80 or more, or 100 roller sets or more are envisioned by the
inventor as being practical and beneficial for many applications.
As indicated in FIG. 54 for example, in some embodiments, referred
to as full complement, the orbital rollers 40, 42 extend entirely
around the annulus formed between the inner race and outer race,
with each and every inner roller 42 being in rolling contact with
two outer rollers 40 and each and every outer roller 40 being in
rolling contact with two inner rollers 42.
[0085] Potential benefits of this principle as applied to the
present speed change device include the following: [0086] Zero
backlash [0087] High torque for size and weight [0088] High
rigidity for size and weight [0089] High precision [0090] High
speed change ratio [0091] Low torque input [0092] Self-locking if
needed [0093] Back-drivable it needed [0094] Integrated emergency
override clutch if needed [0095] Low profile [0096] Light weight
[0097] Integrated air, electric or other type of motor [0098] Large
center through hole [0099] Ease of integrating input and output
encoders [0100] High efficiency
[0101] The contact stress calculations are based on Hertzian line
contact with a correction for the maximum possible traction force
which is based on the coefficient of friction between the two
materials.
[0102] To find the allowable contact stress at one contact point
two formulas are required:
b = 2 .times. F .pi. .times. .times. l ( 1 - v 1 ) 2 E 1 + ( 1 - v
2 ) 2 E 2 1 d 1 + 1 d 2 ( 1 ) P max = 2 .times. F .pi. .times. b
.times. l ( 2 ) ##EQU00001##
Where:
[0103] b=half width of elliptical contact profile P.sub.max=maximum
stress experienced in the material F=applied load 1=length of
contact E.sub.1, E.sub.2=Young's Moduli of respective materials
v.sub.1, v.sub.2=Poisson's ratios of respective materials d.sub.1,
d.sub.2=Diameters of respective cylinders, d is taken as being
negative if the cylinder defines a concave, rather than a convex
surface with respect to the contact area
[0104] P.sub.max is given by the limiting factor, either the
compressive strength of the material or the contact fatigue
strength of the material (if available). Once P.sub.max has been
determined, Fmax can be solved for by iterating between Equations
(1) and (2). Iteration involves guessing one value, F for example,
finding the corresponding b value from Equation (1), plugging the
calculated value of b into Equation (2) and solving for a new F.
This is repeated until the solution converges and F and b values
are found which satisfy both equations.
[0105] Once a theoretical maximum load has been calculated from
above, a traction factor is applied. The traction factor reduces
the load to allow for the increase in contact stress due to surface
shear. The algorithm for determining the increase in Pmax for a
given coefficient of friction comes from TribologyABC.com and can
be summarized by the following code snippet.
TABLE-US-00001 <!-- Interpolation for Tmax --> if (mu>=0.0
&& mu<0.15) {Tmax = 0.387 + (0.41-0.387)*(mu/0.15)1; if
(mu>=0.15 && mu<0.3) {Tmax = 0.41 +
(0.51-0.41)*(mu-0.15)/0.151; if (mu>=0.3 && mu<0.4)
{Tmax = 0.51 + (0.579-0.51)*(mu-0.3)/0.11; if (mu>=0.4
&& mu<0.5) {Tmax = 0.579 + (0.686-0.579)*(mu-0.4)/0.11;
if (mu>=0.5 && mu<0.6) {Tmax = 0.686 +
(0.811-0.686)*(mu-0.5)/0.11; if (mu>=0.6 && mu<0.7)
{Tmax = 0.811 + (0.937-0.811)*(mu-0.6)/0.11; if (mu>=0.7
&& mu<0.8) {Tmax = 0.937 + (1.064-0.937)*(mu-0.7)/0.11;
if (mu>=0.8 && mu<0.9) {Tmax = 1.064 +
(1.19-1.064)*(mu-0.8)/0.11; if (mu>=0.9 && mu<=1.0)
{Tmax = 1.19 + (1.317-1.19)*(mu-0.9)/0.1; if (mu>1) {Tmax =
"Undefined"}; dpc=(Tmax/0.387); dF1= 1/(dpc*dpc); <!--
Interpolation for Tmax -->
[0106] In the above algorithm, dpc represents a coefficient giving
the increase in the maximum stress component experienced by the
material. Similarly dFl, is a coefficient representing the decrease
in the maximum allowable hertzian load. To correct for traction,
one simply multiplies the load calculated from the hertzian contact
formulas by dF1.
[0107] Once the maximum load for one contact has been calculated,
the torque carrying capacity of the actuator is calculated using a
simple moment arm. The torque supplied by one contact point is
given by:
T = Fd ( 3 ) ##EQU00002##
[0108] Where: T=Torque; F=Max load; d=perpendicular distance
between contact point and center of actuator
[0109] The total torque for the actuator is then simply the torque
for a single contact point multiplied by the number of rollers. For
an actuator with rollers at different distances from center, the
contacts which supply the least amount of torque are taken as being
the limit on the torque capacity of the actuator.
[0110] For an actuator with a pre-defined inner diameter, the
effect of using an increasing number of smaller rollers is
illustrated in FIG. 6.
[0111] The graph in FIG. 6 represents a titanium roller on a
titanium inner ring. The diameter of the ring is 40 mm. The
diameter used for the rollers is the largest diameter roller that
can fit around the 40 mm ring for a given number of rollers,
without having the rollers interfering with their neighbors. A
small clearance has been added between each of the rollers, no
correction has been made for surface shear due to traction, thus
actual max torque would be somewhat lower.
[0112] Note, a non-full-complement speed change device is also
possible in some embodiments but requires a means if spacing some
or all of the rollers to achieve adequate preload. Several
embodiments of a non-full-complement version of the present device
are disclosed in this document.
[0113] By matching the coefficient of friction to the traction
angle, the present device can be tuned to increase the roller load
with increased torque until failure, or until a predetermined
maximum traction force or torque.
[0114] One method of limiting the maximum traction is to use
rollers which energize in one rotational direction to preload the
rollers which energize in the other rotational direction. The
angle, preload, and coefficient of friction (CF or cf) between the
materials in this case may be high enough to maintain traction up
to a certain level of torque. The CF and angle is not sufficient on
their own, however, to transmit torque without slipping once the
preload of the opposing roller is reduced due to deformation of the
load bearing low rollers in the opposite direction. At a certain
level of torque, therefore, the preload of the opposing roller/s
will decrease (as the load on the driving direction roller/s
increases) to the point where the opposing direction roller/s no
longer provides adequate preload and the drive rollers are allowed
to slip. The maximum torque before slippage of this configuration
is decreased by increasing the roller angle, and/or by using
material combination with a lower CF, and/or by reducing the
initial preload.
[0115] Many combinations of materials are possible including (but
not limited to) metallic, ceramic, plastic, polyamides, and
elastomers.
[0116] More rollers results in a more consistent load on the races
allowing thinner and lighter races with less deformation. For this
reason, the present device is preferably configured with 10 or
more, 11 or more, 12 or more, 13 or more, 14 or more, 15 or more,
16 or more, 17 or more, 18 or more, 19 or more, 20 or more, 25 or
more, 30 or more, 35 or more, 40 or more, 45 or more, 50 or more,
55 or more, 60 or more, 65, 70 or more, 75 or more, 80 or more, 85
or more, 90 or more, 95 or more, or 100 or more roller sets. A
roller set consisting of one inner roller contacting the OD of the
inner race and one outer roller contacting the ID of the outer
race, with the inner and outer roller of each set contacting each
other.
[0117] The required coefficient of friction to achieve
self-energizing (or camming) is explained as follows. The traction
angle, as defined in this application, is the angle between a
radial line from the center axis of the races through a contact
point of a roller on a race. The other line in the angle
measurement for an inner roller is the line from the inner roller
contact with the inner race to the roller/race contact of outer
roller (which the inner roller is contacting) with the outer
race.
[0118] Referring to FIG. 7, the angle 46 for the outer roller 40
contact is also shown in FIG. 7, but it is typically lower than the
inner roller 42 contact angle, so the limiting angle with regard to
establishing a self-energizing or "camming" traction drive system
will be the larger of the two angles. This larger angle
(18.1.degree. in FIG. 7) will determine the required CF for the
inner roller and race contact to establish a self-energizing
traction angle 34.
[0119] When one of the inner race 38 or outer race 36 is fixed (the
outer race in this example) and a torque is applied to other race
(a counterclockwise torque on the inner race, in this example) the
traction of the inner roller against the inner race (at contact 28)
will create a force that is transferred through the contact with
the outer roller 44 through to the outer roller contact with the
outer race at contact 30.
[0120] If the relative diameters of the inner rollers 42 and outer
rollers 40 and inner races 38 and outer races 36 results in a
preloading of contacts 28, 30 and 44, the amount of this preload
and the coefficient of friction of these contacts will determine
the traction torque capacity of the present device if the traction
angle is not adequate for that coefficient of friction to establish
a self-energizing.
[0121] Referring to FIG. 8 several examples of a range of traction
angles 34 are shown along with the minimum coefficient of friction
limit 48 for each angle. In order to establish a self-energizing
system whereby an increase in the contact force of the rollers will
result from increased torque transmission through the device, the
traction angle must be above the traction angle show. In order to
account for inconsistencies in material prop reduce, and in order
to provide a consistent and predictable result, it is preferable
that the traction angle, for a given minimal CF, the higher than
the traction angle shown in FIG. 8. FIG. 11 shows a further example
of a traction angle. The required coefficient of friction to
produce traction must exceed the minimum coefficient of friction
limit 48 for the associated traction angle 34 for that CF.
[0122] For a speed change device of the present design with each
roller having a single traction contact diameter, the minimum
traction angle achievable if all rollers are the same diameter is
approximately 15.degree.. Practically speaking, for most
conceivable applications, the minimum traction angle will be closer
to approximately 17.degree. with 18.degree. or 19.degree. being
common for many conceivable embodiments of the present speed change
design. For these common geometries for the present device, then, a
coefficient of friction of 0.4 or greater will ensure that the
system is self-energizing under load, while it is also possible to
design a speed change device of the present design to achieve a
self-energizing traction angle with a material coefficient of
friction of as low as 0.34 or lower in some configurations.
[0123] Each of the inner race and outer race are circular and thus
centered on an axis, as shown for example in FIG. 54. The traction
angle o may be defined as follows: for each pair of a first inner
roller that contacts a first outer roller, the traction angle o is
defined as the angle between a first line extending outward from
the axis through a center of the first inner roller and a second
line extending from the contact point of the first outer roller
with the outer race and a contact point of the first inner roller
with the inner race. Thus, according to the geometric
representations in FIG. 8, where the first inner roller contacts
the inner race with a first coefficient of friction cf1 and the
first outer roller contacts the outer race with a second
coefficient of friction cf2, cf1>tan(o) and cf2>tan(o).
Various traction angles may be used, for example: less than or
equal to 45, 40, 39, 38, 37, 36, 35, 34, 33, 32, 31, 30, 29, 28,
27, 26, 25, 24, 23, 22, 21, 20, 19, 18, 17, 16 or 15 degrees, with
corresponding limitations on the coefficients of friction. In some
embodiments, when the inner rollers contact the inner race with a
first coefficient of friction cf1 and the outer rollers contact the
outer race with a second coefficient of friction cf2, at least one
of cf1 and cf2 is 0.27, 0.28, 0.29, 0.30, 0.31, 0.32, 0.33, 0.34,
0.35. 0.36, 0.37, 0.38, 0.39, 0.40, 0.45, 0.50 or 0.60 or
greater.
[0124] For purposes of scale, FIG. 9 shows a configuration of the
present device showing a typical roller diameter for an outer race
36 ID of 6.05''. This example uses 34 roller sets (i.e. 34 inner
rollers 42 and 34 outer rollers 40). One roller set comprises one
inner roller 42 and one outer roller 40.
[0125] All of the inner rollers all preferably contacting two outer
rollers, and all of the outer rollers are preferably contacting two
inner rollers to provide what is referred to here as a "full
complement" assembly.
[0126] As shown in FIG. 10, increasing the size of the outer
rollers 40 relative to the inner rollers 42 will result in a
reduction of the necessary traction angle 34 to will allow the use
of lower CF material combinations.
[0127] The CF for the inner roller/race contact 28 in this case
would need to be greater than 0.31 in order to achieve a
self-energizing camming effect (independent of the initial roller
contact preload).
[0128] Increasing the number of roller sets (using smaller rollers)
will also decrease the necessary traction angle 34 as shown
schematically in FIG. 12, for example, with 100 roller sets,
comprising 100 inner rollers 42 and 100 outer rollers 40.
[0129] An extreme example of a full complement speed change device
of the present design is shown schematically in FIG. 13 with 313
roller sets. With these very small rollers (relative to the
diameter of the races) it is necessary to use the material
combination for the inner roller-to-inner race traction contact
which is higher than 0.28 in order to create a self-energizing
torque transfer multi-element self-energizing device.
[0130] Note: although the coefficient of friction 48 of 0.28 or
higher is required for most speed change device configurations of
the present device, there are certainly benefits to a device of the
present design with a traction angle and CF which does not achieve
a self-energizing effect. Benefits of such a device include a very
predictable breakaway torque which can be useful, for example, but
not limited to, robotic applications where interaction with humans
is expected. For these and other reasons, and to account for other
system variables which may be difficult to predict precisely when
creating a self-energizing speed change device of the present
design, coefficients of friction of 0.2 or higher are included here
as optional design parameters.
[0131] Note: the traction angle for the outer roller against the
outer race is generally lower than the traction angle of the inner
rollers against the inner race. For this reason a slightly lower
coefficient of friction can be used for the outer roller/race
contact. The contact between the inner and outer rollers in each
torque transmitting set will also require a minimum coefficient of
friction. It has been established by experimentation that using
materials with a similar coefficient of friction between the
rollers as the coefficient of friction between the rollers and the
races is adequate to achieve a self-energizing effect with a full
complement system as disclosed in this document.
[0132] Referring to FIG. 14, a further example of the present
design is shown, showing outer rollers 40, inner rollers 42, outer
race 36, inner race 38, contacts 28, 30 and 44, traction angle 34
and angle 46 for the outer roller contact.
[0133] There are many material combinations which achieve a
coefficient of friction of 0.28 or higher. These include but are by
no means limited to the following (the following are, however,
considered to be preferable materials for the reasons listed
below). Many other materials exist or are like to exist in the
future which fulfill the preferable requirements of a coefficient
of friction of 0.28 or higher and good rolling contact
characteristics. Other characteristics such as a reduction of the
coefficient of friction with increased contact load, such as is
exhibited by certain formulations and heat treatments of spinodal
bronze, are also believed to be beneficial for certain
applications.
[0134] Beryllium copper--This material has a high coefficient of
friction when running without lubrication against steel and the low
coefficient of friction when running against itself. A preferable
configuration of the present speed change device would include a
combination of beryllium copper and steel rollers and races such
that that contacts which require traction are beryllium copper
against steel and contacts which require a low coefficient of
friction are beryllium copper against beryllium copper. An example
would be a beryllium copper inner race, steel inner rollers,
beryllium copper outer rollers, and the steel outer race. A more
preferable combination would include a steel inner race, beryllium
copper inner rollers, steel outer rollers, and the beryllium copper
outer race. In this case, and some applications, the inner rollers
could be designed to be nearly contacting each other, and the low
coefficient of friction of the beryllium copper on itself could act
as a simple spacing system for the inner rollers.
[0135] Steel with boron diffusion surface treatment--This
material/surface treatment has a high coefficient of friction when
running without lubrication against steel or itself
[0136] Titanium--Titanium has some unique properties which include
a relatively low modulus of elasticity (which reduces the Hertzian
stress of the contacts) light weight, high strength and a
relatively high coefficient of friction.
[0137] Kevlar reinforced Torlon--There are many different injection
moldable materials which could be used as rollers and or races for
certain configurations up the present device. Kevlar reinforced
Torlon has a relatively high coefficient of friction and exhibits
very low cold flow which is a benefit for a preloaded rolling
contact system. The use of this material, or others with similar
properties, could allow the injection molding of rollers and or
races for low cost production. The use of steel inserts running
against the Torlon in a number of different potential combinations,
may be a preferable combination of materials for certain
applications, where low cost, low weight, and low to medium torque
are required.
[0138] Steel--Many different types of steel will provide the
necessary characteristics in combination with the above materials
and/or many other materials now existing or possibly existing in
the future.
[0139] The above list is in no way necessarily limiting, but
rather, an example of some preferable materials which can be used
with the present speed change device.
[0140] FIG. 15 shows a partial section through the center of an
outer roller 40 and an inner roller 42 in a set and the inner 38
and outer races 36 (some parts not shown, for clarity). The inner
race may be formed of split races 38 and one not shown but the same
item on the other side of the structure, and the outer race may be
formed of split races 36 and one not shown but the same item on the
other side of the structure. The split races may be secured
together by a housing, not shown in this figure. In this embodiment
of the present device, a conical surface 50 on the annular double
conical groove in the outer roller will contact the surface 52 on
the raised double conical ring on the ID of the outer race if the
outer roller 40 strays to the left in this illustration. If the
outer roller 40 strays to the right in this illustration, the
opposite surfaces will contact. Using a groove in the outer roller
with a face on either side of a center plane of the outer roller
will cause the contact to occur at a smaller radial distance from
the center of the outer roller then the radial distance of the
outer roller traction contact with the outer race. This will cause
the roller to slow down on the side of the plane that is further
ahead in the rolling direction than its ideal position such that it
is steering towards the edge of the race, and so by slowing this
end of the roller down, it will cause it to change angle back to a
more aligned position, or if it is already aligned, but simply
displaced axially towards and edge of the race, slowing the
rotation of the end of the roller which is opposite the edge of the
race to which the roller is off center to ward, will change the
angle of the roller such that it will steer back toward the ideal
centered position.
[0141] A raised double conical annular ring 54 on the inner roller
will not contact the conical face of the annular groove on the OD
of the inner races 56 if it moves to the right in this illustration
because the outward facing conical face on the inner roller 58 will
contact first (inner race on the right side with conical groove not
shown). The contact 58 is a greater radial distance from the center
of the inner roller than the cylindrical traction surface of the
inner roller and will, therefore, have a greater surface speed than
the cylindrical traction surface for a given roller rotation
causing the right end of the roller, in this example, to increase
in speed when it is rolling partially on the larger diameter
contact 58 so as to steer it back toward center. The opposite will
occur if the inner roller moves towards the left side of the inner
races.
[0142] Annular grooves, instead of rings, on the inner roller
(similar to the outer roller in this illustration) could also be
used instead of, or in addition to, the rings shown here.
[0143] The groove 60 is preferably designed for clearance so these
conical surfaces between the rollers do not ever come into
contact.
[0144] The cage 62 in this embodiment is an example of a
low-profile, simple to assemble, alignment means for the inner
rollers. It uses a thicker boss 64 which partially protrudes into
the ends of the inner rollers (in this embodiment, although a
similar structure could be used with the outer rollers instead of
or as well) to provide a stable and preferably press fit attachment
for the pins 66 which are used for spacing the inner rotors equally
around the inner race, and/or to align the inner roller center axes
parallel with the axis of the inner and outer races. Because of the
potentially large number of these pins, corresponding to the
potentially large number of rollers, even if these pins are a
relatively small diameter, they will be adequate to maintain the
torsional rigidity of the cage assembly necessary for effect of
alignment, in many applications. As an additional benefit the
smaller the diameter of these pins, the less frictional force they
will exert on the rollers. A material like spinodal bronze or
beryllium copper can have a relatively low coefficient of friction
on itself. For this reason the use of a material like spinodal
bronze or beryllium copper for the inner rollers as well as for the
cage and possibly for the pins 66 is a preferred configuration. A
material like beryllium copper has a much higher coefficient of
friction when running dry against steel then it does against
itself. For this reason it is preferable in the example given here,
to use a harder material such as, but not limited to, hardened
steel for the outer rollers and the inner races. The outer races
are, in this example, preferably of the material such as, but not
limited to, beryllium copper or spinodal bronze.
[0145] Various combinations of these materials (such as the inverse
to what is described here) are anticipated by the inventor. The
purpose of this disclosure is to describe preferred embodiments of
the present device and its various working principles. Many
different variations and combinations of the features disclosed
here are anticipated by the inventor and can be implemented with
various effects without straying from the principles disclosed
here.
[0146] The raised center disks 68 on the inner rollers are used to
provide the traction (or possibly geared) input to the actuator
with an annular ring member (not shown in this illustration). These
larger diameter disks can be on every inner or outer roller or on
every second inner or outer roller or on every third inner or outer
roller, but all of the disks 68 are preferably on rollers in the
same row, for example on the inner rollers exclusively or the outer
rollers exclusively). An outer housing member attaching the two
outer races together, and in inner housing member attaching the two
inner races together are also not shown here. A similar pin
alignment cage structure to cage 62 can also be provided for the
individual outer rotors.
[0147] Referring to FIG. 16, the conical angle of the contacting
alignment faces (for example conical face 50 and 52 in this
embodiment) are preferably different by a large enough angle that
when the roller is off center (to the left in this illustration)
the contact between the ring and groove will begin adjacent to the
cylindrical traction surface 18, and if the misalignment of the
roller 40 is great enough, the contact will cause movement or
deformation of the surfaces 50 and 52, such that the edge of
contact between the conical faces 5 and 6 which is furthest from
the conical traction surface 70 will move progressively further
away from the conical traction surface 70 as the misalignment
increases. This will provide the effect of a progressively greater
alignment/steering effect as the misalignment of the roller
increases. This or similar geometry can be used for any of the
conical alignment surface is disclosed in this document. Conical
surfaces begin to contact adjacent to cylindrical traction surface.
Innermost edge of contact between conical surfaces moves
progressively away from cylindrical traction surface with increased
misalignment of the roller (with leftward movement of the roller in
this illustration). A variation of this preferred embodiment of
this self-steering alignment system uses a smaller convex radius on
the edge 72 then on the concave radius 74. This difference in radii
will also contribute to a smooth transition from cylindrical
rolling contact to partially conical rolling contact during
alignment.
[0148] FIG. 15 shows is a simplified partial cross-section view of
a LiiveDriive speed change device with a preferred roller spacing
and alignment cage 62. The cage is a low-profile construction to
reduce the necessary with of the assembly and includes a preferably
elliptical boss 64 which provides a deeper bore to preferably press
fit a cross member such as a dowel pin 66. The cross member, which
is preferably coaxially assembled through three or more, but
preferably all of the inner rollers, and or three or more but
preferably all of the outer rollers. If a large number of cross
members such as ten or more, for example, the torsional rigidity of
the cage assembly can be high enough to provide angular alignment
stability to the rollers, even with the use of relatively small
diameter crossmembers.
[0149] An advantage of using small diameter crossmembers is a
reduction of friction between the crossmembers and the rollers for
higher speeds and efficiencies. Lower friction is especially
preferable in applications of the LiiveDriive which do not use
lubrication.
[0150] The tapered bores in the ends of the rollers are a
preferable structural element in that they allow for the boss to on
the cage and also provide a centering feature for manufacturing and
potentially, re-machining of the rollers to a smaller size for a
refurbished device after the surface of the rollers have been all
or part way through their useful fatigue life.
[0151] Another example of a preferred cage configuration is shown
in FIG. 17 (as with other figures in this application, the figure
is missing the inner and outer races on the far side and some other
components of a complete speed change device such as the housing
and means of attaching the inner and outer races to a fixed and
output structure). In FIG. 17, the outer rollers 40 are formed of
axially aligned roller sections. Two are shown, but there could be
more. The inner rollers 42 are longer than the outer roller
sections. This arrangement may be reversed, with the inner rollers
being formed in axially aligned sections. The tapered features in
the roller ends are still present, but a larger cylindrical bore in
the ends of the inner rollers (in this example, although the same
could be done on the outer rollers, and the long rollers could be
used as outer rollers, and the short rollers used as the inner
rollers). The cage is designed with thin sections 72 which allow
slight radial movement of the crossmembers. This radial movement
can accommodate for manufacturing variables, but can also allow the
crossmembers to move outwardly cheering high rotational speed
operation. In this case the cross numbers can be brought into
closer proximity to or even contacting the through bore in the
roller thereby increasing the precision of the angular
alignment.
[0152] FIG. 18 shows a partial view of the assembly shows a
preferably elliptical shape of the bosses 64.
[0153] FIG. 19 shows the wider surface 74 of the preferably
elliptical (but other shapes which are narrower along a radius from
the center of the actuator and wider perpendicular to this radius
can also be used) boss 64 is closely situated to the forward and
leading inside cylindrical surface of the ends 76 and 78 of the
rollers.
[0154] A simplified schematic view of a primarily rolling contact
roller spacing system is shown in FIG. 20. The system may be used
on the inside rollers 42 or outside rollers 40, but is shown here
on the inside rollers as an example. The preferred embodiment of
this concept uses a smaller diameter boss on the ends of separated
for example every second (but every fourth or sixth) roller in the
same (inner or outer) row and a circular disk or ring 82 which can
rotate on boss 84. Another boss 80 on the end of an intervening
roller 12 in the same row that is halfway between the two rollers
with the disks 82 cooperates with the circular members 82 to space
the rollers. This boss can be the same size as the roller 12,
smaller than the roller 12, or larger diameter than the roller 12.
The combination of diameters of the boss 80 and the disk 82 are
such that very little if any clearance exists between the disk 82
and the adjacent boss 80. It may even be preferable to provide a
small amount of preload on these contacts and in so doing provide a
method of precisely spacing one of the rows of rollers with a
primarily rolling contact. If the inner rollers can be spaced
equally with this or one of the other methods disclosed in this
document or a variation of these methods or some other method which
is obvious from those described here, or some other practical
means, the inner and outer races can be totally positioned
coaxially without stress on an additional external bearing or
bearings, or may even eliminate the need for a radially loaded
additional bearing altogether.
[0155] Other bearing systems which can be used include external
bearing cages similar to those used for cylindrical roller
bearings. These conventional external bearing cages have the
disadvantage of sliding on the largest diameter of the rollers and
can, for this reason, result in higher friction and wear them in
the preferred systems shown here.
[0156] Another possible cage system would not use a bore through
the center of the rollers for the crossmembers but rather, would
use the gaps between the rollers for the crossmembers, or for
crossmembers in addition to those through the center of the
rollers. Alignment surfaces on the cage could in this case contact
the outer traction surface of the rollers, or a smaller diameter
surface preferably near or at the ends of the rollers.
[0157] There are a number of inherent disadvantages to achieving
high reduction ratios in a single stage, such as with a harmonic
drive. These include decreased efficiency due to a movement of the
high torque output resulting in significant power loss through
friction. Increased wear of these highly loaded and high speed
components is also a potential effect.
[0158] A common reason for avoiding multiple stages is the increase
of backlash that results from more than one stage of conventional
system that will typically geared device which will typically
exhibit some degree of backlash.
[0159] Embodiments of the present device, by contrast, do not
necessarily display any inherent backlash, and so the precision
remains extremely high, and zero backlash us still achievable, even
by combining two or more stages.
[0160] With the backlash removed from the system, the potential
advantages of multiple stages include, but are not limited to, the
following:
[0161] The lower speed of the final stage reduces the frequency
(and potentially the energy level of) output vibrations which could
result from higher speed movement of single high ratio stage.
[0162] The lower speed final stage can be more efficient because
the highest load of the final stage is moving at a lower speed, and
so any efficiency loss at the traction of geared contacts is lower
than if they were moving at higher speeds as with a single stage
high reduction ratio system such as a harmonic drive or compound
planetary or geared differential speed reducer.
[0163] Higher efficiency will, in this case, increase
backdriveability as well, making a higher ratio actuator
backdrivable than a single stage device. Backdrivability is
considered to be beneficial in many applications.
[0164] For the cylindrical rolling elements if the present device,
axial positioning of the rollers is a critical function and is less
challenging with slower moving rollers in the final stage of a
multi-stage embodiment.
[0165] With two or more stages, it us much easier to achieve a low
friction input to the first stage because the output torque of a
preliminary stage will be significantly lower requiring
significantly less traction force and resulting in less friction.
This is very beneficial because using a high speed, low torque
motor or air turbine etc., can allow for a lighter, smaller, and
less expensive speed change device. A preliminary speed change
stage may be designed in accordance with the principles disclosed
here or have a different, conventional, design.
[0166] FIG. 21 shows a simplified partial sectioned example of a
staged embodiment of the present device.
[0167] Referring to FIG. 22, removing the inner race on the right
side reveals a representative sampling of an array of preferably
smaller (than in the final stage) inner rollers 96 and outer
rollers 94 in the first speed change stage. The input ring to the
first stage 88 is preferably a traction drive system which causes
the outer rollers 40 to spin but can also be a geared input similar
to that shown in other embodiments in this disclosure. The outer
rollers 40 self-energize (cam) between A] the first stage inner
race 36 which is preferably one piece with (or a fixed member
relative to) the inner race 90 of the subsequent (and in this
example, final) stage and B] the final stage input ring 86. Outer
rollers 98 and inner rollers 100 of the second stage are also
shown. In this example, inner rollers 100 are formed of axially
aligned roller sections, while the outer rollers 98 are longer.
[0168] Referring to FIG. 23, the final stage input ring 86, is
preferably able to expand radially outward so that the
self-energizing of the first stage will expand it against the outer
rollers 98 of the second (and in this case final) stage when torque
is applied to the input ring.
[0169] Referring to FIG. 24, an example of one of many possible
methods of providing a radially expanding inner ring 86 is shown
here. The ring has two or more (in this case three) interlocking
split rings which can expand individually but always maintain
enough rotational alignment that the discontinuities 92 of all the
rings to not ever line up axially. These discontinuities (or
breaks, or gaps) in the rings are preferably also angulation as
shown here so the force on the rollers from the races is maintained
as consistently as possible.
[0170] FIG. 25 shows a simplified but functional configuration of a
preferred geared input drive embodiment of the present speed change
device. It comprises a fixed member 102, output member 104, an
outer array of traction rollers 122, and an inner array of traction
rollers 124 with geared input members 106, an inner traction race
108 and outer traction race 110. The fixed member 102 and output
member 104 may function as housings to hold split races together.
In these drive configurations, any one of the various roller
designs may be used, including axially aligned roller sections.
[0171] FIG. 26 shows input drive motor 112 with a pinion 116 is
preferably geared (but a traction drive pinion would also work in
some applications) to an input drive ring 114.
[0172] Referring to FIGS. 27 and 28, the outer input drive ring 114
has a geared surface 120 on the ID which is in mesh with one or
more of the rollers in either the outer row 122 or preferably the
inner row 124 of rollers. An inner geared ring 118 is also in mesh
with these roller gears and is analogous to a free spinning sun
gear.
[0173] The inner geared ring 118 does not input any drive torque to
the rollers but serves two other purposes. It provides angular
spacing between the geared rollers without the need for a spacing
cage, and it contributes to the angular alignment of the geared
rollers.
[0174] Referring to FIG. 29, the geared surface 120 on the ID of
the outer input ring, in combination with the free spinning sun
gear ring 118, provides equal spacing between geared inner rollers
as well as angular alignment for the geared rollers.
[0175] Referring to FIG. 30, every second geared roller 126 is
preferably rigidly attached to two gear faces 130, one on either
side of its center plane with the objective of getting these two
gear faces as wide apart as practically possible with an the size
constraints of the speed change device. The effect of this wide
effective gear face is to provide a significant level of angular
stability and alignment to these rollers when mashed together with
the geared surface of the outer and inner geared rings.
[0176] Every first geared roller 128 preferably has only one gear
132 which is staggered with the double gears on every second roller
to allow for the largest possible diameter on these gears. This
large diameter is preferable to provide another level of speed
change in the device.
[0177] Referring to FIGS. 31 and 32, an alternate embodiment uses a
pinion drive 134 on the ID inner geared ring 138 with a free
spinning outer gear ring 140. The geared rollers, in this
configuration are preferably the outer row 122. One or more drive
motors 136 and/or 112 and pinions 134 and/or 116 in the previous
embodiment can be used together or individually for increased power
and or to reduce or eliminate backlash on the geared input to the
rollers.
[0178] Note: axial alignment features are not shown in this and the
previous simplified exemplary illustrations. One or more of the
axial and/or angular alignment systems disclosed elsewhere in this
document may be used in combination with the geared input
configuration shown in this section.
[0179] With a high enough "camming angle", traction camming is no
longer possible with common rigid materials (and will, for this
geared embodiment description of the present speed change device,
be referred to as the contact angle). By using a gear contact angle
that is larger than the inner gear-roller contact angle, but
smaller than the outer gear-roller contact angle with the outer
ring, it will cause the inner roller/ring gear tooth mesh to
separate when load is applied. With a full complement of
gear-rollers with enough backlash to allow relative radial movement
of the inner gear-rollers toward the outer gear rollers (but a
small enough amount of backlash to maintain adequate tooth contact,
the inner gear-rollers will separate from the inner ring rolling
contacts to the point where the gear teeth of each inner gear-
roller gear teeth will come into contact with the gear teeth of the
adjacent outer gear-roller.
[0180] Due to the fact that the inner rollers will be loaded toward
the adjacent outer rollers with a relatively small radial force,
this floating inner roller effect is believed by the inventor to
allow the inner roller in each roller set to find a radial position
where the forces acting on it will be balanced by the forces acting
on the adjacent set, such that the gear tooth load on all rollers
will be very consistent, allowing a high number of gear teeth to
share the applied torque load.
[0181] An additional factor which will needs to be considered is
the effect of the separation force of the outer or inner drive ring
gear (outer shown here as an example in a solid line, inner drive
ring shown as a broken line). If the outer ring gear has too high
of a contact angle, the separation force, will push the inner
gear-roller against the inner race and it will no longer float. If
the contact angle of the outer ring (in this example) is too low,
the inner rollers will all find their best fit position to share
the tooth loading consistently but the outer ring will not contact
the gear-roller drive gear teeth consistently.
[0182] For a traction drive system, as shown in the computer-aided
design (CAD) model of the gear roller embodiment of the present
speed change device on the following pages, the traction on the
largest diameters of the inner roller (in this example, will
preferably allow the inner rollers to move radially and still
maintain traction).
[0183] The contributing factors to determining the best gear
contact angle for the outer (or inner) input gear ring are complex
and will require some experimentation to determine the best balance
of gear separation forces, manufacturing precision, centrifugal
force on the inner rollers, etc.
[0184] FIGS. 33 to 43 show simplified models of the geared system,
showing the inner race 38 with geared rings 156 on the OD of the
inner race, and outer race 36 with geared rings 158 on the ID of
the outer race, and inner geared rollers 152 and outer geared
rollers 154, according to the principles of the present speed
change device described on this page. FIGS. 33-43 also show an
embodiment with split outer races and split inner races.
[0185] Note: A geared speed change system according to present
device using gear-rollers with two different pitch diameters 160
and 162 for the inner rollers and/or outer rollers to create a
compound gear-roller arrangement is possible and similar to that
shown in another section of this document and described as compound
and semi-compound traction roller configurations of the present
speed change device except that one or more of the traction
surfaces on the rollers and or races are replaced with geared
surfaces.
[0186] FIG. 33 is a schematic illustration of an example of a
geared LiiveDriive configuration. Gear teeth not shown. Outer or
inner drive input rings can be traction drive instead of geared for
higher speed operation. Apart from that option, all contacts in
this schematic are geared contacts. The example geared
configuration comprises roller/roller contact angle 48.3 degrees,
outer gear-roller contact angle 51.4 degrees, drive ring contact
angle 45 degrees, inner roller contact angle 45 degrees, drive gear
on inner rollers 142, inner gear ring pitch circle 144 (inner
gear-rollers shown slightly separated from inner ring), alternate
inner drive ring and outer roller gear 146, outer ring pitch circle
148, outer drive ring pitch circle 150 (inner drive ring may also
be used), inner gear rollers 152 and outer gear rollers 154.
[0187] FIG. 34 shows a 2D CAD drawings of a preferred geometry for
a geared configuration of the present speed change device.
Examples of Other Embodiments of the Present Speed Change
Device
[0188] FIGS. 44-46 show an examples of a very large thru-hole, high
torque actuator using an embodiment of the present traction speed
change device. This embodiment uses a two-part inner race 164 and
166 and a single outer race 168. A single traction (or geared)
input ring 170 provides the input torque.
First Plastic Prototype
[0189] Here is example of the first prototype of the change device
which was constructed out of plastic with a compressive strength of
12,000 PSI for the rollers and races and input rings, and aluminum
for the fixed and output arms.
[0190] It has an inner race OD of 4.9'' and an input and output
race width of 0.25''. By selecting a material combination with a
coefficient of friction above 0.34, an output torque of 60 foot
pounds was achieved without damage or slipping.
[0191] FIGS. 47-52 show examples of the CAD models for this
prototype in various states of disassembly, including output arm
176 and fixed arm 178. In FIG. 48, dual input rings 180 spin and
cause inner rollers 174 to roll along inner race, outer rollers 172
are the same width as the outer race 182 and inner race 186, inner
rollers are the width of the races plus the two cage members 184
and the two input members 180. In FIG. 49, simplified UHMW cage 184
achieves consistent annular spacing of the inner rollers 174.
[0192] In FIG. 50, an aluminum sleeve 188 is spline-fit to ID of
inner race 186 and ID of fixed arm thru-hole. FIGS. 51 and 52 show
the inner race 186, outer race 182, inner rollers 176, and outer
rollers 172.
[0193] FIG. 53 shows an example of the present device where the
rollers 42 in the inner row are more than one diameter. FIG. 53
also shows an example of the present device where the rollers 40 in
the outer row are more than one diameter.
[0194] The present device with rollers of the inner and or outer
row spaced closely together may be advantageous in certain
applications. The closer the rollers are to the adjacent rollers in
the same row, the lower the traction angle possible for certain
geometric relationships between inner roller diameter, outer roller
diameter, race diameter and the number of roller sets.
[0195] Another advantage of closely spaced rollers, for certain
applications, is the potential to eliminate the need for a
circumferential spacing means. When the rollers are equally spaced,
it takes very little force to keep them at that spacing relative to
the other rollers in the same row. So if a roller material is used
which has a lower CF against itself than it does against the other
row of rollers and against the race it is in contact traction with,
then in some applications, it is preferable to allow closely spaced
rollers to contact adjacent rollers in the same row if they become
unequally spaced. Material combinations which exhibit these
frictional properties include, but are not limited to, for example,
spinodal bronze or beryllium bronze against itself and against
steel.
[0196] For the present speed change device, if a material
combination is used whereby the outer rollers exhibit a lower
coefficient of friction when sliding against the adjacent outer
rollers as compared to a higher coefficient of friction between the
outer rollers and the outer race, the preferred average maximum gap
for some applications and material combinations will be 0.01'' or
less immediately after assembly.
[0197] Larger gaps between the outer rollers may still work in this
configuration for some applications but are believed to be
impractical if no other spacing elements are used.
[0198] FIG. 54 shows an example of inner rollers 42 and outer
rollers 40 contacting or nearly contacting. It is preferable that
only the inner or only the outer rows of rollers are near
contacting or contacting according to the description above.
[0199] FIGS. 55-59 are examples of five configurations of the
present speed change device as a single stage reduction with an
additional input ring to provide rotational torque and motion to
either the inner row of rollers 42 or outer row of rollers 40.
[0200] Each of these schematic representations of the present speed
change device is shown with an accompanying mathematical formula
which provides the basis for determining the speed change ratio and
output speed rotation relative to the input ring rotation direction
of each configuration. For each of these equations a negative
result indicates the inner race rotating the opposite direction of
the input ring. For clarity of explanation, all of the
configurations shown in FIGS. 55-59 referred to here are presented
with the outer race (or fixed race) 36, the input ring 190 and the
inner race (or output race) 38 as the rotating output.
[0201] It is possible, and in some cases preferable, according to
the principles disclosed in this document, to combine one or more
stages as described in the following drawings and other examples of
configurations in this document.
[0202] FIG. 55 shows an example of a first configuration. Equations
(4)-(6) provide the basis for determining the speed change ratio
and output speed rotation relative to the input ring rotation
direction of the first configuration. The input roller rotates at
same speed as output roller.
.omega. o .omega. i = e 1 .function. ( e 2 - 1 ) ( e 1 - 1 ) ( 4 )
e 1 = r 1 r f .times. r i r 2 ( 5 ) e 1 = r f r o ( 6 )
##EQU00003##
Where:
[0203] .omega..sub.o=output angular velocity .omega..sub.i=input
angular velocity r.sub.f=fixed race diameter r.sub.i=input ring
diameter r.sub.o=output race diameter r.sub.2=input roller diameter
r.sub.1=output roller diameter r.sub.3=idler roller diameter
[0204] FIG. 56 shows an example of a second configuration.
Equations (4), (7) and (8) provide the basis for determining the
speed change ratio and output speed rotation relative to the input
ring rotation direction of the second configuration.
.omega. o .omega. i = e 1 .function. ( e 2 - 1 ) ( e 1 - 1 ) ( 4 )
e 1 = ( r i r 1 .times. r 2 r f ) ( 7 ) e 2 = r f r o ( 8 )
##EQU00004##
Where:
[0205] .omega..sub.o=output angular velocity .omega..sub.i=input
angular velocity r.sub.f=fixed race diameter r.sub.i=input ring
diameter r.sub.o=output race diameter r.sub.2=input roller diameter
r.sub.1=output roller diameter r.sub.3=idler roller diameter
[0206] FIG. 57 shows an example of a third configuration. Equations
(4), (9) and (10) provide the basis for determining the speed
change ratio and output speed rotation relative to the input ring
rotation direction of the third configuration.
.omega. o .omega. i = e 1 .function. ( e 2 - 1 ) ( e 1 - 1 ) ( 4 )
e 1 = r 1 r f ( 9 ) e 2 = r f r o ( 10 ) ##EQU00005##
Where:
[0207] .omega..sub.o=output angular velocity .omega..sub.i=input
angular velocity r.sub.f=fixed race diameter r.sub.i=input ring
diameter r.sub.o=output race diameter r.sub.2=input roller diameter
r.sub.1=output roller diameter r.sub.3=idler roller diameter
[0208] FIG. 58 shows an example of a fourth configuration.
Equations (4), (11) and (12) provide the basis for determining the
speed change ratio and output speed rotation relative to the input
ring rotation direction of the fourth configuration. note: r.sub.4
does not contact r.sub.1.
.omega. o .omega. i = e .function. ( e 2 - 1 ) ( e 1 - 1 ) ( 4 ) e
1 = ( r i r 3 .times. r 4 r f ) ( 11 ) e 2 = ( r f r 4 .times. r 5
r 2 .times. r 1 r o ) ( 12 ) ##EQU00006##
Where:
[0209] .omega..sub.o=output angular velocity .omega..sub.i=input
angular velocity r.sub.f=fixed race diameter r.sub.i=input ring
diameter r.sub.o=output race diameter r.sub.2=input roller diameter
r.sub.1=output roller diameter r.sub.3=idler roller diameter
[0210] FIG. 59 shows an example of a fifth configuration. Equations
(4), (13) and (14) provide the basis for determining the speed
change ratio and output speed rotation relative to the input ring
rotation direction of the fifth configuration.
.omega. o .omega. i = e 1 .function. ( e 2 - 1 ) ( e 1 - 1 ) ( 4 )
e 1 = ( r i .times. r 2 .times. r 4 r 1 .times. r 3 .times. r f ) (
13 ) e 2 = ( r f r 4 .times. r 3 r o ) ( 14 ) ##EQU00007##
Where:
[0211] .omega..sub.o=output angular velocity .omega..sub.i=input
angular velocity r.sub.f=fixed race diameter r.sub.i=input ring
diameter r.sub.o=output race diameter r.sub.2=input roller diameter
r.sub.1=output roller diameter r.sub.3=idler roller diameter
[0212] As shown in FIG. 60, any one of the rollers disclosed in
this patent document may have nearly cylindrical traction surfaces
192 on the rollers 12 such that the surfaces are slightly smaller
in diameter toward the axial ends of the roller traction surfaces
to reduce Hertzian stress near the axial ends of the contact
surfaces as shown in FIG. 60 schematically and not to scale.
Similarly, the inner race 38 may have a slightly decreasing
diameter towards its extremities in the axial direction, and the
outer race 36 may have slightly increasing diameter towards its
extremities in the axial direction. Any of the shown rollers may be
hollow to allow for compressibility and enhance pre-loading. An
example of a hollow roller is shown in FIGS. 15 and 24, also FIG.
62 (inner roller 224).
[0213] In FIG. 61, a speed change device is shown with outer race
210, outer roller 212 (in a set of outer rollers), inner roller
214, inner race 216, and output 218 (in a speed increaser
embodiment, or input for a speed decreaser). Arrows in the section
show traction contacts. The outer rollers 212 are compound with a
larger diameter contacting the race 210 and a smaller diameter
contacting the inner roller 214. The inner roller 214 has the same
diameter contacting the inner race 216 and the outer roller
212.
[0214] In FIG. 62, a speed change device is shown with outer race
220, outer roller 222 (in a set of outer rollers), inner roller 224
and inner race 226. In this embodiment, the outer rollers 222 have
a larger maximum cylindrical surface diameter than the maximum
cylindrical surface diameter of the inner roller 224 and the outer
roller 222 has a larger diameter traction contact with the outer
race 220 than with the inner roller 224, and the inner roller 224
has the same diameter contact with the outer roller 222 as the
inner roller 224 contact with the inner race 226.
[0215] In the embodiment of FIG. 7 for example, contacts of each
inner roller with the inner race and with the outer rollers that
transmit torque all have the same torque transmitting diameter, and
contacts of each outer roller with the outer race and with the
inner rollers that transmit torque have a torque transmitting
diameter and all the torque transmitting diameters are equal.
However, in some embodiments the outer rollers may have a different
diameter contact with the inner surface of the outer race, than the
diameter of contact of the outer rollers with the inner rollers. In
some embodiments, the outer rollers have a larger diameter primary
torque transmitting contact with the inner surface of the outer
race, than the diameter in primary torque transmitting contact with
the inner rollers. In some embodiments, the outer rollers have a
larger diameter primary torque transmitting contact with the
primary torque transmitting contact inner surface of the outer
race, than the primary torque transmitting contact diameter in
contact with the inner rollers, and the inner rollers have the same
primary torque transmitting contact diameter with the outer rollers
as with the primary torque transmitting contact diameter of the
inner race. Further in some embodiments, the inner rollers have
larger diameter primary torque transmitting contact with the outer
rollers and smaller primary torque transmitting contact with the
outer surface of the inner race. In some embodiments, the outer
rollers have two diameters and inner rollers have one diameter such
that rotation of rollers causes orbiting of the rollers in one
direction resulting in rotation of inner race in the same direction
as roller orbit direction when the outer race is fixed. In some
embodiments, the outer rollers have two diameters and the inner
rollers have one diameter such that rotation of rollers causes
orbiting of rollers in one direction resulting in rotation of the
inner race in the opposite direction as the roller orbit direction
when the outer race is fixed. In some embodiments, the outer
rollers have two diameters and inner rollers have two different
diameters such that rotation of rollers causes orbiting of rollers
in one direction resulting in rotation of inner race in the same
direction as roller orbit direction when outer race is fixed. In
some embodiments, the outer rollers have two diameters and inner
rollers with two different diameters such that rotation of the
rollers causes orbiting of the rollers in one direction resulting
in rotation of the inner race in the opposite direction as the
roller orbit direction when outer race is fixed.
[0216] In the claims, the word "comprising" is used in its
inclusive sense and does not exclude other elements being present.
The indefinite articles "a" and "an" before a claim feature do not
exclude more than one of the feature being present. Each one of the
individual features described here may be used in one or more
embodiments and is not, by virtue only of being described here, to
be construed as essential to all embodiments as defined by the
claims.
* * * * *