CN106662230B

CN106662230B - With the stepless transmission for being uniformly input to output speed ratio independent of friction

Info

Publication number: CN106662230B
Application number: CN201480079073.2A
Authority: CN
Inventors: R·R·拉金德兰; P·普拉桑特·阿尔·拉金德兰
Original assignee: Individual
Current assignee: Individual
Priority date: 2014-03-18
Filing date: 2014-03-18
Publication date: 2019-08-20
Anticipated expiration: 2034-03-18
Also published as: CA2962854C; US9970520B2; US20150267794A1; JP6454456B2; CA2962854A1; CN106662230A; JP2017508939A; EP3120046A1; EP3120046A4; WO2015142323A1

Abstract

The main object of the present invention is to provide uniform stable output when input be uniform stable, with disobeying frictionally or the ability of coefficient of friction and transfer high torque.Many stepless transmissions are dependent on friction currently on the market, therefore lack the ability of transfer high torque.Those stepless transmissions for not relying on friction do not have uniform stable output when input is uniform stable.Design help reduces overall size, and economically facilitates mass production.The design is easily incorporated into any system.The design is very versatile, and use scope can be from light load to heavy load.The design allows to replace existing conventional transmission device, and needs very small modification.The design provides the fixed selection coaxially output and input.

Description

Continuously variable transmission with friction-independent uniform input-to-output speed ratio

Applicant

Name:

RAJA RAJENDRAN

nationality: united states of America

And (3) address: 5179SHADY CREEK DRIVE

TROY，MICHIGAN，48085

PRASHANTH RAJENDRAN

Nationality: united states of America

And (3) address: 5179SHADY CREEK DRIVE

TROY，MICHIGAN，48085

Cross Reference to Related Applications

Provisional application

Application No.: 61788563

Name: continuously variable transmission

Background

Patents US 5603240 and US 20100199805 use some of the features used in the present design.

The advantages of the invention include:

the patent US 5603240 has no coaxial input to output and therefore cannot be used for applications requiring this configuration. The output moves as the speed ratio changes. Therefore, this design cannot be used when a fixed output is required. The new invention provides fixed and coaxial input and output shafts. The envelope (envelope) used in the prior art is much larger.

US 20100199805 provides a sinusoidal output and uses several modules only to minimize "ripple" when providing a stable and uniform input. Therefore, this design cannot be used when a stable and uniform output is required. The invention provides a stable and uniform output when the input is stable and uniform. This can be achieved by as few as 3 modules.

Disclosure of Invention

The primary object of the present invention is to provide a uniform and stable output when the input is uniform and stable, with the ability to transmit high torque independent of friction or coefficient of friction. Many continuously variable transmissions on the market today rely on friction and therefore lack the ability to transmit high torques. Those continuously variable transmissions that do not rely on friction do not have a uniform and stable output when the input is uniform and stable. This design helps reduce the overall size and economically facilitates mass production. The design can be easily incorporated into any system. The design is very versatile and can range from light to heavy duty. This design allows for the replacement of existing ordinary transmissions with very little modification. This design provides a fixed choice of coaxial input and output.

Drawings

FIG. 1-perspective view of a sequentially assembled overall assembly of a CVT;

FIG. 2-perspective view of the sequentially assembled general assembly of the CVT with the frame transparent, showing the general arrangement of the component internal sub-assemblies;

figure 3-frame-main housing-two identical parts are bolted together to form one main housing:

A. a perspective view detailing one side of the main housing;

B. a perspective view detailing the other side of the main housing;

FIG. 4-perspective view of the frame-telescoping sleeve Guide (Guide);

FIG. 5-perspective view of the rack-crossed rack guide;

FIG. 6-perspective view of the input shaft;

FIG. 7-perspective view of the intermediate gear shaft;

FIG. 8-perspective view of the power connection shaft;

FIG. 9-supporting shaft perspective;

fig. 10-shows two perspective and orthographic views of a crossed rack assembly, detailing the input shaft slots and crank pin slots, the orientation of the rack, detailing the tines (prongs):

a-top view;

b-perspective view 1;

c-perspective view 2;

d-a main view;

e-side view;

f-rear view;

g-an enlarged view detailing the slave tine;

fig. 11-pinion:

a-a front view;

b-side view;

c-top view;

d-perspective view;

fig. 12-pinion shaft:

a-a front view;

b-side view;

c-perspective view;

figure 13-crank pin holder:

a-a front view;

b-side view;

c-perspective view;

fig. 14-input disc:

a-a front view;

b-side view;

c-perspective view;

fig. 15-gear change lever (lever) -planetary mechanism:

a-a front view;

b-top view;

c-perspective view;

FIG. 16-compression spring perspective;

FIG. 17-static collar bevel ring gear-perspective view;

fig. 18-primary telescopic sleeve:

a-a front view;

b-side view;

c-perspective view;

figure 19-secondary telescope:

a-a front view;

b-side view;

c-top view;

d-perspective view;

FIG. 20-ratio cam:

a-a front view;

b-top view;

c-perspective view;

figure 21-non-circular gear (driven):

a-top view;

b-a front view;

c-perspective view;

figure 22-non-circular gear (active):

a-top view;

b-a front view;

c-perspective view;

figure 23-imitation (Dummy) crank pin:

a-top view;

b-a front view;

c-perspective view;

figure 24-crankpin:

a-top view;

b-a front view;

c-side view;

d-perspective view;

fig. 25-intermediate circular gear C2-C3:

a-a front view;

b-side view;

c-perspective view;

FIG. 26-supporting gears C4a-C5 b:

a-a front view;

b-side view;

c-perspective view;

fig. 27-intermediate circular gear C4-C5:

a-a front view;

b-side view;

c-perspective view;

fig. 28-intermediate circular gear C1:

a-a front view;

b-side view;

c-perspective view;

FIG. 29-Spacer (Spacer):

a-a front view;

b-top view;

c-perspective view;

fig. 30-gear change Lever (Lever) for spiral groove mechanism:

a-a front view;

b-side view;

c-top view;

d-perspective view;

FIG. 31-helical groove:

a-a front view;

b-side view;

c-perspective view;

fig. 32-static differential ferrule:

a-a front view;

b-side view;

c-sectional view;

d-perspective view;

fig. 33-dynamic differential ferrule:

a-a front view;

b-side view;

c-sectional view;

d-perspective view;

FIG. 34-perspective view of sleeve-input-ramp;

fig. 35 to 43-show the movement/position on the rack assembly, the crank pin rotating with the input disc: the various stages are shown:

FIG. 35-crank pin closer to axis, input disc 0;

figure 36-crank pin closer to axis, input disc 45 °;

FIG. 37-crank pin closer to axis, input disc 90;

FIG. 38-crank pin at midpoint, input disk 0;

FIG. 39-crank pin at midpoint, input disc 45;

FIG. 40-crank pin at midpoint, input disc 90;

FIG. 41-crank pin furthest from gear, input disc 0;

FIG. 42-crank pin furthest from gear, input disc 45;

FIG. 43-crank pin furthest from gear, input disc 90;

FIG. 44-exploded view depicting input modification-perspective view detailing the non-circular gear and intermediate gear to input disc arrangement and gear set;

45-46-perspective views of ratio cam, input disc and crankpin, showing the back operation of the cam how the pin position is changed:

FIG. 45-input disk side (ratio cam and input disk shown in transparency for clarity);

FIG. 46-ratio cam side;

fig. 47 to 50, show the operation of the planetary gear change mechanism:

FIG. 47-perspective view of the planetary gear shifting mechanism; for clarity, the mainframe is shown semi-transparent;

FIG. 48-shows a perspective view of the planetary gear change mechanism, detailing the circular slot in the main frame, which is shown semi-transparent (closed) for clarity;

FIG. 49, which shows a front view of the planetary gear shifting mechanism, with the main frame made transparent for clarity;

FIG. 50, which shows a side view of the planetary gear shifting mechanism, with the main frame made transparent for clarity;

FIG. 51-shows an exploded view of the differential mechanism showing the configuration and operation of the components (perspective view);

fig. 52 to 57-depict the ratio shifting operations at different stages of the differential mechanism, showing a partial section to explain the function and internal details:

FIG. 52-differential mechanism (partial cross-section) view 1;

FIG. 53-differential mechanism (partial cross-section) view 2;

FIG. 54-differential mechanism (partial section) view 3;

FIG. 55-differential mechanism (partial cross-section) view 4;

FIG. 56-differential mechanism (partial cross-sectional) view 5;

FIG. 57-differential mechanism (partial section) view 6;

FIG. 58-assembly, showing the operation of the gear change mechanism-the spiral groove mechanism (exploded);

FIG. 59-a top view illustrating the operation of the telescoping guide;

FIG. 60-shows the telescoping mechanism in detail, with one side of the primary and secondary transparent to show detail;

fig. 61 to 62-input disc, cross-rack assembly, crank pin and crank pin holder assembly, illustrating the principle behind the crank pin holder function:

FIG. 61-crank pin and crank pin retainer when in the middle of the input slot;

FIG. 62-crank pin and crank pin holder with crank pin holder exiting the input slot;

FIG. 63-an exploded view of the one-way bearing assembly (pinion gear partially cut away to show internal detail);

FIG. 64-one-way bearing assembly;

FIG. 65-Power connection Assembly;

FIG. 66-assembly, showing the principle of vibration cancellation;

figure 67-vibration cancellation mechanism: a sub-assembly;

fig. 68-complete CVT assembly, showing orientation of module and rack: how 4 modules are placed is illustrated;

FIGS. 69 through 72-selection of non-circular gear positions when a common non-circular drive gear is used with two non-circular driven gears;

FIG. 69-non-circular gear set at 135 °;

FIG. 70-non-circular gears set at 45;

FIG. 71-non-circular gear set at (-)45 °;

FIG. 72-non-circular gear set to (-)135 °;

figures 73 to 75-show in detail how a constant uniform output is achieved:

FIG. 73-Assembly orientation of a single module;

FIG. 74-graph showing individual output and combined total output for each rack, showing overlap

The divided constant and uniform output is carried out;

FIG. 75-a graphical representation of an output with overlapping portions and a complete cycle of the engagement sequence;

FIGS. 76 to 79-depict the helical gear assemblies for forward (forward), reverse (reverse), neutral (neutral) and park (park):

FIG. 76-engagement of the clutch for the forward gear;

FIG. 77-engagement of the clutch for reverse gear;

FIG. 78-engagement of the clutch for the neutral gear;

FIG. 79-engagement of the clutch for "park";

figure 80-principle of using intermediate gears to eliminate multiple joints between non-circular gears:

a-top view;

b-a front view;

figure 81-coaxial output element:

a-a front view;

b-cross-sectional side view;

c-perspective view;

figure 82-shows in detail the arrangement of the coaxial output members in the assembly;

Detailed Description

Briefly described, the present invention is a Continuously Variable Transmission (CVT). Unlike existing CVT designs, this particular design does not rely on friction to transmit power. Most current CVTs rely on friction to transmit power and therefore cannot be used where high power transmission at low speeds is required. Due to this advantage, the invention can be used where high torque transmission is required. Coaxial input and output can be achieved using this design.

The operation of the CVT can be described by the following single sequence operation.

a) The crank pin (fig. 23) rotates about the axis of the input disc (fig. 14) by an offset distance that can be modified. The principle described in this operation is presented in another patent US 20100199805. However, here a completely different approach is adapted in a simpler and compact envelope, how to use the principle, how to modify the offset, etc. ]

b) The offset crank pin 42 snaps (caged) into the input disk 16, alternatively into the crank pin shaft collar, and into a slot in the rack assembly (fig. 10), the crank pin shaft collar sliding over the crank pin shaft, constraining the rack assembly so that the rack can only move in a direction parallel to the rack 64. The crank pin axis is orthogonal to the input axis (fig. 6). By positioning the other slot perpendicular to the direction of motion, the rotational motion of crank pin 42 is translated into a purely linear back and forth motion of rack 64. This mechanism is commonly referred to in the industry as a "scotch yoke" mechanism. The distance (stroke) of this linear back and forth movement is proportional to the radial distance of the crank pin 42 from the axis of the input disc 16.

c) The rack 64 is connected to a pinion (fig. 11) which converts this linear movement of the rack 64 into a rocking vibration of the pinion 47.

d) This rocking vibration is converted to unidirectional rotation through the use of a ratchet mechanism/one-way bearing/computer controlled clutch.

A main object of the present invention is to achieve a constant and uniform output angular velocity when the input angular velocity is constant and uniform. However, with the above steps, this cannot be achieved because the output is sinusoidal. By adjusting the rate of change of the angular displacement of the input disc 16, uniform and stable output can be achieved. The rate of change of angular displacement on the input disc 16 can be varied by using a set of non-circular gears, driving (fig. 22) and driven (fig. 21). The output of the driven non-circular gear 9 is then transmitted to the input disc 16 via some intermediate circular gears.

Given the profile of one of the non-circular gears 8 by the equation, when the radius "r" is expressed as a function of θ: r (θ) ═ R × K × CTR/[ R × K + f (θ) ], where "K" is a constant that depends on all constant gear radii and "R" is the desired speed ratio, which is the ratio between the rate of change of angular displacement of the input on the active non-circular gear 8 and the output on the input disc 16.

The desired value of "R" is generally 1. "K" is generated from the radius of the intermediate gear, which is equal to the product of the driven gear radius divided by the product of the drive gear radius. The ideal value of "K" is generally 1. "CTR" is the center-to-center distance of the two non-circular gears 8 and 9. This is selected based on the available envelope of the assembly.

f (θ) can be sin θ or cos θ. Both equations produce identical and interchangeable profiles except that they are rotated 90.

The profile of the conjugate non-circular gear 9 is given by the formula R (θ) ═ CTR- { R × K × CTR/[ R × K + f (θ) ] }. The generation of these contour shapes and parameters used is explained in detail in the following topics.

To aid in understanding the present invention, a design creates a CAD model and is described below.

The features used here are:

the value of "R" is chosen to be 1.

The value of "K" is chosen to be 1.

A common input shaft (fig. 6) and a driving non-circular gear 8 are used for all four modules.

A common cross rack assembly 44, input disk 16, driven non-circular gear 9, intermediate circular gear, crankpin 42, ratio cam (fig. 20), and ratio varying mechanism are used for both modules.

Two racks 64 are located on the cross-rack assembly 44 with a 180 ° phase shift.

Another identical assembly of modules is placed such that the second assembly of modules is laterally inverted and rotated 90 deg. with respect to the first assembly of modules. The laterally inverted planar selection results in a variety of assembly configurations, such as sequentially assembled assemblies (fig. 1) or conjoined assemblies.

List of components:

1) frame main casing

2) Rack-crossed rack guide

3) Rack-telescopic guide

4) Input shaft

5) Input shaft bearing

6) Intermediate gear shaft

7) Intermediate gear shaft bearing

8) Non-circular gear (initiative)

9) Non-circular gear (driven)

10) Intermediate circular gear C1

11) Intermediate circular gear C2-C3

12) Intermediate circular gear C4-C5

13) Bearings-ring (static and dynamic)

14) Bearing-circular gear C2-C3

15) Bearing-circular gear C4-C5

16) Input disc

17) Bearing input disc

18) Speed ratio cam

19) Bearing-ratio cam

20) Middle supporting circular gear C4a-C5a

21) Supporting axle

22) Bearing-bearing axle

23) Speed ratio change control lever-planetary mechanism

24) Sleeve-input disc-oblique

25) Static differential ferrule

26) Static differential ferrule straight gear shaft bearing

27) Static differential ferrule straight gear shaft

28) a) static differential ferrule bevel pinion

b) Static differential ferrule large bevel gear

29) Static differential ferrule straight gear

30) Gasket

31) Dynamic differential ferrule

32) Dynamic differential ferrule straight gear shaft bearing

33) Dynamic differential ferrule straight gear shaft

34) a) dynamic differential ferrule bevel pinion

b) Dynamic differential ferrule large bevel gear

35) Dynamic differential ferrule straight gear

36) Universal joint

37) Helical groove

38) Slotted disk-input disk

39) Compression spring

40) Thrust bearing

41) Speed ratio change control rod-spiral groove mechanism

42) Crank pin

43) Imitation crank pin

44) Cross rack assembly

45) Primary telescopic sleeve

46) Secondary telescopic sleeve

47) Pinion gear

48) Pinion shaft

49) Pinion bearing

50) One-way bearing

51) Output sprocket/gear

52) Power connecting shaft

53) Power connecting shaft bearing

54) Power coupling sprocket/gear

55) Imitation rack

56) Wheel-vibration cancellation

57) Collar-wheel-vibration cancellation

58) Input shaft for helical bevel gear

59) Bevel gear with helical teeth

60) Clutch-park/neutral/reverse

61) Output shaft

62) Intermediate gear-non-circular gear connector

63) Guide-intermediate-non-circular gear connector

64) Rack bar

65) Coaxial output element

Auxiliary input shaft

Auxiliary input shaft ferrule

Connecting rod

Crank pin shaft

Crank pin shaft sleeve ring

Planetary gear system

Description of the constituent assemblies, sub-assemblies and their function:

description of the overall configuration:

the input shaft (fig. 6) is mounted on two input shaft bearings 5 and is placed in the center of the frame-main housing (fig. 3). Input disc 16 is mounted on input shaft 4 and is sandwiched between a rack assembly (fig. 10) and a ratio cam (fig. 20), with crank pin 42 captured in the slot. The crank pin 42 has a body shaped like a rectangular prism having circular prisms extending on both sides. One of which acts as a cam-follower and causes it to engage the ratio cam and the other of which acts as a crankpin 42 and causes it to engage a rack 64 on cross-rack assembly 44. A driving non-circular gear 8 parallel to the input disc 16 is mounted on the input shaft 4.

The intermediate gear shaft (fig. 7) is mounted on two constant gear shaft bearings 7, one constant gear shaft bearing 7 per main housing 1. The intermediate gear shaft 6 is placed parallel to the input shaft 4 at a distance "CTR" for creating the shape of the non-circular gear. The powertrain transmitted from the input shaft 4 to the input disc 16 follows the table provided below.

The driven non-circular gear 9 and the intermediate gears C2-C3 (fig. 25) are mounted on the input shaft 4, and the intermediate gear C1 (fig. 28) and the intermediate gears C4-C5 (fig. 27) are mounted on the constant gear shaft 6. The driving non-circular gear 8 is directly mounted on the input shaft 4, and the driven non-circular gear 9 together with the intermediate gear C110 is directly mounted on the intermediate gear shaft 6. The other gears are placed in bearings and mounted on their respective shafts.

The rack assembly 44 is free to move only in the direction along the rack 64, the movement of which is constrained by the rack-and-pinion guide 2. A primary and secondary set of telescoping sleeves are placed on either side of the rack assembly 44. This will reduce the overall size required for the rack assembly 44 and the housing main housing 1. A tine is placed on either side of the rack assembly 44 and another tine is placed on the secondary sleeve 46 to pull and extend the telescoping sleeve, which is collapsed by the body of the rack assembly 44 (collapsed). The telescoping sleeves are captured by the frame telescoping-guides (fig. 4).

The rack 64 is coupled with a one-way bearing assembly (fig. 64) consisting of a pinion 47 placed on a pinion shaft (fig. 12). The pinion shaft 48 is mounted on the frame telescoping guide 3 with a pinion bearing 49. A gear or sprocket is mounted on the pinion shaft 48 by a one-way bearing 50 and is placed parallel to the pinion 47. The power connection shaft assembly (fig. 65) is placed parallel to the one-way bearing assembly (fig. 64). The power connection assembly consists of a power connection shaft (fig. 8) mounted on two bearings placed on the frame-telescopic guide 3. A gear or sprocket is placed on each end of the power connecting shaft. The power of the pinion shaft 48 is transmitted to the power connection through the gear or sprocket.

Operation and principle of the primary CVT:

as input disc 16 rotates, crank pin 42, through a "scotch yoke" mechanism, moves the cross-rack assembly in a direction parallel to rack 64. The distance traveled is proportional to the distance between the axis of crank pin 42 and the axis of input disc 16. By varying this distance, the distance traveled by the rack assembly, which is referred to as the "stroke," can be varied. Since work is constant, it is the product of the applied force times the distance traveled (F × stroke). For shorter strokes, the force applied is greater and for longer strokes, the force applied is less. However, the motion is a back and forth swing. This force from the linear back and forth movement of the rack 64 is then transferred to the pinion 47 as a rocking motion. The torque generated by this rocking motion is proportional to the force exerted by the rack 64. This is transferred to the output sprocket/gear for unidirectional rotation by a one-way bearing 50 or a computer controlled clutch or ratchet mechanism. This unidirectional rotation is further transmitted to the wheel.

Power transmission configuration from engine/power source to input disc 16:

the rate of change of angular displacement on the input disc 16 is varied by using a set of non-circular gears, driving (fig. 8) and driven (fig. 9). The output of the input shaft 4 is transmitted through a set of non-circular gears, and then the output of the input shaft 4 is transmitted to the input disc 16 through 5 intermediate circular gears. The non-circular drive gear 8 is mounted directly on the input shaft 4. The driven non-circular gear 9 is mounted on an intermediate gear shaft (fig. 7) which is mounted on two bearings 7 and placed on the two main housings 1.

The intermediate circular gear C110 is mounted on the intermediate gear shaft 6 and is directly connected to the driven non-circular gear 9. Intermediate gears C2-C3 (fig. 25) are mounted on the input shaft 4 to rotate freely with the bearings 14. Intermediate gears C4-C5 (fig. 26) are mounted on the intermediate gear shaft 6, the intermediate gear shaft 6 freely rotates with the bearing 15, and the intermediate gear C5 drives the input disc 16. The radii of these intermediate gears are selected so that one revolution of the input disc 16 is completed by one revolution of the driving non-circular gear (fig. 22). It should satisfy the conditions-rC 2/rCl ═ nl, rC4/rC3 ═ n2, rdisc/rC5 ═ nl ═ n2, the K value will be 1.

The reason behind the need for circular gears between non-circular gears when profiles interfere/make multiple contacts simultaneously:

depending on the values selected for the variables "R", "K", and "CTR", the shape of the non-circular gear may have multiple contact points at any given point in time. As can be seen from the equation for the non-circular gear profile, the driven non-circular gear 9 has a lower radius than the input shaft 4, is mounted over a wide area and reaches zero at two locations. In addition, it is possible that the driven non-circular gear 9 and the driving non-circular gear 8 may have multiple contact points at a given time due to the profile shape. This can be eliminated by inserting the intermittent circular gear 62 between two non-circular gears. This increases the distance between the two non-circular gears and eliminates the problem of multiple contact points at any given time.

The principle behind the ratio change cam is used:

to change the speed ratio input to output, the position of crank pin 42 must be changed. This can be achieved by rotating the ratio cam disc 18, which ratio cam disc 18 has a groove with a profile. This profile forces crank pin 42 to move in a radial direction of the disc axis as ratio cam plate 18 rotates about input disc 16. This is because the axis of crank pin 42 extends through the slots of input plate 16 and the slots of ratio cam plate 18. As the crank pin 42 is closer to the axis of the input disc 16, the stroke is shorter and the force increases as the work done is constant. Likewise, as the crank pin 42 moves away from the axis of the input disc 16, the stroke is longer and the force decreases as work is constant. However, the challenge here is to have the ratio cam plate 18 and input disc 16 rotate synchronously during normal operation, and when a ratio change is required, the input disc 16 and ratio cam plate 18 should have relative angular velocities. By using one of the three mechanisms described below, relative angular velocity between the input disc 16 and the ratio cam disc 18 can be achieved when desired.

The method for changing the speed ratio comprises the following steps:

1. the planetary mechanism:

a set of intermediate support circular gears C4a and C5a (fig. 26) are axially connected and mounted on a common support shaft (fig. 9). C4a is identical to circular gear C4 and C5a is identical to circular gear C5. The movement of this common axis is constrained by the circular grooves/tracks, which are at a constant distance from the axis of rotation of the input disc 16 and the ratio cam disc. Gear 4a is radially connected to gear C3 and gear C5a is radially connected to ratio cam plate 18. A ratio changing lever-planetary mechanism (fig. 37) pivoted on the frame enables the position of the support shaft 21 to move along the slot. When this position is shifted, there is a relative angular displacement between the input disc 16 and ratio cam disc 18.

2. Spiral groove mechanism:

a helical groove input disc collar (fig. 38) having a twisted profile is axially connected to the input disc 16. A groove matching the twisted profile of the helical groove is cut in the ratio cam disc 18 and is positioned coaxially with the input disc 16. Input disc 16 and ratio cam disc 18 rotate in unison while the distance between ratio cam disc 18 and input disc 16 remains constant. As the distance between the input disc 16 and the ratio cam disc 18 is changed, the relative angular velocity between the input disc 16 and the ratio cam disc 18 changes as the ratio cam disc 18 is forced to rotate about the input disc 16. This axial translation is achieved using a ratio change control rod 41-helical groove mechanism that pushes a thrust bearing 40 connected to the ratio cam plate 18 towards the input disc 16. A compression spring (fig. 58) placed between the input disc 16 and ratio cam disc 18 causes it to spring back.

3. A differential mechanism:

the static collar bevel ring gear 28b is axially connected to the input disc 16 by a sleeve-input disc-to-bevel (fig. 32). A static differential collar (fig. 32) spaced coaxially from large bevel gear 28b is free to rotate independently about large bevel gear 28b by thrust bearing 40. The static differential collar 25 is constrained from moving axially with respect to the large bevel gear 28 b. The freely rotating static collar shaft 27 is placed perpendicular to the axis of the static differential collar 25 in a bearing 26, the bearing 26 being placed in the static differential collar 25. The static-ring small bevel gear 128a and the static differential-ring spur gear 29 are axially rigidly connected to the static-ring shaft 27, and the static-ring small bevel gear 128a is paired with the static-ring large bevel gear 28 b.

In the same way as above, the first and second,

a dynamic large bevel gear (fig. 17) is placed coaxially parallel to the ratio cam plates so that they rotate synchronously but allow displacement along the axis between them. A dynamic differential collar (fig. 33) placed coaxially with the dynamic collar bevel gear 28a separated by the thrust bearing 40 is free to rotate independently with respect to the dynamic collar bevel gear 34 b. The dynamic differential collar 31 is constrained from moving axially with respect to the dynamic collar large bevel gear 34 a. A freely rotating dynamic collar shaft 33 is placed in the bearing 32 perpendicular to the axis of the dynamic differential collar, the bearing 32 is placed in the dynamic differential collar 31, the dynamic collar shaft 33 has a universal joint 36 placed on its axis. The dynamic-collar small bevel gear 34a and the dynamic-collar spur gear 35 are axially rigidly connected to the dynamic-collar spur gear shaft 33, and the dynamic-collar small bevel gear 34a is paired with the dynamic-collar large bevel gear 34 b. The universal joint 36 is common to both the dynamic collar spur shaft 33 and the pinion shaft, allowing for a small mismatch.

The spacer keeps the two spur gears in contact. The shim (fig. 29) is free to move axially with respect to the dynamic collar straight gear shaft 33.

The static differential collar 25 and the dynamic differential collar 31 are identical and interchangeable here.

With this configuration, the power transmission path is as follows:

a. the static collar bevel pinion 28a rotates the static collar bevel pinion 28 b.

b. The static collar bevel pinion 28 rotates the static collar shaft 27.

c. The static collar shaft 27 rotates the static collar spur gear 29.

d. The static collar spur gear 29 causes the dynamic collar spur gear 35 to rotate.

e. The dynamic collar spur gear 35 causes the dynamic collar shaft 33 to rotate.

f. The dynamic collar shaft 33 rotates the dynamic collar bevel pinion 34a via the universal joint 36.

g. The dynamic collar bevel pinion 34a causes the dynamic collar bevel pinion 34b to rotate.

h. The dynamic collar large bevel gear 34b rotates the ratio cam plate 18.

Since the two large bevel gears, the two small bevel gears and the spur gears are respectively identical and have the same size, when the dynamic differential ferrule 31 is static, the angular velocity of the ratio cam plate 18 is synchronous with the input disc 16. When the dynamic differential ring 31 is rotated about the static differential ring 25, there will be a relative angular displacement between the input disc 16 and the ratio cam disc 18.

4. Link mechanism

The auxiliary hollow input shaft has a cross-section with a circular bore in the middle and a non-circular shape about an outer boundary. This mates with sliding ferrules with mating holes that are placed coaxially, allowing axial movement while constraining rotational movement with respect to each other. The thrust bearing 40 is placed coaxially in contact with one of the collars, which has a pivot on the other end. One end of the connecting rod is connected to the pivot and the other end is connected to either the crankpin or the crankpin shaft collar, as the case may be. Axial displacement of the collar will produce radial displacement of the crank pin 42 by the connecting rod. This axial translation is achieved using a speed ratio change control rod 41, which speed ratio change control rod 41 pushes the thrust bearing 40 to connect the sliding collar. A compression spring placed between the input disc 16 and the auxiliary input shaft collar causes it to spring back.

The principle behind the use of telescopic sleeves to make the design compact:

for operation of the design, the length of the input slot of the rack assembly must be equal to 2 stroke + input shaft diameter +2 minimum material thickness +2 distance to the rack guide. The entire length must be guided by the rack guide. Since the rack guide must also accommodate the trajectory of the rack 64, the open portion of the rack guide should have a width of at least the diameter of the input disc 16 or it will not reach when the rack 64 travels to one side to the farthest end. The telescopic guide allows the support member to extend, and therefore, the entire length of the rack assembly can reduce the "distance to the rack guide". This also enables the main casing 1 to be shorter by reducing the distance. Tines are provided on the design of the rack assembly and the second sleeve to extend the telescoping sleeve. The body of the rack assembly collapses the telescoping sleeve.

Principle behind the use or operating function of the slide guide:

the crankpin is smaller than the input shaft 4. The crank pin is likely to slide into the input shaft slot due to the intersection of the two slots. This is eliminated by using a slider guide (fig. 13) that is larger than the input shaft slot. Causing it to float in the crankpin slot surrounding (enclosing) the crankpin 42.

The design for realizing the principle is that the power transmission overlapping part:

to ensure a smooth transition from one module to the next, for a short time, both modules are active and join when their outputs reach a constant uniform value. The first module is released while it is still in the functional area, while the second module is well located in the functional area.

Modules and their assembly design and constraints:

all four modules share a common input shaft and a common non-circular drive gear. The two modules share a common input disc 16 and gear change mechanism. The racks are placed 90 ° out of phase with respect to the next. To accommodate this, the driven non-circular gear 9 is oriented at 45 ° and the driven non-circular gear 9 is phase shifted by 45 ° relative to the other non-circular driven gear. And since the non-circular gear is symmetrical it can also be oriented at 135 deg.. This increases the 90 ° phase shift between the racks.

Principle of power transmission or connection between modules:

when the modules are run in sequence, they must be connected before power is transmitted to the wheels. This is achieved by using a power connection shaft 52, the power connection shaft 52 having a gear or sprocket to connect the output of each module so that it has continuous power to the wheels. Power is also transmitted in sequence.

A reverse gear mechanism:

the output of the power connecting shaft 52 is coupled to the input shaft 4 of the bevel gear differential. Therefore, the outputs of these helical gears rotate in opposite directions. The output shaft 61 is coaxially disposed with a gap with respect to the output helically toothed bevel gear if the differential mechanism is free to rotate independently with respect to the output helically toothed bevel gear. Two collars with clutches are placed on the output shaft 61, allowing the two collars to move coaxially. The two collars can be made to connect with either output helical bevel gear rotating in opposite directions. When one of the races is caused to engage a particular output bevel gear by a clutch, the output shaft 61 will rotate in a particular direction. If the connection is switched to the other output bevel gear, its direction will be reversed.

Neutral gear mechanism:

when the collar is not connected to any of the output bevel gears, the collar and the output shaft 61 are not constrained and therefore they are free to rotate in either direction and act as a "neutral" gear.

Parking shelves mechanism:

when the collar is connected to two output bevel gears, the collar is constrained in rotation and acts as a "park" gear.

Features and mechanisms to counteract vibrations:

1. imitation of crank pin: when the input disc 16 rotates, the crank pin is placed off-center. This imbalance can lead to vibration. To counteract this vibration, dummy crankpins are placed 180 ° apart by the same distance. This is cam shifted by the same ratio that moves the crankpin. The motion is the same as that of the crank pin. So that the cam grooves are equally spaced 180 apart.

2. Dead load of counter oscillation: as the input disc 16 rotates, the cross-rack assembly oscillates, which results in vibration. Which is eliminated by the oscillation of the suitable mass in the opposite direction. This is achieved by the fixed wheel being in contact with the rack 64, which will have a back and forth rotation. Contacting the wheel 180 ° apart with a suitable mass will counteract this vibration.

Coaxial input and output selection features:

where coaxial input and output is required, this can be achieved by incorporating an output member 65, the output member 65 having internal gears which are paired with power connection gears. Bearings are placed between the input shaft 4 and the coaxial output member 65 so that they rotate independently.

And (3) constraint:

when K ═ 1 and R ═ 1, the conditions apply:

the number of teeth of the driving non-circular gear (fig. 22) should be the same as the number of teeth of the driven non-circular gear (fig. 21), which means that their circumferences are the same, i.e. they complete 1 revolution at the same time, even though the instantaneous rates may not be the same. Alternatively, the second arrangement of non-circular gears can optionally be used for parallel achievement purposes without following the portion of the desired shape, i.e. using the portion of minimum radius "r".

The application is rc2/rc l n, rc4/rc3 n2, and rdisc/rc5 n 2.

It is desirable, but not mandatory: (rvl + rv2) ═ rc3+ rc4 ═ rc5+ rdisc ═ rcl + rv2 ═ ctr. This will allow the placement of all driving and driven gears on two common shafts, one of them being the input shaft 4.

Mathematical derivation:

the main purpose is to determine the mathematical expression of the shape of the non-circular gear, such that v_{Rack bar}(the linear speed of the rack 64) is constant.

Wherein:

v_{rack bar}＝ω_Dish*r_{Gear wheel}*f(θ)

ω_{Output of}＝ω_Dish*r_{Gear wheel}*f(θ)

Wherein,

ω_{input device}Input angular velocity

Angular velocity of the driving non-circular gear

Angular velocity of driven non-circular gear

Constant angular speed of the gear 1

Constant angular speed of the gear 2

Constant angular speed of the gear 3

Constant angular speed of the gear 4

Constant angular speed of the gear 5

ω_DishAngular velocity of the disc

ω_{Output of}Output angular velocity of the output

Radius of the driving non-circular gear

Radius of driven non-circular gear

Radius of the constant gear 1

Radius of the constant gear 2

Radius of the constant gear 3

Radius of the constant gear 4

Radius of the constant gear 5

r_DishRadius of the disc

r_{Offset of}Radial position of the crank pin

R-input to output angular velocity ratio

K- (ratio of radius product of driven gear and driving gear)

CTR-center distance between two non-circular gears

f (theta) -sin theta or cos theta

Claims

1. A continuously variable transmission comprising:

at least one module, the module comprising:

(a) an input disc having a radial slot extending radially in length, said input disc disposed between (b) a speed ratio cam disc and (c) a cross-rack assembly;

(b) a speed ratio cam plate including a non-radial slot extending at least partially in a non-radial direction;

(c) a cross rack assembly comprising one or more racks having longitudinal axes perpendicular to a longitudinal axis of a first slot that receives (d) a crank pin;

(d) a crank pin, wherein the crank pin is disposed in a radial slot of the input disc, a non-radial slot of the ratio cam disc, and a first slot of the cross-rack assembly, and extends parallel to a longitudinal axis of the input disc;

(e) one or more pinion gears mounted on one or more pinion shafts and coupled with the corresponding one or more racks; and

(f) at least one driven non-circular gear having a functional region and a non-functional region operatively connected to the input disc;

wherein the at least one module is arranged such that at least one driving non-circular gear disposed on the input shaft rotates at a uniform angular velocity about a longitudinal axis and meshes with and drives the at least one driven non-circular gear, thereby causing a non-uniform angular velocity of the input disc about its longitudinal axis, wherein the crankpin reciprocates a cross-rack assembly, wherein the cross-rack assembly only allows movement along the longitudinal axis of the one or more racks, and the reciprocating movement of the cross-rack assembly rotates the one or more pinions, and the rotation of the one or more pinions periodically alternates direction and is converted into unidirectional rotation of the output gear or sprocket, wherein, when the pinions rotate in a particular direction, the output gear or sprocket is driven by the one or more pinion shafts and unidirectional bearings, At least one of a ratchet mechanism, a computer controlled clutch rotates, and at least one of a one-way bearing, a ratchet mechanism, a computer controlled clutch engages the output gear or sprocket with the pinion.

2. The variable transmission of claim 1, wherein the ratio cam plate and the input disc are positioned adjacent and coaxial to each other and are controllable to rotate synchronously or asynchronously by the control mechanism, and the longitudinal axis of the crankpin remains at a constant distance from the longitudinal axis of the input disc when rotating, and the distance from the longitudinal axis of the crankpin to the longitudinal axis of the input disc is varied by the ratio varying mechanism when rotating asynchronously.

3. The continuously variable transmission of claim 2, wherein the control mechanism comprises a first pair of bevel gears comprising: a first drive bevel gear and a first driven bevel gear having different pitch circle diameters, wherein the first drive bevel gear is coaxially connected to the input disc, and the first driven bevel gear is coaxially connected to a drive spur gear which in turn rotates an identical driven spur gear spaced apart from the drive spur gear by a set distance using a spacer, and the driven spur gear is coaxially connected to a second drive bevel gear of a second pair of bevel gears, rotating the second driven bevel gear of the second pair of bevel gears, wherein the first drive bevel gear is identical to the second driven bevel gear, the first driven bevel gear is identical to the second drive bevel gear, and the second driven bevel gear is coaxially connected to the speed ratio cam disc; when there is no relative motion between the longitudinal axes of the driving spur gear and the driven spur gear, the input disc and the speed ratio cam disc rotate synchronously, and when there is relative motion between the longitudinal axes of the driving spur gear and the driven spur gear, the input disc and the speed ratio cam disc rotate asynchronously, and the asynchronous rotation changes the distance between the longitudinal axis of the input disc and the longitudinal axis of the crank pin through the speed ratio changing mechanism, so as to change the linear displacement of the cross rack assembly.

4. The continuously variable transmission of claim 3, wherein a universal joint is provided at an intersection of a longitudinal axis of the driving spur gear and a longitudinal axis of the first driven bevel gear, or a universal joint is provided at an intersection of a longitudinal axis of the driven spur gear and a longitudinal axis of the second driving bevel gear, or both.

5. The continuously variable transmission of claim 2, wherein the control mechanism includes a spiral groove collar coaxially attached to the input disc, and the ratio cam disc defines a bore having a shape matching the spiral groove collar and is positioned coaxially with the spiral groove collar such that the ratio cam and the input disc are separated by a distance, the ratio cam disc and the input disc rotate synchronously while the distance separating the ratio cam disc and the input disc remains constant, the input disc and the ratio cam disc rotate asynchronously while the distance is modified, and asynchronous rotation of the ratio cam disc and the input disc is used to vary the distance between the longitudinal axis of the crankpin and the longitudinal axis of the input disc via the ratio varying mechanism.

6. The continuously variable transmission of claim 2, wherein the input disc and ratio cam disc have identical pitch curve gear profiles at their peripheries, and the control mechanism includes two sets of axially connected pairs of intermediate circular gears, wherein the two gears of each pair of intermediate circular gears have different pitch curves, one gear of each pair of intermediate circular gears has an identical pitch curve, the axes of the two sets of axially connected pairs of intermediate circular gears are parallel to the longitudinal axis of the input disc and the longitudinal axis of the ratio cam disc, the spacing being such that one gear of one set is configured to radially engage the input disc, one gear of the other set having an identical pitch curve is configured to radially engage the ratio cam disc, and the other gear of the two pairs of gears having an identical pitch curve is configured to radially engage the other common intermediate circular gear, the common intermediate circular gear being coaxial with the input disc and the ratio cam disc, and wherein the axially connected intermediate circular gear is constrained to move only along a path of constant distance from the longitudinal axis of the input disc, and during this movement the input disc rotates asynchronously with the ratio cam disc, such asynchronous rotation of the ratio cam disc and the input disc serving to vary the distance between the longitudinal axis of the crank pin and the longitudinal axis of the input disc by the ratio varying mechanism.

7. The continuously variable transmission of claim 2, wherein the speed ratio varying mechanism includes: a crank pin, a ratio cam plate, and an input disc, wherein the crank pin is disposed in a radial slot of the input disc and a non-radial slot of the ratio cam plate such that a relative angular velocity between the input disc and the ratio cam plate causes the crank pin to move radially along the radial slot, changing a distance between a longitudinal axis of the input disc and a longitudinal axis of the crank pin.

8. The variable transmission of claim 1, further comprising a plurality of modules, wherein the plurality of modules are oriented such that the functional zone of the at least one driven non-circular gear is in contact with the functional zone of the driving non-circular gear; and the functional region of the at least one driven non-circular gear is in overlapping engagement with the functional region of another driven non-circular gear such that the functional region of the driving non-circular gear is always in contact with the functional region of the at least one driven non-circular gear during full rotation of the input disc between the sequentially consecutively engaged driven non-circular gears.

9. The continuously variable transmission of claim 8, wherein the amount of overlapping mesh between the functional region of the driving non-circular gear and the continuously meshing driven non-circular gear is the same.

10. The continuously variable transmission of claim 1, wherein said cross-rack assembly further comprises a dummy rack, wherein said dummy rack is disposed adjacent to said cross-rack assembly with the same mass as said cross-rack assembly, said dummy rack moving in an opposite direction of said cross-rack assembly to counteract vibrations caused by imbalance due to reciprocation of said cross-rack assembly.

11. The continuously variable transmission of claim 1, wherein the input disc includes a second radial slot opposite the radial slot in a circumferential direction about a longitudinal axis of the input disc; the speed ratio cam plate includes a second non-radial slot opposite the non-radial slot in a circumferential direction about a longitudinal axis of the speed ratio cam plate; and a dummy crank pin having a weight equal to that of the crank pin and sliding in a direction opposite to the crank pin along the second radial groove of the input disc and the second non-radial groove of the speed ratio cam disc to cancel out vibration caused by imbalance due to eccentric rotation of the crank pin.

12. The variable transmission of claim 1, wherein the cross-rack assembly further comprises a second groove perpendicular to the first groove for receiving the input shaft.

13. The variable transmission of claim 1, wherein the functional region of the at least one driven non-circular gear meshes with the functional region of the at least one driving non-circular gear to move the crossed rack assembly at a constant speed, and the non-functional region of the at least one driven non-circular gear meshes with the non-functional region of the at least one driving non-circular gear to decelerate the crossed rack assembly to a stop and accelerate in the opposite direction to a constant speed.

14. The variable transmission of claim 1, wherein the transmission further comprises a plurality of power connecting shafts connecting the output of each output gear or output sprocket to the next.

15. The variable transmission of claim 1, wherein the cross-rack assembly further comprises at least one telescoping guide sleeve that guides the cross-rack assembly to move in only a single dimension in a slot of the frame, thus allowing for a reduction in frame size.

16. The continuously variable transmission of claim 1, wherein a slider guide having a rectangular slot longer than a width of the slot of the input shaft is placed in the slot of the crank pin of the cross-rack assembly to eliminate the sliding of the crank pin into the slot of the input shaft.

17. The continuously variable transmission of claim 1, wherein the pinion shaft is further coupled with an assembly comprising an input helical gear, a plurality of coaxial output helical bevel gears having through holes in centers positioned opposite to each other such that they rotate in opposite directions to each other, and a through shaft positioned coaxially with the output helical bevel gears.

18. The variable transmission of claim 1, wherein when a collar is connected to both output bevel gears, the collar is rotationally constrained and acts as a "park" gear.

19. A continuously variable transmission as recited in claim 14, wherein a gear or sprocket placed coaxially with the input shaft transmits power from the power connecting shaft to an output member.

20. The variable transmission of claim 19, wherein the power of the output member is connected to either a ring gear, a carrier, or a sun gear of a planetary gear system; the input shaft is connected to one of the remaining two elements of the planetary gear system, and the final output is connected to the third remaining element.

21. The variable transmission of claim 20, wherein the final output of the planetary gear system temporarily stores energy in a flywheel system, then transfers energy back to a wheel and/or directly to a wheel.

22. The variable transmission of claim 1, wherein an auxiliary input shaft having a non-circular cross-section is mated with a sliding collar having a mating bore, the sliding collar being coaxially disposed with the auxiliary input shaft, permitting motion with respect to each other while constraining rotational motion; a thrust bearing, said thrust bearing being coaxially disposed in contact with one end of said ferrule; a pivot on the other end of the ferrule; one end of a connecting rod is connected to the pivot and the other end of the connecting rod is connected to either the crankpin or the crankpin shaft sleeve as the case may be, wherein axial displacement of the collar produces radial displacement of the crankpin by the connecting rod.