WO1992018791A1 - Transmission variable en continu et a regeneration de couple - Google Patents

Transmission variable en continu et a regeneration de couple Download PDF

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Publication number
WO1992018791A1
WO1992018791A1 PCT/US1991/002690 US9102690W WO9218791A1 WO 1992018791 A1 WO1992018791 A1 WO 1992018791A1 US 9102690 W US9102690 W US 9102690W WO 9218791 A1 WO9218791 A1 WO 9218791A1
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WO
WIPO (PCT)
Prior art keywords
pulley
gear
belt
speed
shaft
Prior art date
Application number
PCT/US1991/002690
Other languages
English (en)
Inventor
Emerson L. Kumm
Original Assignee
Kumm Industries, Inc.
Priority date (The priority date is an assumption and is not a legal conclusion. Google has not performed a legal analysis and makes no representation as to the accuracy of the date listed.)
Filing date
Publication date
Application filed by Kumm Industries, Inc. filed Critical Kumm Industries, Inc.
Priority to PCT/US1991/002690 priority Critical patent/WO1992018791A1/fr
Publication of WO1992018791A1 publication Critical patent/WO1992018791A1/fr

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Classifications

    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F16—ENGINEERING ELEMENTS AND UNITS; GENERAL MEASURES FOR PRODUCING AND MAINTAINING EFFECTIVE FUNCTIONING OF MACHINES OR INSTALLATIONS; THERMAL INSULATION IN GENERAL
    • F16H—GEARING
    • F16H37/00—Combinations of mechanical gearings, not provided for in groups F16H1/00 - F16H35/00
    • F16H37/02—Combinations of mechanical gearings, not provided for in groups F16H1/00 - F16H35/00 comprising essentially only toothed or friction gearings
    • F16H37/021—Combinations of mechanical gearings, not provided for in groups F16H1/00 - F16H35/00 comprising essentially only toothed or friction gearings toothed gearing combined with continuous variable friction gearing
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F16—ENGINEERING ELEMENTS AND UNITS; GENERAL MEASURES FOR PRODUCING AND MAINTAINING EFFECTIVE FUNCTIONING OF MACHINES OR INSTALLATIONS; THERMAL INSULATION IN GENERAL
    • F16H—GEARING
    • F16H61/00—Control functions within control units of change-speed- or reversing-gearings for conveying rotary motion ; Control of exclusively fluid gearing, friction gearing, gearings with endless flexible members or other particular types of gearing
    • F16H61/66—Control functions within control units of change-speed- or reversing-gearings for conveying rotary motion ; Control of exclusively fluid gearing, friction gearing, gearings with endless flexible members or other particular types of gearing specially adapted for continuously variable gearings
    • F16H61/662—Control functions within control units of change-speed- or reversing-gearings for conveying rotary motion ; Control of exclusively fluid gearing, friction gearing, gearings with endless flexible members or other particular types of gearing specially adapted for continuously variable gearings with endless flexible members
    • F16H61/66254—Control functions within control units of change-speed- or reversing-gearings for conveying rotary motion ; Control of exclusively fluid gearing, friction gearing, gearings with endless flexible members or other particular types of gearing specially adapted for continuously variable gearings with endless flexible members controlling of shifting being influenced by a signal derived from the engine and the main coupling
    • F16H61/66259—Control functions within control units of change-speed- or reversing-gearings for conveying rotary motion ; Control of exclusively fluid gearing, friction gearing, gearings with endless flexible members or other particular types of gearing specially adapted for continuously variable gearings with endless flexible members controlling of shifting being influenced by a signal derived from the engine and the main coupling using electrical or electronical sensing or control means

Definitions

  • the present invention relates to variable speed transmissions and particularly to an improved transmission which employs a known type of continuously variable transmission in combination with a shifting gearbox to provide increased efficiency and power handling capabilities over a full speed range.
  • CVTs continuously variable transmissions
  • the hydrostatic CVT such as those manufactured by the Eaton Corporation, designated as Eaton Hydrostatic Transmissions Model 33 through Model 76, consist of a variable displacement hydraulic pump driving a fixed displacement hydraulic motor. These are commercially available in capacities up to 200 HP for use on heavy construction equipment and provide full load operating efficiencies of over 85%. The speed ratio is continuously variable from full forward to full reverse. The generally large size requirements, high noise, low efficiency and cost make this type unsuitable for automotive applications.
  • V-belt type of CVT such as one developed in Holland by van Doorne Transmissie, which has been utilized in Fiat, Renault, Subaru and Ford automobiles, is similar to a conventional V-belt drive except that the relatively complex, multi-segmented belt is made wide so that the faces of the pulleys on which it rides can be moved together or apart. Moving the faces of one pulley together and the faces of the other pulley apart causes the belt to ride at a larger radius on the first pulley and at a smaller radius on the second pulley thus causing the speed ratio of the first and second pulleys to increase.
  • This type of transmission is capable of operating at speed ratios in excess of 4:1, but it has no reverse capability.
  • the traction type of transmission typically consists of hardened steel rollers operating against a pair of toroidal discs. An extremely high contact force allows the rollers to transmit considerable power without slippage. Tilting the rollers changes the drive ratio between the discs.
  • the device is capable of an efficiency of over 98% at full forward and 80% in full reverse with power ratings of several hundred horsepower being possible.
  • the high cost of the required high strength materials limits its applicability.
  • continuous operation of the traction type of transmission at a constant speed ratio often leads to wear of the toroidal discs and subsequent control difficulties.
  • the flat belt type of transmission such as described in United States Letters Patent No. 4,295,836 by Emerson L. Kumm, has various advantages relative to the other variable speed transmissions as described above including higher efficiencies over a wider speed ratio range and a more compact size for a given power with a relatively low cost belt.
  • it does not have a reverse capability, and, like V-belt CVTs, its minimum output speed is above zero, increasing with increases in the input speed.
  • One known example of an attempt to improve the use of a CVT by combining it with a gear mechanism is found in United States Letters Patent No. 3,527,119.
  • This device utilizes a unidirectional variable speed motor which can be selectively coupled to one of two plural paths to control the relative speeds.
  • variable speed drive is non-reversible and the device has only limited use for automotive applications.
  • One known example of measuring torques is shown in United States Letters Patent No. 3,253,658.
  • the torque of a shaft is obtained rather than the pulley belt slip which is used to measure torque in the subject transmission.
  • the belt slip while giving a measure of the pulley or shaft torque, gives a more critical and vital measurement of the pulley operation; i.e., the belt efficiency and durability.
  • Objects of the Invention It is therefore a broad object of the present invention to provide an CVT which is more efficient and has greater power capabilities for automotive applications than prior art CVTs. It is another object of this invention to permit use of a CVT over a full speed range, including reverse, with reduced losses and increased capacity.
  • a principal object of this invention is to provide CVT control that optimizes the durability and efficiency of the belt used with the pulleys. It is a further object of this invention to combine a CVT with a shift mechanism to obtain two speed modes, to extend the output speed range and to increase the output torque capability at low speed. It is another object of this invention to utilize the CVT with its control to provide a very wide range of dynamic braking for a vehicle.
  • a still further object of the invention is to insure that all power shift transfer operations occur between shafts rotating at the same speeds.
  • a further object of this invention is to obtain reversal of the output speed without shifting gears or clutches or interrupting the power flow.
  • Another object of the invention is to provide a CVT having the capability of operating at a condition resulting in zero output speed at all possible input speeds.
  • a still further object is to give a control that can optimize the acceleration of a vehicle.
  • An additional object is to provide a CVT system that can operate a vehicle at the minimum fuel consumption for any required power and output speed.
  • an object of this invention is to provide a geared CVT of reduced size and cost.
  • a continuously variable transmission typically, the Kumm flat belt CVT
  • a normal planetary mechanism using measurements and calculations of the belt slip in the rotating pulleys to adjust the operating pulley speed ratio and control the input and output speed.
  • the input power is divided into two paths; viz.: a first path through the planetary gearing to the output and a second path directed regeneratively from the planetary gearing through the CVT back to the input shaft.
  • a clutch on the shafting from the planetary ring gear to the output shaft is disengaged and another clutch is engaged which permits the input power to be transmitted directly through the CVT to the output shaft.
  • the clutch operation occurs with the shafts rotating , at the same speed, and no interruption in power flow or torque occurs since the non-operating clutch is always engaged before the other clutch is disengaged.
  • Only the pulley radius ratio control reverses direction during the change from the low speed to the high speed mode or vice versa.
  • Reverse output speeds are made available by changing the radius ratio control direction in the CVT when in the low speed mode without actuating clutches.
  • the power flow is divided into two paths, and the input power operates regeneratively through the CVT to the planetary gearing and back to the input shaft while supplying power through the planetary gearing to the output shaft.
  • the control of the CVT speed ratio which is based on the belt slip, basically measures the torque passing through the pulley and adjusts the pulley speed ratio as desired to control the pulley torque and hence the regenerative torque transfer, while the engine speed is being regulated. As a result, excessive or unwanted regenerative torque transfers at low or zero output speeds are avoided.
  • Fig. 1 is an edge on view of driving and driven pulley assemblies coupled by a flat belt and representative of a class of continuously variable transmissions in which the present invention finds application; o
  • Fig. 2 is a cross sectional view, taken along the lines 2-2 of Fig. 1, of the pulley assembly system illustrated in Fig. 1;
  • Fig. 3 is a fragmentary perspective view, partially broken away, of a pulley assembly particularly illustrating the relationships between inner and outer guideway disk components and belt engaging element components;
  • Fig. 4 is a simplified cross sectional view of a flat belt CVT illustrating the fundamental aspects of the mechanical components of the control system for establishing the angular relationship between the inner and outer guideway disks of each pulley;
  • Fig 5 is a cross sectional view taken along the lines 5 - 5 of Fig. 4 particularly illustrating a planetary gear assembly component of the control system;
  • FIG. 6A, 6B and 6C are illustrations showing the principle of operation of a harmonic drive, certain co - ponents being shown in an exaggerated elliptical shape in order to more clearly demonstrate the principle;
  • FIG. 7 is a simplified cross sectional view of a flat belt CVT illustrating the fundamental aspects of the mechanical components of a variant of the subject control system, employing a harmonic drive, for establishing the angular relationship between the inner and outer guideway disks of each pulley;
  • Fig. 8 is a simplified block diagram of an exemplary automatic version of a control system for a flat belt CVT of the type shown in Figs. 4 and 7;
  • Fig. 9 is a schematic layout of a flat belt CVT with regeneratively geared two speed mode transmission;
  • FIG. 10 is a cross sectional view taken along the lines 10-10 of Fig. 9;
  • Fig. 11 is a plot representative of input torque versus percent of flat belt slip for a typical mid size vehicle;
  • Fig. 12 is a schematic layout of the control arrangement of the transmission of Fig. 9 as applied to an automotive engine;
  • Fig. 13 is a graph which shows a typical vehicle speed versus the pulley speed ratio using the transmission arrangement of Fig. 9 and control arrangement of Fig. 12;
  • Fig. 14 is a graph which shows the typical belt geometry in the pulleys for a mid size vehicle;
  • Fig. 15 is a graph which shows the acceleration torque advantages of the transmission of Fig. 9 relative to an automatic transmission in a mid-size vehicle;
  • FIG. 16 is a partially broken away view illustrating a variant pulley actuator using hydraulic control elements to establish the effective diameter of a flat belt CVT;
  • FIG. 17 is a cross sectional view taken along the lines 17-17 of Fig. 16 showing a variant torque increasing component of the actuator;
  • Fig. 18 is a block diagram for a transmission control module;
  • Fig. 19 is a chart showing typical operational ranges for the pulley radii and speeds;
  • Fig. 20 is a block diagram of a feedback control system;
  • Fig. 21 is a graph comparing a desired engine speed to a vehicle speed ratio;
  • Fig. 22 is a graph which shows the operational vehicle speed range in the high speed mode versus the accelerator or brake position; and
  • Fig. 23 is a block diagram showing the CPU and hydraulic control elements of the pulley actuators.
  • the Kumm flat belt CVT is especially suited for a regenerative transmission in that a relatively small CVT component can be optimally employed with large input torques and give very high output to input torque ratios.
  • the incorporation of a zero shaft velocity change during a shift to use the flat belt CVT components in a direct drive arrangement without regenerative torque transfer extends the output speed range by a major factor and permits high cruise speeds to be obtained at low engine speed in the application to an automobile.
  • Figs. 1, 2 and 3 fundamental aspects of the Kumm flat belt type of continuously variable transmission are illustrated as embodied in a variable diameter pulley drive system 10 including variable diameter pulley assemblies 11 and 12 connected by a flat drive belt 13.
  • the pulley assembly 11 will be considered as the driving pulley assembly and the pulley assembly 12 as the driven pulley assembly in this discussion, but it will be understood that the roles of these pulley assemblies may be reversed without altering the concepts involved.
  • the pulley assembly 11 is appropriately mounted on a shaft 14, and the pulley assembly 12 is similarly appropriately mounted on a shaft 15 as is well understood in the art.
  • the pulley assemblies 11 and 12 are similar to each other, and only one of them, pulley assembly 11, will be specifically described in this discussion.
  • the belt 13 as shown in Fig. 3 corresponds to the position of the belt 13 of Fig. 2 in the dashed line position.
  • the pulley assembly 11 includes a pair of pulley sheaves 16 and 17 between which there extends a series of belt engaging elements 18, the latter being engaged by the belt 13 for driving, or driven, conditions as will be understood.
  • a series of twenty-four belt engaging elements 18 equally circumferentially distributed such that an angle of fifteen degrees is established between runs of the belt 13 coming off tangentially from one belt engaging element 18 as compared to that of an immediately adjacent belt engaging element 18.
  • Each belt engaging element 18 includes a central shank 28, which engages the belt 13, and bearing regions 29 at each end.
  • the pulley sheave 16 incorporates a pair of pulley guideway disks 19 and 21 which are parallel to and lie immediately adjacent each other in juxtaposition.
  • the pulley sheave 17 comprises a pair of pulley guideway disks 22 and 23 which are parallel to and lie immediately adjacent each other in juxtaposition.
  • the longitudinal spacing between the pulley sheaves 16 and 17 i.e., the axial spacing between the inner guideway disks 21 and 22
  • This spacing is sufficient to accommodate with clearance the belt drive element which supports the belt 13 which is selected to carry the load that the system is designed for as is well understood.
  • the range of radial adjustment or position of the belt 13 on the pulley assembly 11, as may be envisioned by the solid line and dashed line positions of belt 13 in Fig. 2, is achieved by altering the radial positions of the belt engaging elements 18.
  • the belt engaging elements 18 are close to the center of the shaft 14 in the solid line position of the belt 13 on pulley assembly 11; conversely, the belt engaging elements are radially farther out, namely adjacent the periphery, when the belt 13 is in its dashed line position which is also the position shown in Fig. 3.
  • Variations in the radial positions of the belt engaging elements 18 is achieved by relative rotation of the outer guideway disk 19 and the inner guideway disk 21 of pulley sheave 16 to change their angular relationship.
  • the guideway disks 23, 22, respectively, of pulley sheave 17 change of angular relationship is accomplished.
  • the outer guideway disk 19 has a series of logarithmic spiral guideways 24 therein which progress outwardly from adjacent the center at an angle of forty- five degrees with respect to the pulley assembly radius.
  • the inner guideway disk 21 has a series of logarithmic spiral guideways 25 radiating outwardly at an angle of forty-five degrees with respect to the pulley assembly radius, but in the opposite sense to the guideways 24 of guideway disk 19.
  • the guideways 24 and 25 radiate outwardly at angles of forty-five degrees with respect to the pulley radius, but in opposite senses, the intersections of these guideways exist at ninety degrees at all radial positions. This results in a substantially constant geometry at the intersections of the logarithmic spiral guideways 24 and 25 at all radial positions for receiving the bearing region ends 29 of the belt engaging elements 18.
  • the inner guideway disk 22 has a series of logarithmic spiral guideways 26 radiating outwardly identically to the guideways 25 of inner guideway disk 21, and the outer guideway disk 23 includes logarithmic spiral guideways 27 extending outwardly identically to the guideways 24 of outer guideway disk 19.
  • the guideways 26 and 27 intersect at ninety degrees at all radial positions to give a constant intersection geometry identical to the logarithmic spiral guideways 24 and 25 for receiving the other ends of the belt engaging elements 18. While forty-five degree spirals have been shown and give ninety degree intersections, it will be understood that logarithmic spirals of other angularities may be used as desired. Also, minor variations from a particular angularity may be tolerated so long as the belt engaging element bearing ends supported at the guideway intersections will move appropriately when the sheaves are rotated relative to each other to change the angular relationship between the inner and outer guideway disks.
  • FIG. 4 and 5 illustrating a first embodiment of a CVT which incorporates a geared system for precisely establishing the speed ratio between a first pulley assembly 48 and a second pulley assembly 49.
  • the geared speed ratio control mechanism associated with the upper pulley assembly 48 and that associated with the lower pulley assembly 49 are essentially identical with minor exceptions to be particularly noted in the following discussion. For that reason, the essential description will be set forth in conjunction with the upper pulley assembly 48 in Fig. 4 which, for purposes of illustration, may be deemed the driving pulley assembly whereas the lower pulley assembly 49 is considered the driven pulley assembly.
  • the geared speed ratio control mechanism which serves to establish the angular relationship between the inner and outer guideway disks (and hence the radial positions of the belt engaging elements 18), is contained in a stationary housing 50.
  • a reduced diameter region of the shaft 14 extends into the housing 50 and terminates in an outer sun gear 54.
  • the set of inner guideway disks 21, 22 are directly connected to an hollow shaft 52 which encompasses the end region of the shaft 14 for mutual rotation therewith, and the hollow shaft 52 terminates in an inner sun gear 53.
  • the sun gears 53, 54 have the same diameter and the same tooth count which is necessary for a stationary control at specific speed ratios; however, this relationship is not a design constraint of the system which could employ continuously rotating control components.
  • the sun gears 53, 54 are identically configured for simplicity.
  • a planet carrier 51 is disposed for rotation within the stationary housing 50 about the axis of the shaft 14 and carries a plurality of planet gears 55, each supported for rotation on an individual planet gear shaft 57 as shown in Figs. 4 and 5.
  • Each of the planet gears 55 meshes, at its radially inner (with respect to the axis of the shaft 14) peripheral region, with the inner sun gear 53 and at its radially outer peripheral region with a first set of teeth 59A of an internal ring gear 59.
  • the internal ring gear 59 also carries a second set of internal teeth 59B axially offset from the teeth meshing with the planet gears 55.
  • the teeth of the second set 59B are typically identical to the teeth of the first set 59A.
  • a plurality of star gears 56 are supported in fixed positions by star gear shafts 58 which are integral with the stationary housing 50.
  • the star gears 56 are situated to mesh with both the second set of internal teeth 59B of the internal ring gear 59 and the teeth of the outer sun gear 54 which is directly connected to the shaft 14.
  • the planet gears 55 and the star gears 56 are shown in Fig. 4 in face to face juxtaposition; however, this is only for simplicity of illustration, and it will become apparent below that the angular positions of the planet gears 55, relative to the star gears 56, about the axis of the shaft 14 are changed during operation of the control system.
  • the planet carrier 51 is provided with an external ring gear 60 with which there is meshed a spur pinion 63.
  • the pinion 63 is manually driven by a speed control crank 61 via a shaft 62.
  • the crank 61 may be constrained against rotation by permitting a crank retaining pin 64 to enter an aperture 65 provided in the stationary housing 50 for this purpose.
  • the lower pulley assembly 49 does not have a speed control crank, but rather is provided with a load device 66 which is connected by a shaft 62' to a spur pinion 63'.
  • the load device 66 may take any form capable of absorbing and transmitting rotational force.
  • the load device 66 may be a hydraulic motor/pump and its corresponding control system as adapted from the oil motor/pump units described in the above-referenced United States Patents No. 4,768,996 and 4,810,234.
  • Another suitable load device is an electrical motor/generator as will be discussed below in conjunction with the description of an alternate embodiment of the flat belt CVT.
  • a still simpler device for the purpose is a spiral torsion spring which generates a torque when wound in one direction and applied to rotate the planet carrier 51' relative to the housing 50', much in the manner of a clock spring.
  • the speed control crank 61 may be rotated, for example, in the clockwise direction as viewed in the direction indicated by the reference arrow 67.
  • the planet carrier 51 will rotate, at a reduced rate (according to the gear ratio between the spur pinion 63 and the external ring gear 60) in the counter clockwise direction. Since the shaft 14 is deemed to be stationary and the housing 50 is immovable, the internal ring gear 59 is held stationary.
  • the speed control crank 61 may again be fixed in place to maintain the new angular relationship between the planet gears 55 and star gears 56 to accordingly maintain the new pulley speed ratio.
  • all this transient adjustment is simply superimposed on the actual rotation of the several components such that the adjustment to the relative angular positions of the set of inner guideway disks 21, 22 to the outer guideway disks 19, 23 actually takes place during pulley rotation and power transfer.
  • the set of outer guideway disks 19', 23' are also stationary. Therefore, the set of inner guideway disks 21', 22' must have moved counter clockwise as viewed from the direction indicated by the reference arrow 68. As the set of inner guideway disks 21', 22' moves counter clockwise, the inner sun gear 53' also rotates counter clockwise. Therefore, the planet gears 55' rotate clockwise. Since the driven shaft 15 is assumed to be stationary, the star gears 56' cannot rotate such that the internal ring gear 59' is also held stationary. Consequently, the action of the planet gears 55' enmeshed with the teeth of the internal ring gear 59' is such as to cause the planet carrier 51' to rotate counter clockwise.
  • the spur pinion 63' is rotated, at a multiplied rate, in the clockwise direction to drive the load device 66 in the clockwise direction via the shaft 62'.
  • This action takes place in synchronism with the adjustment to the relative positions of the planet gears 55 and star gears 56 of the driving pulley assembly 48; this action ceases and a new steady state condition is assumed when the speed control crank 61 is again fixed in position.
  • the load device 66 will have functioned as such. For example, if the load device 66 is an oil motor/pump it will have functioned transiently as a pump. If the load device 66 is a motor/generator, it will have functioned transiently as a generator.
  • the load device is a spring, it will have been wound to some degree. It will be readily apparent that if the speed control crank 61 is manually rotated in the counter clockwise direction, a string of effects opposite to those discussed above will ensue. Briefly, the planet carrier 51 will rotate in the clockwise direction causing the set of inner guideway disks 21, 22 to rotate in the clockwise direction to decrease the effective diameter of the driving pulley assembly 48. The belt engaging elements 18' of the driven pulley assembly 49 will therefore move radially outwardly, and will be assisted by virtue of the load device 66 operating in an energy delivering mode; i.e., as a hydraulic or electric motor or as an unwinding spring rotating in the clockwise direction.
  • an energy delivering mode i.e., as a hydraulic or electric motor or as an unwinding spring rotating in the clockwise direction.
  • a harmonic drive employs three concentric components to produce high mechanical advantage and speed reduction.
  • the use of nonrigid body mechanics allows a continuous elliptical deflection wave to be induced in a nonrigid external gear, thereby providing a continuous rolling mesh with a rigid internal gear.
  • an elliptical wave generator 130 deflects a flexspline 131 which carries outside teeth and therefore meshes with the inside teeth of a rigid circular spline 132.
  • the elliptical shape of the flexspline and the amount of flexspline deflection is shown greatly exaggerated in Figs. 6A, 6B and 6C in order to demonstrate the principle. The actual deflection is very much smaller than shown and is well within the material fatigue limits.
  • the flexspline rotates counterclockwise.
  • Fig. 6B it will be seen that the tooth 134, after one-quarter revolution of the wave generator 130, has moved counter clockwise one-half of one flexspline tooth position.
  • the tooth 134 is fully disengaged. Full reengagement occurs in the adjacent circular spline tooth space when the major axis of the wave generator 130 has rotated 180° as shown in Fig. 6C, and the tooth 134 has now advanced one full tooth position.
  • the harmonic drive principle can be extended by the addition of a fourth element designated the dynamic spline.
  • the dynamic spline is an internal gear that rotates at the same speed and in the same direction as the flexspline.
  • the dynamic spline has the same number of teeth as the flexspline. Flexspline shape rotation results in tooth engagement/disengagement within the same tooth space of the dynamic spline such that the ratio between the two is one to one.
  • the system therefore, is a flexspline output with the same characteristics as the three element harmonic drive model; i.e., gear reduction ratio tabulated with the direction of rotation opposite to the input. Ultra high dual ratio capability can be obtained by using two circular splines in mesh with the flexspline with each developing a different single-stage ratio.
  • Harmonic drives suitable for use in a flat belt CVT may be obtained from the Harmonic Drive Division of Quincy Technologies, Inc. in Wakefield, Mass. consider now, with reference to Fig. 7, the manner in which a harmonic drive can be advantageously incorporated into a CVT.
  • a harmonic drive component 70 is interposed between the set of inner guideway disks 21, 22 (and the set of outer guideway disks 19, 23) of the driving pulley assembly 78 and the inner sun gear 53 via a hollow shaft 71.
  • another harmonic drive component 70' is interposed between the set of inner guideway disks 21', 22' (and the set of outer guideway disks 19', 23') of the driven pulley assembly 79 and the inner sun gear 53' via a hollow shaft 71'.
  • the driving shaft 14 extends through the hollow shaft 71 and carries the outer sun gear 54 at its end while the driven shaft 15 extend through the hollow shaft 71' and carries the outer sun gear 54' at its end.
  • the remaining structure of the planetary gear assembly is as described in conjunction with Figs. 4 and 5.
  • the set of outer guideway disks 19, 23 are connected to the dynamic spline 72 of the harmonic drive 70, and the set of inner guideway disks 21, 22 are connected to the circular spline 73 of the harmonic drive.
  • the wave generator 74 is directly connected to the hollow shaft 71 and is hence directly coupled to the inner sun gear 53.
  • the sets of inner guideway disks 21, 22 and outer guideway disks 19, 23 are rotating at the same rate. Consequently, there is no relative movement between the dynamic spline 72 and circular spline 73 such that the wave generator 74, the hollow shaft 71 and inner sun gear 53are all rotating at the same rate as the shaft 14 and the outer sun gear 54.
  • the speed control crank 61 may be released and turned as previously described.
  • the driving shaft 14 is stationary to best appreciate the transient movements among the several components of the control mechanism. If the speed control crank 61 is rotated in the clockwise direction, the planet carrier 51 rotates in the counter clockwise direction to also drive the inner sun gear 53 in the counter clockwise direction. Therefore, the hollow shaft 71 and the wave generator 74 rotate counter clockwise. This rotation is communicated through the flexspline 75 to both the dynamic spline 72 and the circular spline 73.
  • the belt engaging elements 18' of the driven pulley assembly 79 will move radially outwardly, and energy will be delivered to the load device 76.
  • the speed control crank 61 and load device 76 will rotate many more turns (equivalent to the gear ratio of the harmonic drive) during the transient speed ratio adjustment operation.
  • the speed control torque and power required in the transient speed ratio adjustment operation are reduced by the gear ratio of the harmonic drive (not considering efficiency losses) . Since the harmonic drive gear ratio is typically of the order of 100;1 or more, the reduction in the control torque and power are very large indeed.
  • harmonic drives 70, 70' into the geared control system serves to greatly magnify the precision of control over the speed ratio between the pulley assemblies of a flat belt CVT.
  • torque requirement for the device driving the pinion 63 is accordingly reduced.
  • This feature permits the use of, for example, a small reversible electrical motor as the device connected to the pinion 63 and this, in turn, permits the use of relatively simple external control structure such as that illustrated in Fig. 8 in an exemplary vehicular environment.
  • normal engine speed and accelerator position information may be analyzed by conventional logic 80 to determine if a change in speed ratio is desirable and, if so, to energize the reversible d-c motors 81, 82 to drive the mechanisms previously described in the appropriate directions until the sought ratio is obtained.
  • this is an ongoing process to obtain the desired engine torque-speed characteristic giving minimum fuel consumption and, as those skilled in the art will understand, other process variables (coolant temperature, atmospheric pressure, "performance" setting, brake actuation, vehicle speed, etc.) are involved in determining the desired instantaneous speed ratio.
  • a flat belt CVT employing the principles described above may be incorporated into a regenerative transmission according to the present invention.
  • the subject invention incorporates a control arrangement for the CVT components which measures and limits the maximum belt slip operating with two pulleys.
  • the maximum pulley torque is specifically related to the belt slip for any specific pulley radius ratio and pulley geometry (pulley radii, pulley center distance, and belt factors) .
  • the belt slip measurement provides a measurement of torque that is used in the control.
  • Other devices could be used for measurements of pulley torque and belt tension to obtain control information for the subject transmission (see, e.g.. United States Letters Patent No. 3,253,658), but belt slip is the critical parameter and would be calculated or inferred in any case.
  • the speed ratio of the two pulleys is uniquely given when operating at zero belt slip.
  • the belt slippage can be determined by measuring the pulley speed ratio during operation and comparing this value to the zero slip speed ratio for specific pulley radii.
  • Second order effects, such as belt wear may be corrected using periodic calibration. Such a calibration or control modification can be accomplished at idle or startup of operation where zero vehicle velocity and low output torque (not sufficient to move the vehicle) occur at a slightly changed pulley speed ratio (as compared to the initial control calibration) .To implement such a control, it is only necessary to measure accurately the two pulley speeds and the effective radii of the belt in the two pulleys.
  • Such measurements can be utilized by electronic solid state equipment of high reliability to give the proper output control signals to limit the maximum belt slippage or pulley torque at the low to zero transmission output speed where large regenerative torque transfers would otherwise occur with any significant input torque.
  • the same basic control would be utilized by the driver of an automobile using this system to obtain an increase or decrease in output torque to change the vehicle speed by using an accelerator to change the engine speed and the speed ratio for the CVT using control techniques well known to designers of CVTs.
  • the geared actuator configurations of Figs. 4 and 7 have been shown to permit a simple, precise, low cost approach to the measurement of the belt radial position in a pulley, many other techniques also exist to obtain such information.
  • the subject invention is primarily concerned with the use of such information to determine the belt slippage during operation and to operate the pulley actuators to give the necessary pulley speed ratio and belt tension that permits the desired torque to be transmitted with belt slippage that optimizes the overall operation and gives reasonable belt life.
  • the percent slip of the flat belt on a rotating pulley may be calculated in the following fashion. It is well known that, at zero slip, the velocity of a conventional flat belt composed of high strength cords in a rubber matrix is larger than the pulley surface that it contacts because substantially all of the belt tension is supported by the cords which are located in the rubber matrix at a radius larger than the pulley surface.
  • VB 2Pi (R+a)N ft/min
  • R the radius of the pulley surface in ft.
  • a the distance from the pulley surface to belt mean load bearing location (typically, to the center of a single line of cords)
  • N the speed of the pulley in RPM.
  • a 3.14159
  • the quantity “a” of the flat belt will decrease during the life of the belt due to wear.
  • a periodic update of the value of "a” for the control can be accomplished for an automotive drive at startup or idle operation where initial low torque assures an essentially zero slip condition.
  • the engine input via the shaft 300 normally supplies torque or power except when the engine is used to slow down the vehicle.
  • This engine torque added to torque from pulley B is transmitted to the planetary carrier, C, in the low speed mode via the E-I and I-C gear meshes, gear I being an idler employed to obtain the same directions of rotation of components E and C.
  • gear I being an idler employed to obtain the same directions of rotation of components E and C.
  • the planetary carrier C drives the internal ring gear R, via the planetary gears P, in the planetary gear set to give torque via the low speed gear L to the output shaft 301 and its output gear 0 driving through the low output shaft gear LO and its clutch LC.
  • the planet carrier also simultaneously drives the sun gear S and hence pulley A which is located on the sun gear shaft 302.
  • the high speed clutch HC is open so that no power passes through the high speed gear H to the high output shaft gear HO. Torque and power from the sun gear S to pulley A is then transmitted via a flat belt 303 to pulley B and through shaft 300 to be added to the engine input power to give torque on the carrier C.
  • the power transmitted through the CVT acts in an input regenerative fashion increasing the torque and power transmitted through the planetary and pulley system relative to the engine power.
  • a significant engine input power can, through regenerative torque transfer, result in excessive pulley and shaft torques which has previously substantially limited the use of such regenerative transmission arrangements.
  • the subject transmission system measures a small practical belt slip to sense pulley torque at belt tensions suitable for the employed belt to continuously adjust the pulley speed ratio (using actuators A and B) while regulating engine speed to prevent excessive torques from occurring. Excessive regenerative torque transfer can only occur at low vehicle speeds, but is dependent on the magnitude of engine torque.
  • the output torque is always proportional to the pulley torques and gear torques at any operating point, differing only due to small gear and shaft bearing efficiency losses.
  • the input torque to the planet carrier C is variable dependent on the amount of regenerative torque buildup.
  • an output torque necessary to give a specific output speed and/or acceleration- deceleration i.e., for a vehicle
  • the previous problem relating to the possibility of excessive torques at zero or low output speeds is solved in this system by measuring belt slip in the pulleys as previously described to obtain a value which is directly related to the torques through the system and responsively using an actuator to change the pulley radius ratio or speed ratio, i.e., to program an output speed that requires less torque while the engine speed is being simultaneously regulated.
  • the pulley radius ratio is set accordingly to give as low a pulley torque as desired for the automotive application for any specific engine speed.
  • Fig. 11 shows the slip characteristics typical of a specific belt and pulley arrangement as determined from tests. The percent slip depends on the belt tension, and the input torque can be increased with increases in belt tension up to the maximum desired belt tension.
  • the maximum operating belt tension and maximum percent slip depend on many parameters - operation duration, belt ultimate strength and its change with operation, specific belt-pulley characteristics such as the friction coefficient and its variation with operation, etc. Typically, the maximum belt tension is limited to about 10 - 15% of the initial ultimate belt strength with a maximum belt slip of 1 - 1.5%.
  • the control schematic as shown in Fig. 12 would use the belt slip characteristics of Fig. 11 as applied to the transmission arrangement of Fig. 9.
  • the engine speed could be programmed to change with vehicle speed; i.e., low engine speed at zero vehicle speed increasing as the vehicle speed increases to give the desired torque and power for the vehicle acceleration characteristic as set by the accelerator up to the shift point going into the high speed mode.
  • a large depression of the accelerator calling for a large vehicle acceleration would set a high engine speed to vehicle speed ratio which would then result in an acceleration in the low speed mode up to a higher vehicle and engine speed at the shift point for the high speed mode as compared to the shift point speeds resulting from a smaller depression of the accelerator.
  • the resulting vehicle speed versus engine speed for a typical transmission is shown in Fig. 13 which will be discussed subsequently.
  • a transmission control module 310 is used to regulate the engine speed through the engine control module 311 and operate the actuators 312, 313 on pulleys B and A respectively, to give the desired pulley speed ratio and pulley torque.
  • Inputs to the transmission control module include the operation mode, the accelerator position, the engine speed or pulley B speed, pulley A speed, and measurements of the belt radial positions in pulleys A and B.
  • the speed and belt radial position inputs permit the belt slip to be directly measured in the transmission control module.
  • the outputs from the transmission control module 310 include a clutch control corresponding to the operation mode and the vehicle to engine speed ratio, a control that positions the belt radial position in pulley B, another output that sets the belt tension as desired for the regulated belt slip in pulley A as shown by the operating characteristic of Fig. 11, and an output to the engine control module 311 that regulates the engine speed.
  • Various specific control logics can be employed to regulate and control the belt slip to the desired levels.
  • One control logic consists in part of measuring the belt slip as previously discussed and comparing it to the desired belt slip given as a function of the accelerator position, engine speed and vehicle speed to determine a belt slip error. This error measurement is then used to operate the actuators on pulley A and pulley B to change belt tension and/or pulley radius ratio. However, change in the pulley speed ratio will also change the engine speed for a specific vehicle speed. Hence, the engine fuel control will change the fuel flow to maintain the scheduled engine speed, thus modifying the torque supplied by the engine to the transmission. A change in engine torque changes the pulley torque reducing the belt slip error as desired. Another control logic could use the belt slip error to directly operate the engine fuel control to change engine torque to reduce this error in one control loop.
  • the available pulley speed ratio permits vehicle operation from reverse to forward speeds in the low speed mode (power passing through the planetary gearing) and, after the shift, operation at forward speeds in a high speed mode (no power passing through the planetary gearing) .
  • the engine speed may be varied at zero vehicle speed with an engaged low speed mode clutch and that the maximum reverse speed as well as the forward speed at the shift point (low to high speed mode) and maximum forward speed all increase with increasing engine speed. Due chiefly to the regenerative gearing arrangement, the engine can easily give the maximum design vehicle acceleration at low vehicle speeds with low engine speeds.
  • the engine speed control can operate to call for an engine speed that is initially low at low vehicle speed but increasing with increasing vehicle speed in the maximum vehicle acceleration condition.
  • the path typical of such a maximum acceleration is shown in Fig. 13 and it is noted that the maximum engine speed of 4500 RPM may be obtained only on reaching a vehicle speed of 20 mph.
  • the engine transient speed acceleration can be reduced from the conventional automotive transmission characteristic where the engine speed is not dependent on the vehicle speed.
  • a significant advantage is obtained in that the actuation speed requirement of the variable speed pulleys is reduced with this transmission arrangement.
  • the speed ratio, NA/NB or RB/RA requires the vehicle to accelerate from zero to 36 mph while changing from 1.8 to 0.60, This will normally take several seconds for a conventional mid size vehicle.
  • the engine speed may be reduced to a minimum value corresponding to a pulley speed ratio as given in this design of 2.45.
  • the possible engine acceleration rate is again limited by the vehicle speed to result in reasonably low pulley actuator rates.
  • Lower actuator rates are desirable to reduce the actuator power requirement.
  • the maximum vehicle acceleration is still available at maximum engine torque in the high speed mode which normally occurs at engine speeds substantially lower that the maximum engine speed. It is possible to obtain a very rapid pulley radius change in the high speed mode at a given vehicle speed by using a reversal in the actuator control pressure employing a directional control valve. This would give a very rapid increase in engine speed suitable for a high speed passing maneuver.
  • the transmission operation must encompass many conditions. Consider operation with the transmission in the forward drive position, but stopped going up a hill.
  • the accelerator may be depressed slightly calling for increased output torque, thus providing adequate engine speed and torque to prevent the vehicle from moving backward.
  • the regenerative torque buildup is limited partly by the low engine torque and transmission losses but also by components which would measure belt slip at known belt tensions to give the desired pulley torque by adjusting the pulley radius ratio.
  • the actuator on pulley B had called for too high an output speed, the regenerative torque transfer will give pulley torques resulting in excessive belt slip. But if the belt slip is being monitored, the actuator on pulley A can then be controlled to give the maximum permissible belt slip giving the maximum output torque possible from the engine and allowable without excessive belt slip throughout the critical low speed range.
  • the output speed varies from zero to a shift speed point while the speed ratio in the pulleys varies correspondingly from a low value of NB/NA or RA/RB to the maximum value of NB/NA or RA/RB.
  • the maximum pulley speed ratio NB/NA and gears may all be chosen so that, at the shift point, the complete planetary gear set operates at the same speed; i.e., the sun gear S speed equals the carrier gear C speed equals the ring gear R speed. Hence, at the shift point, the high speed gear H will be rotating at the same speed as the low speed gear L. If the high speed gear H and low speed gear, L, were chosen to have different sizes, the shift point giving the same output speed is changed. While this may be useful in some applications, the maximum pulley radius ratio change is adequately utilized using the H and L gears of the same size.
  • the high speed mode clutch HC is engaged before the low speed mode clutch LC is disengaged so that the drive can continuously accelerate the vehicle without hesitation.
  • Operation in the high speed mode occurs with torque transfer occurring from pulley B to pulley A rather than from pulley A to pulley B as in the low speed mode.
  • the planetary gears do not transfer any torque to the output shaft in the high speed mode.
  • the high speed mode output wheel-to-pulley torque ratio at the shift point equals the low speed mode output wheel-to-pulley torque ratio.
  • Pulley B is designed to be larger in size than pulley A as shown in Fig. 14 since the maximum torque required to be taken from the engine occurs in the high speed mode just after the shift.
  • the relatively large minimum diameter of pulley B permits very large input torques to be used with comparatively small overall pulley geometry, a major advantage of this arrangement.
  • the pulley minimum torque capacity varies with the square of the diameter of the driving pulley for a belt tension that is a constant percent of the belt breaking strength.
  • the typical belt geometry as shown in Fig. 14 was chosen for operation with an engine giving a maximum output torque of 200 Newton-meters at 2700 RPM.
  • the belt radius in pulley B varies from 53 mm to 98 mm while the pulley A belt radius is changing from 88 mm to 40 mm.
  • a flat belt of 50 mm width can transmit 200 Newton-meters at a pulley radius of 53 mm with slippage of less than 1.3% at 2700 RPM in the Kumm flat belt CVT.
  • An unusually compact configuration is available using a pulley center distance of 203 mm giving basic advantages in size and cost of this system.
  • the overall scheme permits a vehicle to achieve maximum speeds considerably above 100 MPH yet operate when desired at comparatively low engine speeds for high vehicle cruise speeds as shown in Fig. 13.
  • the actuator control for the pulleys in the high speed mode is somewhat reversed from the control in the low speed mode.
  • the belt drive radius of pulley A is increased and the radius of pulley B is decreased to increase the output vehicle speed.
  • the belt drive radius of pulley A is decreased and the radius of pulley B is increased to increase the output vehicle speed. Consequently, the actuator control direction must be shifted simultaneously with the clutch operation. However, the belt tension as applied in one pulley by the actuator would continue tc operate in the same direction. Reverse speed can be obtained from this transmission starting from zero output speed when operating in the low speed mode without operating the usual transmission clutches or disengaging the drive from the wheels.
  • the power passes through the planetary gears and the low (L) to low output (LO) gear mesh with the low speed mode clutch (LC) engaged.
  • the actuator control is identical to that used in the high speed mode.
  • the pulley geometry permits the belt drive radius of pulley A to be decreased and the radius of pulley B to be increased to increase the speed in the reverse output speed direction.
  • Regenerative torque transfer requires monitoring of the belt slip and adjustment by the actuator to change the pulley B radius to limit the amount of belt slip for any regulated input engine speed.
  • the output torque is related directly to the belt slip that is permitted. It is possible to achieve only relatively modest, though reasonable, reverse speeds since it is desirable to use most of the available pulley speed ratio range in the forward low speed mode. However, design reverse speeds of 9-10 MPH appear easily achievable at engine speeds of 2500 RPM as shown on Fig. 13.
  • the output wheel torque is limited by the maximum pulley torque for the slip that can be tolerated. At low reverse speeds, very large output torques are available at the wheels with maximum pulley belt slip, due in part to the planetary gearing arrangement.
  • the electronic transmission control also permits using the engine for dynamic braking by compressing engine air at all vehicle speeds. While conventional hydraulic braking would be employed as a backup, the control system would permit obtaining high engine speeds for low vehicle speeds in the low speed mode that would give substantial engine dynamic braking.
  • the overall acceleration performance of the transmission is compared in Figure 15 to a conventional automatic transmission which employs a three speed transmission with overdrive and torque converter. The engine speed is assumed to be limited to 4500 RPM. The design as given herein shows a significant acceleration advantage over the entire operating range.
  • the drive with the transmission can be completely disconnected from the vehicle wheels by simply disengaging both the high and low speed mode clutches shown in Fig. 9 to give a neutral mode of operation called for in Fig. 12.
  • the "parking" mode of operation could then consist of locking the output shaft with disengaged high and low speed mode clutches.
  • the parking mode can give the desired minimum torque for "updating" the calibration of the slip measurement when starting the vehicle.
  • An attractive pulley actuator arrangement using hydraulics alternate to that shown in Figs. 4 and 7 is shown in Figs. 16 and 17. These Figs, show an actuator similar to those described in conjunction with Figs. 4 and 7.
  • the differential gearing arrangement employed is a planetary geared assembly in place of a "harmonic gear" assembly.
  • the inner guideway discs 320 are connected to a ring gear 321 of the differential planetary gear assembly 322 .
  • the outer guideway discs 323 are fastened to a circumferential ring 324 which acts as the planet carrier of the differential planetary gear assembly.
  • the circumferential ring supports shafts 325 on which the planet gears 326 are mounted with their bearings.
  • the planet gears connect with a sun gear 327 on an annular shaft 328 to another sun gear 329 in the actuator gear housing.
  • the differential planetary gear assembly thus operates in the same basic fashion as the "harmonic drive" arrangement previously described, but with a much smaller gear ratio.
  • the harmonic drive may have a gear ratio of 100 to one or more whereas the planetary gear ratio shown here is approximately 8.5 to one.
  • any movement of the sun gear on the annular shaft relative to the pulley shaft connected to the outer guideway disc will rotate the inner guideway disc with respect to the outer guideway disc in the differential motion well known to designers of planetary gear systems.
  • the inner guideway disc 320 will rotate at a speed of 1/8.5 or 0.1176 times the speed of the sun gear 327 on the annular shaft 328 relative to the speed of the pulley shaft connected to the outer guideway disc 323.
  • the direction of rotation of the inner guideway disc 320 relative to the outer guideway disc 323 is opposite to the direction of rotation of the sun gear 327 on the annular shaft 328 relative to the pulley shaft 300.
  • the actuator gear housing previously described consists of two sun gears normally of the same size with star and planet gears of equal size with a ring gear.
  • the rotational positions of the star gears are stationary due to their shafts being fixed to the actuator gear box housing. If the planet carrier in the actuator gear box is held stationary, there can be no movement of the inner guideway discs relative to the outer guideway discs and the belt radial position is stationary in the pulley.
  • the planet carrier to the actuator gear box may be rotated in either direction by the hydraulic power unit 331 operating through the gear reduction assembly. Any positional change of the planetary carrier results in a precise angular change in the position of the inner guideway disc 320 relative to the outer guideway disc 323, irrespective of the pulley operating speed.
  • the angular direction change of the inner guideway disc relative to the outer guideway disc is dependent on the direction of rotation of the hydraulic unit 331.
  • the angular position of the inner guideway disc relative to the outer guideway disc specifies the belt radial position with a given guideway geometry.
  • measurement of the rotation of any of the gears in the gear train from the hydraulic power unit to and including the planetary carrier can be used to specify precisely the belt radial position in the pulley.
  • This information with input and output pulley speeds can now be easily used to determine the belt slip with substantial accuracy at any operating point as previously described.
  • the hydraulic control elements shown in Fig. 16 were chosen to minimize the control operating power requirement.
  • the pressure control valve 332 is operated to maintain only that hydraulic pressure (sensed by sensor 336) necessary for the actuator torques required for a specific output torques. Little if any flow passes through the hydraulic power unit 331 for any pulley operation at a constant pulley speed ratio where no rotation of the hydraulic power unit is required.
  • the hydraulic power unit and oil pump are sized for the maximum movement speed required in the pulley system to change speed ratio. This maximum movement speed normally occurs in the automotive application during the maximum vehicle acceleration.
  • the regenerative geared layout as shown in Figure 9 together with the control schematic of Figure 12 permits maximum vehicle output torque to be achieved changing the pulley speed ratio much more slowly than possible in previous continuously variable speed drives.
  • a two position directional control valve 334 is used in the pulley B actuator to permit reversing the pressure differential on the hydraulic unit to obtain very rapid engine accelerations at a given vehicle speed for high speed passing.
  • the hydraulic control arrangement is shown with only one of the two pulleys in the flat belt CVT arrangement. A similar unit is used on the other pulley.
  • the letter A after a reference numeral refers to pulley A and the letter B after a reference numeral refers to pulley B.
  • the signal or control on the pressure control valve 332 is normally used at one pulley to set the pulley belt drive radius as desired while the signal or control on the pressure control valve at the other pulley operates to give the desired belt tension and belt slip for the operating torque.
  • Other arrangements including electric motors, etc. could also be used in the control of the pulley speed ratio and belt slip. While the preceeding description has given the control functions for the proposed operation of the continuously variable regenerative transmission, a specific description of a typical control and its elements corresponding to the above control functions is desirable. Accordingly, using an engine or Pulley B speed pickup 306B and Pulley A speed pickup 306A as shown on Fig.
  • the transmission control module 310 is arranged to operate one pulley to have specific operating belt radius and the other pulley to tension the belt to operate at a specific percent belt slip.
  • the accelerator position input sets both the desired engine speed as a function of the vehicle speed, the specific relationship changing from the low speed mode to the high speed mode, as well as the desired percent belt slip, both functions increasing with increased accelerator position.
  • the transmission clutches as shown in Fig.
  • FIG. 9 An electronic control diagram for the transmission control module 310 of Fig. 12 is given in Fig. 18 which shows that it functions as a digital computer controller and has a central processing unit (hereafter referred to as CPU) which carries out arithmetic and logic processing means on a repetitive sequential basis for all input signals. Simple connections to the CPU are referred to interfaces (I/F) .
  • the specific vehicle mode through its interface calls for specific operational routines to be used in the CPU.
  • the speed pickups on Pulley A and B being digital in nature send their signals directly through an I/F to the CPU where a speed measurement is derived which then stored and updated in speed registers in the random access memory (RAM) .
  • the accelerator and brake use analogue to digital converters (A/D) for inputs to the CPU for processing by the CPU to be stored and updated in registers in the RAM.
  • A/D analogue to digital converters
  • the rotation of a gear in the actuator of Pulley B and the rotation of a gear in the actuator of Pulley A can accurately specify the radius of the belt surface in each pulley through analogue to digital outputs (A/D) 335B and 335A. This information is supplied for processing by the CPU for storage and updating in appropriate RAM registers for the pulley radii.
  • the output signals from the CPU to change the actuator torque in Pulley A or B operate by changing the valve flow areas in 332A or 332B, Fig. 16.
  • digital to analogue (D/A) converters are used to operate the valves.
  • Pressure sensors, 336A and 336B, Fig. 16 are used in the actuator controls on the two pulleys to send their signals through analogue to digital converters to the CPU.
  • the directional control valve 33 B on the actuator of Pulley B is operated by the CPU through an interface to reverse momentarily the normal actuator torque to obtain a special short duration very rapid pulley radius change for accelerating the engine when operating in the high speed mode, i.e., for a rapid vehicle acceleration in passing.
  • the transmission clutch controls are off-on hydraulic valves that are triggered by the CPU. If no current is supplied to solenoid 307 or 309, Fig. 9, spring centering of valve 308 causes no pressure on either the HC or HL clutch leaving them open with no connection between the wheels and the transmission.
  • the type of the output from the CPU to the engine depends on the nature of the entire control module. It may be capable to responding to a digital input or it may require an analogue input. A digital to analogue output is indicated in Fig. 18.
  • a read only memory is included in the transmission control for the various design constants, i.e., gear ratios and needed algorithms such as the specific relation of actuator control gear rotation to pulley belt radius in each pulley and the desired engine speed versus vehicle speed for various accelerator or brake input positions in the two speed modes.
  • the ROM would also include the essential algorithm giving the desired operating belt slip as a function of the accelerator or brake input position.
  • Typical operational ranges for the pulley radii and speeds are listed in Fig. 19 for the various vehicle modes and accelerator or brake position, for a medium size vehicle giving a maximum torque of 165 ft. lb. The manner in which each pulley varies in radius relative to the other for each operating condition is specified.
  • FIG. 20 A block diagram of the feedback control system is given in Fig. 20.
  • the driver initiates the operation and is the over-riding feedback element as shown in the schematic diagram.
  • the vehicle's engine is started with the vehicle mode at park (P) .
  • the CPU using a self contained clock and battery, is triggered by the starter switch S (Fig. 18) to begin a repetitive sampling of the registers for all bus connected elements.
  • the pressures on the control valve 332A of Pulley A and control valve 332B of Pulley B, Fig. 16 are set initially so that with both transmission clutches HC and LC, (Fig.
  • This interrupt service routine actuates a specific interrupt service routine suitable for the vehicle drive mode in the low speed mode condition.
  • This interrupt service routine uses the previous data stored in the registers, then calculates the desired percent belt slip as described previously and compares it to the operating belt slip value (as stored in the RAM) . The difference (+ or -) is used as an input to increase or decrease the control pressure on Pulley A actuator through a command to the pressure control valve 332A, Fig. 16. For an increase in the desired belt slip, the pressure to Pulley A hydraulic unit 331A, Fig. 16, is increased thereby trying to increase the belt radial position in Pulley A. The interrupt service routine may then continue by comparing the desired engine or Pulley B speed to the vehicle speed ratio as given on Fig.
  • the difference is then used to decrease or increase the control pressure on Pulley B actuator through a command to the pressure control valve 332B, Fig. 16.
  • the vehicle speed is easily computed by the CPU knowing the speed mode and Pulley A and B speeds using the ROM. For the desired increase in vehicle speed, the Pulley B radius will be controlled to decrease while the Pulley A radius increases in the low speed drive mode as indicated in Fig. 19.
  • the interrupt service routine may then continue by comparing the desired engine speed (Pulley B speed) with the existing engine speed as stored in the RAM and using the accelerator position input call for a specific speed change by an input to the engine control module.
  • the engine control module changes the engine speed by varying the fuel flow to the engine changing the engine torque.
  • the engine control module is not a direct part of the transmission control module
  • the electronic signals from the transmission control module basically control the engine speed.
  • the stability of the overall control depends on the specific feedback loops gains but the scheduling of the interrupt service routines recognize the need for updating and controlling the percent belt slip more frequently than the belt radius ratio change and a signal to change the engine speed.
  • the vehicle speed change due to the vehicle inertia is much slower than the possible pulley speed ratio and engine speed changes.
  • the transmission speed ratio change essentially limits how fast the engine speed is permitted to change by the speed of changes in the belt radius ratio.
  • the computer processing unit can access very rapidly all sensors and update all registers in the RAM repetitively every 0.1 millisecond.
  • the hydraulic controllers each consist of a micro computer controller for the hydraulic pressure as commanded by an error command from the central CPU with its interrupt service routines based on a programmed timer.
  • the hydraulic control with its feedback loop can operate quite rapidly - order of a few milliseconds for large pressure changes at hydraulic unit 331 for the actuator.
  • the internal hydraulic micro controller functions much faster than the time intervals programmed for the interrupt service routines that can send a new command to the micro controllers. This permits a stable system operation.
  • the CPU can initiate and cause all three controls - for the percent belt slip, the Pulley B radius (or the pulley radius ratio) and the engine speed - to function simultaneously using the high speed digital capabilities of a CPU.
  • the engine speed is controlled to increase and belt radius in Pulley B decreases to its minimum position.
  • an interrupt service routine causes the engagement of the transmission clutch control 309 and disengages the clutch control 307, Fig. 9, if the accelerator calls for additional vehicle speed.
  • this clutch actuation occurs with the LO gear and HO gear at the same speed.
  • the control now operates in the high speed mode and uses a different algorithms for setting the engine speed as a function of the accelerator position as shown on Fig. 22.
  • the control for the hydraulic pressure changes at the Pulley B control valve 332B is reversed from the low speed mode operation in that an accelerator input calling for increased vehicle speed will increase the Pulley B radius using the actuator hydraulic unit 331B, Fig. 16 rather than a decrease in the Pulley B radius as in the low speed mode.
  • the high speed mode operation has one variant in that a high acceleration vehicle passing maneuver is significantly aided by a momentary reversing of the hydraulic control pressure to the actuator of Pulley B. This is accomplished in a special interrupt service routine for Pulley B belt radius control that reverses Pulley B actuator torque by reversing the inlet and outlet flows to the Pulley B hydraulic unit 33IB Fig. 16 using the directional control valve 334B.
  • the momentary pressure reverse causes the actuator in pulley to aid in reducing the Pulley B radius thus allowing the engine speed to accelerate rapidly from a low cruise speed to generate the desired power for the vehicle passing maneuver.
  • the triggering for this special interrupt service routine depends on the accelerator forward rate of movement and its duration would be limited and end by returning the directional control valve 334B to its original position.
  • the more rapid sequence of the interrupt service routines for the pressure to the actuator hydraulic unit 331A on Pulley A will operate to maintain, within limits, the desired percent belt slip in the transmission or pulley torque giving a desired output torque.
  • the engine acceleration can be tailored to avoid a significant vehicle hesitation lag in the driver demand for greatly increased power.
  • the operational vehicle speed range in the high speed mode vs the low speed mode is shown in the typical case on Fig. 13 versus the pulley speed ratio and on Fig. 21 and Fig. 22 versus the accelerator or brake position.
  • the shift from the high speed mode to the low speed mode occurs whenever the accelerator position calls for a reduction in vehicle speed when Pulley A has reached its maximum radius position (or Pulley B has reached its minimum radius, or if L gear and H gear (Fig. 9) are sensed to have the same speed) .
  • the transmission clutch control 307 is switched on and control 309 is switched off Fig. 9 and the CPU is now operated in the low speed mode.
  • This control arrangement can be adapted to permit the engine to give significant dynamic braking by compressing air in either speed mode.
  • the control for the hydraulic pressure changes at the Pulley B control valve 332B is reversed from the low speed mode operation in that an accelerator input calling for increased reverse vehicle speed will increase rather than decrease the Pulley B radius using the actuator hydraulic unit 3IB.
  • This is the same operation as in the high speed mode except that the power flow is through the planetary gear system and low speed mode clutch LC (Fig. 9) .
  • the essential % belt slip control basically functions to vary the pulley torque and consequently the torque to the wheels under all operating conditions.

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Abstract

Transmission régénératrice variable en continu dans laquelle un ensemble de transmission variable en continu (A, B) est associé à un ensemble d'engrenages à régénération (H, L, C, I, E) incorporant des embrayages supérieur (HC) et inférieur (LC) et situé entre une machine motrice (300) telle que le moteur d'un véhicule, et une sortie (O) telle qu'une transmission reliée aux roues du véhicule. Par un contrôle de certains états du système (306B, 306B) (par exemple la vitesse et le sens de rotation voulus à la sortie, la vitesse de rotation du moteur et le glissement de la courroie) et simultanément par un réglage du régime moteur, un sous-système de commande règle les diamètres réels des poulies d'entrée (B) et de sortie (A) de l'ensemble de transmission variable en continu de manière à obtenir un rapport entre le couple de la poulie et la tension de courroie produisant une valeur de glissement de courroie menant à une rotation de la sortie dans le sens voulu à une vitesse qui correspond à un programme de rapports prédéterminé entre la vitesse de rotation du moteur et la vitesse de rotation en sortie dans les conditions instantanées mesurées et eu égard à la vitesse de sortie voulue.
PCT/US1991/002690 1991-04-19 1991-04-19 Transmission variable en continu et a regeneration de couple WO1992018791A1 (fr)

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Cited By (3)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
WO1995027159A1 (fr) * 1994-04-02 1995-10-12 Robert Bosch Gmbh Regulateur de glissement pour transmission a variation continue
WO1998043003A1 (fr) * 1997-03-25 1998-10-01 Robert Bosch Gmbh Dispositif et procede pour reduire le patinage lors de la commande d'une transmission a changement de vitesses continu dans un vehicule a moteur
EP1403560A3 (fr) * 2002-09-26 2009-08-12 JATCO Ltd Appareil et procedé de controle pour vehicules automobiles dans lequel une transmission à variation continue du type à courroie est equipé avec un dispositif de prevention du glissement de la courroie

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US4824419A (en) * 1988-07-05 1989-04-25 Kumm Industries, Inc. Flat belt continuously variable transmission with geared speed ratio control system
US5011458A (en) * 1988-11-09 1991-04-30 Kumm Industries, Inc. Continuously variable transmission using planetary gearing with regenerative torque transfer and employing belt slip to measure and control pulley torque

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US4295836A (en) * 1979-06-01 1981-10-20 Kumm Emerson L Flat belt transmission with rotary actuator and integrated control system
US4665773A (en) * 1984-03-13 1987-05-19 Mitsubishi Jidosha Kogyo Kabushiki Kaisha Continuously variable transmission apparatus for automobile
US4627308A (en) * 1984-04-30 1986-12-09 Aisin Warner Kabushiki Kaisha Automatic transmission for a vehicle
US4619629A (en) * 1984-05-03 1986-10-28 Toyota Jidosha Kabushiki Kaisha Hydraulic pressure control apparatus for a continuously variable transmission
US4673378A (en) * 1984-05-03 1987-06-16 Toyota Jidosha Kabushiki Kaisha Hydraulic control apparatus for a continuously variable transmission
US4591351A (en) * 1985-06-07 1986-05-27 Kumm Emerson L Variable ratio pulleys for flat belt transmission system
US4714452A (en) * 1986-06-06 1987-12-22 Kumm Emerson L Oriented flat belt continuously variable transmission using pulleys with guideways
US4824419A (en) * 1988-07-05 1989-04-25 Kumm Industries, Inc. Flat belt continuously variable transmission with geared speed ratio control system
US5011458A (en) * 1988-11-09 1991-04-30 Kumm Industries, Inc. Continuously variable transmission using planetary gearing with regenerative torque transfer and employing belt slip to measure and control pulley torque

Cited By (5)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
WO1995027159A1 (fr) * 1994-04-02 1995-10-12 Robert Bosch Gmbh Regulateur de glissement pour transmission a variation continue
US5871411A (en) * 1994-04-02 1999-02-16 Robert Bosch Gmbh Slip controller for a continuously variable transmission
WO1998043003A1 (fr) * 1997-03-25 1998-10-01 Robert Bosch Gmbh Dispositif et procede pour reduire le patinage lors de la commande d'une transmission a changement de vitesses continu dans un vehicule a moteur
US6409627B2 (en) 1997-03-25 2002-06-25 Robert Bosch Gmbh Device and method for reducing slip in the control system of a CVT in a motor vehicle
EP1403560A3 (fr) * 2002-09-26 2009-08-12 JATCO Ltd Appareil et procedé de controle pour vehicules automobiles dans lequel une transmission à variation continue du type à courroie est equipé avec un dispositif de prevention du glissement de la courroie

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