WO2021046942A1 - 全齿轮无级自动变速与速比主动控制系统 - Google Patents

全齿轮无级自动变速与速比主动控制系统 Download PDF

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Publication number
WO2021046942A1
WO2021046942A1 PCT/CN2019/109084 CN2019109084W WO2021046942A1 WO 2021046942 A1 WO2021046942 A1 WO 2021046942A1 CN 2019109084 W CN2019109084 W CN 2019109084W WO 2021046942 A1 WO2021046942 A1 WO 2021046942A1
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Prior art keywords
gear
speed ratio
differential
automatic transmission
planetary
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PCT/CN2019/109084
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English (en)
French (fr)
Inventor
何耀华
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何耀华
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Application filed by 何耀华 filed Critical 何耀华
Priority to JP2020555150A priority Critical patent/JP7250815B2/ja
Priority to EP19928291.4A priority patent/EP3819521A4/en
Priority to US17/033,689 priority patent/US11060588B2/en
Publication of WO2021046942A1 publication Critical patent/WO2021046942A1/zh

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    • 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
    • F16H3/00—Toothed gearings for conveying rotary motion with variable gear ratio or for reversing rotary motion
    • F16H3/44—Toothed gearings for conveying rotary motion with variable gear ratio or for reversing rotary motion using gears having orbital motion
    • F16H3/72—Toothed gearings for conveying rotary motion with variable gear ratio or for reversing rotary motion using gears having orbital motion with a secondary drive, e.g. regulating motor, in order to vary speed continuously
    • F16H3/724—Toothed gearings for conveying rotary motion with variable gear ratio or for reversing rotary motion using gears having orbital motion with a secondary drive, e.g. regulating motor, in order to vary speed continuously using external powered electric machines
    • 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
    • F16H37/06—Combinations of mechanical gearings, not provided for in groups F16H1/00 - F16H35/00 comprising essentially only toothed or friction gearings with a plurality of driving or driven shafts; with arrangements for dividing torque between two or more intermediate shafts
    • F16H37/08—Combinations of mechanical gearings, not provided for in groups F16H1/00 - F16H35/00 comprising essentially only toothed or friction gearings with a plurality of driving or driven shafts; with arrangements for dividing torque between two or more intermediate shafts with differential gearing
    • F16H37/0806—Combinations of mechanical gearings, not provided for in groups F16H1/00 - F16H35/00 comprising essentially only toothed or friction gearings with a plurality of driving or driven shafts; with arrangements for dividing torque between two or more intermediate shafts with differential gearing with a plurality of driving or driven shafts
    • 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
    • F16H37/06—Combinations of mechanical gearings, not provided for in groups F16H1/00 - F16H35/00 comprising essentially only toothed or friction gearings with a plurality of driving or driven shafts; with arrangements for dividing torque between two or more intermediate shafts
    • F16H37/08—Combinations of mechanical gearings, not provided for in groups F16H1/00 - F16H35/00 comprising essentially only toothed or friction gearings with a plurality of driving or driven shafts; with arrangements for dividing torque between two or more intermediate shafts with differential gearing
    • F16H37/0833—Combinations of mechanical gearings, not provided for in groups F16H1/00 - F16H35/00 comprising essentially only toothed or friction gearings with a plurality of driving or driven shafts; with arrangements for dividing torque between two or more intermediate shafts with differential gearing with arrangements for dividing torque between two or more intermediate shafts, i.e. with two or more internal power paths
    • 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
    • F16H37/06—Combinations of mechanical gearings, not provided for in groups F16H1/00 - F16H35/00 comprising essentially only toothed or friction gearings with a plurality of driving or driven shafts; with arrangements for dividing torque between two or more intermediate shafts
    • F16H37/08—Combinations of mechanical gearings, not provided for in groups F16H1/00 - F16H35/00 comprising essentially only toothed or friction gearings with a plurality of driving or driven shafts; with arrangements for dividing torque between two or more intermediate shafts with differential gearing
    • F16H37/0833—Combinations of mechanical gearings, not provided for in groups F16H1/00 - F16H35/00 comprising essentially only toothed or friction gearings with a plurality of driving or driven shafts; with arrangements for dividing torque between two or more intermediate shafts with differential gearing with arrangements for dividing torque between two or more intermediate shafts, i.e. with two or more internal power paths
    • F16H37/084—Combinations of mechanical gearings, not provided for in groups F16H1/00 - F16H35/00 comprising essentially only toothed or friction gearings with a plurality of driving or driven shafts; with arrangements for dividing torque between two or more intermediate shafts with differential gearing with arrangements for dividing torque between two or more intermediate shafts, i.e. with two or more internal power paths at least one power path being a continuously variable transmission, i.e. CVT
    • 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
    • F16H48/00—Differential gearings
    • F16H48/06—Differential gearings with gears having orbital motion
    • F16H48/08—Differential gearings with gears having orbital motion comprising bevel gears
    • 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
    • F16H37/04—Combinations of toothed gearings only
    • F16H2037/049—Forward-reverse units with forward and reverse gears for achieving multiple forward and reverse gears, e.g. for working machines
    • 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
    • F16H3/00—Toothed gearings for conveying rotary motion with variable gear ratio or for reversing rotary motion
    • F16H3/02—Toothed gearings for conveying rotary motion with variable gear ratio or for reversing rotary motion without gears having orbital motion
    • F16H3/08—Toothed gearings for conveying rotary motion with variable gear ratio or for reversing rotary motion without gears having orbital motion exclusively or essentially with continuously meshing gears, that can be disengaged from their shafts
    • F16H3/14—Gearings for reversal only
    • 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
    • F16H3/00—Toothed gearings for conveying rotary motion with variable gear ratio or for reversing rotary motion
    • F16H3/44—Toothed gearings for conveying rotary motion with variable gear ratio or for reversing rotary motion using gears having orbital motion
    • F16H3/46—Gearings having only two central gears, connected by orbital gears
    • F16H3/60—Gearings for reversal only

Definitions

  • the invention relates to the technical field of power transmission systems, in particular to a full gear continuously variable automatic transmission power transmission system.
  • the power transmission system (or device) is the most basic and important part of all machinery. Its function is to transmit the power from the power source (such as engine, steam turbine, water turbine or electric motor) to the working mechanism or operating machinery (such as various types of The driving wheels of vehicles, the propellers of ships, the spindles/cutter heads of machine tools, rotary tillers, hydraulic pumps, compressors, blowers, etc.). Since the dynamic characteristics of the power source (such as power, torque, etc. change with speed) are often very different from the requirements of the working mechanism or working machinery, the power transmission device between the power source and the working mechanism or working machinery must have Decelerating or shifting function, this is the reason why many mechanical transmission devices use reducer or transmission (transmission device with variable transmission ratio).
  • the power source such as engine, steam turbine, water turbine or electric motor
  • the working mechanism or operating machinery such as various types of The driving wheels of vehicles, the propellers of ships, the spindles/cutter heads of machine tools, rotary tillers, hydraulic pumps, compressors, blowers, etc
  • the transmissions in use in various types of machinery mainly include manual mechanical stepped transmission (MT), electronically controlled mechanical stepped automatic transmission (AMT), belt type continuously variable automatic transmission (CVT), and hydraulic automatic transmission (AT) , Dual-clutch automatic transmission (DCT), etc.
  • manual mechanical stepped transmission AMT
  • AMT electronically controlled mechanical stepped automatic transmission
  • CVT belt type continuously variable automatic transmission
  • AT hydraulic automatic transmission
  • DCT Dual-clutch automatic transmission
  • MT manual mechanical stepped transmission
  • the mechanically stepped transmission with gears greater than 6 usually adopts the structure of main and auxiliary transmissions.
  • the electronically controlled mechanically stepped automatic transmission is an improved version of the manual mechanically stepped transmission.
  • the variable speed transmission part is almost the same as the original mechanical manual stepped transmission. The only difference is that the shift operation is controlled by electronic hydraulic pressure or electronically controlled air pressure.
  • the original manual mechanical stepped transmission's inherent disadvantages such as "large speed ratio difference between each gear, step change in speed ratio during gear shifting, and unsatisfactory power transmission characteristics" inherent in the original manual mechanical stepped transmission also exist, but shifting does not require manual operation. operating. It is precisely because both manual mechanical stepped transmission (MT) or electronically controlled mechanical stepped automatic transmission (AMT) have many of the above shortcomings, so people have always hoped to replace the mechanical stepped transmission with a continuously variable transmission.
  • the biggest difference between it and the electronically controlled mechanically stepped automatic transmission (AMT) is:
  • the mechanically stepped transmissions are used in parallel, using two clutches and electronically controlled hydraulic pressure or electronically controlled air pressure or electronically controlled electric shifting operating mechanism to complete the shifting operation. Since two mechanically stepped transmissions are used, the number of gears of the transmission can be Increased, the speed ratio difference between each gear can be reduced; due to the use of two clutches to cooperate with the shift, the shifting process can be shortened, and the shifting frustration is reduced.
  • the dual-clutch automatic transmission is also a mechanically stepped transmission
  • its shortcomings of "the speed ratio of each gear has a step difference, the speed ratio of the gear shifting process, and the power transmission characteristics are not ideal" are still Exist, but the above shortcomings are reduced to a certain extent.
  • the structure and process of the dual-clutch automatic transmission (DCT) are relatively complicated, the manufacturing cost is high and the adaptability to the use environment is poor.
  • the technical problem to be solved by the present invention is to completely overcome the shortcomings of various transmissions in the prior art, and provide a full-gear stepless automatic transmission and speed ratio active control system with simple structure, small size and light weight, high transmission efficiency and speed ratio
  • the wide range of change and the ability to transmit torque can meet the requirements of all mechanical transmissions. It not only effectively realizes continuously variable automatic transmission within the entire range of speed ratio changes, but also realizes active control of the speed ratio to meet the needs of various mechanical transmissions.
  • the present invention adopts the following technical solutions:
  • a full-gear stepless automatic transmission and speed ratio active control system which is characterized in that: a differential mechanism that distributes torque proportionally is connected in series at the front end of the planetary gear mechanism to form a continuously variable automatic transmission device; An active speed ratio control mechanism is set between the speed mechanism and the planetary gear mechanism; the differential mechanism that distributes torque proportionally includes two differential bevel gears arranged at the power and motion output ends: the first differential bevel gear and the second differential Bevel gear, the first differential bevel gear is rigidly connected to the ring gear of the planetary gear mechanism through the hollow first differential bevel gear shaft; the second differential bevel gear shaft passes through the hollow first differential bevel gear shaft, the second The differential bevel gear is connected to the sun gear of the planetary gear mechanism through the second differential bevel gear shaft;
  • the speed ratio active control mechanism includes a speed ratio adjusting motor, a speed ratio adjusting driving gear in a constant meshing state, and a speed ratio adjusting driven gear; the speed ratio adjusting driven gear is rigidly connected with the first differential bevel gear shaft, and the speed ratio adjusting is active The gear is installed on the output shaft of the speed ratio adjustment motor;
  • the two input ends of the planetary gear mechanism are respectively the sun gear located in the center and the outermost ring gear.
  • the planetary gears mesh with the sun gear and the ring gear at the same time and output power to the outside through the planet carrier.
  • the differential mechanism for distributing torque proportionally includes a differential housing, the input shaft of the continuously variable automatic transmission device is rigidly connected to the differential housing at the front end of the differential housing; inside the differential housing, the first differential cone Both the gear and the second differential bevel gear are rotatably supported in the shaft hole of the differential housing extending along the front and rear ends or left and right; the second differential bevel gear and the first differential bevel gear are connected to the planets at the front and rear ends respectively.
  • the bevel gear meshes; the front end of the first differential bevel gear shaft is rigidly connected with the first differential bevel gear; the front end of the second differential bevel gear shaft is rigidly connected with the second differential bevel gear, and the rear end is from the hollow first differential
  • the bevel gear shaft passes through and passes through the differential case to be rigidly connected to the sun gear or form an integral structure with the sun gear shaft.
  • the planetary bevel gear is provided with a circumferential annular groove around the center line of the planetary bevel gear shaft, one end of the planetary bevel gear shaft is fixedly mounted on the differential case, and the other end is mounted on the ring through a bearing.
  • the rotation center of the planetary bevel gear is the central arc line of the annular groove when the planetary bevel gear rotates, that is, the centerline of the planetary bevel gear shaft is the circumferential motion track in the annular groove relative to the planetary bevel gear.
  • the distance from the meshing point A of the planetary bevel gear and the second differential bevel gear to the centerline of the planetary bevel gear shaft is S1
  • the meshing of the planetary bevel gear and the first differential bevel gear is S 2
  • the ratio of S 1 to S 2 is always the set ratio.
  • the differential mechanism for proportionally distributing torque uses at least two planetary bevel gears.
  • At least two planetary gears are provided in the planetary gear mechanism.
  • a deceleration mechanism is provided before the sun gear to increase the speed ratio change range of the all-gear stepless automatic transmission and the gear ratio active control system, and the deceleration mechanism is a symmetrical fixed-axis gear reduction mechanism , Or one of an asymmetric fixed-axis gear train gear reduction mechanism or a planetary gear reduction mechanism, the second differential bevel gear shaft is decelerated by the deceleration mechanism and connected to the sun gear, and the deceleration mechanism is connected in series before the sun gear.
  • the front or rear end of the continuously variable automatic transmission device is connected in series with a reversing mechanism; the switching of at least three modes of forward rotation, reverse rotation and interruption of power and motion transmission is realized; the reversing mechanism is a combined structure of a clutch and a synchronizer , Or double-joint multi-disc clutch type reversing mechanism, or double clutch type reversing mechanism.
  • the clutch input shaft and the transmission input shaft are arranged in a straight line, the input shaft gear is arranged on the clutch input shaft, and the left end or front end of the input shaft of the continuously variable transmission device is installed on the input through a bearing.
  • the reversing driving gear is constantly meshed with the input shaft gear, and at the same time is also constantly meshing with the previous set of bridge gears; the latter set of bridge gears are coaxial with the previous set of bridge gears and both
  • the bridge gear shaft is rigidly connected, and the latter group of bridge gears are constantly meshed with the reversing driven gear.
  • the reversing driven gear is sleeved on the input shaft of the continuously variable automatic transmission using needle bearings or sliding bearings.
  • the input shaft gear and A lock ring or lock pin synchronizer is installed between the reversing driven gears.
  • the clutch adopts one of a friction type die spring clutch, a friction type peripheral cylindrical coil spring clutch, a friction type central cone coil spring clutch, a multi-plate dry friction clutch, a multi-plate wet friction clutch, and an electromagnetic clutch.
  • the operation of the clutch It adopts one of the operation modes of electronic control hydraulic pressure, electronic control air pressure, electronic control electromagnetic, electronic control servo motor, and electronic control stepping motor.
  • an input shaft gear is provided on the clutch input shaft, and the reversing driven gear is sleeved on the input shaft of the continuously variable automatic transmission through a needle bearing or a sliding bearing.
  • a dual multi-disc clutch is installed between the shaft gear and the reversing driven gear; the left or front end of the input shaft of the stepless automatic transmission is installed in the bearing seat hole at the right end of the clutch input shaft gear through a bearing; the reversing driving gear and the input
  • the shaft gear is constantly meshed, and at the same time it is also constantly meshing with the bridge gear; the front and rear two sets of bridge gears are rigidly connected with the bridge gear shaft, and the rear bridge gear is constantly meshed with the reversing driven gear.
  • the dual multi-disc clutch adopts a dry dual multi-disc clutch or a wet dual multi-disc clutch; the operation of the dual multi-disc clutch adopts one of electronic control hydraulic pressure, electronic control air pressure, and electronic control electromagnetic operation mode.
  • the dual clutch reversing mechanism is composed of a dual clutch and a planetary gear mechanism with a lock-up mechanism.
  • the dual clutch is composed of a forward gear clutch and a reverse gear clutch.
  • the forward gear clutch passes through a hollow forward gear.
  • the input shaft is rigidly connected with the planet carrier of the reversing mechanism
  • the reverse clutch is rigidly connected with the sun gear of the reversing mechanism through the reverse input shaft passing through the center of the forward gear input shaft
  • the gear ring of the shifting mechanism is rigidly connected to the input shaft of the continuously variable automatic transmission device.
  • the left end or front end of the gearbox is rigidly connected to transmit power and movement to the differential mechanism that distributes the torque proportionally; the reverse gear input shaft and the planet carrier of the reversing mechanism are respectively equipped with a sun gear locker and a planet carrier locker.
  • the dual clutch adopts at least one of a friction type die spring clutch, a multi-plate dry friction clutch, a multi-plate wet friction clutch, and an electromagnetic clutch; the operation of the clutch adopts electronically controlled hydraulic pressure, electronically controlled air pressure, electronically controlled electromagnetic, and electronically controlled servo One of the operation modes of motor and electronically controlled stepping motor.
  • the full-gear stepless automatic transmission and speed ratio active control system of the present invention is provided with a speed ratio active control mechanism between the differential mechanism that distributes torque proportionally and the planetary gear mechanism, and the torque is proportionally distributed.
  • the differential mechanism includes two differential bevel gears arranged at the power and motion output ends.
  • the two differential bevel gears are respectively connected with the ring gear and the sun gear of the planetary gear mechanism as the input shaft of the planetary gear mechanism;
  • the speed ratio active control mechanism includes Speed ratio adjusting motor, ratio adjusting driving gear and speed ratio adjusting driven gear in constant meshing state;
  • speed ratio adjusting driven gear is rigidly connected with one of the differential bevel gear shafts, speed ratio adjusting driving gear and speed ratio adjusting motor
  • the output shaft is connected; the planetary gear meshes with the sun gear and the ring gear at the same time and outputs power to the outside through the planet carrier.
  • All-gear continuously variable automatic transmission and speed ratio active control system All transmission components of the system are gears and shafts with extremely high efficiency, so it has the same high efficiency as a mechanical manual transmission.
  • the speed ratio has a wide range of change, which can meet the requirements of various mechanical systems for the speed ratio change range.
  • a deceleration mechanism is provided before the sun gear to increase the speed ratio change range of the full gear continuously variable automatic transmission and the speed ratio active control system, or to combine two or more of the present invention's continuously variable automatic transmission and speed ratio active control system. Used in tandem, it can obtain a continuously variable automatic transmission and a speed ratio active control system with a speed ratio change range of any size. For example, when a large speed ratio change range is required, a planetary gear mechanism with a deceleration mechanism set before the sun gear can significantly reduce the radial size of the ring gear while increasing the full-gear stepless automatic transmission and active speed ratio control.
  • the speed ratio change range of the system For example, the gear ratio ⁇ between the ring gear and the sun gear is reduced from the original 7 to 5, the stepless automatic transmission range is increased from 1 to 4 to 1.143-12, and the stepless automatic transmission range is greatly increased to nearly 3 times the original. .
  • Stepless automatic transmission can be effectively realized in the whole range of speed ratio change.
  • Fig. 1 is a schematic diagram of the structure of the full-gear stepless automatic transmission and speed ratio active control system of the present invention.
  • Figure 2 is a schematic diagram of the planetary bevel gear structure.
  • Figure 3 is a schematic diagram of the differential mechanism that distributes torque proportionally.
  • Fig. 4 is a schematic diagram of the structure of a full-gear stepless automatic transmission and speed ratio active control system with a reversing mechanism of the present invention.
  • Figure 5 is a schematic diagram of the structure of the reversing mechanism.
  • Fig. 6 is a structural diagram of another embodiment of the planetary gear mechanism of the acceleration and deceleration mechanism at the input end of the sun gear.
  • Figure 7 is a schematic diagram of the structure in the forward gear state (three implementations of the reversing mechanism).
  • Figure 8 is a schematic diagram of the structure in the reverse gear state (three implementations of the reversing mechanism).
  • Figure 9 is a schematic diagram of the structure in the neutral state (three implementations of the reversing mechanism).
  • I Downifferential mechanism that distributes torque proportionally; II—Active speed ratio control mechanism; III—Planetary gear mechanism; IV—Reversing mechanism; 1—Infinitely variable automatic transmission input shaft; 2—Planetary bevel gear; Planetary bevel gear shaft; 4—differential housing; 5—differential bevel gear shaft; 6, 7—differential bevel gear; 8—speed ratio adjustment motor; 9—speed ratio adjustment driving gear; 10—speed ratio adjustment slave Moving gear; 11—sun gear; 12—ring gear; 13—planetary gear; 14—power output shaft; 15—planet carrier; 16—sun gear shaft (differential bevel gear shaft); 17—clutch input shaft; 18—clutch 19—input shaft gear; 20—bridge gear I; 21—bridge gear shaft; 22—bridge gear II; 23—reversing driven gear; 24-synchronizer; 25—reversing driving gear; 26— Reversing driving gear shaft; 27—reversing input shaft; 28—double multi-disc clutch; 29—reverse gear clutch; 30—
  • the specific structure of the all-gear continuously variable automatic transmission and speed ratio active control system implemented in accordance with the present invention is shown in Figure 1.
  • the all-gear stepless automatic transmission and active ratio control system includes at least three parts connected in sequence: the differential mechanism I which distributes the torque proportionally, the active speed ratio control mechanism II, and the planetary gear mechanism III.
  • the differential mechanism I and the planetary gear mechanism III which distribute torque proportionally, together form a continuously variable automatic transmission.
  • the differential mechanism I that distributes torque proportionally includes two power and motion output end differential bevel gears: the first differential bevel gear 6 and the second differential bevel gear 7; the first differential bevel gear 6 passes through the first differential
  • the speed bevel gear shaft 5 is connected to the ring gear 12 of the planetary gear mechanism III, and the connection mode can be a key or spline;
  • the second differential bevel gear 7 is connected to the planetary gear mechanism III sun gear 11 through the second differential bevel gear shaft 16 ,
  • the connection mode can be key or spline;
  • the second differential bevel gear shaft 16 is used as the central shaft in the differential mechanism I and the speed ratio active control mechanism II that distribute torque proportionally, and the second differential bevel gear shaft 16 Passing through the hollow first differential bevel gear shaft 5 (coaxial instead of coaxial), the first differential bevel gear shaft 5 penetrates through the active speed ratio control mechanism II and respectively extends to distribute the
  • the structure of planetary gear mechanism III is as follows:
  • the planetary gear mechanism III is input by the continuously variable automatic transmission, the second differential bevel gear shaft (or sun gear shaft) 16, the sun gear 11, the planet carrier 15, the planet
  • the gear 13 and the ring gear 12 are composed of the sun gear 11 and the ring gear 12 being the two input ends of the planetary gear mechanism III, and the power of the planetary gear mechanism III is output by the planet carrier 15.
  • the sun gear 11 is rigidly connected to the second differential bevel gear shaft (sun gear shaft) 16 through keys or splines
  • the ring gear 12 is rigidly connected to the first differential bevel gear shaft 5 of hollow structure through keys or splines.
  • two planetary gears 13 are provided.
  • the differential mechanism I that distributes the torque proportionally is connected in series at the front end of the planetary gear mechanism III, including the differential housing 4 that distributes the torque proportionally, outside the differential housing 4, and a stepless automatic transmission device.
  • the input shaft 1 is rigidly connected to the differential housing 4 at its front end.
  • the first differential bevel gear 6 and the second differential bevel gear 7 are both supported on the front and rear ends of the differential housing 4 through bearings. Extend the shaft hole.
  • the second differential bevel gear 7 and the first differential bevel gear 6 mesh with the planetary bevel gear 2 at the front and rear ends, respectively.
  • the front end of the hollow first differential bevel gear shaft 5 is rigidly connected to the first differential bevel gear 6 through a key or spline; the first differential bevel gear 6 passes through the first differential bevel gear shaft 5 and the planetary gear mechanism III
  • the ring gear 12 is rigidly connected.
  • the second differential bevel gear shaft 16 passes through the hollow first differential bevel gear shaft 5.
  • the front end is rigidly connected to the second differential bevel gear 7 through keys or splines, and the rear end passes through the differential bevel gear.
  • the speed gear housing 4 is rigidly connected to the sun gear 11 through a key or spline (the sun gear 11 can also be made into an integral structure with the sun gear shaft 16).
  • the planetary bevel gear 2 is provided with an annular groove 46, one end of the planetary bevel gear shaft 3 is fixedly installed in the hole of the differential case 4 (see Figure 1, not shown in Figure 2), and the other end It is installed in the annular groove 46 of the planetary bevel gear 2 through a bearing 45 (which can be a cylindrical roller bearing, a needle roller bearing or a sliding bearing), as shown in FIG. 2(a).
  • a bearing 45 which can be a cylindrical roller bearing, a needle roller bearing or a sliding bearing
  • the structure of the active speed ratio control mechanism is as follows:
  • the speed ratio active control mechanism II is composed of a speed ratio adjusting motor 8, a speed ratio adjusting driving gear 9 and a speed ratio adjusting driven gear 10.
  • the speed ratio adjusting driven gear 10 is rigidly connected with the hollow structure of the first differential bevel gear shaft 5 through keys or splines, and the speed ratio adjusting driving gear 9 is fixedly installed on the output shaft of the speed ratio adjusting motor 8 through keys or splines.
  • the speed ratio adjusting driving gear 9 and the speed ratio adjusting driven gear 10 are in a constant meshing state.
  • the structure of the reversing mechanism IV is as follows: In engineering, many working mechanisms or operating machinery often need to reverse rotation during the working process, that is, sometimes forward and sometimes reverse. There are two ways to realize its forward rotation and reverse rotation, namely: 1 let the power source run forward and reverse; 2 add a reversing mechanism to the transmission. Method 1 Only power sources such as electric motors can be realized, and most other types of power sources are difficult to realize.
  • a reversing mechanism IV is added to the all-gear continuously variable automatic transmission and speed ratio active control system of the present invention.
  • the reversing mechanism can adopt the reversing mechanism of clutch + synchronizer (see Figure 5 (a)) or the reversing mechanism of double multi-disc clutch (see Figure 5 (b)) or the double clutch reversing mechanism (see Figure 5 (see Figure 5 (b)).
  • c) A variety of solutions with different structures. Respectively corresponding to Figure 4 4 (a) ⁇ 4 (c) three kinds of full gear continuously variable automatic transmission with reversing mechanism and the active control system structure of the speed ratio.
  • 4(a) is the structure diagram of the first embodiment of the full gear stepless automatic transmission and speed ratio active control system with the reversing mechanism of the present invention
  • 4(b) is the full gear stepless automatic transmission with the reversing mechanism of the present invention
  • 4(c) is the third embodiment of the full gear continuously variable automatic transmission and speed ratio active control system with a commutation mechanism of the present invention. Therefore, it is convenient to change the direction of the motion transmission, and realize the switching of the forward gear and the reverse gear.
  • the reversing mechanism IV can be connected in series to the front end of the continuously variable automatic transmission device or the rear end of the continuously variable automatic transmission device. Its function is to realize the three modes of forward rotation, reverse rotation and interruption of power and motion transmission. Switch.
  • the reversing mechanism IV composed of clutch + synchronizer is composed of clutch input shaft 17, clutch 18, input shaft gear 19, reversing driving gear 25, reversing driving gear shaft 26, synchronization
  • the left side is a cross-sectional view perpendicular to the axial direction between the input shaft gear 19, the bridge gear 20 and the reversing driving gear 25; in the right panel, the continuously variable transmission input shaft 1
  • the left end or front end of the shaft is mounted in the bearing housing hole at the right end of the input shaft gear 19 through a bearing;
  • the reversing driving gear 26 is constantly meshed with the input shaft gear 19, and at the same time is also constantly meshing with the bridge gear 20;
  • the bridge gear 20 and the bridge gear 22 It is rigidly connected to the bridge gear shaft 21 through a key or spline, and the bridge gear 22 is constantly meshed with the reversing driven gear 23.
  • the reversing driven gear 23 uses needle bearings or sliding bearings to be input in the stepless automatic transmission.
  • a lock ring type (or lock pin type) synchronizer 24 is installed between the input shaft gear 19 and the reversing driven gear 23.
  • the function and working process of the reversing mechanism IV composed of clutch + synchronizer is: the forward rotation of the reversing mechanism is realized by the synchronizer 24 (the direction of rotation of the input shaft 1 of the stepless automatic transmission is the same as the direction of rotation of the clutch input shaft 17), Reverse rotation (the direction of rotation of the input shaft 1 of the continuously variable automatic transmission is opposite to the direction of rotation of the clutch input shaft 17) and the switching of neutral.
  • the clutch 18 When a reversing operation is required, the clutch 18 is first disengaged and the synchronizer 24 is withdrawn from the original gear position, and then the clutch 18 is engaged after the gear is engaged.
  • the synchronizer 24 When the synchronizer 24 is in the middle position, the reversing mechanism terminates the transmission of power, that is, the all-gear continuously variable automatic transmission and the active speed ratio control system are in neutral gear; the synchronizer 24 is moved to the leftmost end, and the reversing mechanism rotates forward; The synchronizer 24 moves right to the rightmost end, and the reversing mechanism is reversed.
  • the clutch 18 in the all-gear continuously variable automatic transmission and speed ratio active control system of the present invention can adopt a friction type die spring clutch, a friction type peripheral cylindrical spiral spring clutch, a friction type central cone spiral spring clutch, and a multi-plate dry friction clutch. , Multi-disc wet friction clutches, electromagnetic clutches and other different types of clutches.
  • the operation of the clutch can adopt various operation modes such as electronic control hydraulic pressure, electronic control air pressure, electronic control electromagnetic, electronic control servo motor, electronic control stepping motor, etc.
  • the dual multi-disc clutch reversing mechanism consists of a dual multi-disc clutch 28, a clutch input shaft 17, an input shaft gear 19, a reversing driving gear 25, and a reversing driving gear shaft 26 ,
  • the bridge gear and shaft including bridge gears 20 and 22, bridge gear shaft 21) and reversing driven gear 23 and other components, as shown in Figure 5 (b).
  • the left side of Figure 5(b) is a cross-sectional view perpendicular to the axial direction between the input shaft gear 19, the bridge gear 20 and the reversing driving gear 25; in the right panel, the continuously variable automatic transmission input shaft 1
  • the left end or front end of the shaft is mounted in the bearing housing hole at the right end of the input shaft gear 19 through a bearing;
  • the reversing driving gear 25 is constantly meshed with the input shaft gear 19, and at the same time is also constantly meshing with the bridge gear 20;
  • the bridge gear 20 and the bridge gear 22 is rigidly connected with the bridge gear shaft 21 through a key or spline, the bridge gear 22 is constantly meshed with the reversing driven gear 23, and the reversing driven gear 23 is sleeved in the stepless automatic transmission device through needle bearings or sliding bearings.
  • a dual multi-disc clutch 28 is installed between the input shaft gear 19 and the reversing driven gear 23.
  • the function and working process of the reversing mechanism of the dual multi-disc clutch is: the forward rotation of the reversing mechanism is realized through the dual multi-disc clutch 28 (the rotation direction of the input shaft 1 of the continuously variable automatic transmission is the same as the rotation direction of the clutch input shaft 17) , Reversal (the direction of rotation of the input shaft 1 of the continuously variable automatic transmission is opposite to the direction of rotation of the clutch input shaft 17) and the switching of neutral.
  • the reversing mechanism rotates forward; when the right clutch is engaged and the left clutch is disengaged in the double multi-disc clutch, the reversing mechanism is reversed ; When both the left and right clutches of the dual multi-disc clutch are disengaged, the terminal power transmission of the reversing mechanism is the neutral gear of the all-gear continuously variable automatic transmission.
  • the dual multi-disc clutch 28 in the full-gear continuously variable automatic transmission and speed ratio active control system of the present invention may adopt a dry dual multi-disc clutch or a wet dual multi-disc clutch.
  • the operation of the dual multi-disc clutch can adopt different operation modes such as electronically controlled hydraulic pressure, electronically controlled air pressure, and electronically controlled electromagnetic operation.
  • the dual clutch reversing mechanism consists of a dual clutch and a planetary gear mechanism with a lock mechanism.
  • the specific structure is as follows: the forward clutch 30 in the dual clutch is rigidly connected to the planet carrier 35 of the reversing mechanism through the hollow forward input shaft 31.
  • the reverse clutch 29 of the clutch is rigidly connected to the sun gear 32 of the shift mechanism through the reverse input shaft 38 passing through the center of the forward input shaft 31, and the shift mechanism ring gear 34 is connected to the left end of the continuously variable automatic transmission input shaft 1.
  • the front end is rigidly connected to transfer the power and movement to the differential mechanism I that distributes the torque proportionally.
  • a sun gear locker 36 and a planet carrier locker 37 are respectively installed on the reverse gear input shaft 38 and the planet carrier 35 of the reversing mechanism.
  • the function and working process of the dual-clutch reversing mechanism is: the forward rotation of the reversing mechanism is realized by the alternate combination and separation of dual clutches 29 and 30 (the direction of rotation of the input shaft 1 of the stepless automatic transmission and the direction of rotation of the input shaft 17 of the dual clutch) Same), reverse rotation (the direction of rotation of the input shaft 1 of the continuously variable automatic transmission is opposite to the direction of rotation of the dual clutch input shaft 17) and the switching of neutral.
  • the forward clutch 30 is engaged and the reverse clutch 29 is disengaged, at the same time the sun gear locker 6 is locked and the planet carrier locker 37 is unlocked, the power and movement will pass through the forward clutch 30, the forward gear input shaft 31, and the planetary gear shift mechanism.
  • the carrier 35, the planetary gear 33 of the reversing mechanism, and the gear ring 34 of the shift mechanism are transmitted to the input shaft 1 of the continuously variable automatic transmission.
  • the rotation direction of the input shaft 1 of the continuously variable automatic transmission is the same as the direction of the clutch input shaft 17, and the reversing mechanism rotates forward; if the reverse clutch 29 is engaged, the forward clutch 30 is disengaged, and the planet carrier lock 37 is locked at the same time.
  • the wheel locker 6 is unlocked, and the power and movement are transmitted to the continuously variable automatic transmission input via the reverse clutch 29, the reverse input shaft 38, the reversing mechanism sun gear 32, the reversing mechanism planetary gear 33, and the shift mechanism ring gear 34.
  • Axis 1 The rotation direction of the input shaft 1 of the continuously variable automatic transmission is opposite to the direction of the clutch input shaft 17, the reversing mechanism is reversed; the forward clutch and the reverse clutch are both separated, and the reversing mechanism interrupts the power transmission, that is, the full-gear stepless automatic Neutral gear of the transmission.
  • the dual clutches (forward gear clutch 30 and reverse gear clutch 29) in the full gear continuously variable automatic transmission and speed ratio active control system of the present invention can adopt friction type diaphragm spring clutches, multi-plate dry friction clutches, and multi-plate wet friction clutches. , Electromagnetic clutch and other different types of clutches.
  • the operation of the clutch can adopt various operation modes such as electronic control hydraulic pressure, electronic control air pressure, electronic control electromagnetic, electronic control servo motor, electronic control stepping motor, etc.
  • the planetary gear mechanism is composed of four elements: sun gear, planet gear, planet carrier and ring gear. It is a typical two-degree-of-freedom mechanism, in which the sun gear, planet carrier and ring gear can all be used for power and movement.
  • the input terminal can also be used as the output terminal of power and movement.
  • the planetary gear mechanism has the following three different transmission schemes, namely: 1 two power input ends and one power output end; 2 one power input end and two power output ends; 3 one power input end and one power output end End, a fixed end.
  • n T , n H , n Q —respectively the rotational speeds of the sun gear T, the planet carrier H and the ring gear Q;
  • the definite movement at the input includes the following two situations, namely:
  • i TH The speed ratio of the sun gear as the input end of the motion, the ring gear is fixed, and the planet carrier as the output end of the motion.
  • the number of teeth Z Q of the ring gear must be greater than the number of teeth Z T of the sun gear, that is, ⁇ >1, so:
  • i QH 1 continuously change, that is, stepless speed change can also be realized.
  • the speed ratio change range that the sun gear input can reach by adjusting the speed of the ring gear is larger than the speed ratio change range that the ring gear input can reach by adjusting the speed of the sun gear.
  • the sun gear input has a significant advantage of a large speed ratio variation range by adjusting the speed of the ring gear.
  • mechanical systems that require power driving such as driving wheels of various vehicles, propellers of ships, spindles/cutter heads of machine tools, rotary tillers, hydraulic pumps, compressors, blowers, etc.
  • power source is not only one, but more importantly: the reason why various mechanical systems that require power drive require variable speed transmission is because the dynamic characteristics of the power source (such as power, torque, etc. change with speed) are often
  • the variable transmission device needs to be used to make up for its shortcomings. That is, to realize the continuously variable transmission characteristics of the planetary gear mechanism, a second power source that can easily adjust its speed is necessary.
  • the sun gear and the planet carrier are the two motion input ends of the planetary gear mechanism, and the ring gear is the motion output end of the planetary gear mechanism, whether it is the sun gear or the planet carrier, if the motion of any one of them changes, it will be used as the output
  • the movement of the end gear ring changes accordingly. Definite movement at the input also includes the following two situations, namely:
  • i HQ The speed ratio of the planet carrier as the input end of the motion, the sun gear is fixed, and the ring gear as the output end of the motion.
  • the maximum transmission ratio of the planetary gear mechanism is The minimum transmission ratio is 3
  • the input end of the planetary gear mechanism is the sun gear and the output end is the ring gear. If the speed n H of the planet carrier continuously changes within the range of 0 and n T , the speed ratio of the planetary gear mechanism will be at the maximum speed ratio accordingly.
  • the sun gear and the planet carrier are the two motion input ends of the planetary gear mechanism, and the ring gear is the motion output end of the planetary gear mechanism
  • the planetary gear mechanism can also achieve continuously variable transmission, but due to the above serious shortcomings, and In order to realize its continuously variable transmission characteristics, it is necessary to have a second power source that can easily adjust its speed. Obviously, this is not only impossible to achieve, but also of no use value.
  • the two power output ends can be output at various speeds Power, this is how the planetary gear mechanism can be used as a differential.
  • the torque at the two input ends of the wheel and the ring gear is completely mismatched, so the symmetrical planetary gear differential cannot meet the requirements of the present invention;
  • the Torson differential and the helical tooth type limited slip differential also cannot meet the requirements of the present invention.
  • the present invention specially invents a special differential mechanism I that distributes torque proportionally (as shown in Figure 1-3), which divides the power of a single power source into two powers according to a set ratio For external output, this ratio is the gear ratio ⁇ between the ring gear and the sun gear in the planetary gear mechanism III.
  • the differential mechanism I that distributes the torque proportionally is connected in series to the front end of the planetary gear mechanism III to form the continuously variable automatic transmission device of the present invention.
  • the specific continuously variable automatic transmission principle is as follows:
  • the index circle diameters of the first differential bevel gear 6 and the second differential bevel gear 7 must be equal. If the index circle diameter of the first differential bevel gear 6 and the second differential bevel gear 7 is R, the torques M 6 and M 7 output by the differential bevel gears 6 and 7 are respectively:
  • the gear ratio of the ring gear to the sun gear in the planetary gear mechanism III is n 6 and n 7 , respectively.
  • the rotational speeds of the ring gear and the sun gear in the planetary gear mechanism III are n Q and n T , respectively. Due to the first difference
  • the speed bevel gear 6 and the second differential bevel gear 7 are rigidly connected with the ring gear 12 and the sun gear 11 in the planetary gear mechanism III through their respective shafts, namely:
  • n 6 , n 7 respectively the speeds of the first differential bevel gear 6 and the second differential bevel gear 7.
  • the maximum speed ratio (also called transmission ratio) i max of the continuously variable automatic transmission device is:
  • the minimum speed ratio i min of the continuously variable automatic transmission device is:
  • the gears with a transmission reduction ratio (also known as transmission ratio) of 1 are called direct gears.
  • the transmission of a motor vehicle is provided with an overdrive gear (a gear with a reduction ratio of less than 1).
  • the all-gear stepless automatic transmission and speed ratio active control system of the present invention can not only realize any overspeed gear of the reduction ratio, but also in the entire speed ratio change range (from the smallest overspeed gear to the maximum speed ratio) without any changes to the system structure. Stepless automatic transmission can be realized.
  • the specific implementation method is: increase the speed ratio to adjust the speed of the motor 8, so that the speed ratio is adjusted through the speed ratio of the main and driven gears 9 and 10 to the planetary gear mechanism III ring gear 12 n
  • Q is greater than the rotational speed n T of the sun gear 11 (ie, n Q > n T )
  • the required overdrive gear reduction ratio can be obtained.
  • the speed of the planet carrier in this state is n H0.5 and the reduction ratio i 0.5 of the stepless automatic transmission device.
  • the required larger speed ratio can be effectively obtained.
  • the following two effective methods can be used, namely: 1 Use two or more infinitely automatic transmissions of the present invention in series with the active speed ratio control system, so that a continuously variable automatic transmission and a speed ratio change range of any size can be obtained.
  • Active speed ratio control system 2 The planetary gear mechanism of the acceleration and deceleration mechanism at the input end of the sun gear, which is specially invented for it in the present invention, as shown in Figure 6, replaces the planetary gear mechanism III in Figures 1 to 5 .
  • the reduction mechanism V may be a fixed-axis gear train gear reduction mechanism or an asymmetric fixed-axis gear train gear reduction mechanism or a planetary gear reduction mechanism with a symmetrical structure as shown in FIG. 6.
  • the two powers transmitted to the planetary gear mechanism are respectively transmitted to the ring gear 40 and transmitted to the sun gear 42 by the input shaft 39 after being decelerated by the deceleration mechanism V.
  • the power is output through the output shaft 43 rigidly connected to the planet carrier 44.
  • the maximum speed ratio i bV max is:
  • the stepless automatic transmission range of the all-gear continuously variable automatic transmission and speed ratio active control system of the present invention is increased from the original 1 to 4 to 1.143 ⁇ 12. While the radial size of the present invention is greatly reduced, the stepless automatic transmission range is greatly increased to nearly three times the original.
  • the gear ratio (transmission ratio) of the all-gear continuously variable automatic transmission device is the largest, which is very beneficial to the working mechanism or The start-up operation of the working machine; as the output power and speed of the power source increase, the ring gear 12 starts to rotate, and the speed ratio (transmission ratio) of the all-gear continuously variable automatic transmission device changes with the change in the rotation speed of the ring gear 12, namely Automatic shifting starts. Since the rotational speed of the ring gear 12 must be continuously changed and cannot be changed in jumps, the speed ratio (transmission ratio) of the all-gear continuously variable automatic transmission device of the present invention must automatically and continuously change continuously, that is, continuously variable automatic transmission.
  • the speed ratio of the all-gear continuously variable automatic transmission device of the present invention decreases (or increases) accordingly, that is, when the resistance is not
  • the speed ratio of the all-gear continuously variable automatic transmission decreases (or increases) with the increase (or decrease) of the power, and the rotation speed of the working mechanism or working machine Synchronous rise (or fall) to achieve a good acceleration (or deceleration) effect
  • the full gear stepless automatic transmission device of the present invention The speed ratio increases (or decreases) accordingly, and the all-gear stepless automatic transmission just outputs a larger (or smaller) torque to overcome the increased (or decreased) resistance of the working mechanism or the working machinery.
  • the present invention is connected between the differential mechanism I and the planetary gear mechanism III that distribute the torque proportionally
  • the first differential bevel gear 6 and the hollow shaft 5 of the ring gear 12 are equipped with a speed ratio active control mechanism II. Because the power output end of the first differential bevel gear 6 in the differential mechanism I that distributes the torque proportionally, the ring gear 12 in the planetary gear mechanism III, and the other power output in the differential mechanism I that distributes the torque proportionally
  • the end second differential bevel gear 7 and the sun gear in the planetary gear mechanism III are always in a dynamic equilibrium state.
  • the first differential bevel gear shaft 5 is the connecting shaft of the first differential bevel gear 6 and the ring gear 12
  • a speed ratio active control mechanism II composed of a speed ratio adjusting motor 8, a speed ratio adjusting driving gear 9, and a speed ratio adjusting driven gear 10 is added. Adjusting the speed ratio according to actual needs and adjusting the speed of the motor 8 can realize stepless automatic transmission between the maximum speed ratio and the minimum speed ratio as required.
  • the working process of the full-gear stepless automatic transmission and speed ratio active control system the commutation mechanism shown in Figure 5 and the stepless automatic transmission device with speed ratio active control shown in Figure 1 are connected in series to form a
  • the full-gear stepless automatic transmission and speed ratio active control system of the present invention with the function of continuously variable automatic transmission and free switching of forward, reverse and neutral functions, refer to FIG. 4.
  • the forward rotation of the transmission is often called the forward gear
  • the reverse rotation of the transmission is often called the reverse gear
  • the interruption of the power transmission is called the neutral gear.
  • the commutation mechanism and the stepless automatic transmission device can be connected in series in two connection modes, namely: string the commutation mechanism on the front end (or input end) of the continuously variable automatic transmission device or string the commutation mechanism on the continuously variable automatic transmission device.
  • the reversing mechanism is connected to the front end (or input end) of the continuously variable automatic transmission device to achieve a good weight reduction effect.
  • only the reversing mechanism is connected to the front end (or input end) of the continuously variable automatic transmission device.
  • the switching process and principle of forward gear, reverse gear and neutral gear of the full-gear continuously variable automatic transmission and speed ratio active control system of the present invention are introduced.
  • the control system of the all-gear stepless automatic transmission and speed ratio active control system performs the following operations: the clutch 18 is disconnected, the synchronizer 24 is moved to the far left end, so that the synchronizer sleeve is engaged with the input shaft gear The meshing teeth on the right side of 19 are engaged (see Figure 7(a)), the clutch 18 is engaged, and the power and movement are transmitted to the reversing input shaft through the clutch 18 27 ⁇ synchronizer 24 ⁇ continuously variable automatic transmission input shaft 1 ⁇ differential Shell 4 ⁇ planetary bevel gear shaft 3 ⁇ planetary bevel gear 2 ⁇ differential bevel gears 6 and 7 ⁇ ring gear 12 and sun gear 11 (power and movement are divided into two paths by the differential mechanism I that distributes the torque in proportion They are respectively transmitted from the differential bevel gears 6 and 7 to the ring gear 12 and the sun gear 11 of the planetary gear mechanism III) ⁇ the planet carrier 15 ⁇ the output shaft 14 of the stepless automatic transmission. If it is necessary to perform automatic transmission (
  • the control system of the full gear stepless automatic transmission and the speed ratio active control system performs the following operations: the speed ratio adjustment motor 8 is stationary, the clutch 18 is disconnected, and the synchronizer 24 is moved to the rightmost end.
  • the synchronizer sleeve is engaged with the meshing teeth on the left side of the reversing driven gear 23 (see Figure 8(a)), the clutch 18 is combined, and the power and movement are transmitted to the reversing input shaft 27 ⁇ the input shaft gear 19 through the clutch 18 ⁇ Reversing driving gear 25 ⁇ Bridge gear 20 ⁇ Bridge gear shaft 21 ⁇ Bridge gear 22 ⁇ Reversing driven gear 23 ⁇ Synchronizer 24 ⁇ Continuously variable automatic transmission input shaft 1 ⁇ Differential housing 4 ⁇ Planet Bevel gear shaft 3 ⁇ planetary bevel gear 2 ⁇ differential bevel gears 6 and 7 ⁇ ring gear 12 and sun gear 11 (power and motion are divided into two paths by a differential mechanism that distributes torque proportionally. 6 and 7 are transmitted to the ring
  • the control system of the all-gear stepless automatic transmission and the speed ratio active control system performs the following operations: the clutch on the left side of the dual multi-disc clutch 28 is engaged, the clutch on the right side is disengaged, and the input shaft gear 19 is fixedly connected to the input shaft 1 of the continuously variable automatic transmission via the dual multi-disc clutch 28 (see Figure 7(b)), and the power and movement are transmitted to the dual multi-disc clutch 28 through the clutch input shaft 17 ⁇ the continuously variable automatic transmission Input shaft 1 ⁇ differential housing 4 ⁇ planetary bevel gear shaft 3 ⁇ planetary bevel gear 2 ⁇ differential bevel gears 6 and 7 ⁇ gear 12 and sun gear 11 (differential mechanism in which power and movement are proportionally distributed torque Divide it into two paths from the differential bevel gears 6 and 7 to the ring gear 12 and the sun gear 11 of the planetary gear mechanism III) ⁇ planet carrier 15 ⁇ continuously variable automatic transmission output shaft 14.
  • the control system of the full-gear stepless automatic transmission and the speed ratio active control system performs the following operations: the speed ratio adjustment motor is stationary and does not rotate, the clutch on the right side of the dual multi-disc clutch 28 is engaged, and the left The clutch on the side is disengaged, and the reversing driven gear 23 is fixedly connected to the input shaft 1 of the continuously variable automatic transmission via the double-linked multi-plate clutch 28 (see Figure 8(b)), and the power and movement are transmitted to the reversing through the clutch input shaft 17.
  • the control system of the all-gear stepless automatic transmission and the active ratio control system performs the following operations: the forward clutch 29 is engaged in the dual clutch, the reverse clutch 30 is disengaged, and the sun gear locker 36 Unlock, ring gear locker 37 is locked (see Figure 7(c)), power and movement are transmitted to the forward gear clutch 29 through the clutch input shaft 17 ⁇ forward gear input shaft 31 ⁇ reversing mechanism sun gear 32 ⁇ reversing mechanism Planetary gear 34 ⁇ shift mechanism planet carrier 35 ⁇ continuously variable automatic transmission input shaft 1 ⁇ differential housing 4 ⁇ planetary bevel gear shaft 3 ⁇ planetary bevel gear 2 ⁇ differential bevel gears 6 and 7 ⁇ ring gear 12 and sun Wheel 11 (power and motion are divided into two paths by the differential mechanism that distributes the torque in proportion to each other and are respectively transmitted from the differential bevel gears 6 and 7 to the ring gear 12 and the sun gear 11 of the planetary gear mechanism III) ⁇ Planet carrier 15 ⁇ Continuously variable automatic transmission output shaft 14.
  • the control system of the full-gear stepless automatic transmission and the speed ratio active control system performs the following operations: the speed ratio adjustment motor 8 is stationary and does not rotate, the reverse clutch 30 in the dual clutch is combined, and the forward clutch is engaged. 29 is separated, the ring gear locker 37 is unlocked, and the sun gear locker 36 is locked (see Figure 8(b)).

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Abstract

一种全齿轮无级自动变速与速比主动控制系统，在按比例分配转矩的差速机构（Ⅰ）和行星齿轮机构（Ⅲ）之间设置速比主动控制机构（Ⅱ），按比例分配转矩的差速机构（Ⅰ）包括两个差速锥齿轮（6、7），两个差速锥齿轮（6、7）分别和行星齿轮机构（Ⅲ）的齿圈（12）和太阳轮（11）相连；速比主动控制机构（Ⅱ）包括速比调节电机（8）、处于常啮合状态的速比调节主动齿轮（9）和速比调节从动齿轮（10）；速比调节从动齿轮（10）与其中一个差速锥齿轮轴（5、16）刚性联结，速比调节主动齿轮（9）连接速比调节电机（8）；行星齿轮（13）分别与太阳轮（11）和齿圈（12）啮合并通过行星架（15）向外部输出动力。全齿轮无级自动变速与速比主动控制系统的速比变化范围大、能有效实现无级自动变速，还能对速比主动控制。

Description

全齿轮无级自动变速与速比主动控制系统 技术领域
本发明涉及动力传动系统技术领域，具体涉及一种全齿轮无级自动变速动力传动系统。
技术背景
动力传动系统(或称装置)是一切机械中最基础的重要组成部分，其功用是将动力源(如发动机、汽轮机、水轮机或电动机等)发出的动力传给工作机构或作业机械(如各类车辆的驱动车轮、舰船的螺旋桨、机床的主轴/刀盘、旋耕机、液压泵、压缩机、鼓风机等)。由于动力源的动力学特性(如功率、转矩等随转速的变化)往往与工作机构或作业机械的需求存在交大的差异，为此动力源与工作机构或作业机械间的动力传动装置需具有减速或变速功能，这就是众多机械中的传动装置采用减速器或变速器(传动比可变的传动装置)的原因。
目前，各类机械中在用的变速器主要有手动机械有级变速器(MT)、电控机械有级自动变速箱(AMT)、带式无级自动变速器(CVT)、液力自动变速器(AT)、双离合器自动变速器(DCT)等多种，其中：⑴手动机械有级变速器(MT)的缺点是，变速器的档位数(变速的级数)有限，最常用的是4～6档，对于档位数大于6的机械有级变速器，通常采用主、副变速器的结构。对于4～6档的机械有级变速器，不仅每一个档位间的速比有较大的级差，使得动力源不可能始终工作在高效区，且换档过程速比阶梯变化，因此其动力传动特性不够理想；对于档位数大于6的主副变速器，尽管动力源工作在高效区的机率有所增大、每一个档位间的速比级差得以减小，但不仅换档过程速比阶梯变化、动力传动特性不够理想等问题依然存在，而且由于档位数的增加，还使得换档操作更加复杂，对操作者的技术水平提出了更高的要求。⑵电控机械有级自动变速器(AMT)是手动机械有级变速器的改进版，变速传动部分和原机械手动有级变速器几乎完全相同，唯一不同的是换档操作由电控液压或电控气压执行机构来完成，原手动机械有级变速器所固有的“每一个档位间速比级差大、换档过程速比阶梯变化、动力传动特性不够理想”等缺点都同样存在，只是换档无需手动操作。正因为无论是手动机械有级变速器(MT)还是电控机械有级自动变速器(AMT)都具有上述诸多缺点，所以人们一直希望用无级变速器取代机械有级变速器。⑶带式无级自动变速器(CVT)中主从动带轮直径大小的自动调节和液力自动变速器(AT)中的液力变矩器具有无级变速的功能，但二者的共同缺点是：①速比变化范围小，常需与另一个有级机械变速器配合使用才能满足各类机械中对速比变化范围的要求；②传动效率低、能耗大；③结构复杂、制造成本高；④对于CVT还存在传递转矩能力十分有限、不能传递大转矩的缺点，因此只是在小型车辆及小型设备上得到应用。⑷近些年才发展起来的“双离合器自动变速器(DCT)”，事实上也是一种机械有级自动变速器，它和电控机械有级自动变速器(AMT)的最大不同是：它将两个机械有级变速器并联起来使用，利用两个离合器和电控液压或电控气压或电控电动换档操纵机构完成换档操作，由于用到了两个机械有级变速器，因此变速器的档位数得以增加，每一个档位间的速比级差得以减小；由于利用两个离合器配合换档，因此换档过程得以缩短，换档顿挫感有所降低。尽管如此，但由于双离合器自动变速器(DCT)也是一种机械有级变 速器，因此其“每一个档位间速比有级差、换档过程速比阶梯变化、动力传动特性不够理想”的缺点仍然存在，只是上述缺点在一定程度上有所减小。此外，还由于双离合器自动变速器(DCT)的结构与工艺都比较复杂，因此制造成本高、使用环境的适应性较差。
发明内容
本发明要解决的技术问题是彻底克服现有技术中各类变速器的不足，提供一种全齿轮无级自动变速与速比主动控制系统，结构简单、体积小重量轻、传动效率高、速比变化范围大、传递转矩的能力可以满足所有机械传动的要求，不仅在整个速比变化范围内有效实现无级自动变速，还可以实现速比的主动控制满足各类机械传动的需要。
为解决上述技术问题，本发明采用如下技术方案：
一种全齿轮无级自动变速与速比主动控制系统，其特征在于：按比例分配转矩的差速机构串联在行星齿轮机构的前端构成无级自动变速装置；在按比例分配转矩的差速机构和行星齿轮机构之间设置速比主动控制机构；按比例分配转矩的差速机构包括两个设置在动力与运动输出端的差速锥齿轮：第一差速锥齿轮和第二差速锥齿轮，第一差速锥齿轮通过空心的第一差速锥齿轮轴与行星齿轮机构的齿圈刚性相连；第二差速锥齿轮轴穿过空心的第一差速锥齿轮轴，第二差速锥齿轮通过第二差速锥齿轮轴与行星齿轮机构的太阳轮相连；
速比主动控制机构包括速比调节电机、处于常啮合状态的速比调节主动齿轮和速比调节从动齿轮；速比调节从动齿轮与第一差速锥齿轮轴刚性联结，速比调节主动齿轮安装在速比调节电机的输出轴上；
行星齿轮机构的两个输入端分别为位于中心的太阳轮和最外围的齿圈，行星齿轮同时与太阳轮和齿圈啮合并通过行星架向外部输出动力。
进一步的，按比例分配转矩的差速机构包括差速器壳，无级自动变速装置输入轴在差速器壳前端与差速器壳刚性连接；差速器壳内，第一差速锥齿轮和第二差速锥齿轮均可旋转地支承在差速器壳沿着前后端或者左右方向延伸的轴孔中；第二差速锥齿轮和第一差速锥齿轮分别在前后端与行星锥齿轮啮合；第一差速锥齿轮轴的前端与第一差速锥齿轮刚性联结；第二差速锥齿轮轴前端与第二差速锥齿轮刚性联结，后端从空心的第一差速锥齿轮轴中穿过，并穿过差速器壳与太阳轮刚性相连或形成与太阳轮轴一体的结构。
进一步的，行星齿轮机构中，行星锥齿轮上开有绕行星锥齿轮轴中心线的周向环形槽，行星锥齿轮轴的一端固定安装在差速器壳上，另一端通过轴承安装在环形槽内，行星锥齿轮自转时转动中心是环形槽的中心圆弧线，即行星锥齿轮轴中心线是相对于行星锥齿轮在环形槽中的周向运动轨迹。
进一步的，当行星锥齿轮自转时，设行星锥齿轮与第二差速锥齿轮的啮合点A至行星锥齿轮轴中心线的距离为S 1，行星锥齿轮与第一差速锥齿轮的啮合点B至行星锥齿轮轴中心线的距离为S 2，S 1和S 2之比始终为设定的比例。
进一步的，按比例分配转矩的差速机构采用至少2只行星锥齿轮。
进一步的，所述的行星齿轮机构中设置至少2个行星齿轮。
进一步的，两个或两个以上所述全齿轮无级自动变速与速比主动控制系统串联连接来增 大全齿轮无级自动变速与速比主动控制系统的速比变化范围；
或者在所述的行星齿轮机构中，在太阳轮之前设置减速机构来增大全齿轮无级自动变速与速比主动控制系统的速比变化范围，所述减速机构为对称定轴轮系齿轮减速机构、或非对称定轴轮系齿轮减速机构、或行星齿轮减速机构中的其中一种，第二差速锥齿轮轴经减速机构减速后连接太阳轮，所述减速机构串联在太阳轮之前。
进一步的，无级自动变速装置的前端或后端串接换向机构；实现动力与运动传输的正转、反转和中断至少三种模式的切换；换向机构为离合器和同步器的组合结构、或双联多片离合器形式的换向机构、或双离合器形式的换向机构。
进一步的，采用离合器和同步器组合的换向机构时，离合器输入轴和变速器输入轴成直线设置，离合器输入轴上设置输入轴齿轮，无级变速装置输入轴的左端或前端通过轴承安装在输入轴齿轮右端的轴承座孔内；换向主动齿轮与输入轴齿轮常啮合，同时与前一组过桥齿轮也是常啮合；后一组过桥齿轮与前一组过桥齿轮同轴且均与过桥齿轮轴刚性联结，后一组过桥齿轮与换向从动齿轮常啮合，换向从动齿轮利用滚针轴承或滑动轴承空套在无级自动变速装置输入轴上，输入轴齿轮和换向从动齿轮之间装有锁环式或锁销式同步器。
离合器采用摩擦式模片弹簧离合器、摩擦式周置圆柱螺旋弹簧离合器、摩擦式中央圆锥螺旋弹簧离合器、多片干式摩擦离合器、多片湿式摩擦离合器、电磁离合器中的其中一种，离合器的操作采用电控液压、电控气压、电控电磁、电控伺服电机、电控步进电机操作方式中的一种。
进一步的，采用双联多片离合器形式的换向机构时，离合器输入轴上设置输入轴齿轮，换向从动齿轮通过滚针轴承或滑动轴承空套在无级自动变速装置输入轴上，输入轴齿轮和换向从动齿轮之间装有双联多片离合器；无级自动变速装置输入轴的左端或前端通过轴承安装在离合器输入轴齿轮右端的轴承座孔内；换向主动齿轮与输入轴齿轮常啮合，同时与过桥齿轮也是常啮合；前后两组过桥齿轮均与过桥齿轮轴刚性联结，后端的过桥齿轮与换向从动齿轮常啮合。
双联多片离合器采用干式双联多片离合器或湿式双联多片离合器；双联多片离合器的操作采用电控液压、电控气压、电控电磁操作方式中的一种。
进一步的，采用双离合器换向机构时，双离合器换向机构由双离合器和带锁止机构的行星齿轮机构组成，双离合器由前进档离合器和倒档离合器组成，前进档离合器通过空心的前进档输入轴与换向机构行星架刚性连接，倒档离合器通过从前进档输入轴中心穿过的倒档输入轴与换向机构太阳轮刚性连接，换档机构齿圈与无级自动变速装置输入轴的左端或前端刚性联结，将动力和运动传给按比例分配转矩的差速机构；在倒档输入轴和换向机构行星架上分别装有太阳轮锁止器和行星架锁止器。
双离合器采用摩擦式模片弹簧离合器、多片干式摩擦离合器、多片湿式摩擦离合器、电磁离合器中的至少一种；离合器的操作采用电控液压、电控气压、电控电磁、电控伺服电机、电控步进电机操作方式中的一种。
相对于现有技术，本发明全齿轮无级自动变速与速比主动控制系统，在按比例分配转矩 的差速机构和行星齿轮机构之间设置速比主动控制机构，按比例分配转矩的差速机构包括两个设置在动力与运动输出端的差速锥齿轮，两个差速锥齿轮分别和行星齿轮机构的齿圈和太阳轮相连作为行星齿轮机构的输入轴；速比主动控制机构包括速比调节电机、处于常啮合状态的速比调节主动齿轮和速比调节从动齿轮；速比调节从动齿轮与其中一个差速锥齿轮轴刚性联结，速比调节主动齿轮与速比调节电机的输出轴连接；行星齿轮同时与太阳轮和齿圈啮合并通过行星架向外部输出动力。速比变化范围大、传递转矩的能力可以满足所有机械传动的要求，不仅在整个速比变化范围内有效实现无级自动变速，还可以实现速比的主动控制，满足各类机械传动的需要。
具有如下有益效果：
(1)全齿轮无级自动变速与速比主动控制系统所有传动部件都是效率极高的齿轮与轴，因此具有和机械式手动变速器一样的高效率。
(2)速比变化范围大，可以满足各类不同机械系统对速比变化范围的要求。行星齿轮机构中在太阳轮之前设置减速机构来增大全齿轮无级自动变速与速比主动控制系统的速比变化范围或者将两个或两个以上的本发明无极自动变速与速比主动控制系统串联使用，可以得到速比变化范围为任意大小的无级自动变速与速比主动控制系统。比如，在需要速比变化范围较大时，采用在太阳轮之前设置减速机构的行星齿轮机构，在显著减小齿圈径向尺寸的同时还可以增大全齿轮无级自动变速与速比主动控制系统的速比变化范围。如齿圈与太阳轮的齿数比α由原来的7减小到5，无极自动变速范围却由原来的1～4增大到1.143～12，无极自动变速范围却大幅增加到了原来的近3倍。
(3)在整个速比变化范围内，可有效实现无级自动变速。
(4)不仅可以自动无级变速，而且还可以根据使用过程中的实际需要，实现速比的主动控制。
(5)传递转矩的能力强，其传递转矩的能力仅与结构尺寸的大小有关，因此其传递转矩的能力可以满足所有机械传动的要求。
(6)结构和工艺都很简单，易于制造、生产成本低、使用环境的适应性好。
(7)体积小、重量轻。
本发明“全齿轮无级自动变速与速比主动控制系统”的上述诸多优点，正好可以彻底克服现有各类变速器的不足，可满足各类机械传动的需要。
附图说明
图1为本发明全齿轮无级自动变速与速比主动控制系统的结构示意图。
图2为行星锥齿轮结构示意图。
图3为按比例分配转矩的差速机构结构示意图。
图4为本发明具有换向机构的全齿轮无级自动变速与速比主动控制系统结构示意图。
图5为换向机构结构示意图。
图6为太阳轮输入端加减速机构的行星齿轮机构另一实施例结构示意图。
图7为前进挡状态下结构示意图(换向机构三种实施方式)。
图8为倒挡状态下结构示意图(换向机构三种实施方式)。
图9为空挡状态下结构示意图(换向机构三种实施方式)。
附图标记说明如下：
Ⅰ—按比例分配转矩的差速机构；Ⅱ—速比主动控制机构；Ⅲ—行星齿轮机构；IV—换向机构；1—无级自动变速装置输入轴；2—行星锥齿轮；3—行星锥齿轮轴；4—差速器壳；5—差速锥齿轮轴；6、7—差速锥齿轮；8—速比调节电机；9—速比调节主动齿轮；10—速比调节从动齿轮；11—太阳轮；12—齿圈；13—行星齿轮；14—动力输出轴；15—行星架；16—太阳轮轴(差速锥齿轮轴)；17—离合器输入轴；18—离合器；19—输入轴齿轮；20—过桥齿轮I；21—过桥齿轮轴；22—过桥齿轮II；23—换向从动齿轮；24—同步器；25—换向主动齿轮；26—换向主动齿轮轴；27—换向输入轴；28—双联多片离合器；29—倒档离合器；30—前进档离合器；31—前进档输入轴；32—换向机构太阳轮；33—换向机构行星轮；34—换向机构齿圈；35—换档机构行星架；36—太阳轮锁止器；37—行星架锁止器；38—倒档输入轴；39—减速器输入轴；40—齿圈；41—行星轮；42—太阳轮；43—输出轴；44—行星架；45—轴承；46—环形槽；47—环形槽中心圆弧线。
具体实施方式
根据本发明实施的全齿轮无级自动变速与速比主动控制系统的具体结构见图1。全齿轮无级自动变速与速比主动控制系统至少包括顺次连接的三大部分：按比例分配转矩的差速机构Ⅰ、速比主动控制机构Ⅱ、行星齿轮机构Ⅲ。
其中，按比例分配转矩的差速机构Ⅰ、行星齿轮机构Ⅲ两大部分一起共同构成无级自动变速装置。按比例分配转矩的差速机构Ⅰ包括两个动力与运动输出端差速锥齿轮：第一差速锥齿轮6和第二差速锥齿轮7；第一差速锥齿轮6通过第一差速锥齿轮轴5与行星齿轮机构Ⅲ的齿圈12相连，连接方式可以是键或花键；第二差速锥齿轮7通过第二差速锥齿轮轴16与行星齿轮机构Ⅲ太阳轮11相连，连接方式可以是键或花键；第二差速锥齿轮轴16作为中心轴在按比例分配转矩的差速机构Ⅰ、速比主动控制机构Ⅱ中使用，第二差速锥齿轮轴16从空心的第一差速锥齿轮轴5中穿过(同轴心线而不同轴)，第一差速锥齿轮轴5贯穿于速比主动控制机构Ⅱ且分别延伸至按比例分配转矩的差速机构Ⅰ和行星齿轮机构Ⅲ。在无级自动变速装置的基础上加上速比主动控制机构Ⅱ形成具有速比主动控制的无级自动变速装置，如图1所示，则可实现无级自动变速和速比主动控制功能。
行星齿轮机构Ⅲ结构如下：
图1具有速比主动控制的无级自动变速装置中，行星齿轮机构Ⅲ由无级自动变速装置输入轴第二差速锥齿轮轴(或太阳轮轴)16、太阳轮11、行星架15、行星齿轮13和齿圈12组成，其中太阳轮11和齿圈12是行星齿轮机构Ⅲ的两个输入端，行星齿轮机构Ⅲ的动力由行星架15输出。太阳轮11通过键或花键与第二差速锥齿轮轴(太阳轮轴)16刚性联结，齿圈12通过键或花键与空心结构的第一差速锥齿轮轴5刚性联结。图2的实施例中设置了2只行 星齿轮13。基于提高行星齿轮机构传递转矩能力的考虑，可以采用设置3只行星齿轮13、4只行星齿轮13、5只行星齿轮13或6只行星齿轮13等多种不同的结构设计。本领域技术人员可以根据传递转矩能力使用需求选用。
按比例分配转矩的差速机构Ⅰ结构如下：
如图1所示，按比例分配转矩的差速机构Ⅰ串联在行星齿轮机构Ⅲ的前端，包括按比例分配转矩的差速器壳4，差速器壳4外，无级自动变速装置输入轴1在其前端与差速器壳4刚性连接，差速器壳4内，第一差速锥齿轮6和第二差速锥齿轮7均通过轴承支承在差速器壳4的前后端延伸的轴孔中。第二差速锥齿轮7和第一差速锥齿轮6分别在前后端与行星锥齿轮2啮合。空心的第一差速锥齿轮轴5的前端通过键或花键与第一差速锥齿轮6刚性联结；第一差速锥齿轮6通过第一差速锥齿轮轴5与行星齿轮机构Ⅲ的齿圈12刚性相连。
第二差速锥齿轮轴16(太阳轮轴16)从空心的第一差速锥齿轮轴5中穿过，前端通过键或花键与第二差速锥齿轮7刚性联结，后端穿过差速器壳4并通过键或花键与太阳轮11刚性相连(太阳轮11也可以做成与太阳轮轴16一体的结构)。
如图2所示，行星锥齿轮2上开有环形槽46，行星锥齿轮轴3的一端固定安装在差速器壳4的孔中(参见图1，图2中未有显示)，另一端通过轴承45(可以是圆柱滚子轴承、滚针轴承或滑动轴承)安装在行星锥齿轮2的环形槽46内，如图2(a)所示。行星锥齿轮2自转时，其转动中心不是行星锥齿轮2的中心线，而是环形槽46的中心圆弧线47(图2(b)中点划线表示)。为了实现差速器按照设定的比例将转矩分配给第一差速锥齿轮6和第二差速锥齿轮7，当行星锥齿轮2自转时，设行星锥齿轮2与第二差速锥齿轮7的啮合点A至行星锥齿轮轴3中心线的距离S 1、行星锥齿轮2与第一差速锥齿轮6的啮合点B至行星锥齿轮轴3中心线的距离S 2，S 1和S 2之比始终为设定的比例(见图3)。基于提高差速器传递转矩能力的考虑，可以采用2只行星锥齿轮、3只行星锥齿轮、4只行星锥齿轮、5只行星锥齿轮或6只行星锥齿轮等多种不同的结构设计，本领域技术人员可以根据转矩荷载能力匹配。
速比主动控制机构结构如下：
速比主动控制机构Ⅱ由速比调节电机8、速比调节主动齿轮9和速比调节从动齿轮10组成。速比调节从动齿轮10通过键或花键与空心结构的第一差速锥齿轮轴5刚性联结，速比调节主动齿轮9通过键或花键固定安装在速比调节电机8的输出轴上，速比调节主动齿轮9和速比调节从动齿轮10处于常啮合状态。
换向机构IV结构如下：工程上，许多工作机构或作业机械在工作过程中经常需要换向运转，即有时正转、有时反转。实现其正转与反转有两种方法，即：①让动力源正向与反向运转；②在变速器上增设一换向机构。方法①只有电动机这类动力源可以实现，绝大多数其它类型的动力源都难以实现。
作为优选项，本发明全齿轮无级自动变速与速比主动控制系统增设了换向机构IV。换向机构可以采用离合器+同步器的换向机构(见图5(a))或双联多片离合器的换向机构(见图5(b))或双离合器换向机构(见图5(c))等多种不同结构的方案。分别对应图4的4(a)～4(c)三种具有换向机构的全齿轮无级自动变速与速比主动控制系统实施结构。4(a)为本 发明具有换向机构的全齿轮无级自动变速与速比主动控制系统第一种实施方式结构示意图，4(b)为本发明具有换向机构的全齿轮无级自动变速与速比主动控制系统第二种实施方式结构示意图，4(c)为本发明具有换向机构的全齿轮无级自动变速与速比主动控制系统第三种实施方式结构示意图。由此能方便改变运动传动方向，实现前进档和倒退档的切换。换向机构IV可以串接在无级自动变速装置的前端也可串接在无级自动变速装置的后端，其功能是实现动力与运动传输的正转、反转和中断等三种模式的切换。
具有离合器+同步器组成的换向机构的全齿轮无级自动变速与速比主动控制系统实施例结构如下：
参见图4(a)和5(a)，离合器+同步器组成的换向机构IV由离合器输入轴17、离合器18、输入轴齿轮19、换向主动齿轮25、换向主动齿轮轴26、同步器24、过桥齿轮与轴(包括过桥齿轮20和22、过桥齿轮轴21)、换向从动齿轮23和变速器输入轴27等组成，如图4(a)和5(a)所示。图5(a)中左侧为输入轴齿轮19、过桥齿轮20和换向主动齿轮25三者间沿着与轴向方向垂直截面图；右侧图幅中，无级变速装置输入轴1的左端或前端通过轴承安装在输入轴齿轮19右端的轴承座孔内；换向主动齿轮26与输入轴齿轮19常啮合，同时与桥齿轮20也是常啮合；过桥齿轮20和过桥齿轮22通过键或花键与过桥齿轮轴21刚性联结，过桥齿轮22与换向从动齿轮23常啮合，换向从动齿轮23利用滚针轴承或滑动轴承空套在无级自动变速装置输入轴1上，输入轴齿轮19和换向从动齿轮23之间装有锁环式(或锁销式)同步器24。离合器+同步器组成的换向机构IV的功能和工作过程是：通过同步器24实现换向机构正转(无级自动变速装置输入轴1的转动方向与离合器输入轴17的转动方向相同)、反转(无级自动变速装置输入轴1的转动方向与离合器输入轴17的转动方向相反)及空挡的切换。当需要进行换向操作时，先将离合器18分离并将同步器24退出原来所在的档位，挂入待进档位后再将离合器18结合。同步器24处在中间位置时，换向机构终断动力的传递，即全齿轮无级自动变速与速比主动控制系统处于空档；同步器24左移到最左端，换向机构正转；同步器24右移到最右端，换向机构反转。
本发明全齿轮无级自动变速与速比主动控制系统中的离合器18可以采用摩擦式模片弹簧离合器、摩擦式周置圆柱螺旋弹簧离合器、摩擦式中央圆锥螺旋弹簧离合器、多片干式摩擦离合器、多片湿式摩擦离合器、电磁离合器等各种不同类型的离合器。离合器的操作可以采用电控液压、电控气压、电控电磁、电控伺服电机、电控步进电机等各种不同的操作方式。
具有双联多片离合器换向机构的全齿轮无级自动变速与速比主动控制系统实施例结构如下：
参见图4(b)和5(b)，双联多片离合器换向机构由双联多片离合器28、离合器输入轴17、输入轴齿轮19、换向主动齿轮25、换向主动齿轮轴26、过桥齿轮与轴(包括过桥齿轮20和22、过桥齿轮轴21)和换向从动齿轮23等组成，如图5(b)所示。图5(b)左侧为输入轴齿轮19、过桥齿轮20和换向主动齿轮25三者间沿着与轴向方向垂直截面图；右侧图幅中，无级自动变速装置输入轴1的左端或前端通过轴承安装在输入轴齿轮19右端的轴承座孔内；换向主动齿轮25与输入轴齿轮19常啮合，同时与过桥齿轮20也是常啮合；过桥齿轮 20和过桥齿轮22通过键或花键与过桥齿轮轴21刚性联结，过桥齿轮22与换向从动齿轮23常啮合，换向从动齿轮23通过滚针轴承或滑动轴承空套在无级自动变速装置输入轴1上，输入轴齿轮19和换向从动齿轮23之间装有双联多片离合器28。双联多片离合器换向机构的功能和工作过程是：通过双联多片离合器28实现换向机构正转(无级自动变速装置输入轴1的转动方向与离合器输入轴17的转动方向相同)、反转(无级自动变速装置输入轴1的转动方向与离合器输入轴17的转动方向相反)及空挡的切换。双联多片离合器28中左侧的离合器结合、右侧的离合器分离时，换向机构正转；双联多片离合器中右侧的离合器结合、左侧的离合器分离时，换向机构反转；双联多片离合器中左、右侧的离合器都分离时，换向机构终端动力的传递，即全齿轮无级自动变速装置的空档。
本发明全齿轮无级自动变速与速比主动控制系统中双联多片离合器28可以采用干式双联多片离合器或湿式双联多片离合器。双联多片离合器的操作可以采用电控液压、电控气压、电控电磁操作等不同的操作方式。
具有双离合器换向机构的全齿轮无级自动变速与速比主动控制系统实施例结构如图4(c)和5(c)所示：
双离合器换向机构由双离合器和带锁止机构的行星齿轮机构组成，具体结构如下：双离合器中的前进档离合器30通过空心的前进档输入轴31与换向机构行星架35刚性连接，双离合器中的倒档离合器29通过从前进档输入轴31中心穿过的倒档输入轴38与换向机构太阳轮32刚性连接，换档机构齿圈34与无级自动变速装置输入轴1的左端或前端刚性联结，将动力和运动传给按比例分配转矩的差速机构I。为了实现前进档与倒档的自由切换，在倒档输入轴38和换向机构行星架35上分别装有太阳轮锁止器36和行星架锁止器37。双离合器换向机构的功能和工作过程是：通过双离合器29和30的交替结合与分离实现换向机构正转(无级自动变速装置输入轴1的转动方向与双离合器输入轴17的转动方向相同)、反转(无级自动变速装置输入轴1的转动方向与双离合器输入轴17的转动方向相反)及空挡的切换。若前进档离合器30结合、倒档离合器29分离，同时太阳轮锁止器6锁止、行星架锁止器37解锁，动力和运动经前进档离合器30、前进档输入轴31、换向机构行星架35、换向机构行星轮33、换档机构齿圈34传给无级自动变速装置输入轴1。无级自动变速装置输入轴1的转动方向与离合器输入轴17的方向相同，换向机构正转；若倒档离合器29结合、前进档离合器30分离，同时行星架锁止器37锁止、太阳轮锁止器6解锁，动力和运动经倒档离合器29、倒档输入轴38、换向机构太阳轮32、换向机构行星轮33、换档机构齿圈34传给无级自动变速装置输入轴1。无级自动变速装置输入轴1的转动方向与离合器输入轴17的方向相反，换向机构反转；前进档离合器和倒档离合器都分离，换向机构中断动力的传递，即全齿轮无级自动变速装置的空档。
本发明全齿轮无级自动变速与速比主动控制系统中的双离合器(前进档离合器30和倒档离合器29)可以采用摩擦式膜片弹簧离合器、多片干式摩擦离合器、多片湿式摩擦离合器、电磁离合器等各种不同类型的离合器。离合器的操作可以采用电控液压、电控气压、电控电磁、电控伺服电机、电控步进电机等各种不同的操作方式。
本发明全齿轮无级自动变速原理介绍：
本发明全齿轮无级自动变速与速比主动控制系统的无级自动变速原理如下：
行星齿轮机构由太阳轮、行星轮、行星架和齿圈等四大元件组成，是一种典型的二自由度机构，其中太阳轮、行星架和齿圈三个元件均既可作动力和运动的输入端也可作动力和运动的输出端。如此，行星齿轮机构就有如下三种不同的传动方案，即：①二个动力输入端、一个动力输出端；②一个动力输入端、二个动力输出端；③一个动力输入端、一个动力输出端、一个固定端。
二个动力输入端、一个动力输出端的行星齿轮机构:
只有当行星齿轮机构的三个动力和运动的输入/输出元件中的任意二个元件有确定的运动输入(包括零输入)时，另一个元件才有确定的运动输出，且此运动的输出随二个运动输入端中任何一个输入端运动的变化而变化，其原理式如下：
n T+αn Q-(1+α)n H＝0     (1)
式中：n T,n H,n Q——分别是太阳轮T、行星架H和齿圈Q的转速；
α——齿圈与太阳轮的齿数比，
Figure PCTCN2019109084-appb-000001
z T,z Q——分别是太阳轮T和齿圈Q的齿数。
1、若设太阳轮和齿圈作为行星齿轮机构的两个运动输入端、行星架为行星齿轮机构的运动输出端(本发明中的行星齿轮机构Ⅲ属这种类型)，无论是太阳轮还是齿圈，其中任何一个的运动发生变化，则作为输出端行星架的运动随之发生变化。输入端有确定的运动包括如下两种情况，即：
1)二者之一的转速为零，另一个输入端的转速自由变化。此种情况下又有n Q＝0,n T和n T＝0,n Q二种不同的模式。
(1)n Q＝0,n T，将其带入式(1)
Figure PCTCN2019109084-appb-000002
式中：i TH——太阳轮为运动的输入端、齿圈固定、行星架为运动的输出端的速比。
(2)n T＝0,n Q，将其带入式(1)得：
Figure PCTCN2019109084-appb-000003
式中：i QH——齿圈为运动的输入端、太阳轮固定、行星架为运动的输出端的速比。
对于行星齿轮机构，齿圈的齿数Z Q必然大于太阳轮的齿数Z T，即α＞1，因此：
i TH＞i QH    (4)
2)两个输入端的转速相同，均为n(n T＝n Q＝n i＝n)，此时，行星架的转速亦必然和太阳轮及齿圈的转速相等，即n H＝n Q＝n T＝n，将其带入式(1)得：
Figure PCTCN2019109084-appb-000004
Figure PCTCN2019109084-appb-000005
比较式(2)、(3)、(4)、(5)、(6)得：①行星齿轮机构可以有三个固定的速比(又称传动比)，分别是
Figure PCTCN2019109084-appb-000006
i TH＝1、i QH＝1；②由于齿圈与太阳轮的齿数比必然大于1(α＞1)，因此行星齿轮机构的最大传动比为
Figure PCTCN2019109084-appb-000007
最小传动比为i TH＝1；③设行星齿轮机构的输入端为太阳轮、输出端为行星架，若齿圈的转速n Q在0和n T的范围内连续变动，则行星齿轮机构的速比随之在最大速比
Figure PCTCN2019109084-appb-000008
和最小速比i TH＝1之间连续变动；设行星齿轮机构的输入端为齿圈、输出端为行星架，若太阳轮的转速n T在0和n Q的范围内连续 变动，则行星齿轮机构的速比随之在
Figure PCTCN2019109084-appb-000009
和最小速比i QH＝1之间连续变动。由此可见：若动力和运动由太阳轮输入、行星架输出，只要齿圈能在0和n T的范围内连续调节其转速，行星齿轮机构的速比就可在最大速比
Figure PCTCN2019109084-appb-000010
和最小速比i TH＝1之间连续变化，即实现无级变速；若动力和运动由齿圈输入、行星架输出，只要太阳轮能在0和n Q的范围内连续调节其转速，行星齿轮机构的速比就可在
Figure PCTCN2019109084-appb-000011
和i QH＝1之间连续变化，即同样可以实现无级变速。由于齿圈的齿数是一定大于太阳轮的齿数的，即
Figure PCTCN2019109084-appb-000012
因此，太阳轮输入通过调节齿圈的转速所能达到的速比变化范围比齿圈输入通过调节太阳轮的转速所能达到的速比变化范围要大。如：若设齿圈与太阳轮的齿数比α＝4，太阳轮输入通过调节齿圈的转速所能达到的速比变化范围是i TH＝1□5(最大速比i TH＝1+α＝1+4＝5)；齿圈输入通过调节太阳轮的转速所能达到的速比变化范围是i QH＝1□1.25(最大速比
Figure PCTCN2019109084-appb-000013
)。由此可见，对于同一个行星齿轮机构，太阳轮输入通过调节齿圈的转速具有速比变化范围大的显著优势。但对于各类需要动力驱动的机械系统(如各类车辆的驱动车轮、舰船的螺旋桨、机床的主轴/刀盘、旋耕机、液压泵、压缩机、鼓风机等)，绝大多数情况下，其动力源不仅只有一个，而且更重要的是：各类需要动力驱动的机械系统之所以需要变速传动，其原因在于动力源的动力学特性(如功率、转矩等随转速的变化)往往与工作机构或作业机械的需求存在交大的差异，需借用变速传动装置弥补其不足，即若要实现行星齿轮机构的无级变速特性就必须要有可以十分方便调节其速度的第二动力源。
2、若设太阳轮和行星架为行星齿轮机构的两个运动输入端、齿圈为行星齿轮机构的运动输出端，无论是太阳轮还是行星架，其中任何一个的运动发生变化，则作为输出端齿圈的运动随之发生变化。输入端有确定的运动同样包括如下两种情况，即：
1)二者之一的转速为零，另一个输入端的转速自由变化。此种情况下又有n H＝0,n T和n T＝0,n H二种不同的模式。
(1)n H＝0,n T，将其带入式(1)得：
Figure PCTCN2019109084-appb-000014
式中：i TQ——太阳轮为运动的输入端、行星架固定、齿圈为运动的输出端的速比，
Figure PCTCN2019109084-appb-000015
前面的“-”是指输出端齿圈的转速与输入端太阳轮的转速方向相反。
(2)n T＝0,n H，将其带入式(1)得：
Figure PCTCN2019109084-appb-000016
式中：i HQ——行星架为运动的输入端、太阳轮固定、齿圈为运动的输出端的速比。
2)两个输入端的转速相同，均为n i(n T＝n Q＝n i＝n)，此时，行星架的转速亦必然和太阳轮及齿圈的转速相等，即n Q＝n Q＝n T＝n，将其带入式(1)得：
Figure PCTCN2019109084-appb-000017
Figure PCTCN2019109084-appb-000018
比较式(7)、(8)、(9)、(10)得：①行星齿轮机构可以有三个固定的速比(又称传动比)，分别是
Figure PCTCN2019109084-appb-000019
i TQ＝1、i HQ＝1；②行星齿轮机构的最大传动比为
Figure PCTCN2019109084-appb-000020
最小传动比为
Figure PCTCN2019109084-appb-000021
③设行星齿轮机构的输入端为太阳轮、输出端为齿圈，若行星架的转速n H在0和n T的范围内 连续变动，则行星齿轮机构的速比随之在最大速比
Figure PCTCN2019109084-appb-000022
和i TQ＝1之间连续变动；设行星齿轮机构的输入端为行星架、输出端为齿圈，若太阳轮的转速n T在0和n H的范围内连续变动，则行星齿轮机构的速比随之在
Figure PCTCN2019109084-appb-000023
和i HQ＝1之间连续变动。由此可见：若动力由太阳轮输入、齿圈输出，只要行星架能在0和n T的范围内连续调节其转速，行星齿轮机构的速比就可在最大速比
Figure PCTCN2019109084-appb-000024
和i TQ＝1之间连续变动，但这种速比变化完全不符合机械传动系统的要求，因为在变速的过程中速度输出端的转动方向要发生变化；若动力由行星架输入、齿圈输出，只要太阳轮能在0和n H的范围内连续调节其转速，行星齿轮机构的速比就可在
Figure PCTCN2019109084-appb-000025
和i HQ＝1之间连续变动，但速比变化范围非常有限。由此可见，尽管太阳轮和行星架为行星齿轮机构的两个运动输入端、齿圈为行星齿轮机构的运动输出端的行星齿轮机构也可实现无级变速，但由于存在上述严重之不足，且若要实现其无级变速特性还必须要有可以十分方便调节其速度的第二动力源，显然这不仅无法实现也无使用价值。
一个动力输入端、二个动力输出端的行星齿轮机构
若动力由行星齿轮机构三个元件中任意一个元件(如行星架)输入，动力从另二个元件(如齿圈和太阳轮)输出，则二个动力输出端可以以各种不同的转速输出动力，这就是行星齿轮机构可以用作差速器的原理。
由前面的分析知，要想利用行星齿轮机构实现无级自动变速：①太阳轮和齿圈两个输入、行星架输出的方案具有速比变化范围大的显著优势，其最大传动比为1+α，最小传动比为1；②必须要有两个可供输入的动力。
对于需要变速传动的各类动力机械系统，绝大多数情况下只有一个动力源，然而要想利用行星齿轮机构实现无级变速，必须要有两个可供输入的动力，要想解决这一问题，很容易想到差速器，但到目前为止尚没有任何一款能符合本发明要求的差速器。因为：①对于对称式行星齿轮差速器，尽管具有传动效率高的显著优点，但由于两个输出端只能输出数值相等的转矩，而由前面的分析可知，较适合用于实现无级变速的是太阳轮和齿圈两端输入、行星架输出的行星齿轮机构，此行星齿轮机构齿圈输入端的转矩应为太阳轮输入端转矩的α倍 (即M Q＝αM T，α—齿圈与太阳轮的齿数比，M Q、M T—分别是齿圈和太阳轮端的输入转矩)，由于对称式行星齿轮差速器两个输出端的转矩与行星齿轮机构Ⅲ中太阳轮和齿圈两输入端的转矩完全不匹配，因此对称式行星齿轮差速器不能满足本发明的使用要求；②尽管托森差速器和螺旋齿式限滑差速器具有输出两个不相等转矩的功能，但此两种差速器都是利用增大内摩擦方法(如利用摩擦力矩大的蜗轮蜗杆或大螺旋角的螺旋齿轮)实现转矩的不等量输出，因此托森差速器和螺旋齿式限滑差速器不仅传动效率低，而且两输出端的转矩差十分有限，因此托森差速器和螺旋齿式限滑差速器同样不能满足本发明的使用要求。为了有效解决这一问题，本发明专门发明了一种按比例分配转矩的特殊差速机构Ⅰ(如图1-3所示)，将单一动力源的动力按设定的比例分成两个动力对外输出，此比例即为行星齿轮机构Ⅲ中齿圈与太阳轮的齿数比α。将按比例分配转矩的差速机构Ⅰ串联在行星齿轮机构Ⅲ的前端就构成了本发明的无级自动变速装置，其具体的无级自动变速原理如下：
1)由于按比例分配转矩的差速机构I中行星锥齿轮2不是绕固定轴线转动，而是绕如图2所环形槽中心圆弧线转动，因此行星锥齿轮2与第二差速锥齿轮7的啮合点A至行星锥齿轮轴3的距离S 1和行星锥齿轮2与第一差速锥齿轮6的啮合点B至行星锥齿轮轴3的距离S 2不相等。若设第二差速锥齿轮7作用在行星锥齿轮2上的啮合力为F 1(A点处的啮合力)，第一差速器锥齿轮6作用在行星锥齿轮2上的啮合力为F 2(B点处的啮合力)。作用在行星锥齿轮2上的力矩应平衡，即：
F 1S 1＝F 2S 2
则：
Figure PCTCN2019109084-appb-000026
若设行星锥齿轮2作用在第二差速锥齿轮7上的啮合力为F 1′，行星锥齿轮2作用在第一差速锥齿轮6上的啮合力为F 2′，根据作用力与反作用力原理，则有：
F 1′＝F 1，F 2′＝F 2
由于第一差速锥齿轮6和第二差速锥齿轮7的齿数和模数均相等，第一差速锥齿轮6和第二差速锥齿轮7的分度圆直径必然相等。若设第一差速锥齿轮6和第二差速锥齿轮7的分度圆直径为R，则差速锥齿轮6和7输出的转矩M 6和M 7分别为：
Figure PCTCN2019109084-appb-000027
M 7＝RF 1′＝RF 1     (13)
比较式(12)和(13)得：
Figure PCTCN2019109084-appb-000028
若
Figure PCTCN2019109084-appb-000029
正好等于行星齿轮机构Ⅲ中齿圈与太阳轮的齿数比α，如此便可正好满足行星齿轮机构Ⅲ中齿圈的转矩M Q是太阳轮转矩M Tα倍的要求，即：
M Q＝αM T     (15)
如此，在按比例分配转矩的差速机构Ⅰ和行星齿轮机构Ⅲ之间便可实现转矩的平衡，即完全满足转矩的有效传递关系。
若设按比例分配转矩的差速机构Ⅰ的输入转速为n i＝n，行星齿轮机构Ⅲ中齿圈与太阳轮的齿数比
Figure PCTCN2019109084-appb-000030
第一差速锥齿轮6和第二差速锥齿轮7的转速分别为n 6和n 7，行星齿轮机构Ⅲ中的齿圈和太阳轮的转速分别为n Q和n T，由于第一差速锥齿轮6和第二差速锥齿轮7通过各自的轴分别与行星齿轮机构Ⅲ中的齿圈12和太阳轮11刚性联结，即：
n Q＝n 6     (16)
n T＝n 7     (17)
由差速器的原理知，差速机构的输入转速n i与第一差速锥齿轮6和第二差速锥齿轮7两个差速锥齿轮输出的转速始终满足如下关系式：
n 6+n 7＝2n i    (18)
式中：n i——差速器输入转速，n i＝n；
n 6、n 7——分别是第一差速锥齿轮6和第二差速锥齿轮7的转速。
当n 6＝0时，由式(17)、(18)、(19)得：
n Q＝n 6＝0
n T＝n 7＝2n
将n Q＝0、n T＝2n、α＝7代入式(1)便可计算出无级自动变速装置的输出 转速n H：
2n+α×0-(1+α)n H＝0
Figure PCTCN2019109084-appb-000031
无级自动变速装置的最大速比(又称传动比)i max为：
Figure PCTCN2019109084-appb-000032
当n 6＝n时，由式(16)、(17)、(18)得：
n Q＝n 6＝n
n T＝n 7＝n
将n Q＝n、n T＝n、α＝7代入式(1)便可计算出无级自动变速装置的输出转速n H：
n+7×n-(1+7)n H＝0
n H＝n
无级自动变速装置的最小速比i min为：
Figure PCTCN2019109084-appb-000033
由前面的计算可以看出，若齿圈12的转速在0～n间连续变化，则无级自动变速装置的速比可在4～1之间无级自动变化，即可以实现无级自动变速。
在机动车辆(汽车、拖拉机、坦克、可行走的工程机械的总称)领域，将变速器减速比(又称传动比)为1的档位称为直接档，为了提高机动车辆的燃料经济性，很多机动车辆的变速器设置有超速档(减速比小于1的档)。本发明全齿轮无极自动变速与速比主动控制系统无需对系统结构作任何改变便不仅可以实现任何减速比的超速档，而且在整个速比变化范围(从速比最小的超速档到最大速比)都可以实现无级自动变速，其具体实施方法是：提高速比调节电机8的转速，使之通过速比调节主、从动齿轮9和10传到行星齿轮机构Ⅲ齿圈12 上的转速n Q大于太阳轮11的转速n T(即n Q＞n T)就可得到所需的超速档减速比。还是以前面述及α＝7的例子为例，若设差速器输入转速n i＝n、n Q＝n 6＝1.5n、n T＝n 7代入式(18)计算得n T＝0.5n。设此状态下行星架的转速为n H0.5、无极自动变速装置的减速比i 0.5，将相关参数代入式(1)计算得n H0.5＝1.375n、
Figure PCTCN2019109084-appb-000034
若设n Q＝1.2n，可计算得到n H0.5＝1.15n，i 0.5＝0.8696。计算结果表明，只要连续调节速比调节电机的转速，同样可在超速档范围实现无级自动变速。
由行星齿轮机构的原理式n T+αn Q-(1+α)n H＝0知，当齿圈12的转速n Q＝0、太阳轮的转速n T＝n时，行星齿轮机构的减速比最大，其值为i max＝1+α。但对于本发明全齿轮无级自动变速与速比主动控制系统，由于在行星齿轮机构Ⅲ的前端或左端串联的是按比例分配转矩的差速机构Ⅰ，当n Q＝0时，太阳轮11的转速n T≠n，而是n T＝2n。
将n Q＝0,n T＝2n代入(1)式得本发明全齿轮无级自动变速与速比主动控制系统的实际最大速比
Figure PCTCN2019109084-appb-000035
本发明全齿轮无级自动变速与速比主动控制系统的实际最大速比i b max只有行星齿轮机构最大速比i max的1/2。
由前述的示例知，当行星齿轮机构Ⅲ中齿圈12与太阳轮11的齿数比
Figure PCTCN2019109084-appb-000036
时，本发明全齿轮无级自动变速与速比主动控制系统的实际最大速比i b max＝4。若要想获得更大的速比变化范围，尽管可以通过继续增大齿圈与太阳轮的齿数比α来获得，但α的增大必然带来行星齿轮机构径向尺寸的增大，即本发明全齿轮无级自动变速与速比主动控制系统的速比变化范围受齿圈与太阳轮齿数比α的限制，为了突破速比变化范围受α的限制，有效获得所需更大的速比变化范围，可以采用如下两个有效方法，即：①将二个或多个本发明无极自动变速与速比主动控制系统串联使用，如此可以得到速比变化范围为任意大小的无级自动变速与速比主动控制系统；②采用如图6所示本发明中专门为其发明的径向尺寸较小 的太阳轮输入端加减速机构的行星齿轮机构替代图1-图5中的行星齿轮机构Ⅲ。
减速机构V可以是如图6所示对称结构的定轴轮系齿轮减速机构或非对称定轴轮系齿轮减速机构或行星齿轮减速机构。传给行星齿轮机构的两路动力分别传给齿圈40和由输入轴39经减速机构V减速后传给太阳轮42，动力经过刚性连接在行星架44上的输出轴43输出。若用此太阳轮输入端加减速机构的行星齿轮机构代替本发明全齿轮无级自动变速与速比主动控制系统中的行星齿轮机构Ⅲ，设减速机构V的减速比为i V、齿圈与太阳轮的齿数比为α、无级自动变速装置输入轴1的转速为n i＝n，若齿圈的转速n Q＝0，则太阳轮的转速
Figure PCTCN2019109084-appb-000037
将前述参数代入(1)式得行星架输出转速
Figure PCTCN2019109084-appb-000038
最大速比i bV max为：
Figure PCTCN2019109084-appb-000039
若齿圈的转速n Q＝n，则太阳轮的转速
Figure PCTCN2019109084-appb-000040
将其代入(1)式得行星架输出转速
Figure PCTCN2019109084-appb-000041
最小速比i bV min为：
Figure PCTCN2019109084-appb-000042
作为一个减速比优选例，若设减速机构V的减速比为i V＝4、齿圈与太阳轮的齿数比为α＝5，将其分别带入式(19)和式(20)得：
Figure PCTCN2019109084-appb-000043
Figure PCTCN2019109084-appb-000044
尽管齿圈与太阳轮的齿数比α由原来的7减小到5，但本发明全齿轮无级自动变速与速比主动控制系统的无极自动变速范围却由原来的1～4增大到1.143～12。本发明的径向尺寸 大幅减小的同时，无极自动变速范围却大幅增加到了原来的近3倍。
当动力源的动力接入本发明全齿轮无级自动变速与速比主动控制系统的无级自动变速装置输入轴1时，由于此前机械系统处于静止状态，且与本发明全齿轮无级自动变速与速比主动控制系统输出轴14相连的工作机构或作业机械(如各类车辆的驱动车轮、舰船的螺旋桨、机床的主轴/刀盘、旋耕机、液压泵、压缩机、鼓风机等)由静止到运动过程中的阻力比较大，因此行星齿轮机构Ⅰ中的齿圈12仍处于静止状态，此时全齿轮无级自动变速器装置的速比(传动比)最大，十分有利于工作机构或作业机械的启动运行；随着动力源输出动力和转速的增加，齿圈12开始转动，全齿轮无级自动变速装置的速比(传动比)随着齿圈12转动速度的变化而变化，即开始自动变速。由于齿圈12的转速一定是连续变化不可能跳跃式变化，因此本明发全齿轮无级自动变速装置的速比(传动比)必然自动连续无级变化，即无级自动变速。若工作机构或作业机械的阻力不变、动力源输出的动力增大(或减小)，本发明全齿轮无级自动变速装置的速比随之减小(或增大)，即当阻力不变输入的动力增大(或减小)时，全齿轮无级自动变速装置的速比随着动力的增大(或减小)而减小(或增大)，工作机构或作业机械的转速同步上升(或下降)，达到良好的加速(或减速)效果；若动力源输出的动力不变、工作机构或作业机械的阻力增大(或减小)，本发明全齿轮无级自动变速装置的速比随之增大(或减小)，全齿轮无级自动变速装置正好输出更大(或更小)的转矩，以克服工作机构或作业机械增大(或减小)的阻力。
速比主动控制：为了超越全齿轮无极自动变速装置固有的自动变速规律，以便更好地满足实际使用的要求，本发明在按比例分配转矩的差速机构I与行星齿轮机构Ⅲ之间连接第一差速锥齿轮6和齿圈12的空心轴5上加装了速比主动控制机构Ⅱ。由于按比例分配转矩的差速机构I中的动力输出端第一差速锥齿轮6与行星齿轮机构Ⅲ中的齿圈12及按比例分配转矩的差速机构I中的另一动力输出端第二差速锥齿轮7与行星齿轮机构Ⅲ中的太阳轮之间始终处于动态平衡状态。从理论上讲，要想改变二者其中之一的运动十分简单，仅需给其中之一一个运动输入就可以任意调节其运动速度，且由于改变二者中任意一个的运动其速比便自动跟随发生变化，为此，为了实现对其速比主动控制，并获得大的速比变化范围，在第一差速锥齿轮6和齿圈12的连接轴第一差速锥齿轮轴5处增设了由速比调节电机8、速比调节主动齿轮9、速比调节从动齿轮10组成的速比主动控制机构Ⅱ。根据实际需要调节速比调节电机8的转速就可以实现按需在最大速比和最小速比间无级自动变速。
全齿轮无级自动变速与速比主动控制系统的工作过程：将如图5所示的换向机构与图1所示的具有速比主动控制的无级自动变速装置串联连起来便构成了具有无级自动变速和正转、反转、空档自由切换功能的本发明全齿轮无级自动变速与速比主动控制系统，参照图4。工程上，变速器的正转常将其称为前进档；变速器的反转常将其称为倒档；中断动力传动称为空档。换向机构与无级自动变速装置串联连接可以有如下两种连接方式，即：将换向机构串在无级自动变速装置的前端(或称输入端)或将换向机构串在无级自动变速装置的后端(或称输出端)。换向机构串在无级自动变速装置的前端(或称输入端)可以达到良好的减重效果，在此仅以换向机构串在无级自动变速装置前端(或称输入端)的结构方案为例，结合图7-9， 介绍本发明全齿轮无级自动变速与速比主动控制系统前进档、倒档和空档的切换过程与原理。
采用离合器+同步器换向机构的全齿轮无级自动变速与速比主动控制系统：
前进档：
将变速操纵杆置于D档位置，全齿轮无级自动变速与速比主动控制系统的控制系统执行如下操作：离合器18分离，同步器24移至最左端，使同步器啮合套与输入轴齿轮19右侧的啮合齿相啮合(见图7(a))，离合器18结合，动力和运动经离合器18传给换向输入轴27→同步器24→无级自动变速装置输入轴1→差速器壳4→行星锥齿轮轴3→行星锥齿轮2→差速锥齿轮6和7→齿圈12和太阳轮11(动力与运动经按比例分配转矩的差速机构I将其分成两路分别由差速锥齿轮6和7传给行星齿轮机构Ⅲ的齿圈12和太阳轮11)→行星架15→无级自动变速装置输出轴14。若需要按照实际使用要求进行自动变速(调节速比)，则启动速比主动控制系统的速比调节电机8，按照所需要求调节速比调节电机8的转速便可实现速比的主动控制。
倒档：
将变速操纵杆置于R档位置，全齿轮无级自动变速与速比主动控制系统的控制系统执行如下操作：速比调节电机8静止不转动，离合器18分离，同步器24移至最右端，使同步器啮合套与换向从动齿轮23左侧的啮合齿相啮合(见图8(a))，离合器18结合，动力和运动经离合器18传给换向输入轴27→输入轴齿轮19→换向主动齿轮25→过桥齿轮20→过桥齿轮轴21→过桥齿轮22→换向从动齿轮23→同步器24→无级自动变速装置输入轴1→差速器壳4→行星锥齿轮轴3→行星锥齿轮2→差速锥齿轮6和7→齿圈12和太阳轮11(动力与运动经按比例分配转矩的差速机构将其分成两路分别由差速锥齿轮6和7传给行星齿轮机构Ⅲ的齿圈12和太阳轮11)→行星架15→无级自动变速装置输出轴14。
空挡：
将变速操纵杆置于N档位置，离合器24分离，同步器24拨到中间位置(图9(a)所示的位置)，离合器24结合。
采用双联多片离合器换向机构的全齿轮无级自动变速与速比主动控制系统；
前进档：
将变速操纵杆置于D档位置，全齿轮无级自动变速与速比主动控制系统的控制系统执行如下操作：双联多片离合器28左侧的离合器结合、右侧的离合器分离，输入轴齿轮19经由双联多片离合器28与无级自动变速装置输入轴1固联(见图7(b))，动力和运动经离合器输入轴17传给双联多片离合器28→无级自动变速装置输入轴1→差速器壳4→行星锥齿轮轴3→行星锥齿轮2→差速锥齿轮6和7→齿圈12和太阳轮11(动力与运动经按比例分配转矩的差速机构将其分成两路分别由差速锥齿轮6和7传给行星齿轮机构III的齿圈12和太阳轮11)→行星架15→无级自动变速装置输出轴14。若需要按照实际使用要求进行自动变速(调节速比)，则启动速比主动控制系统的速比调节电机8，按照所需要求调节速比调节电机8的转速便可实现便速比的主动控制。
倒档：
将变速操纵杆置于R档位置，全齿轮无级自动变速与速比主动控制系统的控制系统执行如下操作：速比调节电机静止不转动，双联多片离合器28右侧的离合器结合、左侧的离合器分离，换向从动齿轮23经由双联多片离合器28与无级自动变速装置输入轴1固联(见图8(b))，动力和运动经离合器输入轴17传给换向主动齿轮25→过桥齿轮20→过桥齿轮轴21→过桥齿轮22→换向从动齿轮23→离合器28→无级自动变速装置输入轴1→差速器壳4→行星锥齿轮轴3→行星锥齿轮2→差速锥齿轮6和7→齿圈12和太阳轮11(动力与运动经按比例分配转矩的差速机构将其分成两路分别由差速锥齿轮6和7传给行星齿轮机构Ⅲ的齿圈12和太阳轮11)→行星架15→无级自动变速装置输出轴14。
空档：
将变速操纵杆置于N档位置，双联多片离合器36分离(见图9(b))。
采用双离合器换向机构的全齿轮无级自动变速与速比主动控制系统：
前进档
将变速操纵杆置于D档位置，全齿轮无级自动变速与速比主动控制系统的控制系统执行如下操作：双离合器中前进档离合器29结合、倒档离合器30分离、太阳轮锁止器36解锁、齿圈锁止器37锁止(见图7(c))，动力和运动经由离合器输入轴17传给前进档离合器29→前进档输入轴31→换向机构太阳轮32→换向机构行星轮34→换档机构行星架35→无级自动变速装置输入轴1→差速器壳4→行星锥齿轮轴3→行星锥齿轮2→差速锥齿轮6和7→齿圈12和太阳轮11(动力与运动经按比例分配转矩的差速机构将其分成两路分别由差速锥齿轮6和7传给行星齿轮机构Ⅲ的齿圈12和太阳轮11)→行星架15→无级自动变速装置输出轴14。若需要按照实际使用要求进行自动变速(调节速比)，则启动速比主动控制系统的速比调节电机8，按照所需要求调节速比调节电机8的转速便可实现便速比的主动控制。
倒档：
将变速操纵杆置于R档位置，全齿轮无级自动变速与速比主动控制系统的控制系统执行如下操作：速比调节电机8静止不转动，双离合器中倒档离合器30结合、前进档离合器29分离、齿圈锁止器37解锁、太阳轮锁止器36锁止(见图8(b))，动力和运动经由离合器输入轴17传给倒档离合器29→倒档输入轴38→换向机构齿圈33→换向机构行星轮34→换档机构行星架35→无级自动变速装置输入轴1→差速器壳4→行星锥齿轮轴3→行星锥齿轮2→差速锥齿轮6和7→齿圈12和太阳轮11(动力与运动经按比例分配转矩的差速机构将其分成两路分别由差速锥齿轮6和7传给行星齿轮机构Ⅲ的齿圈12和太阳轮11)→行星架15→无级自动变速装置输出轴14。
空档：
将变速操纵杆置于N档位置，双离合器中的前进档离合器29和倒档离合器30都分离(见图9(c))。

Claims (9)

  1. 一种全齿轮无级自动变速与速比主动控制系统，其特征在于：按比例分配转矩的差速机构串联在行星齿轮机构的前端构成无级自动变速装置；在按比例分配转矩的差速机构和行星齿轮机构之间设置速比主动控制机构，按比例分配转矩的差速机构包括两个设置在动力与运动输出端的差速锥齿轮：第一差速锥齿轮和第二差速锥齿轮；第一差速锥齿轮通过空心的第一差速锥齿轮轴与行星齿轮机构的齿圈刚性相连；第二差速锥齿轮轴穿过空心的第一差速锥齿轮轴，第二差速锥齿轮通过第二差速锥齿轮轴与行星齿轮机构的太阳轮相连；
    速比主动控制机构包括速比调节电机、处于常啮合状态的速比调节主动齿轮和速比调节从动齿轮；速比调节从动齿轮与第一差速锥齿轮轴刚性联结，速比调节主动齿轮安装在速比调节电机的输出轴上；
    行星齿轮机构的两个输入端分别为位于中心的太阳轮和最外围的齿圈，行星齿轮同时与太阳轮和齿圈啮合并通过行星架向外部输出动力。
  2. 根据权利要求1所述的全齿轮无级自动变速与速比主动控制系统，其特征在于：按比例分配转矩的差速机构包括差速器壳，无级自动变速装置输入轴在差速器壳前端与差速器壳刚性连接；差速器壳内，第一差速锥齿轮和第二差速锥齿轮均可旋转地支承在差速器壳沿着前后端或者左右方向延伸的轴孔中；第二差速锥齿轮和第一差速锥齿轮分别在前后端与行星锥齿轮啮合；第一差速锥齿轮轴的前端与第一差速锥齿轮刚性联结；第二差速锥齿轮轴前端与第二差速锥齿轮刚性联结，后端从空心的第一差速锥齿轮轴中穿过，并穿过差速器壳与太阳轮刚性相连或形成与太阳轮轴一体的结构。
  3. 根据权利要求1或2所述的全齿轮无级自动变速与速比主动控制系统，其特征在于：行星锥齿轮上开有绕行星锥齿轮轴中心线的周向环形槽，行星锥齿轮轴的一端固定安装在差速器壳上，另一端通过轴承安装在环形槽内，行星锥齿轮自转时转动中心是环形槽的中心圆弧线，即行星锥齿轮轴中心线是相对于行星锥齿轮在环形槽中的周向运动轨迹。
  4. 根据权利要求3所述的全齿轮无级自动变速与速比主动控制系统，其特征在于：当行星锥齿轮自转时，设行星锥齿轮与第二差速锥齿轮的啮合点A至行星锥齿轮轴中心线的距离为S 1、行星锥齿轮与第一差速锥齿轮的啮合点B至行星锥齿轮轴中心线的距离为S 2，S 1和S 2之比始终为设定的比例。
  5. 根据权利要求3所述的全齿轮无级自动变速与速比主动控制系统，其特征在于：按比例分配转矩的差速机构采用至少2只行星锥齿轮。
  6. 根据权利要求1所述的全齿轮无级自动变速与速比主动控制系统，其特征在于：所述的行星齿轮机构中设置至少2个行星齿轮。
  7. 根据权利要求1或2或4-6任一项所述的全齿轮无级自动变速与速比主动控制系统，其特征在于：两个或两个以上所述的全齿轮无级自动变速与速比主动控制系统串联连接；或者所述的行星齿轮机构中在太阳轮之前设置减速机构来增大全齿轮无级自动变速与速比主动控制系统的速比变化范围；所述减速机构为对称定轴轮系齿轮减速机构、或非对称定轴轮系齿轮减速机构、或行星齿轮减速机构中的其中一种，第二差速锥齿轮轴经减速机构减速后连接太阳轮，所述减速机构串联在太阳轮之前。
  8. 根据权利要求1或2或4-6任一项所述的全齿轮无级自动变速与速比主动控制系统，其特征在于：无级自动变速装置的前端或后端串接换向机构；实现动力与运动传输的正转、反转和中断至少三种模式的切换；换向机构为离合器和同步器组合结构的换向机构、或双联多片离合器形式的换向机构、或双离合器形式的换向机构。
  9. 根据权利要求8所述的全齿轮无级自动变速与速比主动控制系统，其特征在于：采用离合器和同步器组合结构的换向机构时，离合器输入轴和变速器输入轴成直线设置，离合器输入轴上设置输入轴齿轮，无级变速装置输入轴的左端或前端通过轴承安装在输入轴齿轮右端的轴承座孔内；换向主动齿轮与输入轴齿轮常啮合，同时与前一组过桥齿轮也是常啮合；后一组过桥齿轮与前一组过桥齿轮同轴且均与过桥齿轮轴刚性联结，后一组过桥齿轮与换向从动齿轮常啮合，换向从动齿轮利用滚针轴承或滑动轴承空套在无级自动变速装置输入轴上，输入轴齿轮和换向从动齿轮之间装有锁环式或锁销式同步器；
    采用双联多片离合器形式的换向机构时，离合器输入轴上设置输入轴齿轮，换向从动齿轮通过滚针轴承或滑动轴承空套在无级自动变速装置输入轴上，输入轴齿轮和换向从动齿轮之间装有双联多片离合器；无级自动变速装置输入轴的左端或前端通过轴承安装在离合器输入轴齿轮右端的轴承座孔内；换向主动齿轮与输入轴齿轮常啮合，同时与过桥齿轮也是常啮合；前后两组过桥齿轮均与过桥齿轮轴刚性联结，后端的过桥齿轮与换向从动齿轮常啮合；
    采用双离合器换向机构时，双离合器换向机构由双离合器和带锁止机构的行星齿轮机构组成，双离合器由前进档离合器和倒档离合器组成，前进档离合器通过空心的前进档输入轴与换向机构行星架刚性连接，倒档离合器通过从前进档输入轴中心穿过的倒档输入轴与换向机构太阳轮刚性连接，换档机构齿圈与无级自动变速装置输入轴的左端或前端刚性联结，将动力和运动传给按比例分配转矩的差速机构；在倒档输入轴和换向机构行星架上分别装有太阳轮锁止器和行星架锁止器。
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Cited By (3)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
CN113309834A (zh) * 2021-07-02 2021-08-27 浙江恒友机电有限公司 一种无级变速装置及无级变速器
CN113978237A (zh) * 2021-10-22 2022-01-28 奇瑞汽车股份有限公司 混合动力系统
CN117823581A (zh) * 2023-02-24 2024-04-05 江吉彬 电机调速的混合动力无级变速器及其工作方法

Families Citing this family (9)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
CN111368422B (zh) * 2020-03-03 2021-08-13 凯博易控车辆科技(苏州)股份有限公司 一种犬牙式离合器进齿控制方法
CN113027999B (zh) * 2021-04-28 2024-05-28 广东信稳能控技术研究有限公司 球伞弧锥自动变速器
CN113339467B (zh) * 2021-05-24 2024-10-18 深圳市雷圣特科技有限公司 一种摩托车无级调速变速箱
CN113653779B (zh) * 2021-07-07 2023-05-30 东风汽车集团股份有限公司 一种无级变速器及车辆
CN113565939B (zh) * 2021-08-20 2024-10-11 山西荣盛志达科技有限责任公司 一种全范围调速的机电一体化行星调速装置
CN113864411B (zh) * 2021-10-26 2024-04-12 中南大学 一种无动力中断无级变速箱
CN114704601B (zh) * 2022-03-09 2024-05-03 浙江金欣传动控制有限公司 一种速比可调的活齿减速机及其控制方法
CN114992302B (zh) * 2022-04-22 2024-09-24 南京工诺科技有限公司 一种农机用大功率多功能变速箱
CN115055083A (zh) * 2022-06-16 2022-09-16 山东大学 一种变速变向旋转搅拌装置及工作方法

Citations (6)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
US4696205A (en) * 1986-02-04 1987-09-29 Deere & Company Three-speed planetary transmission
CN201170296Y (zh) * 2007-12-18 2008-12-24 张耀新 具有无级变速和离合器两种功能的行星齿轮离合装置
CN202914648U (zh) * 2012-12-04 2013-05-01 何晓荷 电机调速的无级变速装置
CN103335075A (zh) * 2013-07-05 2013-10-02 武汉理工大学 电控调速无级变速系统与控制方法
CN103775579A (zh) * 2014-02-28 2014-05-07 张宝山 双差式无极变速器
CN108136889A (zh) * 2015-10-27 2018-06-08 舍弗勒技术股份两合公司 Cvt差速器

Family Cites Families (12)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
US3204476A (en) * 1960-04-05 1965-09-07 William S Rouverol Variable speed transmission
JPH01216145A (ja) * 1988-02-22 1989-08-30 Takashi Takahashi 無段変速装置
US5409425A (en) 1992-02-10 1995-04-25 Honda Giken Kogyo Kabushiki Kaisha Torque distributing mechanism in differential
JP3933081B2 (ja) 2003-04-03 2007-06-20 日産自動車株式会社 ハイブリッド変速機の変速制御装置
DE102006019131B4 (de) * 2006-04-21 2008-01-24 Getrag Driveline Systems Gmbh Aysmmetrisches, aktives Achsgetriebe
RU2397390C1 (ru) 2008-12-31 2010-08-20 Виктор Михайлович Кузеванов Дифференциал
JP2011252564A (ja) 2010-06-03 2011-12-15 Honda Motor Co Ltd 駆動制御装置
CN203453399U (zh) * 2013-07-27 2014-02-26 赵良红 双离合变速器传动机构
CN103836126B (zh) * 2014-03-31 2016-06-01 奇瑞汽车股份有限公司 一种汽车双离合器变速器
EP3380352B1 (en) * 2015-11-24 2023-12-20 BorgWarner Sweden AB A vehicle driveline system
CN109630628A (zh) * 2019-01-03 2019-04-16 钟跃荣 一种机械式无级变矩变速装置
CN110030336B (zh) * 2019-04-23 2021-01-15 合肥工业大学 一种含带传动的多挡电动汽车电力驱动系统

Patent Citations (6)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
US4696205A (en) * 1986-02-04 1987-09-29 Deere & Company Three-speed planetary transmission
CN201170296Y (zh) * 2007-12-18 2008-12-24 张耀新 具有无级变速和离合器两种功能的行星齿轮离合装置
CN202914648U (zh) * 2012-12-04 2013-05-01 何晓荷 电机调速的无级变速装置
CN103335075A (zh) * 2013-07-05 2013-10-02 武汉理工大学 电控调速无级变速系统与控制方法
CN103775579A (zh) * 2014-02-28 2014-05-07 张宝山 双差式无极变速器
CN108136889A (zh) * 2015-10-27 2018-06-08 舍弗勒技术股份两合公司 Cvt差速器

Non-Patent Citations (1)

* Cited by examiner, † Cited by third party
Title
See also references of EP3819521A4 *

Cited By (5)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
CN113309834A (zh) * 2021-07-02 2021-08-27 浙江恒友机电有限公司 一种无级变速装置及无级变速器
CN113309834B (zh) * 2021-07-02 2024-06-04 浙江恒友机电有限公司 一种无级变速装置及无级变速器
CN113978237A (zh) * 2021-10-22 2022-01-28 奇瑞汽车股份有限公司 混合动力系统
CN113978237B (zh) * 2021-10-22 2023-09-26 奇瑞汽车股份有限公司 混合动力系统
CN117823581A (zh) * 2023-02-24 2024-04-05 江吉彬 电机调速的混合动力无级变速器及其工作方法

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