Eight/sixteen forward gear planetary gear mechanism automatic transmission series

Abstract

An eight and sixteen forward gear planetary gear mechanism automatic transmission series, where it includes eight forward speeds realized by combining three simple planetary gear sets, five clutches, plus improved and adapted sixteen forward speeds realized by combining four simple planetary gear sets, six clutches. In this solution, a major type of synchronizer clutch being used corresponding to each planetary gear set. The gear train also uses ratchet and pawl pair clutch structure as an alternative option, so that the automatic transmission has a more compact overall structure, the gearshift efficiency is high, the synchronous responsiveness and stability are excellent. The transmission series has wide and uniform transmission gear ratios, making-capacity maximum to achieve 26:1; and the shift unit is simple and reliable, which can significantly enhance the maneuverability of the vehicle driving and reduce fuel consumption and greenhouse gas emissions.

Claims

1. An sixteen forward gear planetary gear mechanism automatic transmission series, wherein it comprises; sixteen forward speeds and four reverse speeds realized by combining four simple planetary gear sets with six torque engagement elements, a first planetary gear set P1, a second planetary gear set P2, a third planetary gear set P3, and a fourth planetary gear set P4 are all single pinion type planetary gears and have first to fourth sun gears S1-S4, first to fourth planet carriers C1-C4, and first to fourth ring gears R1-R4 respectively; the four planetary gear sets are sequentially arranged in the axial direction, wherein the first planetary gear set P1 and the second planetary gear set P2 both have a fixed mode of the first sun gear S1 and the second sun gear S2, the sun gear S1 is fixedly mounted on a front end cover of a case body, and the sun gear S2 is separately fixed on a positioning bracket of the inner housing of the transmission body; the third ring gear R3 of the third planetary gear set P3 is fixedly mounted on the other side of the positioning bracket of the second sun gear S2; the fourth planet carrier C4 and the fourth ring gear R4 of the fourth planetary gear set P4 may be selectively fixedly locked respectively or simultaneously idled, reverse gears are directly obtained by adjusting the internal variable relationship of the single-row planetary gear assembly, and on the premise that the fourth sun gear S4 is input, either the fourth planet carrier C4 or the fourth ring gear R4 can be separately locked by a bidirectional double clutch to undertake output functions of forward gears or reverse gears; in the forward gear state, the fourth ring gear R4 is always kept in a locked and fixed state, while a first clutch CL1, a second clutch CL2, a third clutch CL3, a fourth clutch CL4, a fifth CL5, and a sixth clutch CL6 are sequentially arranged in an input shaft direction.

2. The sixteen forward gear planetary gear mechanism automatic transmission series according to claim 1, wherein the first planetary gear set P1 and the second planetary gear set P2 share a first cage for shifting and transmission, and the first sun gear S1 and the second sun gear S2 are always operated as fixed members; the first clutch CL1 is arranged and sleeved on a longitudinal beam of the first cage on a periphery of the first planetary gear set P1; the second clutch CL2 is arranged and sleeved on the longitudinal beam of the first cage on an outer layer of the second planetary gear set P2; Each coaxially arranged single-stage planetary gear train independently generate an (in/out) gear ratio; moreover, a power input end of the planetary gear train may also be directly used as an output end to transmit torque at a constant speed of 1:1, so that each simple planetary gear mechanism has two fixed-value shifting output degrees of freedom; three to four simple planetary gear sets are sequentially arranged in the axial direction, with different characteristic parameter values (ring-to-sun ratios), and are selectively engaged and adapted with the clutches, thus generating sixteen gear ratio series with overall uniform distribution.

3. The sixteen forward gear planetary gear mechanism automatic transmission series according to claim 1, wherein a first connecting member, is a first cage (I), is configured as a driver for variable-speed synchro drive between the first planetary gear set P1 output and the second planetary gear set P2 input; a second connecting member, is a central main shaft, is configured for conducting variable-speed and torque between the second planetary gear set P2 and the third planetary gear set P3; a third connecting member, is a second cage (II), is configured for variable-speed synchro drive between the second planetary gear set P2 output through a main shaft to the third planetary gear set P3 output, and next to the fourth planetary gear set P4 input; a fourth connecting member, is a third cage (III), is configured for conducting variable-speed and torque between the fourth planetary gear set P4 and an output shaft.

4. The sixteen forward gear planetary gear mechanism automatic transmission series according to claim 1, a torque transmission mechanism is at least one of a synchronizer clutch, a dog clutch and ratchet and pawl coupling element clutch, a ratchet ring gear sleeve is a shift driving unit; the first clutch CL1 is configured to output different torques of the first planet carrier C1 and the first ring gear R1 of the first planetary gear set P1 to a first cage (I) by selectively connecting outer peripheral teeth of their respective members; the second clutch CL2 is configured to input different torques, transmitted by the first planetary gear set P1 via the first cage by selectively connecting outer peripheries of respective members of the second ring gear R2 and the second planet carrier C2 of the second planetary gear set P2; the third clutch CL3 is configured to selectively engage and transmit-receive via a second cage (II) for a fixed reduced speed output of the third planetary gear set P3-C3 or the variable-speed transmitted from the second ring gear R2 of the second planetary gear set P2 via a central main shaft; the fourth clutch CL4 is configured to lock in the fourth sun gear S4 to the fourth planet carrier C4 when the planetary gear set P4 is idling and have no fixed output, so that the fourth planetary gear set P4 rotates at a constant speed of 1:1; the fifth clutch CL5 is configured to selectively lock up and fix the fourth planet carrier C4 or the fourth ring gear R4 of the fourth planetary gear set P4, so that the fourth planetary gear set P4 has a clear internal variation operating mode, or the fourth planet carrier C4 and the fourth ring gear R4 are not fixed and cause idling together, the sixth clutch CL6 is configured to finally transmit different torques of the fourth planetary gear set P4 to the output shaft of the transmission via a third cage (III) by selectively connecting the output ends of the ring gear R4 and the carrier C4 of the fourth planetary gear set P4 through an engagement sleeve.

5. The sixteen forward gear planetary gear mechanism automatic transmission series according to claim 4, the synchronizer clutch can use double cone and triple cone type in order to facilitate the connectivity of the connection-oriented make-break operation; each planetary component of the gear set, has a synchronizing cone fixed on its sidewall as a collar rim drive, these collar cone cups and the baulk ring from the synchronizing clutch, as synchromesh counterparts to match, with smooth butt and collar joint getting low energy cost and high load capacity.

6. The sixteen forward gear planetary gear mechanism automatic transmission series according to claim 4, wherein in the ratchet and pawl clutch member transmission pair, the ratchet ring gear itself is a clutch; for a ratchet and pawl coupling, relative to the pawl elements and a smooth hollow sleeve wall surface without engagement are provided on the inner ring surface of the ratchet part; the grooves of the ratchet locking part and the smooth hollow sleeve part have different geometries and are equally spaced on the inner ring surface of the ratchet ring; through the parallel sliding of the ratchet on the cage, the engagement transmission path between the pawls mounted on different rotating members of the planetary gear, and the grooves on the inner ring surface of the ratchet ring gear are allowed to be connected and disconnected, thus generating new torque shifting, in the ratchet and pawl transmission pair, one is active input and the other is driven output, so as to transmit power and torque in turn; the automatic transmission uses a ratchet ring gear type clutch, so that the clutch may be disengaged from another gear when the clutch is engaged in one gear during gear shifting, and there is no power interruption during synchronous gear shifting and there is no need to reduce engine power during gear shifting.

7. The sixteen forward gear planetary gear mechanism automatic transmission series according to claim 1, wherein drives of the clutches are adjusted by at least one of a hydraulic assembly, SVA electric assembly, and pneumatic actuator through reciprocating driving accordingly.

Description

BRIEF DESCRIPTION OF DRAWERS

(1) FIG. 1 is a schematic diagram of the first embodiment (three single-stage planetary gear sets) of an eight-speed transmission according to the invention;

(2) FIG. 2 is a schematic diagram of the first and third embodiment which has a series connection of compound planetary gear train in the P1 gear set according to the invention;

(3) FIG. 3 is a schematic diagram of the relationship between the grooves and the idling distribution on the inner ring surface of the ratchet clutch, and pawl install position on each planetary component collar;

(4) FIG. 4 is a schematic diagram of the shift process between teeth on the inner ring surface of the ratchet and pawls and ratchet coupling engagement;

(5) FIG. 5 is a structural diagram of a sixteen forward gears transmission (four single-stage planetary sets) according to the fourth and fifth embodiment;

(6) FIG. 6 is a structural diagram of the sixteen-speed transmission which has two sub gear-sets jointly cascade amplification gear ratio in the compound P1 planetary gear train according to the invention;

(7) In the drawings, 1transmission body, 2input shaft, 3central main shaft, 4output shaft, 5first roller cage, 6second cage, 7third cage, 8first sun gear, 9first planet carrier, 10first ring gear, 11second sun gear, 12second planet carrier, 13second ring gear, 14fixing bracket of second sun gear, 15third sun gear, 16third planet carrier, 17third ring, 18fourth sun gear, 19fourth planet carrier, 20fourth ring, 21first clutch CL1, 22second clutch CL2, 23third clutch CL3, 24fourth clutch CL4, 25fifth clutch CL5, 26sixth clutch CL6, 27stationary cage, 28pawl A, 29pawl B, 30pawl C, 31planetary pinion, 32pawl mounting ring seat, 33ratchet sleeve, 34ratchet sleeve inside groove, 35smooth idling end face in the middle of inner ratchet, 36collar flange and mounting hole of ratchet sleeve, 37first compound planetary gear train P1, 38synchro clutch sleeve, 39synchronizing ring, 40synchronizing cone; 41spline hub;

DETAILED DESCRIPTION OF THE PREFERRED EMBODIMENTS

(8) To make the transmission coherent feature and benefit more clear, the technical solutions will be described clearly and completely with reference of the data sheet and schematic drawings in the embodiments of the invention.

(9) (I) An eight forward gear planetary gear train automatic transmission consists of the transmission body 1, three single-stage planetary gear sets, five shift clutches, most are bidirectional double-acting type; three internal connecting members, shift actuator drives, and input shafts 2 plus output shafts 4 at both ends of the transmission body.

(10) TABLE-US-00001 Pgs-1/CL1 Pgs-2/CL2 Pgs-3/CL3 Pgs-3/CL4 Pgs-3/CL5 Synchronizer Pawl A Pawl B Pawl C Synchronizer Synchronizer C1 R1 R2 C2 C3-S3 C3 R3 R3 C3 output output input input lockup (lock) (lock)

(11) As shown in Table 1, in the five clutches, CL1, CL4 & CL5 is synchronizer clutch. CL2, CL3 is pawl ratchet coupling type. Pawl A and B are respectively provided on the pawl rings on the outer peripheral surfaces of the ring gear member R2 and planet carrier C2 of the planetary gear P2, and pawl C is provided on the central main shaft between the planetary gear sets P2 and P3, correspond to the carrier C3 of planetary gear set P3. CL4 is locating on the stationary cage exterior of P3, can selectively engage the collar flange teeth of the planet carrier and the ring gear of P3 and it is a brake function.

(12) Each clutch can be selectively operated. The synchronizer type CL1 and ratchet sleeve of the second clutch CL2 are plug-in mounted on the longitudinal beam of the first cage (I), and the four-gear ratio shift action process is completed synchronously by adjusting the clutches CL1 and CL2 to selectively lock and engage with different output/input engagement end faces of the respective rotating members edge joint of the planetary gear sets P1-P2. The pawl C on the central main shaft 3 is then adjusted by using the clutch CL3 to accordingly engage and interlock the P3 planet carrier C3 with sun gear S3, and the on-off mode of the CL4 to change the internal variable relationship of the planetary gear mechanism P3 by individually lock up the ring R3 or carrier C3. Finally, by the fifth clutch CL5 selective engagement to the output shaft to forma full series of shifting course. Thus, eight forward and two reverse gears are achieved by the adaptation between the transmission members and the respective internal variable rotating elements of the planetary gear sets P1-P2-P3.

(13) As shown in FIG. 1, the planetary gear sets P1, P2 and P3 are all single pinion type planetary gears composed of a sun gear, a planet carrier supporting the planetary pinions and a ring gear. The three planetary gear sets are sequentially arranged in the axial direction, and the planetary gear sets P1 and P2 are always a fixed sun gear mode. The sun gear S1 is fixedly mounted on the front end cover of the case body, and the sun gear S2 is separately fixed on a positioning bracket of the inner housing of the case body.

(14) Input shaft rigidly connected to the first ring gear R1, the second ring gear R2 of P2 rigidly connected to the main shaft as the second planetary gear set output end. The planet carrier C3 and the ring gear R3 of the planetary gear set P3 can be selectively fixedly locked respectively or simultaneously idled. In the forward gear state, the ring gear R3 is always kept in a locked and fixed state.

(15) The input shaft (2) passes through the front end cover bearing of the transmission body and the inner hole of the sun gear S1 of the first planetary gear set and is directly fixedly connected to the first ring gear R1 to input the power. The output shaft (4) is connected to the third connecting member (6) to complete the power output. The front input shaft is usually a turbine shaft of the hydraulic torque converter and it transmits the power of the engine to the transmission, and the output shaft subsequently transmits the power to the left and right driving wheels through differentials.

(16) The first connecting member, i.e., the first outer ring Cage (I), is configured as a hollow cylindrical structure with openings on both sides, and the fixed longitudinal beams with open slots at equal spacing are distributed on the surface of the housing. It is provided on the exterior layers of the planetary gear sets P1-P2, and the left and right end faces are supported and installed on the base of positioning steadier on the front end cover of the transmission body and the circular bead, stepped boss circumference of the fixing bracket for the sun gear S2 of the second planetary gear set.

(17) The second connecting member, i. e., the central main shaft (3) of the transmission body, is nested with the inner holes of the sleeve shaft (quill) on the end faces of the input and output shafts as spigot and socket joint supported at the shaft shoulders. With the input shaft, the ring R2 and planet carrier C2, the second sun gear S2 and the fixing bracket, the third planet carrier C3, the third sun gear, the second cage (II), and the output shaft in turn in axial direction; and fixedly connected to the second ring gear R2, the third clutch CL3 (pawl C), and the third sun gear S3 to undertake power flow. The pawl seat of the clutch CL3 is connected to the spline keyway of the main shaft and can slide axially to complete the lock-in synchronism on-off adjustment of the planetary gear set P3.

(18) The third connecting member, i.e., the second clutch cage (II), marked as 6; is configured as a hollow shell structure or cylindrical prong structure with one side open and the other closed. It is provided at the end of the transmission body and the fifth clutch CL5 is sleeve plug-in mounted on the longitudinal beam, or the skewer of the cage. The clutch CL5 can also be a dog type from a compact perspective and can move transversely along the cage to alternatively engage with the power output end face R3-C3 of the planetary gear set P3 for transmission, forming forward to reverse gear shift transmission. The output end face of the ring gear R3 is the reverse gear output side.

(19) The synchronizer CL4, the sliding sleeve is connected by through-holes on the rib of the collar flange to the longitudinal beams of an annular stationary cage which is around the planetary gear set P3; it can axially move to selectively engage with the respective outward circular end faces of the front rotating arm of the planet carrier C3 and the third ring gear R3 of the third planetary gear set. The CL4 sliding sleeve moves transversely along the longitudinal beam of the cage to lock and control the different rotating elements, the planet carrier C3 or the ring gear R3 to be stopped respectively. The clutch can be freely switched that allows the forward rotation (forward direction) of P3 and prevents the reverse rotation (backward direction), or prevents the forward rotation and allows the reverse rotation, or to be in neutral state wherever the P3 planetary gear members are simultaneously idling without exact output.

(20) The eight forward gears automatic transmission according to the embodiments are described clearly and completely with reference to the schematic drawings in the invention. As shown in FIG. 1, the first embodiment of the invention provides the 1-OD high gear option. In the embodiment, the first to third planetary gear mechanisms P1-P3 are simple planetary gear sets.

(21) TABLE-US-00002 Gear Ring-to- Input Output Stationary Calculation Ratio Sun ratio P1 Ring (R) Planet Sun (1 + S1/R1) 1.56 1.7857 Carrier P2 Planet Ring Sun 1/(1 + 0.76923 3.333 Carrier S2/R2) P3 Sun Planet Ring 1 + (R3/S3) 3.2258 2.2258 Carrier Ring Planet (R3/S3) 2.2258 Carrier
In this example, the configuration of each planetary gear set is shown in Table 2 above. P1 is R1 input source power, and C1 is speed-down output, or direct R1 constant speed output mode; P2 is C2 input, and R2 is speed-up output or R2 input and direct constant speed output; P3 is always S3 input, and the ring gear R3, the planet carrier C3 are alternately blockaded by the CL4 stopper to form C3 speed-down output or R3 output reverse gear. Or C3 and R3 are unlocked simultaneously and C3 interlocked with S3 by the third clutch CL3, then P3 produces overall 1:1 constant speed output.

(22) Specific gearshift ratio-kinematic logic: P1(1.56/1)P2(0.76923/1)P3(3.2258/1), in parentheses are the two gear ratios value of each single-stage planetary gear mechanism: internal variable speed gear ratio and constant speed gear ratio. The gear ratio of each gear path is comprehensively determined by the characteristic parameters of individual planetary gear set, i.e., the ratio for the number of teeth of the ring gear to that of the sun gear. In the first embodiment, the fixed gear ratios of gear-set can be obtained when (R1/S1)1.7857, (R2/S2)3.333, and (R3/S3)2.225. The specific gear distribution from lowest to top is as follows: 5.032_3.87_3.2258_2.481_1.56_1.2_1_0.76923. The wide overall gear ratio range for the forward gears is obtained, and (5.032/0.76923) reaches 6.5416.

(23) The following is a description of the power shift by P1-P2-P3 gang control of eight forward and two reverse gears. To realize a specific gear output, it is necessary to adjust the detail engagement and release relative elements of each planetary gear mechanism. The related operation process of gears is as follows:

(24) 1st gear D1, simultaneous kinematic equation of gear ratio:
P1(1.56)P2(1)P3(3.2258)

(25) The synchro clutch CL1 moves leftward on the first cage, so that the output end of P1-C1 locking teeth engages with CL1 sliding sleeve. The clutch CL2 ratchet moves rightward on the first cage and pawl B at the P2-C2 radical end is contracted and idled by the smooth belt in the middle of the inner ring surface of the ratchet, and pawl A at P2-R2 collar rim engages with the left grooves of the CL2 ratchet inner surface to input torque. The clutch CL3 is in neutral. The synchro clutch CL4 moves rightward on the stationary cage, and the R3 rim cross-linked and engages with sleeve, ring gear R3 is locked and fixed. Clutch CL5 moves rightward on second cage, engages and locks the output end of P3-C3, and realizes integral transmission with the output shaft through longerons of the second cage.

(26) Transmission path: Input shaft.fwdarw.ring gear R1.fwdarw.C1 rim output end.fwdarw.CL1 sleeve.fwdarw.first cage.fwdarw.CL2 ratchet.fwdarw.pawl A at input end of R2.fwdarw.R2 ring gear.fwdarw.main shaft (3).fwdarw.third sun gear S3.fwdarw.third planet carrier C3.fwdarw.shaft sleeve outward output end face of C3 rear rotating arm.fwdarw.CL5 engagement sleeve.fwdarw.second cage.fwdarw.Output shaft.

(27) 2nd gear D2, simultaneous kinematic equation of gear ratio: P1(1.56)P2(0.76923)P3(3.2258)

(28) {According to the 1st gear} The second clutch CL2 moves leftward so that pawl A at the input end of P2-R2 is contracted and idled by the smooth inner ring surface of the ratchet, and pawl B at the rim end of the planet carrier C2 is released to engage with and lock the ratchet teeth at the right end of the inner ring of the CL2 ratchet, thus realizing shifting and transmission;

(29) Transmission path: Input shaft.fwdarw.ring gear R1.fwdarw.C1 rim output end.fwdarw.CL1 sleeve.fwdarw.first cage.fwdarw.CL2 ratchet.fwdarw.pawl B at input end of planet carrier C2.fwdarw.R2 ring gear.fwdarw.main shaft (3).fwdarw.third sun gear S3.fwdarw.third planet carrier C3.fwdarw.shaft sleeve outward output end face of C3 rear rotating arm.fwdarw.CL5 sleeve.fwdarw.second cage.fwdarw.Output shaft

(30) 3rd gear D3, simultaneous kinematic equation of gear ratio: P1(1)P2(1)P3(3.2258)

(31) {According to the 2nd gear} The clutch CL1 moves rightward so that the output end of P1-R1 is locking to engage with the sleeve. The second clutch CL2 moves rightward so that pawl B at the input end of P2-C2 is contracted and idled, and pawl A at the rim input end of P2-R2 is released to engage with the groove on the inner ring surface at the left end of the ratchet, thus realizing overall transmission.

(32) Transmission path: Input shaft.fwdarw.ring gear R1.fwdarw.R1 rim output end.fwdarw.CL1 sleeve.fwdarw.first cage.fwdarw.CL2 ratchet.fwdarw.pawl A at input end of ring gear R2.fwdarw.R2 ring gear.fwdarw.main shaft (3).fwdarw.third sun gear S3.fwdarw.third planet carrier C3.fwdarw.shaft sleeve outward output end face of C3 rear rotating arm.fwdarw.CL5 sleeve.fwdarw.second cage.fwdarw.Output shaft

(33) 4th gear D4, simultaneous kinematic equation of gear ratio:
P1(1)P2(0.76923)P3(3.2258)

(34) {According to the 3rd gear} The clutch CL2 moves leftward so that pawl A at the input end of P2-R2 is contracted and idled by the smooth belt in the middle of the inner ring surface of the ratchet, and pawl B at the input end of P2-C2 is released to engage with and lock the grooves of the inner ring at the right end of the ratchet, thus realizing transmission.

(35) Transmission path: Input shaft.fwdarw.ring gear R1.fwdarw.R1 rim output end.fwdarw.CL1 sleeve.fwdarw.first cage.fwdarw.CL2 ratchet.fwdarw.pawl B at input end of planet carrier C2.fwdarw.R2 ring gear.fwdarw.main shaft (3).fwdarw.third sun gear S3.fwdarw.third planet carrier C3.fwdarw.shaft sleeve outward output end face of C3 rear rotating arm.fwdarw.CL5 sleeve.fwdarw.second cage.fwdarw.Output shaft

(36) 5th gear D5, simultaneous kinematic equation of gear ratio: P1(1.56)P2(1)P3(1) {According to the 4th gear} The synchro clutch CL1 moves leftward to engage with the output end of P1-C1. The clutch CL2 moves rightward so that pawl B is contracted and idled, and pawl A at the constant speed input end of P2-R2 is released to engage with the ratchet teeth on the inner ring surface. The synchro clutch CL4 moves leftward to the neutral position along the stationary cage, so that the ring gear R3 is unlocked, the planetary gear set P3 idled and has no fixed output. Meanwhile, the clutch CL3 from a neutral state is moving rightward along the keyway of the central main shaft, so that push pawl C engages with of the ratchet ring gear which is fixed on the sidewall of the front rotating arm of the planet carrier C3. thus C3 and S3 is locked together to form a rigid body rotation of the planetary gear set P3 in constant speed 1:1, thus realizing overall transmission.

(37) Transmission path: Input shaft.fwdarw.ring gear R1.fwdarw.C1 rim output end.fwdarw.CL1 sleeve.fwdarw.first cage.fwdarw.CL2 ratchet.fwdarw.pawl A at input end of R2.fwdarw.R2 ring gear.fwdarw.main shaft (3).fwdarw.pawl C.fwdarw.clutch CL3 ratchet.fwdarw.front rotating arm of planet carrier C3.fwdarw.shaft sleeve output end face of C3 rear rotating arm.fwdarw.CL5 sliding sleeve.fwdarw.second cage.fwdarw.output shaft

(38) 6th gear D6, simultaneous kinematic equation of gear ratio: P1(1.56)P2(0.76923)P3(1)

(39) {According to the 5th gear} The second clutch CL2 moves leftward so that pawl A at the input end of P2-R2 is contracted and idled by the smooth inner ring surface of the ratchet, and pawl B at the input end of the planet carrier C2 is released to engage with and lock the groove at the right end of the inner ring of the CL2 ratchet, thus realizing shifting and transmission;

(40) Transmission path: Input shaft.fwdarw.ring gear R1.fwdarw.C1 rim output end.fwdarw.CL1 sleeve.fwdarw.first cage.fwdarw.CL2 ratchet.fwdarw.pawl B at input end of C2.fwdarw.R2 ring gear.fwdarw.main shaft (3).fwdarw.pawl C.fwdarw.clutch CL3 ratchet.fwdarw.front rotating arm of planet carrier C3.fwdarw.shaft sleeve output end face of C3 rear rotating arm.fwdarw.CL5 engagement sleeve.fwdarw.second cage.fwdarw.Output shaft

(41) 7th gear D7, simultaneous kinematic equation of gear ratio: P1(1)P2(1)P3(1)

(42) {According to the 6th gear} The clutch CL1 moves rightward so that the output end of P1-R1 is locking to engage with the sleeve. The second clutch CL2 moves rightward so that pawl B at the input end of P2-C2 is contracted and idled, and pawl A at the input end of P2-R2 is released to engage with the groove on the inner ring surface at the left side of the ratchet, the gear is the 1:1 direct gear mode. Thus realizing overall transmission.

(43) Transmission path: Input shaft.fwdarw.ring gear R1.fwdarw.R1 rim output end.fwdarw.CL1 sleeve.fwdarw.first cage.fwdarw.CL2 ratchet.fwdarw.pawl A at input end of R2.fwdarw.R2 ring gear.fwdarw.main shaft (3).fwdarw.pawl C.fwdarw.clutch CL3 ratchet.fwdarw.front rotating arm of planet carrier C3.fwdarw.shaft sleeve output end face of C3 rear rotating arm.fwdarw.CL5 engagement sleeve.fwdarw.second cage.fwdarw.Output shaft 8th gear D8, simultaneous kinematic equation of gear ratio: P1(1)P2(0.76923)P3(1)

(44) {According to the 7th gear} The clutch CL2 moves leftward on the first cage so that pawl A on the end face of the ring gear R2 is contracted, unlocked and idled, and pawl B at the input end of P2-C2 is released to tilt to engage with the grooves on the inner ring surface at the right side of the ratchet, thus realizing overdrive transmission. The gear is the highest gear of total gear train.

(45) Transmission path: Input shaft.fwdarw.ring gear R1.fwdarw.R1 rim output end.fwdarw.CL1 sleeve.fwdarw.first cage.fwdarw.CL2 ratchet.fwdarw.pawl B at input end of C2.fwdarw.R2 ring gear.fwdarw.main shaft (3).fwdarw.pawl C.fwdarw.clutch CL3 ratchet.fwdarw.front rotating arm of planet carrier C3.fwdarw.shaft sleeve output end face of C3 rear rotating arm.fwdarw.CL5 engagement sleeve.fwdarw.second cage.fwdarw.Output shaft

(46) Reverse gear R1, simultaneous kinematic equation of gear ratio: -P1(1.56)P2(1)P3(2.2258)

(47) The synchro clutch CL1 sliding sleeve moves leftward to engage with the output edge of C1. The clutch CL2 moves rightward so that pawl B is idled, and pawl A engages with the CL2 ratchet for input from R2. Clutch CL3 is idled. The synchro type CL4 moves leftward on the fixed cage, so the carrier C3 circumferential edge fixed and blockage with the CL4 sleeve, ring gear R3 released and rotating reversely.

(48) The clutch CL5 moves leftward on the second cage to engage with the reverse output end of P3-R3, thus realizing reverse transmission to output shaft through the second cage.

(49) Transmission path: Input shaft.fwdarw.ring gear R1.fwdarw.output end of C1.fwdarw.CL1 sleeve.fwdarw.first cage.fwdarw.CL2 ratchet.fwdarw.pawl A at input end of R2.fwdarw.R2 ring gear.fwdarw.main shaft (3).fwdarw.third sun gear S3.fwdarw.third ring gear R3.fwdarw.R3 output end.fwdarw.CL5 engagement sleeve.fwdarw.second cage.fwdarw.Output shaft.

(50) Reverse gear R2, simultaneous kinematic equation of gear ratio: P1(1.56)P2(0.76923)P3(2.2258)

(51) {According to the reverse 1st gear} The clutch CL2 moves leftward so that pawl A on P2-R2 rim is contracted and idled, and pawl B at the rim input end of P2-C2 released outward to engage with the groove on the right inner ring surface of CL2. The clutches CL1, CL3, CL4 and CL5 remain in the original engaged state, thus realizing overall transmission.

(52) Transmission path: Input shaft.fwdarw.ring gear R1.fwdarw.output end of C1.fwdarw.CL1 sleeve.fwdarw.first cage.fwdarw.CL2 ratchet.fwdarw.pawl B at input end of C2.fwdarw.R2 ring gear.fwdarw.main shaft (3).fwdarw.third sun gear S3.fwdarw.third ring gear R3.fwdarw.R3 output end.fwdarw.CL5 engagement sleeve.fwdarw.second cage.fwdarw.Output shaft.

(53) TABLE-US-00003 TABLE 3 Sequential Logic Diagram of 1-OD Overdrive Shift Control Keys according to the First Embodiment of the invention: Engage P3-CL4 P1-CL1 P2-CL2 CL3 synchronizer P3-CL5 synchronizer pawl A paw1 B pawl C C3 R3 synchronizer Gears C1 R1 R2 C2 C3-S3 (Lock) (Lock) R3 C3 1.sup.st .Math. .Math. .Math. .Math. 2.sup.nd .Math. .Math. .Math. .Math. 3.sup.rd .Math. .Math. .Math. .Math. 4.sup.th .Math. .Math. .Math. .Math. 5.sup.th .Math. .Math. .Math. .Math. 6.sup.th .Math. .Math. .Math. .Math. 7.sup.th .Math. .Math. .Math. .Math. 8.sup.th .Math. .Math. .Math. .Math. Reverse 1 .Math. .Math. .Math. .Math. Reverse 2 .Math. .Math. .Math. .Math.

(54) In the table, .Math. indicates that the clutch is engaging with the planetary element. Others without marks in empty, indicate that the released idling state or is not engaged as a freewheel.

(55) As shown in FIG. 2, for the 1-OD overdrive gear series, the first planetary gear set uses two sub single planetary gear set to get a refurbished compound gear ratio list of, i. e., P1a (1.333)P1b(1.25)=1.666, so the gear ratio calculation logic of powertrain is P1(1.666/1)P2(0.76923/1)P3(3.2258/1). The shift logic is the same as in the table 3 above. Gear distribution: 5.376_4.135_3.2258_2.481_1.666_1.282_1_0.76923. The gear ratios can also be expressed as (out/in) version: 0.186_0.2418_0.31_0.403_0.6_0.78 1_1.3. The total gear ratio ranges of forward gears (5.376/0.76923) reach 6.988.

(56) TABLE-US-00004 TABLE 4 Logic Diagram for Engagement Sequence of Eighth Forward Gears (2-OD). Engage CL4 CL1 CL2 CL3 C3 R3 CL5 Gears C1 R1 R2 C2 C3-S3 (Lock) (Lock) R3 C3 1.sup.st .Math. .Math. .Math. .Math. 2.sup.nd .Math. .Math. .Math. .Math. 3.sup.rd .Math. .Math. .Math. .Math. 4.sup.th .Math. .Math. .Math. .Math. 5.sup.th .Math. .Math. .Math. .Math. 6.sup.th .Math. .Math. .Math. .Math. 7.sup.th .Math. .Math. .Math. .Math. 8.sup.th .Math. .Math. .Math. .Math. Reverse 1 .Math. .Math. .Math. .Math. Reverse 2 .Math. .Math. .Math. .Math.

(57) In the second embodiment, as shown in Table 4, 2-OD overdrive gears for three single-stage planetary gear sets and exemplary configurations of gear ratios logic forms for planetary gear sets:

(58) P1(1.2987/1)P2(0.65789/1)P3(3.125/1). The values in parentheses are two gearshift ratios of each single-row planetary gear set produced. For this, when (R1/S1)3.3478, (R2/S2)1.923, and (R3/S3)2.125 in the planetary gear sets, the gear ratios layout can be obtained. The distribution of eight forward gears: 4.058_3.125_2.67_2.056_1.298_1_0.854_0.6578. The total gear ratio range (4.058/0.6578)=6.169; In the third embodiment, the highest gear is the constant speed direct gear, and the gear ratios of each planetary gear mechanism is configured as P1(1.56/1)P2(1.3/1)P3(3.225/1).

(59) As reference shown in FIG. 1, gears distribution: 6.54_5.03_4.194_3.225_2.028_1.56_1.3_1;

(60) Taken as (out/in) version: 0.1528_0.1987_0.2384_0.31_0.49309_0.641_0.76923_1.

(61) As shown in FIG. 2, in the 0 overdrive gear configuration, the planetary gear set P1 uses two single-row gear sets as subassembly, the first subset planet carrier C1 and the second subgrouping ring gear R1 they are integrally formed in series, with continuing total reduction ratio P1a (1.5)P1b (1.388)=2.0833; eight gears calculation logic as:
P1(2.083/1)P2(0.388/1)P3(4.5454/1);
Gear ratio distribution: 13.152_9.469_6.313_4.545_2.893_2.083_1.388_1; gear ratio range of forward gears is 13.1523.
The compound P1 gear-set characteristic parameters of the 2 subdivided planetary gears can be adjusted, P1a (1.5)P1b (1.333)=2, and the overall gearshift configuration is P1(2/1)P2(1.333/1)P3 (4.1666/1);
gear ratio distribution: 11.111_8.333_5.555_4._166_2.666_2_1.333_1

(62) It should be noted that in the third embodiment, the shift logic sequence of 0 overdrive gear sequence is the same as that in the 1-OD gear logic list, but for second planetary gear train P2, power shifting outputs substantially through direct connection of the arm of the planet carrier C2 to the central main shaft (3). The R2 ring gear bracket is freely mounted and positioned on the central main shaft.

(63) TABLE-US-00005 TABLE 5 Three Typical Configurations with Different OD Gears of an Eight-speed Transmission. Type 0 Overdrive gear 1 Overdrive gear 2 Overdrive With With gear compound compound 3 Simple PGS1 PGS1 gear sets Ratio Ratio Ratio Gears In/out Step In/out Step In/out Step 1.sup.st 11.111 5.376 4.11 2.sup.nd 8.333 1.333 4.135 (1.3) 3.125 (1.315) 3.sup.rd 5.555 1.5 3.2258 (1.28) 2.67 (1.1704) 4.sup.th 4.166 1.333 2.481 (1.3) 2.029 (1.315) 5.sup.th 2.666 1.5625 1.666 (1.48) 1.315 (1.542) 6.sup.th 2 1.333 1.282 (1.3) 1 (1.315) 7.sup.th 1.333 1.5 1 (1.28) 0.854 (1.1704) 8.sup.th 1 1.333 0.76923 (1.3) 0.6493 (1.315) R1 8.444 3.71 2.928 R2 6.333 1.333 2.85 (1.3) 2.25 (1.315)

(64) To sum up, in the eight-speed transmission according to the invention, the planetary gear members in the planetary gear sets P1-P2-P3 are normally independently engaged to transmit torque input/output. The transmission can obtain eight forward torque ratios and two reverse gear ratios through the gang adjustment of clutches CL1-CL5 to transmit the torque from the input shaft to the output shaft via different power train. As shown in Table 5 above, the maximum reduction gear ratio of the first gear can reach 13.1 based on the parameter configuration of the planetary gear sets. In combination with shifting of the clutch, each clutch performs to complete the corresponding displacement to change the engagement path. The load capacity and efficiency is high, the drag friction loss during overall operation is small, and the gear setting is reasonable, thus achieving more prominent economic applicability.

(65) The structure according to the invention can be easily modified into an eight-speed transmission for the pure electric vehicle. After the reverse gear clutch is omitted, only three single-stage planetary gear mechanisms and three shift elements are required. With more gears and simpler shifting, the real-time and comprehensive efficiency of the electric vehicle can be flexibly exerted. Driven by the motor in forward and reverse rotation, three synchronizers shift configuration of engagement sleeve structure being used for the automatic shifting of multi-stage gears, to meet and match requirements for optimal efficiency of the drive motor and the battery pack of the electric vehicle under different operating conditions from low to high speed. The configuration has the characteristics of simple structure, flexible and reliable gear shifting, good operating stability, low power consumption, and high reliability.

(66) As shown in FIG. 3, the pawl and ratchet mechanism produces specific shifting action due to mutual movement and displacement. A ratchet sliding sleeve (33) is plug-in mounted on collar beams of the cage and can move axially, a smooth inside surface (35) in the middle and ring grooves (34) on both sides are provided on the ratchet sleeve. The grooves can be locked and engaged with the pawl elements on the interior planetary gear set. While no engagement, free idling between the ratchet and pawl is formed on the smooth ring internal surface to disconnect the transmission path with the clutch. The ratchet clutch shift torque transmission by allowing the connection and disconnection of the engagement path between the pawls mounted on different members of the planetary gear train and the grooves distributed on the inner ring surface of the ratchet through translocating.

(67) The ratchet clutch is connected to the longeron of the cage by using the outer collar flange (36) of its sleeve, and the ratchet ring gear is driven by an actuator to move transversely along the longitudinal beams of the cage as a whole to adjust different pawls on two rotating elements of each planetary gear train in retracted idling or released engagement transmission states respectively.

(68) The pawls are mounted on the outer rim surface seats of different rotating members of each planetary gear set and can rotate at an angle under the action of a preloaded spring. Due to the action of spring preload and rotating centrifugal force, the pawls always keep their front tip faces the outer layer and are locked by engaging with the grooves on the inner ring surface of the ratchet, thus transmitting torque as a whole. The number of pawls in each group of shifting paths can be increased accordingly on the planetary gear members, such as 6, 7, and 8. The increase in the number of pawls may realize the advantages of tight locking between the pawl and the outer ratchet, low operating noise, large transmission load, small structural wear, and strong impact resistance.

(69) FIG. 4 is a schematic diagram of the shifting process between two gear ratios of the single-row planetary gear mechanism. A ratchet sleeve clutch 33 is used for the transmission. When the transmission operates for shifting, since the pawl can rotate freely and the front tip is always in elastic contact with the inner ring surface of the ratchet, the degree of freedom of spatial profiling of the clutch is increased. When the contact interface between the inner ring smooth surface 35 or the grooves 34 of the ratchet and the pawls translocating, the pawls in the idling state can be directly translated without resistance and stably into the grooves on the inner ring surface of the ratchet and locked with each other. Moreover, the original engagement gear is automatically discharged. There is no restraint or hysteresis due to runout impact caused by shift engagement, and the adaptability is good and stable. When the clutch is engaged in one gear, it is synchronously disengaged from another gear, so that there is no need to reduce engine power for shifting. In the ratchet and pawl locking engagement clutch transmission, there is practically no friction loss caused by drag resistance in the pawl unlocked idling state as compared with the friction plate clutch, and it has rigid fixed engagement characteristics to realize large load power transmission.

(70) (II) New Sixteen-Speed Transmission Series

(71) The sixteen-speed structure is further expanded and upgraded from eight-speed structure. It has four simple planetary gear trains (P1-P4) and six clutches for transmission. The first to sixth clutches CL1, CL2, CL3, CL4, CL5, and CL6 are arranged in this order along the input and main shaft direction.

(72) The shift clutch according to the invention is compatible with the synchronizer gearshift or dog clutch. Plus pawl and ratchet coupling as a full line combination in the frame in operation. The clutches CL1 and CL2 on the first roller cage selectively move to adjust the mutual engagement of different end faces of individual rotating element between the planetary gear sets P1 and P2 to complete four-gear shifting sequences. The gear shifting of P3 is then adjusted by the clutch CL3 to form eight gear ratios and output them through the second cage. Further, via CL4 cooperates with CL5, to modulate the planetary gear mechanism P4 for another extra 2 gear ratio, then sum up total sixteen-speed transmission series. By clutch CL6 derives the gear ratios through the third cage to output shaft. As a result, the rotation of the respective internal elements of planetary gear sets P1-P4 and the different transmission paths realize sixteen forward gears and four to five reverse gears, with simple and reliable shift adjustment.

(73) TABLE-US-00006 TABLE 6 Gear Ring-to-Sun Input Output Stationary Calculation Ratio ratio P1 Ring (R) Planet Sun (1 + S1/R1) Carrier P2 Planet Ring (R) Sun 1/(1 + S2/R2) Carrier P3 Sun Planet Ring 1 + (R3/S3) 2.5 1.5 Carrier P4 Sun Planet Ring 1 + (R4/S4) 5 4 Carrier Ring Planet (R4/S4) 4 Carrier

(74) Table 6 shows the basic configuration and parameters of the sixteen forward gear planetary gear set. The characteristic parameters of P3 and P4 have fixed values in the whole structural sequence. The ring gear R3 of P3 is fixed on the backside of the fixing bracket of S2. The planet carrier C4 and the ring gear R4 of gear set P4 can be locked respectively to generate forward and reverse output or being idle simultaneously. The specific planetary gear mechanism characteristic parameters (ring-to-sun ratios) of P1-P2 may be adjusted to obtain the gear ratios of gearshift with different distribution characteristics and the complete series of gear ratios.

(75) In the fourth embodiment, four simple planetary gear mechanisms and six clutch transmissions form a sixteen-speed (2-0D) transmission series.

(76) As shown in FIG. 5, the gear ratio configurations of the planetary gear sets in the sixteen forward gear 2-OD high gear sequence: P1 is R1 source power input, and C1 is speed-down output or R1 input and R1 direct output; P2 is C2 input, and R2 is speed-up output or R2 input and direct constant speed output; P3 is S3 input, and from C3 is speed-down output, or idling and instead of direct 1:1 model that equal with bypass of P2-R2 output via the main shaft; P4 is always S4 input, and C4 is speed-down output or R4 reverse output reverse gear; or P4 overall constant speed 1:1 output.

(77) TABLE-US-00007 TABLE 7 Logic Diagram of Gear Engagement for Sixteen Forward Gear (2-OD) Ratio Series. Engage P4_CL5 P1_CL1 P2_CL2 P3_CL3 CL4 C4 R4 P4_CL6 Gears C1 R1 R2 C2 C3 R2 C4-S4 (Lock) (Lock) R4 C4 1.sup.st .Math. .Math. .Math. .Math. .Math. 2.sup.nd .Math. .Math. .Math. .Math. .Math. 3.sup.rd .Math. .Math. .Math. .Math. .Math. 4.sup.th .Math. .Math. .Math. .Math. .Math. 5.sup.th .Math. .Math. .Math. .Math. .Math. 6.sup.th .Math. .Math. .Math. .Math. .Math. 7.sup.th .Math. .Math. .Math. .Math. .Math. 8.sup.th .Math. .Math. .Math. .Math. .Math. 9.sup.th .Math. .Math. .Math. .Math. .Math. 10.sup.th .Math. .Math. .Math. .Math. .Math. 11.sup.th .Math. .Math. .Math. .Math. .Math. 12.sup.th .Math. .Math. .Math. .Math. .Math. 13.sup.th .Math. .Math. .Math. .Math. .Math. 14.sup.th .Math. .Math. .Math. .Math. .Math. 15.sup.th .Math. .Math. .Math. .Math. .Math. 16.sup.th .Math. .Math. .Math. .Math. .Math. Reverse .Math. .Math. .Math. .Math. .Math. 1 Reverse .Math. .Math. .Math. .Math. .Math. 2 Reverse .Math. .Math. .Math. .Math. .Math. 3 Reverse .Math. .Math. .Math. .Math. .Math. 4

(78) Logic formula of gear ratios for planetary gear sets: P1(1.25/1)P2(0.666/1)P3(2.5/1)P4(5/1).

(79) By adjusting ring-to-sun ratios of planetary gear mechanisms P1-P2. When (R1/S1)=4 and (R2/S2)=2,

(80) the distribution of sixteen forward gears:

(81) 15.625_12.5_10.416_8.333_6.25_5_4._166_3.333_3._125_2.5_2.083_1.666_1.25_1-0._833_0._666

(82) With reference to the (out/in) gear ratio form, the above arrangement of gear ratios is based on P1(1/0.8)P2(1/1.5)P3(1/0.4)P4(1/0.2).

(83) Gears layout: 0.064_0.08_0.096_0.12_0.16_0.2_0.24_0.3_0.32_0.4_0.48_0.6_0.8_1_1.2_1.5.

(84) The (input/output) characteristic parameters of the planetary gear sets P1 and P2 can be adjusted separately as formula: P1(1.282/1)P2(0.666/1)P3(2.5/1)P4 (5/1), gears layout:

(85) 17.8_13.888_11.876_9.259_6.41_5_4.2735_3.56_3.333_2.777_2.374_1.851_1.28_1-0.854_0.666

(86) As shown in FIG. 5, four single-stage planetary gear mechanisms P1, P2, P3 and P4 as well as six shift clutches CL1, CL2, CL3, CL4, CL5 and CL6 are provided in line. Four connecting members are the first, second and third cages and the central main shaft respectively. This transmission mode forms a clear power flow path that transfers the input torque at the front end of the transmission to the internal elements in parallel distribution, so that the load can be evenly distributed to the planetary gear mechanisms and engagement members, to effectively improve torque transmission efficiency and durability under output conditions. It meets the synchronous shifting action during upshift and downshift, with simple and fast gear shifting, small momentum, and good stability.

(87) In the fifth embodiment, for the shift path/logic of the sixteen forward gear 1-OD gear sequence, four single-stage planetary gear trains are arranged axially in sequence. When the ring-to-sun ratios of P1-P2 are calculated as 2 and 4, the kinematic equation P1(1.5/1)P2(0.8/1)P3(2.5/1)P4(5/1),

(88) Gear ratios layout: 18.75_16_12.5_10_7.56_5_4_3.75_3_2.5_2_1.5_1.2_1_0.8

(89) TABLE-US-00008 TABLE 8 Sequential Logic Diagram of Shift Control Keys for Sixteen Gear (1-OD) Ratio Series. Engage CL5 CL1 CL2 CL3 CL4 C4 R4 CL6 Gears C1 R1 R2 C2 C3 R2 C4-S4 (Lock) (Lock) R4 C4 1.sup.st .Math. .Math. .Math. .Math. .Math. 2.sup.nd .Math. .Math. .Math. .Math. .Math. 3.sup.rd .Math. .Math. .Math. .Math. .Math. 4.sup.th .Math. .Math. .Math. .Math. .Math. 5.sup.th .Math. .Math. .Math. .Math. .Math. 6.sup.th .Math. .Math. .Math. .Math. .Math. 7.sup.th .Math. .Math. .Math. .Math. .Math. 8.sup.th .Math. .Math. .Math. .Math. .Math. 9.sup.th .Math. .Math. .Math. .Math. .Math. 10.sup.th .Math. .Math. .Math. .Math. .Math. 11.sup.th .Math. .Math. .Math. .Math. .Math. 12.sup.th .Math. .Math. .Math. .Math. .Math. 13.sup.th .Math. .Math. .Math. .Math. .Math. 14.sup.th .Math. .Math. .Math. .Math. .Math. 15.sup.th .Math. .Math. .Math. .Math. .Math. 16.sup.th .Math. .Math. .Math. .Math. .Math. Reverse .Math. .Math. .Math. .Math. .Math. 1 Reverse .Math. .Math. .Math. .Math. .Math. 2 Reverse .Math. .Math. .Math. .Math. .Math. 3 Reverse .Math. .Math. .Math. .Math. .Math. 4

(90) FIG. 6 shows a schematic structure of the first planetary gear train of the transmission with two sub member planetary gear sets of simply cascade connecting in series to obtain one large reduction ratio. Two sub gear sets have same dynamic frame. The lower part of the symmetrical rotation center is omitted from the figure. The clutches CL3 and CL4 share the second cage (II) of load-supporting. Because CL6 has few shift frequency and lower engagement speed when forward and reverse gear exchanged, the dog type synchronization can also keep shift accuracy and stability, the clutch action being ensured.

(91) If the first planetary gear set P1 is a compound speed reduction structure with two sub planetary gears in series, P1a (1.333)P1b (1.25)=1.6666, and the sixteen-speed transmission of 1-OD gear ratio kinematic sequence is calculated as follows: P1 (1.666/1)P2(0.8/1)P3(2.5/1)P4(5/1),

(92) Gears layout: 20.833_16.66_12.5_10_8.33_6.66_5_4.166_4_3.333_2.5_2_1.666_1.333_1_0.8

(93) Layout (out/in): 0.048_0.06_0.08_0.1_0.12_0.15_0.2_0.24_0.25_0.3_0.4_0.5_0.6_0.75_1_1.25

(94) In the 6.sup.th embodiment, a 3-OD high gear sequence configuration is provided, where the input shaft is directly connected to the first planet carrier. The first planetary gear set P1 can increase the speed of the input shaft by using the ring gear R1 and transmit the increased speed to the planetary gear set P2 via the first cage (I), or the planet carrier C1 directly conducts the input speed. With three overdrive gears, the sixteen-speed transmission can achieve a gear ratio span of 23.4375, which ensures uniform optimization of step ratios among gearshift while efficiently performing multi-gear synchro synergism, thus improving drivability, such as acceleration performance before and after shifting according to engine speed stability. Moreover, the transmission maintains the high load characteristics of low gear and large reduction gear ratio, thus maximizing the engine driving efficiency

(95) TABLE-US-00009 TABLE 9 Example of 3-OD Gear Sequence Configuration: P1 (0.8/1) P2 (0.666/1) P3 (2.5/1) P4 (5/1). 1.sup.st 2.sup.nd 3.sup.rd 4.sup.th 5.sup.th 6.sup.th 7.sup.th 8.sup.th 9.sup.th 10.sup.th 11.sup.th 12.sup.th 13.sup.th 14.sup.th 15.sup.th 16.sup.th 12.5 10 8.333 6.666 5 4 3.333 2.666 2.5 2 1.666 1.333 1 0.8 0.666 0.533 0.08 0.1 0.12 0.15 0.2 0.25 0.3 0.375 0.4 0.5 0.6 0.75 1 1.25 1.5 1.875

(96) The second row in the table lists the (output/input) ratios reference, and the calculation logic formula is P1(1/1.25)P2(1/1.5)P3(0/0.4)P4(1/0.2).

(97) In the 7.sup.th embodiment, the optimal 0-OD (the top gear is the direct gear) for heavy-duty low gears applications configuration can obtain for sixteen-speed. When P1(1.5/1)P2(1.25/1)P3(2.5/1)P4(5/1) is configured and the ring-to-sun ratios of P1-P2 are calculated as 2 and 4, gear ratios as follow:
23.4375_18.75_15.625_12.5_9.375_7.5_6.25_5_4.6875_3.75_3.125_2.5_1.875_1.5_1.25_1

(98) The transmission here is generally applicable to various heavy-duty machines power trains. It can be used to perform operations in industrial fields such as mining, construction, agriculture, and transportation. For example, it is especially suitable for construction machinery, tractors, wheel loaders, excavators, trucks, and material handlers. It can be used to achieve a variety of complex powertrain tasks.

(99) With reference to FIG. 6, in the 0 overdrive gear configuration, the planetary gear set P1 uses two sub single-row planetary gear mechanisms, and the first sub planet carrier and the second sub ring gear are integrally forged in series, with a total reduction ratio P1a(1.333)P1b(1.25)=1.666, for example, gears kinematic equation built as P1 (1.666/1)P2(1.25/1)P3(2.5/1)P4(5/1), where the ring-to-sun ratio of P2 is still 4, and P2 power shifting output achieved by directly connecting the rotating arm of the planet carrier C2 to the central main shaft (3). The R2 ring gear bracket is freely mounted and positioned on the central main shaft. Thus more robust and uniform lower gear ratios can be derived.

(100) TABLE-US-00010 TABLE 10 0-OD Gear sequence: The values in the second row are reference of the (out/input) manner. 1.sup.st 2.sup.nd 3.sup.rd 4.sup.th 5.sup.th 6.sup.th 7.sup.th 8.sup.th 9.sup.th 10.sup.th 11.sup.th 12.sup.th 13.sup.th 14.sup.th 15.sup.th 16.sup.th 26.04 20.38 15.625 12.5 10.416 8.33 6.25 5.2 5 4.16 3.125 2.5 2.083 1.66 1.25 1 0.038 0.048 0.064 0.08 0.096 0.12 0.16 0.192 0.2 0.24 0.32 0.4 0.48 0.6 0.8 1

(101) In this sequence, the 8th and 9th gears can combine into one gear, that is, the full range of gear ratios is of a 15 forward gear compact mode. The capability of low gears and large gear ratios is conducive to the flexible determination of differential, main reduction ratio of the drive axle main reducer of the power train for heavy-duty wheeled vehicles, which fully improves the power performance of various types of vehicles. For off-road vehicles, engineering vehicles and trailers, the normal additional system reduction transmission components (e.g., 2 speeds rear axle/differential) need to be provided to generate additional tractive effort, which corresponding provides the possibility of streamlining the powertrain and realizes realistic cost savings by this option.

(102) In this way, in the sixteen forward gear transmission, the clutches CL1 and CL2 are reciprocated on the first cage (I), so that the different rotating members of the interior planetary gear sets P1-P2 of the first cage may be selectively adjusted to generate different input/output ratios initially. Further, with CL3-CL4-CL5 in conjunction and CL6 in the end, the torque from the input shaft to the output shaft getting variable speeds through different engagement with the clutch dispatch, relying on the power on-off adjustment generated by the morphological switch of the corresponding planetary gear elements, so that as a gang connected unit to jointly transmit power/torque, or disengaged as a freewheel are easily achieved one by one.

(103) TABLE-US-00011 TABLE 11 Type 0 Overdrive gear 1 Overdrive gear 2 Overdrive gear 3 Overdrive gear Gear Ratio Step Ratio Step Ratio Step Ratio Step 1.sup.st 23.4375 18.75 15.625 12.5 2.sup.nd 18.75 (1.25) 16.666 (1.25) 12.5 (1.25) 10 (1.25) 3.sup.rd 15.625 (1.2) 12.5 (1.333) 10.416 (1.2) 8.333 (1.2) 4.sup.th 12.5 (1.25) 10 (1.25) 8.333 (1.25) 6.666 (1.25) 5.sup.th 9.375 (1.333) 8.333 (1.2) 6.25 (1.333) 5 (1.333) 6.sup.th 7.5 (1.25) 6.666 (1.25) 5 (1.25) 4 (1.25) 7.sup.th 6.25 (1.2) 5 (1.333) 4.166 (1.2) 3.333 (1.2) 8.sup.th 5 (1.25) 4.1666 (1.2) 3.333 (1.25) 2.666 (1.25) 9.sup.th 4.6875 (1.0665) 4 (1.0416) 3.125 (1.066) 2.5 (1.066) 10.sup.th 3.75 (1.25) 3.333 (1.2) 2.5 (1.25) 2 (1.25) 11.sup.th 3.125 (1.2) 2.5 (1.333) 2.083 (1.2) 1.666 (1.2) 12.sup.th 2.5 (1.25) 2 (1.25) 1.666 (1.25) 1.333 (1.25) 13.sup.th 1.875 (1.333) 1.5 (1.2) 1.25 (1.333) 1 (1.333) 14.sup.th 1.5 (1.2) 1.2 (1.25) 1 (1.25) 0.8 (1.25) 15.sup.th 1.25 (1.2) 1 (1.333) 0.833 (1.2) 0.666 (1.2) 16.sup.th 1 1.25 0.8 (1.25) 0.666 (1.25) 0.5333 (1.25)

(104) Table 11 shows the 0-OD, 1-OD, 2-OD and 3-OD overdrive configuration transmission series formed by the internal variable structures of four simple planetary gear sets with different characteristic parameters (ring-to-sun ratios). The first planetary gear set P1 may reduce (or increase) the speed of the input shaft and transmit the reduced (or increased) speed to the second planetary gear set P2 via the first cage (I) or transmit the input speed at constant speed. The second planetary gear set P2 selectively receives the upshift or downshift speed from the first planetary gear set P1 or the constant speed from the input shaft, and subsequently outputs them through the central main shaft via the ring gear R2 (or the planet carrier C2) to the third and fourth planetary gear sets P3-P4 at a variable or constant speed by the second cage (II), to establish and obtain more suitable gear ratios and ratio progressions, which diversifies the gear ratio sequence of the transmission, thus meeting various vehicle driving needs.

(105) To sum up, four planetary gear trains operating independently include the planetary gear units P1, P2, P3 and P4 arranged sequentially in the direction from the input shaft to the output shaft. Based on the characteristic parameters (ring-to-sun ratios) of planetary gears, the gear ratios of each gear and the ratio step between gears is in a uniform range. The transmission efficiency is high, and the double-pointed clutches control is simple and fast. The drives of the clutches are adjusted by the hydraulic assembly, SVA electric assembly, or pneumatic actuator through reciprocating driving accordingly.

(106) The torque transmission mechanism is a synchronizer, or ratchet and pawl coupling clutch, with an optional dog clutch. The transmission has a wide range of gear ratios, a high degree of freedom and flexibility in setting the gear train, and a more uniform upshift/downshift amplitude. Multi-gear interval overpass operations can be preferred according to real-time jump operating conditions. The transmission mechanism is capable of achieving both heavy-duty performance with low gear ratios and reduced fuel consumption of high-speed driving with high gear ratios. It contributes to the reduction of the size, weight and manufacturing cost of the automatic transmission and has excellent overall performance.

(107) As described above, compared with the existing gearshift products, according to the invention, the first, second, and third (fourth) simple planetary gear sets are configured and independently cascade connected in series to form an integrated planetary gear transmission system. In different single-stage planetary gear trains, clutch is selectively engaged and locked with different components members, so that the input/output transmission is carried out by different rotating members in different planetary gear sets, thus changing the final output gear ratio of the power flow to form different gears distribution.

(108) The transmission assembly structure according to this embodiment is novel in design. The internal gear ratios of gang connected single-stage planetary gear mechanisms are changed through synchronizer and the ratchet/pawl coupling clutch to integrate and combine different gear ratios to complete output. It is different structural design from the conventional automatic transmission in terms of power transmission path and framework. The transmission assembly is characterized by smaller size, compact structure, simple operating principle, high efficiency, and good operating reliability and durability.

(109) The embodiments above are only the preferred embodiments of the invention. The associated operation of gear sets and the calculation of gear ratios will be readily understood by those skilled in the art. The invention can be modified and varied in various ways. Any modifications, equivalent replacements and improvements made within the invention or the design spirit and principle of the invention are intended to be covered by the protection scope of the invention.