Power Transmission

Abstract: A power transmission has a planetary gear arrangement (10) controlled by a plurality of selectively engageable clutches (56,58,64) and brakes (62,82) to provide forward and reverse speed ratios. A selectively operable input gear arrangement (66,80) is utilized to establish a negative ratio of substantially 1:1 between the transmission input (60) and the planetary gear arrangement. The selective use of the negative ratio input permits substantially equal forward and reverse drive ratios. The negative ratio input arrangement also permits extended coverage when employed with some of the other forward ratios established by the planetary gear arrangement.

Description: This invention relates to power transmissions and more particularly to planetary gear power transmissions having a substantially equal forward and reverse gear ratio. EP-A-0,239,205 discloses a power transmission in accordance with the pre-amble of Claim 1. A power transmission in accordance with the present invention is characterised over EP-A-0,239,205 by the features specified in the characterising portion of Claim 1. It is an object of this invention to provide an improved power transmission having a low forward drive ratio substantially equal to a low reverse drive ratio. It is another object of this invention to provide an improved multi-speed power transmission having an input gear train for establishing a speed reversal input drive to a planetary gear mechanism which cooperates therewith to establish a forward speed ratio between the input and output members of the transmission. It is yet another object of this invention to provide an improved multi-speed power transmission as described in the preceding object wherein the speed reversal gear train is utilized during at least three forward speed ratios and also wherein the lowest forward speed ratio is substantially equal to a reverse speed ratio. The present invention will now be described, by way of example, with reference to the following description, and the accompanying drawings, in which:- Figures 1 to 4 are schematic representations of power transmission gear arrangements incorporating the present invention. Referring to the drawings, wherein like characters represent the same or corresponding parts throughout the several views, there is seen in Figure 1 a planetary gear arrangement, generally designated 10, which is comprised of two simple planetary gear units 12 and 14 (first and second planetary gear means). The planetary gear unit 12 has a sun gear 16 (first member), a ring gear 18 (third member) and a planet carrier 20 (second member), on which is rotatably mounted a plurality of pinion gears 22 meshing with the sun gear 16 and ring gear 18. The planetary gear unit 14 includes a sun gear 24 (first member), a ring gear 26 (second member) and a planet carrier 28 (third member) which has rotatably supported thereon a plurality of pinion gears 30 which mesh with the sun gear 24 and ring gear 26. The sun gear 16 and ring gear 26 are interconnected by a hub 32 which, in turn, is connected to a planetary input shaft 34. The planet carrier 28 and ring gear 18 are interconnected by a hub 36 and are also drivingly connected with a planetary input shaft 38. The sun gear 16 is drivingly connected with another planetary input shaft 40. The planet carrier 20 of planetary gear unit 12 is drivingly connected to a transmission output shaft 42 (output means) which is adapted to provide a rotary drive for a vehicle, not shown. An additional planetary gear arrangement 44 is shown in dashed lines and includes a sun gear 46, a ring gear 48 and a planet carrier 50 on which is rotatably supported pinion gears 52 meshing with the sun gear 46 and ring gear 48. The planet carrier 50 of planetary gear arrangement 44 is operatively connected with a (selectively engageable) brake 54. The sun gear 46 is drivingly connected with the transmission output shaft 42 and the ring gear 48 is drivingly connected to the ring gear 18. The planetary gear arrangement 44 is shown in dashed lines because it is an optional gear which will permit a "deep low" gear ratio and a "deep reverse" ratio. The planetary input shafts 34 and 38 are operatively connected with respective clutch arrangements 56 (first clutch means) and 58 (second clutch means). The clutch arrangements 56 and 58 are drivingly connected with a transmission input shaft 60 (input means). The planetary input shaft 40 is operatively connected to a brake 62 and a clutch 64 (third clutch means). The clutch 64 is drivingly connected with the transmission input shaft 60. The transmission input shaft 60 is drivingly connected to a torque converter and engine, not shown. A reverse gear or negative unity mechanism 66 (reverse gear means) is operatively connected between the transmission input shaft 60 and the planetary input shaft 40. The reverse gear mechanism 66 includes a sun gear 68 which is drivingly connected with the planetary input shaft 40, a ring gear 70 which is drivingly connected with the transmission input shaft 60, and a plurality of pinion gears 72 which are rotatably mounted on a planet carrier 74. The pinion gears 72 have a first toothed portion 76 meshing with the ring gear 70 and a second toothed portion 78 meshing with the sun gear 68. The planet carrier 74 is connected with a (selectively operable) friction brake 80 (torque transmitting means). The ring gear 18 and planet carrier 28 are selectively controllable by a friction brake 82. The friction brakes 54, 62, 80 and 82 are conventional units which are preferably selectively fluid operated to connect respective gear members to a stationary portion of the transmission housing 84. When using the optional planetary gear arrangement 44, this transmission will provide eight forward gear ratios and two reverse gear ratios. The lowest forward gear ratio is established by the selective engagement of brakes 80 and 54. When the brake 80 is engaged, the reverse gear mechanism 66 will cause the input rotation of the transmission input shaft 60 to be reversed prior to causing rotation of the planetary input shaft 40. The reverse gear mechanism 66 is sized to provide a substantially 1:1 negative gear ratio between the transmission input shaft 60 and the planetary input shaft 40. With the planetary input shaft 40 rotating opposite to the input shaft 60 and the brake 54 engaged, the transmission output shaft 42 will rotate in the same direction as the transmission input shaft 60 but at a greatly reduced speed and increased torque. For example, if gears 70, 76, 78 and 68 have teeth 81, 23, 13 and 45, respectively, a negative or reverse ratio of 1.017 will result. To establish the maximum reverse ratio, the brake 54 is engaged while the clutch 64 is engaged, and the remaining clutches and brakes are disengaged. With the clutch 64 engaged, the planetary input shaft 40 will rotate in the same direction as the transmission input shaft 60. With the planetary input shaft 40 rotating in the forward direction and the brake 54 engaged, the transmission output shaft 42 will rotate opposite or in the reverse of the transmission input shaft 60 at a reduced speed and increased torque. Since the planetary input shaft 40 rotates at essentially the same value during these two ratios, the transmission output shaft 42 will also rotate at the same numeric values relative to the transmission input shaft 60 during these ratios. The first forward speed ratio is established by the engagement of brake 80 and brake 82. With these devices engaged, the planetary input shaft 40 will drive the sun gear 24 in a direction opposite to the rotation of transmission input shaft 60. Since the planet carrier 28 is stationary, the ring gear 26 will rotate in a forward direction driving with it the sun gear 16, which in turn will cause forward rotation of the planet carrier 20 and therefore transmission output shaft 42. The reverse ratio is obtained by engaging the clutch 64 and brake 82 while the other clutches and brakes are disengaged. In this reverse ratio, the planetary input shaft 40 is driven forwardly. The forward rotation of planetary input shaft 40 will cause reverse rotation of ring gear 26 and therefore sun gear 16, which will result in reverse rotation of transmission output shaft 42. The only difference between the first forward ratio and the reverse ratio is the use of the reverse gear mechanism 66 which, as previously explained, is substantially 1:1. Therefore, the first forward ratio and the reverse ratio will be substantially equal in numeric value. The second forward ratio is established by disengaging brake 80 and engaging clutch 56. With clutch 56 and brake 82 engaged, the planetary gear arrangement 10 will provide a reduced forward speed ratio. To establish the third forward speed ratio, the brake 82 is disengaged while the brake 80 is engaged. With the brake 80 engaged, the planetary input shaft 40 will be driven opposite to the engine or transmission input shaft 60. This will permit the planet carrier 28 and therefore ring gear 18 to rotate forwardly thereby increasing the speed of transmission output shaft 42 relative to the second ratio. The fourth forward ratio is established by engaging the brake 62 and disengaging the brake 80. With the brake 62 engaged, and the ring gear 26 driven forwardly, the planet carrier 28 and therefore ring gear 18 will be driven forwardly at a speed ratio greater than that accomplished in the third ratio which will result in a higher speed ratio to the transmission output shaft 42. To establish the fifth forward speed ratio, the clutch 58 is engaged while the brake 62 is disengaged. Since both clutches 58 and 60 are connected to both planetary gear arrangements 12 and 14, a 1:1 ratio is established between the transmission input shaft 60 and the transmission output shaft 42. To establish the sixth forward speed ratio, the brake 62 is engaged while the clutch 56 is disengaged. With these friction devices engaged, an overdrive ratio is established in the planetary gear arrangement 14 which is imposed on the ring gear 26 and therefore sun gear 16. The ring gear 18 is driven at the same speed as the transmission input shaft 60 and the sun gear 16 is driven faster than the transmission input shaft 60 resulting in an overdrive ratio to the transmission output shaft 42. To establish the seventh forward speed ratio, the brake 80 is engaged while the brake 62 is disengaged. With these friction devices established, the sun gear 24 is driven in a direction opposite to the transmission input shaft 60. This results in a higher overdrive ratio being established in the planetary gear arrangement 14. It should be appreciated that a higher overdrive ratio at planetary gear arrangement 14 will result in an increase in the speed of sun gear 16 and therefore transmission output shaft 42. The power transmission arrangement shown in Figure 2 utilizes the same planetary gear units 12 and 14 as those utilized in Figure 1, and therefore has the same clutches 56 and 58 and the same brakes 62 and 82 associated therewith. A reverse gear mechanism 66A is utilized in Figure 2 which replaces the reverse gear mechanism 66 in Figure 1. The reverse gear mechanism 66A has an input gear 90 and an output gear 92. These input and output gears 90 and 92 are interconnected by a reverse idler 94 and a transfer countershaft 96. The output gear 92 is drivingly connected with the planetary input shaft 40. The input gear 90 is selectively connectible through a synchronizer clutch 97 with a transfer shaft 98, which in turn is connectible with the transmission input shaft 60 through a friction clutch 64A. The synchronizer clutch 97 is also movable to selectively connect the transfer shaft 98 with the planetary input shaft 40. The synchronizer clutch 97 is moved leftward to connect input gear 90 with transfer shaft 98 during low, first, third and seventh gear ratios, and is moved to the right to connect transfer planetary input shaft 98 with shaft 40 during the reverse and low reverse ratios. The clutch 64A is engaged in the low forward, first, third and seventh ratios, and in the low reverse and reverse ratios. The remaining clutches and brakes are engaged to establish the same ratios, as explained above for Figure 1. The power transmission shown in Figure 3, includes the same planetary gear units 12 and 14, as described for Figures 1 and 2, but has omitted therefrom the optional planetary gear arrangement 44. If additional ratio coverage is desired, the optional planetary unit can be included. The gear arrangement shown in Figure 3 has a reverse gear or negative unity mechanism 66B which includes an input shaft 100, an input gear 102, an idler gear 104 and an output gear 106. The input shaft 100 is drivingly connected with a (selectively engageable) friction clutch 108 which is also drivingly connected with a transmission input shaft 60A. The transmission input shaft 60A is drivingly connected through a pair of transfer gears 110 and 112 to an input shaft 60B, which is drivingly connected with the clutches 56, 58 and 64. The power transmission arrangement in Figure 3 will provide seven forward speed ratios and one reverse speed ratio. The clutch 108 is engaged during the first, third and seventh forward speed ratios, such that the reverse gear arrangement 66B is utilized to provide an input drive to the planetary gear unit 14. The remaining clutches and brakes are engaged, as described above, for Figure 1 to establish the seven forward drive ratios and the reverse ratio. The reverse ratio provided by input and output gears 102 and 106 is substantially equal to the ratio provided by the transfer gears 110 and 112. The power transmission shown in Figure 4 is very similar to that shown in Figure 3, with the exception that a reverse gear mechanism 66C replaces the reverse gear mechanism 66B. The reverse gear mechanism 66C includes a synchronizer clutch 114 which is selectively operable to connect the input shaft 100 to a forward input gear 116 and a reverse input gear 118. The forward input gear 116 is drivingly connected through a transfer gear 120 to the planetary input shaft 40, and the reverse input gear 118 is drivingly connected through an idler gear 122 and a transfer gear 124 to the planetary input shaft 40. The synchronizer clutch 114 essentially replaces the friction clutch 64 of Figure 3. For the first, third and seventh forward gear ratios, the synchronizer clutch 114 is moved rightward to connect the input shaft 100 with the reverse input gear 118 and for the reverse gear ratio, the synchronizer clutch 114 is moved leftward to connect the forward input gear 116 with the input shaft 100. The clutch 108 is engaged during first, third and seventh forward ratios and during the reverse ratio. The remaining clutches and brakes are engaged to establish the gear ratios within the transmission, as described previously. Each of the friction clutches and brakes is preferably a disc type unit which is engaged by a fluid operated piston, such devices are well known. Also, the control of such units is understood by those skilled in the art. Likewise, the synchronizer clutches 97 and 114 are conventional mechanisms which may be constructed in accordance with any of the currently available mechanisms which are known to provide this function. It should be appreciated that the first forward ratio and the reverse ratio in each of the power transmissions described herein are substantially identical with the exception of the reverse gear mechanism utilized for the forward ratio. It should also be appreciated that this same reverse gear mechanism provides input drives for the planetary gear unit 14 during the seventh forward speed ratio. By utilizing the reversing ratio during the third and seventh forward speed ratios, additional ratio coverage is provided which is useful in establishing smooth acceleration of a vehicle from a standing start to maximum vehicle speed. It should also be noted, when the optional planetary gear arrangement 44 is utilized, the low forward ratio and the low reverse ratio are substantially identical since the only difference is the reverse gear mechanism utilization. The addition of the planetary gear arrangement 44 increases the overall ratio coverage of the transmission and is found to be most useful in large hauling type vehicles which must be able to accelerate while carrying very heavy loads.

Claims: A power transmission comprising input means (60) for delivering power; first and second selectively engageable clutch means (56,58) operatively connected with the input means for receiving power therefrom; output means (42) for delivering power from the power transmission; first planetary gear means (12) for establishing various drive ratios within the power transmission including a first member (16) connected to the first clutch means, a second member (20) connected to the output means, and a third member (18); second planetary gear means (14) for establishing various ratios within the power transmission including a first member (24), a second member (26) connected to the first member of the first planetary gear means and a third member (28) connected with the second clutch means and with the third member of the first planetary gear means; reverse gear means (66) for selectively providing a negative ratio between the input means and the first member of the second planetary gear means and including a first gear member (70) drivingly connectable with the input means and a second gear (68) member drivingly connected with the first member of the second planetary gear means; selectively engageable torque transmitting means (80) to establish the negative ratio; and selectively engageable brake means (62,82) cooperating with the first and second planetary gear means, the reverse gear means, and the first and second clutch means to establish a plurality of forward drive ratios in the power transmission; characterised by third selectively engageable clutch means (64) operatively connectable with the input means (60) for receiving power therefrom; in that the first member (24) of the second planetary gear means (14) is connected to the third clutch means; and in that the brake means (62,82) also cooperates with the third clutch means. A power transmission as claimed in Claim 1, wherein the selectively engageable torque transmitting means comprises selectively engageable reverse clutch means (97,108) for connecting the reverse gear means (66A,B,C) with the input means (60,60A) to establish the negative ratio input drive to the second planetary gear means (14). A power transmission as claimed in Claim 2, wherein the selectively engageable reverse clutch means includes a synchronizer clutch (97, 114). A power transmission as claimed in any one of Claims 1 to 3, having a first forward gear ratio, a highest forward gear ratio and a reverse gear ratio. A power transmission as claimed in Claim 4, wherein the reverse gear means (66) is selectively controlled for providing an input drive to the first member (24) of the second planetary gear means (14) during at least two forward gear ratios including the first of said forward gear ratios. A power transmission as claimed in Claim 5, wherein the reverse gear means (66) is selectively controlled to provide an input drive during the highest of the forward gear ratios. A power transmission as claimed in Claim 5, wherein the third selectively engageable clutch means (64) is operable to control an input drive to the first member (24) of the second planetary gear means (14) to provide the reverse gear ratio.

Inventors: KLEMEN DONALD F, KLEMEN, DONALD F.
Applicants: GEN MOTORS CORP, GENERAL MOTORS CORPORATION
Application year: 1990
Main classification: F16H3
IPC codes: F16H3