Power transmission

Abstract: A power transmission has a housing comprised of an end support, an outer casing, and an end cover; an input shaft; an output shaft; and a pair of identically dimensional countershafts (42). Each countershaft has ratio gears (56,125) rotatably disposed thereon and meshing with respective output gears rotatable with the output shaft and a friction clutch (132) is provided for each ratio gear. Each countershaft has two friction clutches mounted on centrally disposed surfaces (74,76,78) thereon, in opposite facing relation. The four friction clutches which are disposed on the countershafts are identical in construction.

Description: This invention relates to power transmissions and more particularly to power transmissions having multiple countershafts as specified in the preamble of claim 1. Such a power transmission is known from, for example, US-A-3,138,965. Power transmissions with multiple countershafts generally have a housing assembly formed with at least two members that are aligned by dowel pins. These members generally include a casing and an end cover or bell housing. The bell housing generally encloses a torque converter or friction clutch. The countershafts are supported in the casing at one end and in the end cover at the other end. The input shaft is generally supported in the end cover and the output shaft is supported in the casing. The alignment and parallelism of the shafts is determined by the doweling between the casing and cover and the bearing bores in each of the casing and end cover. The machining of the required dowel and bearing bores is subject to manufacturing tolerances which must be held within close limits for quiet, efficient and durable operation of the power transmission. As is well known, close tolerance machining adds to the overall cost of the product. In the prior art multiple countershaft transmissions, each countershaft in the power transmission has a different or distinct configuration. The bearing mounting surfaces and the gear mounting surfaces are machined to specific tolerances and dimensions for each bearing and gear. Lube passages are unique from one countershaft to another. These and other differences require that manufacturing and inventory capabilities must be available for each individual shaft configuration. The number of repair and replacement parts proliferates with each countershaft configuration. The end result is an increase in cost for the power transmission. A power transmission in accordance with the present invention is characterised over US-A-3,138,965 by the features specified in the characterising portion of Claim 1. The present invention seeks to overcome the disadvantage of a proliferation of countershafts known in multiple countershaft prior art power transmissions. In the preferred embodiment, the present invention has an input shaft, two countershafts and an output shaft. Gear members are rotatably supported on the countershafts and are selectively connectible therewith through hydraulically operated friction clutch assemblies. Each of the countershafts are identical in physical dimensions, that is, they have identical length, diameters, cross drilled and longitudinally drilled holes. Each of the clutch assemblies is identical and includes an annular housing, which is secured to the countershaft by welding, an annular piston, which is slidably disposed on the countershaft and within the annular housing, a plurality of friction plates alternately splined to the annular housing and to an output hub and a return spring assembly. Each of these individual items are identical within the clutch assembly reducing the number of inventory parts and part selection necessary during manufacturing of the countershaft assemblies. Each countershaft assembly is comprised of two centrally located clutch assemblies, a head gear splined thereto, ratio gear members rotatably supported thereon and bearings for supporting the gear members. Each of the annular pistons slidably engages the respective countershaft as previously mentioned. Each annular piston has an inner annular skirt portion which partially covers a cooling passage in the countershaft for the respective clutch assembly. The skirt portion cooperates with the cooling passages in the countershaft to provide a flow restriction to limit the amount of fluid flow distributed to the clutch assembly whenever the clutch assembly is not engaged. During engagement, the skirt portion is moved sufficiently to permit a secondary opening in the skirt portion to allow full cooling flow from the lube passage in the countershaft to the friction plates to assist in maintaining the internal temperature of the friction plates at or below the desired level. This invention preferably provides an improved power transmission wherein a pair of identical countershafts are rotatably supported in an end cover and have ratio gears disposed thereon meshing with respective head gears on an input shaft and an output shaft, and further wherein the ratio gears are selectively connectible with the countershafts to provide a plurality of forward ratio gears and a reverse gear ratio with said reverse gear ratio being provided by one gear member which is rotatably mounted on one of the countershafts and connectible therewith by a friction clutch, meshing with another gear which is rotatably mounted on the other countershaft and drivingly connected to the output through the ratio gear set which provides the lowest forward gear ratio. This invention preferably provides an improved power transmission wherein a pair of parallel countershafts are rotatably supported in a transmission casing assembly and where each of the countershafts has identical diametral and length dimensions from end to end, and wherein a pair of friction clutch assemblies are secured to each of the countershafts with the friction clutch assemblies each having identical structure including an annular housing secured to respective countershafts, an annular piston slidably disposed on the countershaft and in the annular housing, and a plurality of interleaved friction plates which are alternately splined and spaced with the annular housing and an output hub, and further wherein each of the output hubs are secured with a respective gear member or with a synchronizer clutch assembly. The present invention will now be described, by way of example, with reference to the accompanying drawings, in which:- Figure 1 is a block diagram showing the interconnection of Figures 2a through 2d; Figures 2a through 2d are cross-sectional elevational views of various components within the power transmission, which when assembled as shown in Figure 1, will provide a cross-sectional representation of the preferred embodiment of the power transmission; Figure 3 is an end view of the power transmission taken with the end cover removed to show the positioning of the input shaft and the countershafts and some of the gearing within the outer casing; Figure 4 is block diagram showing the positioning of Figures 5a and 5b; Figures 5a and 5b, when combined, depict the one typical countershaft assembly with the gearing and clutch members assembled thereto; Figure 6 is an end view taken along line 6--6 of Figure 2d; Figure 7 is a plan view taken along line 7--7 of Figure 6; and Figure 8 is a plan view of a countershaft prior to assembly. Figures 2a through 2d, when combined as shown in Figure 1, depict a cross-sectional elevational view of a multiple countershaft power transmission having a multiple piece housing comprised of an end support 10, an outer casing 12, a bell housing or engine mounting member 14 (which with end support 10 defines a first end cover), and a second end cover 16. The end support 10 defines an assembly support plate and is a substantially annular disc member having bearing support recesses 18 and 20 for countershaft members and recess 22 for input shaft support. The outer casing 12 is located on the bell housing 14 by dowel pins, not shown, and the second end cover 16 is located on the outer casing 12 by dowel pins, not shown. The outer casing 12 is a substantially cylindrical member with an annular semi-closed end portion having recesses 24 and 26 formed therein for countershaft bearing support and a recess or opening 28 formed therein for output shaft bearing support. Preferably the bearings 30 through 40 which are disposed in the recesses or openings 18 through 28, respectively, are identical and therefore selection of any particular bearing to be placed in a specific recess or opening is eliminated. The bearings 30-40, however, can be of different sizes if the conditions and assembly warrant. The bearings 30 and 36 support a countershaft 42 while the bearings 32 and 38 support a countershaft 44. The bearing 34 rotatably supports and input shaft 46 and the bearing 40 rotatably supports an output shaft 48. The output shaft 48 is also rotatably supported by a needle bearing 50 and a thrust bearing assembly 52 which cooperate with the input shaft 46. In particular, the neddle bearing 50 and thrust bearing assembly 52 cooperate with a (first) head gear 54, which is formed integral with or otherwise secured to the input shaft 46. The head gear 54 meshes with a (second) head gear 56 splined to countershaft 42 and a third head gear 58 splined to countershaft 44. The countershafts 42 and 44 are dimensionally identical from end-to-end, however, their assembly structure is considerably different as seen in Figures 2a through 2d. The assembly of countershaft 42 is shown in Figures 5a and 5b. In Figure 5a, it can be seen that the countershaft 42 has reduced diameter ends 60 and 62 and a pair of bearing surfaces 64 and 66 adjacent the reduced diameter ends 60 and 62, respectively. Looking along the countershaft 42 from reduced diameter end 60 toward reduced diameter end 62, the outer surface of countershaft 42 has a spline portion 68, a constant diameter portion 70, a slightly larger diameter portion 72, a piston mounting surface or diameter 74, a central clutch housing support surface 76, a piston mounting surface 78, a pair of constant diameter surfaces 80 and 82, and a reduced diameter surface 84 adjacent the bearing surface 66. The countershaft 44 has an identical structure and the corresponding surfaces are given the same numerical designation with an "a" suffix. The countershaft 42 has a pair of longitudinal passages 86 and 88 that extend inwardly from the reduced diameter ends 60 and 62, respectively, and terminate at 90 and 92, respectively, aligned with the central clutch housing support surface 76. Each passage 86 and 88 has disposed therein tubes 94 and 96, respectively, which cooperate with the respective passages 86 and 88 to form a pair of annular lube or lubrication passages 98 and 100, and have formed therein a respective clutch feed passage 102 and 104. The annular lube passage 98 has a pair of radially extending fluid passages 106 and 108 which communicate with the larger diameter portion 72 and piston mounting surface 74, respectively. A radial passage 110 communicates fluid from the clutch feed passage 102 to a position on piston mounting surface 74 adjacent central clutch housing support surface 76. The annular lube passage 100 has communicating therewith three passages 112, 114 and 116 which are effective to communicate fluid with the piston mounting surface 78 and constant diameter surfaces 80 and 82, respectively. A radial passage 118 provides fluid communication from clutch feed passage 104 to the piston mounting surface 78 adjacent the central clutch housing support surface 76. As will become more apparent in the following description, the passages 106, 108, 112, 114 and 116 provide cooling and/or lubrication fluid for various gear members and clutch assemblies which are disposed on the countershaft 42 while the radial passages 110 and 118 are effective to distribute clutch engagement pressure to clutch assemblies disposed on the countershaft 42. The countershaft 44 has identical passages and tube assemblies secured thereto. These corresponding components have been given the same numerical designation with an "a" suffix. The head gear 56 is splined to the spline portion 68 in abutment with the bearing 30 which is positioned on the countershaft 42 by a locking ring 120. Adjacent head gear 56 is a spacer sleeve 122 which is disposed in abutment with the head gear 56 and a thrust bearing assembly 124. The thrust bearing assembly 124 is in abutment with a ratio gear member 125 which is rotatably supported on the countershaft 42 by needle bearings 126. The ratio gear member 125 is in abutment with a thrust bearing 128 which is positioned by a shoulder 130 formed adjacent the large diameter portion 72. A clutch assembly 132 (friction clutch means) is disposed on the countershaft 42 and includes an annular housing 134 which is welded or otherwise secured to the countershaft 42, an annular piston 136 which is slidably disposed in the annular housing 134 and on the countershaft 42, a plurality of friction plates 138 and 140 which are alternately splined to the annular housing 134 and an output hub 142, respectively. The output hub 142 is welded or otherwise secured to the ratio gear member 125. A return spring assembly 144 is effective to urge the annular piston 136 rightward as viewed in Figure 5a to a disengaged position. When fluid pressure is supplied to clutch feed passage 102 and radial passage 110, the annular piston 136 will move leftward to cause frictional engagement between the alternately spaced friction plates 138 and 140, whereby the ratio gear member 125 will be drivingly connected to the countershaft 42. A second clutch assembly 132a is secured to the countershaft 42 adjacent the clutch assembly 132 and is a mirror image thereof. The corresponding components of clutch assembly 132a are given the same numerical designation as the components in clutch assembly 132 with an "a" suffix. The output hub 142a is secured to a gear assembly 145 which includes a reverse gear 146, a sleeve 148 and a ratio gear member 150 (first ratio gear). The gear assembly 145 is positioned on the countershaft 42 by a pair of thrust bearings 152 and 154 and is rotatably supported on the countershaft 42 by needle bearings 156 and 158. The thrust bearing 154 is disposed in abutment with the bearing 36 which is positioned on the countershaft 42 by a locking ring 160. The reduced diameter ends 60 and 62 have formed therein seal grooves 162 which cooperate with a dowel and seal sleeve 164 at reduced diameter end 60 and on a sleeve 164a at reduced diameter end 62. The reduced diameter end 60 has a feed passage 166 which is in communication with the annular lube passage 98 and the reduced diameter end 62 has a fluid passage 166a which is in communication with the annular lube passage 100. The dowel and seal sleeve 164 is pressfit in the bell housing 14 and has a dowel fit with the end support 10 thereby providing for alignment between the end support 10 and the bell housing 14. The sleeve 164a is pressfit in the second end cover 16. The bearing support recesses 18 and 20 are held in close tolerance with the bores in the end support for the dowel and seal sleeve 164 and another dowel and seal sleeve 164b. Thus, the bearing locations for the input end of the countershafts 42 and 44 and input shaft 46 are closely controlled to the bell housing 14. The countershaft 44 has assembled thereon the bearing 32 and head gear 58. A spacer sleeve 168 is disposed between the head gear 58 and a thrust bearing 170 which cooperates with a thrust bearing 172 to position a ratio gear 174 which is rotatably supported on the countershaft 44 by needle bearings 176. The ratio gear 174 is selectively connectible with the countershaft 44 through controlled operation of a clutch assembly 132b which is identical in structure with clutch assembly 132 and has an output hub 142b secured to the ratio gear 174. The remaining components of clutch assembly 132b are identical with the components of the clutch assembly 132 and have therefore been given the same numerical designation with a "b" suffix. Adjacent the clutch assembly 132b is a clutch assembly 132c which is a mirror image of clutch assembly 132b and is comprised of components identical with clutch assembly 132 such that the same numerical designates with a "c" suffix are utilized. The output hub 142c is secured to a sleeve 180 which is positioned on the countershaft 44 by a pair of thrust bearings 182 and 184, and rotatably supported by needle bearings 186. The sleeve 180 is splined to a conventional synchronizer assembly 187 (synchronizer clutch means) which is selectively operable to connect a ratio gear 188 (second ratio gear) or a reverse gear 190 with the output hub 142c of clutch assembly 132c. Therefore, selective positioning of the synchronizer assembly 187 and actuation of the clutch assembly 132c will establish a drive relationship between countershaft 44 and either ratio gear 188 or reverse gear 190. Reverse gear 190 is a member of the reverse gear train and is disposed in meshing relation with the gear assembly 146 which is rotatably supported on countershaft 42. The ratio gear 188 is supported on a needle bearing 192 and is positioned longitudinally by the thrust bearing 184, a thrust bearing 194, a sleeve 196 and the bearing 38. As previously mentioned, the assembly of countershaft 44 has components identical with the countershaft 42 and these components have been given an "a" suffix. In particular, the tubes 94a and 96a are effective to establish the cooling and lubrication passages and the clutch apply passages for fluid which is to be delivered through the dowel and seal sleeve 164b and a sleeve 164c in a manner similar to that described above for dowel and seal sleeve 164 and the sleeve 164a, respectively. The head gear 54 has secured thereto a hub member 142d which is a component in a clutch assembly 132d which is identical with the clutch assembly 132. In the alternative, the clutch assembly 132d can be constructed with fewer friction plates 138d and 140d, and the annular housing 134d can be axially shortened. This has the benefit of reducing the axial length of the power transmission. This axial shortening of the clutch assembly 132d results in less torque capacity which is permitted because the engine torque is not increased by the head gear set 54-56 or 54-58 prior to delivery to the clutch assembly 132d. While this clutch assembly would then have to be treated differently in production, the overall material savings and axial length savings may overshadow this drawback. The remaining gear ratios 1, 2, 3, 5 and reverse all have the engine torque increased, as a result of the respective head gear sets, prior to delivery to the respective clutch assemblies which establish these ratios. The remaining components of clutch assembly 132d have been given the same numerical designation as the corresponding components of clutch assembly 132. The annular housing 134d of clutch assembly 132d is secured to the output shaft 48 such that selective engagement of the clutch assembly 132d, by fluid pressure in a passage 200, will enforce a direct drive connection between the input shaft 46 and the output shaft 48. The output shaft 48 has splined thereto a plurality of ratio output gear members 202, 204, 206 and 208 which are in constant mesh with respective ratio gears 174 on countershaft 44, and ratio gear member 150 on countershaft 42. The ratio gear member 150 and 208 will provide a low forward gear ratio between the input shaft 46 and the output shaft 48 upon selective engagement of the clutch assembly 132a. Second gear operation is provided by ratio gear 174 and ratio output gear member 202 upon selective engagement of clutch assembly 132b. A third forward gear ratio is provided by rightward selection of the synchronizer assembly 187 and selective engagement of the clutch assembly 132c. A fourth forward gear ratio is provided by selective engagement of clutch assembly 132d as described above. This is a direct drive ratio. A fifth and overdrive forward gear ratio is provided by selective engagement of the clutch assembly 132. Quite obviously, only one clutch assembly 132 through 132d can be engaged at any one time. The selective engagement of these clutch assemblies and their interchange can be controlled by many of the currently available and well known hydraulic control circuits. To establish a reverse drive ratio in the transmission, the synchronizer assembly 187 is established in its leftward position and the clutch assembly 132c is selectively engaged. With this power flow path engaged, power flow is from the input shaft 46 through head gears 54, 58, countershaft 44, reverse gears 190 and 146, and ratio gear members 150 and 208 and head gear set 54-58 to the output shaft 48. The reverse gears 146 and 190 provide for reversing of the rotational direction during a reverse ratio and provide first and second members respectively of reverse gear means. The meshing ratio gear members 150, 208 and head gear set 54-58 establish the speed ratio of the reverse gear ratio which is identical with the low forward gear ratio. Since the gear reversing or reverse idling function is provided by reverse gears 146 and 190, a separate idler gear axis is not required with the present invention. A park gear 210 is secured to the reverse gear 146 and therefore is connectible with the output shaft 48 through the low forward ratio of ratio gear members 150 and 208. A conventional parking pawl mechanism can be utilized to engage the park gear 210 such that a mechanical advantage is gained between the park gear 210 and the output shaft 48. This advantage is equal to the ratio of the low forward ratio gear members 150, 208 without benefit of the head gear set 54-56. During assembly of the power transmission, the positioning of the countershafts 42 and 44 and the output shaft 48 is controlled by cam shim pack assemblies 212, 214 and 216, respectively. These cam shim pack assemblies 212-216 are identical in structure and include an inner cam ring 218, an outer cam ring 220 and a locking ring 222. The locking ring 222 prevents rearward axial displacement of the outer cam ring 220 relative to the outer casing 12. The inner cam ring 218 is nonrotatably supported in the outer casing 12 by a plurality of radially extending tabs 224. The inner cam ring 218 is disposed in abutment with bearings 36, 40, and 38, respectively, for each of the cam shim pack assemblies. During assembly, the outer cam ring 220 is rotated relative to the inner cam ring 218 thereby causing relative axial movement between the outer casing 12 and the inner cam ring 218 to enforce longitudinal positioning of the countershafts 42 and 44 and the output shaft 48 through the respective bearing assemblies which support the shafts in the outer casing 12. A tabbed washer 225 is connected at the outer diameter with the outer casing 12, and the inner diameter of the tabbed washer 225 is displaceable into an opening 227 in the outer cam ring 220 to prevent movement between the inner and outer cam rings 218, 220 after assembly. The adjustment mechanisms can be constructed in accordance with the bearing adjustment mechanism shown in US Patent No. 3,901,568. The countershafts 42, 44, input shaft 46 and output shaft 48 and their respective components, can be assembled onto the end support 10 as the end support is vertically positioned on an assembly line or bell housing 14. This assembly can be competed with the outer casing 12 having been installed over the gear assemblies. Thus, as the power transmission assembly progresses along the assembly line, continuous visual inspection is available to ensure that proper gear positioning and gear meshing is occurring. The outer casing 12 can then be installed over the power transmission assembly and secured to the bell housing 14 or in the alternative, to the end support 10. When this assembly is completed, the adjustment mechanisms provided by cam shim pack assemblies 212, 214 and 216 are manipulated to ensure proper placement of the shaft assemblies and then the second end cover 16 can be positioned and secured to the outer casing 12. It should be noted that the end support 10 and bell housing 14 are both aligned with the countershafts 42 and 44 through the dowel and seal sleeve components 164 and 164b. This eliminates the need for additional dowel locations being machined through the various components which would have to rely on some dimensional location on the end support 10 and outer casing 12 for proper alignment. The commonalty of the countershafts 42 and 44 greatly reduce the inventory and machining requirements for these components. Further, material savings are established through the use of the common clutch assemblies. The savings are occasioned not only by the reduction in the number of different components to be utilized, but also in the fact that the likelihood of misassembly of the clutch assemblies is greatly reduced. Each annular piston 136 through 136d has an inner skirt portion 226 which is slidably disposed on the respective portions of the countershafts 42, 44 and output shaft 48. The inner skirt portion 226 partially or completely covers the respective lube and cooling passage, such as 108, for clutch assembly 132 when the clutch is disengaged. During movement of the annular piston 136 to engage the clutch, an opening 228 in the inner skirt portion 226 will be placed in fluid communication with the respective lube passage thereby permitting unrestricted fluid flow from the annular lube passage 98 to the alternately spaced friction plates 138 and 140. This permits controlled cooling flow during clutch engagement when it is desirable to have such flow but limits or eliminates fluid flow in this circuit when the clutch assemblies are disengaged. With the reduced cooling flow requirements, the system flow requirements are also reduced which makes more fluid available for lubrication circuits and reduces the volume of the control pump required to supply the fluid to the control system. It should be noted that the arrangement of the seals, clutch feed passages and lube passages minimize oil leakage and, the oil leakage that may occur from the clutch feed is directed to the lube circuit. This feature contributes to a reduction in the required volume of the supply pump. Also, as pointed out previously, the extended skirts on the annular pistons also contribute to a reduced pump requirement by reducing the amount of continuous oil flow to the system. Typically, countershaft power transmissions are manually shifted and range changes are accomplished with synchronizer clutches. A power transmission designed in accordance with the present invention will shift automatically with hydraulically applied clutch packs. Due to the increased axial length of the clutch packs over synchronizer clutches, it becomes impractical to maintain the same number of ranges with a single countershaft power transmission. The additional countershaft allows the clutch assemblies to be assembled side-by-side, and thus minimize shaft length and reduce shaft deflection. The present power transmission configuration uses identical countershafts at both locations. This provides a cost savings in inventory by the elimination of a part number, unit cost reduction through increased production volume, and design simplicity with use of common parts. The countershaft 42 is shown in Figure 8. As mentioned above, countershafts 42 and 44 are identical so that the following description of countershaft 42 is equally valid for countershaft 44. The countershaft 42 includes the seal grooves 162, feed passage 166 and the longitudinal passages 86 and 88 at each reduced diameter end 60, 62 which combine with the (oil transfer) tubes 94, 96 (shown in Figures 2a, 2b) to provide lube oil to the bearings 30-40 and clutch assemblies 132 and also clutch apply pressure to each clutch assembly. Identical bearings and retaining rings are assembled at grooves 244 in bearing surface 64, within which the locking ring 120 is positioned. Identical head gears 56, 58 are assembled to spline portion 68. Constant diameter portion 70 provides a shoulder for location of the head gear 56 and grinding relief of the adjacent bearing surface of larger diameter portion 72. Larger diameter portion 72 is a journal for needle radial bearings which support input gears (ratio gear members 125) on each countershaft 42, 44. Radially extending passage 106 provides lube to needle bearings 126 supported on the surface of larger diameter portion 72. Identical clutch assemblies 132 - 132c, consisting of a retaining ring, retrun spring assembly 144, annular piston 136, O-ring, and annular housing 134, are assembled on the piston mounting surfaces 74, 78, and on central clutch housing support surface 76, with the retaining ring positioned in a groove 246. Radially extending fluid passages 106, 108, 110, 118 and 112 provide lube to the respective friction plates 138,140 and apply chambers. Constant diameter surface 80 is a journal for needle bearing(s) 156,186 which support reverse gear 146 on the countershaft 42 and the sleeve 180 on the countershaft 44. Radially extending passage 116 provides lube to the bearing surfaces on constant diameter surface 80. Extended groove 247 is a clearance diameter for the thrust bearing 184 on the countershaft 44. Constant diameter surface 82 is a journal for the thrust bearing 182 for the ratio gear 188 on the countershaft 44. Radially extending passage 116 provides lube to the needle bearings 192 supported on constant diameter surface 82. Reduced diameter surface 84 is a journal for a needle bearings 158 which supports the ratio gear member 150 on the countershaft 42. Identical bearings and retaining rings are assembled at bearing surface 66 using a groove 249. A further savings in material and inventories is also provided through the use of commonalty between various thrust bearings and needle bearings which are components within the power transmission assembly. It should further be appreciated that wide latitude is permitted in the selection of the various gear ratios and that the changing of the gear ratios after assembly can be accomplished without significant changes in the other components of the transmission. One other advantage of the present invention is the provision for identical low and reverse drive ratios. Attention is drawn to related patent application nos. 91200386.0 (EP-A-0,449,343) and 91200387.8 (EP-A-0,449,344), filed the same day as the present application.

Claims: A power transmission having an input shaft (46); an output shaft (48); first and second countershafts (42,44) rotatably supported in housing means (10,12,16); a plurality of dissimilar ratio gear means (125,146,150,174,188,190) rotatably supported on each countershaft and being selectively connectable therewith; input gear means (54,56,58) rotatable by the input shaft; and output gear means (202,204,206,208) on the output shaft; the input gear means and the output gear means cooperating with the countershafts, clutch assemblies and the ratio gear means to selectively establish a plurality of drive ratios between the input shaft and the output shaft; characterised in that the countershafts are identical in length and diametral dimensions from end to end and have centrally disposed surface (74,76,78) for mounting two clutch assemblies (132,132a,132b,132c) on each countershaft in opposite facing relation; and in that the ratio gear means are selectively connectable with the countershafts by selective engagement of the two said clutch assemblies. A power transmission as claimed in Claim 1, comprising a housing having an outer casing (12), a first end cover (10,14), and a second end cover (16); the input shaft (46) being rotatably mounted in the first end cover; the output shaft (48) rotatably supported on the input shaft and in the outer casing; the countershafts (42,44) being rotatably supported in one of the end covers and in the outer casing for rotation on respective axis parallel with the input shaft and the output shaft; the input gear means comprising a first head gear (54) continuously rotatable with the input shaft, and second and third head gears (56,58) meshing with the first head gear and continuously rotatable with the first and second countershafts respectively; the ration gear means comprising a plurality of ratio gears rotatably supported on each of the countershafts including a first ratio gear (150) on the first countershaft and a second ratio gear (188) on the second countershaft, and reverse gear means including a first member (146) rotatable with the first ratio gear and a second member (190) rotatably supported on the second countershaft; synchronizer clutch means (187) selectively and alternatively connectible with the second ratio gear and the second member of the reverse gear means; the output gear means being drivingly connected with the output shaft and including a first ratio output gear (208) meshing with the first ratio gear and a second ratio output gear (206) meshing with the second ratio gear; and the two said clutch assemblies comprising a first selectively engageable friction clutch means (132a) for connecting the first ratio gear with the first countershaft for completing a low forward drive ratio between the input and output shafts, a second selectively engageable friction clutch means (132c) for selectively connecting the synchronizer clutch means to the second countershaft and cooperating therewith to selectively alternatively complete a reverse drive ratio between the input shaft and the output shaft and another forward drive ratio between the input and output shafts, and third and fourth selectively engageable clutch means (132,132b) for selectively establishing two additional forward drive ratios between the input and output shafts.

Inventors: KLEMEN DONALD, RASZKOWSKI JAMES ALLEN, KLEMEN, DONALD, RASZKOWSKI, JAMES ALLEN
Applicants: GEN MOTORS CORP, GENERAL MOTORS CORPORATION
Application year: 1991
Main classification: F16H3
IPC codes: F16D25, F16H3