T-drive powertrain assembly

Abstract: A T-drive power transmission mechanism comprising an engine with a crankshaft disposed perpendicularly with respect to the axis of the transmission gearing, the transmission including a hydrokinetic torque converter and a right angle drive connecting the torque output element of the gearing with transverse axle half shafts, the torque input element of the gearing being connected drivably to a crankshaft gear located at the midpoint of the crankshaft.

Description: Technical Field The present invention relates to power transmission mechanisms for automotive vehicles having an internal combustion engine wherein the engine and the transmission mechanism form a common powertrain assembly with the axis of the engine arranged transversely with respect to the geometric axis of the transmission mechanism to form a "T" configuration. Background of the Invention In the case of front-wheel drive vehicles, it is common practice to mount an internal combustion engine in transverse disposition with respect to the fore-and-aft center axis of the vehicle within a forward engine and transmission compartment. Multiple ratio planetary gearing is used to deliver torque from the engine to axle half-shafts disposed in generally parallel disposition with respect to the engine axis. An example of an arrangement of this kind is the transaxle assembly disclosed in U.S. Patent 4,509,389. The transaxle assembly of the '389 patent is adapted to be used with an internal combustion engine mounted in a so-called transverse or "east-west" direction within an engine compartment for a front-wheel drive vehicle. The transaxle includes a hydrokinetic torque converter mounted concentrically with respect to the axis of the engine crankshaft. The turbine of the hydrokinetic torque converter is connected to a torque input shaft for a multiple ratio planetary gear assembly arranged on an axis that is spaced from and parallel to the turbine shaft axis. A torque transfer drive comprising a chain-and-sprocket assembly is adapted to transfer torque from the turbine shaft to the input shaft of a multiple ratio gear system. Fluid pressure operated clutch and brake assemblies are used to establish selectively the various ratios in the gear system to permit adequate torque ratio coverage for the powertrain. A final drive assembly located on the torque output side of the gearing multiplies the output torque of the planetary gearing and delivers it to a differential carrier, the side gears of the carrier being connected to axle half-shafts through universal joints. An example of a transaxle having a transverse engine is shown in U.S. Patent 4,368,649. Arrangements of this kind are referred to in the automotive industry as so-called "U-drives". They usually are found in front-wheel drive vehicles with transversely mounted engines, but they may be adapted also for various other powertrain arrangements for vehicles presently used in the automotive industry. The torque delivered from the transaxle differential assembly in an arrangement of this kind may be transferred to a torque output shaft concentrically disposed within the planetary gearing. The output shaft extends in a transverse direction to the torque input ends of each of the two axle half-shafts for the traction wheels. Another example of a prior art transaxle and engine assembly is shown in U.S. Patent 4,607,541 wherein the hydrokinetic torque converter is disposed concentrically with respect to the multiple ratio planetary gearing on the axis of the crankshaft of the internal combustion engine. As in the case of the arrangement of the '389 patent, the torque output from the planetary gearing is delivered to a driven shaft coaxially disposed with respect to the planetary gearing. A final drive assembly located between the torque converter and the planetary gearing is adapted to deliver torque to the carrier of a differential assembly. Axle shafts are connected to the side gears of the differential assembly, but these half-shafts are in spaced, parallel disposition with respect to the planetary gearing. This is in contrast to the arrangement of the '389 patent where the axle shafts are coaxially disposed with respect to the planetary gearing. An example of an engine and transaxle assembly wherein driving axle half-shafts are disposed perpendicularly with respect to the axis of the engine crankshaft is shown in U.S. Patent 4,056,988. In this case, a differential gear assembly has its side gears disposed perpendicularly with respect to the axis of the planetary gearing, the latter being spaced from and parallel to the axis of the engine crankshaft. Engine and transaxle assemblies of the kind described in the preceding paragraphs require a relatively large engine and transaxle compartment in the vehicle. This restricts the ability of the automotive designer to develop an optimum profile for improved aerodynamics and for reduced vehicle weight. To minimize the space required for the engine and transaxle, and thus increase the space dedicated to the passengers and cargo, attempts have been made to arrange the engine and the transmission gearing with the axis of the engine crankshaft generally perpendicular to the axis of the gearing in a T-drive configuration without a significant offset between the axes. This requires a direct driving connection between the engine crankshaft and the torque input drive for the multiple ratio gearing. An example of an attempt to develop such a configuration is described in an automotive publication entitled "Auto Notizie", published by the Cizeta Moroder Company of Italy, in May 1989. That publication describes a V-16 engine with two independent crankshafts arranged end-to-end. The adjacent ends of the crankshafts are connected at their ends to a drive gear, and engine crankshaft torque is delivered through an idler to a bevel gear which serves as a torque input shaft for a multiple ratio transmission having an axis disposed perpendicularly with respect to the crankshaft axis. Another prior art example of an attempt to locate the transmission axis perpendicularly with respect to the engine crankshaft axis is shown in Japanese Patent Application Publication 56-20861, dated February 26, 1981. That publication shows a drive gear located on one end of an engine crankshaft. The drive gear is connected through a friction clutch to the input shaft of a multiple speed transmission located on an axis parallel to the crankshaft axis. The multiple speed transmission, in turn, is connected to a transversely disposed output shaft that is geared to the torque output shaft of a multiple ratio transmission through torque transfer gearing. A T-drive arrangement having cross-axis gears located externally of the engine is described in French Patent 933,078, issued April 9, 1948. This disclosure describes a driveline having a central clutch located at the torque output side of the cross-axis gearing in a cross-axis housing. The output side of the neutral clutch is connected to a concentric torque input shaft for a multiple ratio transmission. The input side of the transmission is connected to a conventional driveshaft for the rear vehicle wheels. This arrangement has an obvious space penalty in both the transverse direction and the fore-and-aft direction. Another example of a cross-axis drive somewhat similar to the design of the French patent is shown in U.S. Patent 1,991,575. In the case of the design of the '575 patent, right angle drive bevel gears are located on the torque output side of a neutral clutch for the engine. No attempt is made in the design of either the '575 patent or the French patent to provide a cross-axis drive connection having an integrated engine crankshaft and gearing arrangement. Each of these prior art teachings of T-drives is characterized by an undesirable complexity because of the additional gearing required to transfer torque from the crankshaft to the input side of the multiple ratio gearing. They are characterized also by an undesirable space penalty which would make it impossible to use such a design approach in a vehicle design that requires a low profile in the forward engine compartment and which requires a maximum cargo and passenger space. Summary of the invention According to the present invention, there is provided a T-drive vehicle drivetrain assembly as hereinafter set forth in Claim 1 or the appended claims. The features of claim 1 except for the last two features concerned with the location of the drive gear and the final transaxle gear drive are known from EP-A-439761. The powertrain assembly of the preferred embodiment of the present invention, comprises an internal combustion engine with a crankshaft axis situated in a transverse direction with respect to the center plane of the vehicle. The assembly is mounted within the forward engine compartment. It includes also multiple ratio gearing having a gearing axis disposed perpendicularly with respect to the crankshaft axis wherein the axis of the gearing defines a "T" configuration with respect to the engine crankshaft axis. A geared connection is provided between the torque input side of the multiple ratio gearing and the engine crankshaft, the geared connection including a crankshaft bevel gear located between the crankshaft ends. The crankshaft gear engages directly a torque input pinion for the transmission, thus making it possible to reduce to a minimum the overall longitudinal dimensions in the north-south direction for the powertrain package. For any given road clearance, the profile of the powertrain package is also reduced with respect to the corresponding profile of a conventional engine and transaxle assembly of the kind shown, for example, in prior art U.S. Patent No. 4,509,389. The vehicle length for a vehicle incorporating the improved powertrain of the present invention can be reduced relative to prior art designs while maintaining energy absorbing characteristics that are necessary during a front-end vehicle crash. The improved T-drive design of the preferred embodiment of the present invention includes a bevel gear situated at a strategic position on the engine crankshaft. A relatively large pitch diameter crankshaft gear and transmission input pinion are arranged in meshing cross-axis driving relationship in such a way that gear tooth loading is reduced while providing acceptable pitch line speeds at the mesh points for the gear teeth. The crankshaft gear is located at a position displaced from the midpoint of the crankshaft, the amount of the displacement being approximately equal to the pitch radius of the torque input bevel gear for the transmission with which the crankshaft gear meshes. The crankshaft gear is located at an effective nodal point for the first natural torsional frequency for the two crankshaft portions situated on either side of the crankshaft gear. It is possible to arrange the engine crankshaft in a transverse direction with the axis of the transmission gearing extending in a longitudinal direction with respect to the center plane of the vehicle without requiring the crankshaft gear to be located off the mid-center position. The transmission mechanism of the powertrain assembly in the preferred embodiment includes multiple planetary gear elements situated on a main transmission axis and a hydrokinetic torque converter situated on an axis in parallel disposition with respect to the transmission axis, the latter being perpendicularly disposed with respect to the engine crankshaft axis. A torque transfer drive is used to transfer turbine torque of the hydrokinetic torque converter from one axis to the other as the impeller of the hydrokinetic torque converter is driven by the transmission input gear that meshes with the crankshaft gear. Displacing the hydrokinetic torque converter axis from the mid-position of the crankshaft gear in this manner will not result in an undesirable increase in the overall transverse dimension of the powertrain housing because the offset between the axis of the planetary gear assembly and the axis of the hydrokinetic torque converter compensates for the offset between the converter axis and the crankshaft gear. In one embodiment of the powertrain assembly, a front-wheel drive differential is situated between the engine and the hydrokinetic torque converter. Front-wheel drive half shafts are connected drivably to side gears of the differential. The offset between the converter and the planetary gearing makes it possible for the final drive torque output shaft to extend across the center plane of the converter from a torque output pinion of the multiple ratio planetary gearing to the torque input ring gear of the front axle differential without interference and with an economy of space. This does not result in an increase of the transverse dimensions of the powertrain housing. In a further preferred feature of the invention, provision is made for locating a transaxle differential gear system for the powertrain assembly at a location between the engine and the torque converter of the automatic transmission that forms a part of the powertrain assembly. This makes it possible to reduce the fore-and-aft distance between the engine crankshaft axis and the axis of the torque output shafts for the final drive differential of the transaxle. This reduces front-end overhang for the vehicle and permits a shortening of the overall vehicle length while permitting a reduced profile height. Provision may also be made in this embodiment of the invention for transferring torque through a torque transfer drive between the power output element of the multiple ratio gearing and the torque input ring gear of the final drive differential assembly of the transaxle while avoiding interference between the torque converter structure and the torque transfer members that distribute torque to the final drive differential. Brief Description of the Drawings The invention will now be described further, by way of example, with reference to the accompanying drawings, in which: Figures 1A, 1B and 1C, taken together, show a cross-sectional view of a forward portion of a powertrain assembly adapted for a front-wheel drive vehicle embodying the improved T-drive configuration of my invention. Figure 1D is a cross-sectional view of a two-speed transfer drive between the torque converter turbine and the multiple ratio gearing for the powertrain assembly. Figures 1E, IF and 1G, taken together, show a cross-sectional view of the rearward geared position of the powertrain assembly. Figure 2 is a cross-sectional assembly view of a portion of the planetary gearing shown in Figures 1E, IF and 1G as seen from the plane of section line 2-2 of Figures 1E, IF and 1G. Figure 3 is a schematic representation of an engine seen generally in cross-section and a torque input bevel gear for the transmission that forms a part of the powertrain assembly of my invention. Figure 4 is an illustration of a crankshaft, shown partly in elevation, for a six-cylinder engine wherein the crankshaft gear is situated at a mid-position on the crankshaft axis. Figure 5 is a four-cylinder engine crankshaft showing a mid-position crankshaft gear, together with a schematic illustration of a piston and piston rod which cooperate with a crank portion of the crankshaft. Figure 6 is a schematic line diagram of the planetary gear portion of the gearing assembly illustrated in Figures 1E, 1F and 1G. the gearing assembly includes alternate torque output shafts Figure 7 is a chart that shows the clutch and brake engagement and release pattern for the planetary gearing shown schematically in Figure 6. Figures 8A, 8B and 8C, taken together, show an alternate embodiment of my invention which includes a four-speed ratio planetary gear system wherein the axis of the gearing coincides with the axis of the transmission input gear and wherein the transaxle differential gearing is located between the hydrokinetic torque converter of the transmission and the engine. Figure 9 is a schematic representation of the transmission planetary gearing of Figures 8A, 8B and 8C. Figure 10 is a chart that shows the clutch and brake engagement and release pattern for the planetary gearing of Figure 9. Particular Description of the Invention In Figures 1A and 1B, reference numeral 10 designates a hydrokinetic torque converter which is disposed in a torque transfer path between an internal combustion engine, schematically shown at 12 in Figure 1A, and multiple ratio gearing for a planetary transmission designated generally by reference character 14 in Figure 1F. In Figures 1A and 1B, the impeller of the torque converter 10 is a bladed toroidal fluid flow member 16 having an impeller housing 18. A drive plate 20 is bolted at its outer periphery to the radially outward portion of the impeller housing 18 as shown at 22. A radially inward portion of the drive plate 20 is bolted at 24 to transmission input drive shaft 26, the axis of which is situated perpendicularly with respect to the axis of the crankshaft of the engine 12. The torque converter 10 includes a turbine 28 and a bladed stator 30 located between the flow exit section of the turbine 28 and the impeller 16. Torque converter 10 is situated in a transmission housing portion 32, which is bolted to the engine cylinder housing. Impeller housing 18, which is located within the transmission housing portion 32, encloses a lockup clutch disc assembly 34 having annular friction surface 36 situated adjacent an annular clutch plate 38 secured to the inner wall of the housing 18. A damper assembly 40 establishes a resilient connection between the clutch disc assembly 34 and clutch plate hub 42, which is splined at 44 to turbine hub 46. Hub 46, in turn, is splined at 50 to turbine sleeve shaft 52. The converter 10 multiplies torque delivered to the impeller 16. When the speed ratio is high, turbine torque developed by the converter is transferred directly to the turbine sleeve shaft 52. When the lockup clutch is applied, a frictional driving connection is established between friction disc 38 and clutch disc 34. The torque applied to the clutch disc 34 is transferred through tangentially mounted springs of the damper assembly 40 to the clutch plate hub 42 which, as explained, is splined to the turbine hub. This establishes a bypass torque flow path around the hydrokinetic torque flow path of the torque converter 10. Stator 30 is mounted on stator sleeve shaft 54, and overrunning brake 56 establishes a torque reaction for the stator 30 during operation of the converter in the torque multiplication mode and accommodates freewheeling motion of the stator 30 in the direction of motion of the impeller and the turbine when the torque converter 10 achieves a high speed ratio coupling point. A two-speed torque transfer drive mechanism is shown at 58 in Figure 1D. This includes a first torque transfer gear 60 and a large pitch diameter torque transfer gear 62 which are mounted on drive shaft 64. Large diameter gear 62 is secured at 66 to the drive shaft 64 and is journalled at 68 to end wall 70 of housing portion 72. Clutch housing 74 is splined at 76 to the shaft 64, which drives gear 62. Clutch housing 74 is journalled on member 78 by bearing 80. Member 78 is connected to turbine sleeve shaft 52 as seen in Figure 1B. Externally splined clutch discs of a multiple disc clutch assembly 82 are carried by the clutch housing 74 and cooperate with internally splined friction discs 84 secured to clutch member 86. Member 86 is secured to the gear 60, the latter being journalled by tapered roller thrust bearings 88 on shaft 64 as seen in Figure 1D. Clutch housing 74 defines an annular cylinder 90 in which is positioned clutch piston 92, which cooperates with cylinder 90 to define a pressure chamber. When fluid pressure is supplied to the pressure chamber through radial passage 94, a hydrostatic force is applied to the piston which engages the friction clutch assembly, thus connecting the shaft 64 to the gear 62. Gears 60 and 62 mesh, respectively, with gears 96 and 98. Gear 98 is connected directly to torque input shaft 100 for the planetary gear system. An overrunning coupling in the form of a roller clutch is disposed between gear 98 and hub 102, the overrunning coupling being designated by reference numeral 104 as well as the designation OWC3. The planetary gear system shown in Figure 1 is based upon the planetary gear system shown in U.S. Patent 4,509,389. Reference may be made to that patent for the purpose of supplementing this disclosure in order to assist in understanding the mode of operation of the planetary elements of the gear system in Figure 1. Torque input sleeve shaft 100 in Figures 1E, 1F and 1G is splined at 106 to clutch sleeve shaft 108. Sleeve shaft 108 is splined at 110 to forward clutch cylinder 112 and is connected directly also to cylinder housing 114 for direct clutch 116. Housing 114 serves as a cylinder housing also for intermediate clutch 118. The forward clutch, which is designated generally by reference character 120, comprises discs carried by the clutch housing 112 and companion discs carried by overrunning clutch race 122. The companion inner race 124 is secured to brake drum 126, which is adapted to be anchored by friction brake band 128. Overrunning clutch races 122 and 124 form a part of an overrunning clutch assembly generally designated by reference character 130. The friction discs of the forward clutch 120 are engaged as pressure is distributed behind annular piston 132 disposed in a cylinder defined by the cylinder housing 112. The direct clutch 116 comprises friction discs carried by the clutch housing 114 and cooperating friction discs carried by outer overrunning clutch race 134, which forms a part of overrunning coupling 136. The inner race 140 of the overrunning clutch 136 is connected by means of clutch member 142 to the brake drum 126. Torque transfer member 144, seen in Figures 1E and 1F, extends axially and surrounds intermediate clutch 146 and distributes torque to sun gear 148. Sun gear 148 forms a part of a first planetary gear unit 150 which includes also a ring gear 152, carrier 154 and planet pinions 156 carried by the carrier 154 in meshing engagement with the ring gear 152 and the sun gear 148. Carrier 154 is secured to sleeve shaft 158 which is splined to clutch member 160 of the intermediate clutch 146. The clutch 146 includes friction discs carried at its outer periphery. These discs register with clutch discs carried by the cylinder housing 114. Housing 114 defines an intermediate clutch cylinder which receives annular piston 162. When the pressure chamber defined in part by the piston 162 is pressurized, the piston engages the friction discs of the clutch 146. Similarly, cylinder member 114 defines an annular cylinder that receives annular piston 164. When pressure is admitted to the pressure chamber partly defined by the piston 164, the clutch discs of the direct clutch assembly 116 are frictionally engaged. Carrier 154 of the planetary gear unit 150 is adapted to be braked by brake disc assembly 166. Brake disc assembly 166 is engaged during reverse drive. It is defined in part by brake cylinder 168 secured to the transmission housing portion 170. Brake cylinder 168 receives annular brake piston 172, which defines a pressure chamber that may be pressurized to effect frictional engagement of the discs of the brake assembly 166. Carrier 154 is connected drivably to ring gear 174 of the second simple planetary gear unit 176. Ring gear 152 of gear unit 150, shown in Figures 1F and 1G, is connected directly to carrier 178 of the gear unit 176. Sun gear 180 of gear unit 176 is connected to brake drum 182, which is adapted to be anchored by brake band 184 during low and intermediate speed ratio operations. Pinions 186 are carried by carrier 178 in meshing engagement with the ring gear 174 and sun gear 180. Sun gear 174 is journalled on support sleeve shaft 188 and is adapted to be connected to brake member 190 through overrunning brake 192. Brake member 190 carries low speed ratio friction brake discs 194 which cooperate with friction brake discs 196 for a low speed ratio friction brake identified generally by reference character 198. A low speed ratio brake piston 200 is received in brake cylinder 202 formed at one end of the transmission housing 170. When the pressure chamber defined in part by the piston 200 is pressurized, the low speed ratio brake 198 is energized, thus anchoring the sun gear 180. Carrier 178 is splined at 204 to sleeve 206, which in turn is connected directly to power output shaft 208 extending through the planetary gear units and the transmission clutch and brake assemblies. The left end of output shaft 208, as seen in Figures 1B and 1C, is journalled by needle bearing assembly 210, seen in Figure 1E. The left end of shaft 208 carries power output gear 212 which meshes with output gear 214 mounted on shaft 216. Shaft 216 is journalled in bearing assembly 218 and is adapted to rotate on an axis that is offset sufficiently from the axis of the torque converter to avoid interference with the impeller housing 18. The left end of the shaft 216 carries output bevel gear 220, which engages drivably ring gear 222 of a front axle differential assembly 224. Bevel ring gear 222 is carried by differential housing 226 which encloses differential side gears 228 and 230. These mesh with spaced differential pinions 232 and 234. Side gear 228 is splined to a first axle half-shaft 236, and side gear 230 is splined to a companion axle half-shaft 238. Half-shaft 238 extends through sleeve shaft 240 in a direction transverse to the axis of the torque converter and is located between the engine 12 and the drive plate 20 of the torque converter. The outboard end of the sleeve 240 is journalled by bearing 242 in bearing support 244 carried by the transmission housing. One front traction wheel for the vehicle is connected by means of the universal joint, schematically illustrated at 246, to the outboard end of the shaft 238, a splined connection 248 being provided for this purpose. Shafts 236 and 238 are journalled in their respective sleeve shafts by suitable bushings. A drive plate 20 of the torque converter, as mentioned previously, is connected to the torque input shaft shown at 26 in Figure 1B. A torque input bevel gear 252 connected to the shaft 26 meshes with crankshaft gear 254, shown schematically in Figures 1A and 1B. In the embodiment of Figures 1A through 1G, the engine has eight cylinders, as schematically illustrated. Gear 254 is located at the crankshaft center with four cylinders on one side thereof and four cylinders on the opposite side thereof. Crankshaft gear 254 has an axis that coincides with the axis of the crankshaft. It meshes with the gear 252 to form a right angle drive as illustrated. Thus, the offset of the axes of the torque converter and the planetary gearing makes it possible for the gear 254 to be located directly on the center of the crankshaft. In Figure 3, we have shown in generally schematic cross-sectional form the relationship of the gear 252 to the crankshaft. When the gear is assembled in this embodiment, the gear 252 is straddle mounted in the transmission housing wall 256. Straddle mounted bearings shown at 258 and 260 are situated on opposite sides of the gear 252. Bearing 260 is located in a bearing opening 262 in the engine housing wall 256. The hub of gear 252 is splined at 264 to the shaft 250, the inboard end 266 of the shaft 250 being supported by the bearing 258, which is located in a bearing opening formed in engine housing boss 268. The crankshaft, as will be explained with reference to Figures 3, 4 and 5, is comprised of crank portions, one of which is shown in Figure 3 at 270. One end 272 of a piston rod 274 is secured to the crank portion 270 by a crankshaft bearing cap 276. Engine piston 278, which is connected to the piston rod 274, reciprocates in engine cylinder 280, thus creating a combustion chamber of variable volume as shown at 282. Valve ports 284 and 286 communicate with the engine combustion chamber in known fashion. The flow of air-fuel mixture to the combustion chamber and the flow of exhaust gasses from the combustion chamber is accommodated by the intake and exhaust ports that register with their respective port openings. Valve operating camshaft and rocker arm assemblies 288 control the operation of the valves (not shown). The circular path followed by the crankshaft end of the piston rod 274 is illustrated schematically by the family of crankshaft positions illustrated generally by reference character 290. Shown in Figure 4 is a six-cylinder engine crankshaft. The crankshaft ends 292 and 294 are supported in known fashion in the engine housing. Bearing portions 296, 298 and 300 are provided at intermediate locations between the crankshaft ends. These register with and are supported by bearing support walls for the engine cylinder block, a portion 302 of the engine support walls being shown in Figure 4, and a companion bearing support housing portion 304 being located on the opposite side of the crankshaft. Support wall portion 302 registers with bearing portion 306 of the crankshaft, and support wall portion 304 registers with crankshaft bearing support portion 308. The gear 254 for the crankshaft illustrated in Figure 4 is secured to the crankshaft generally in the transverse plane of bearing portion 298, which is at the center of the crankshaft. The plane of the pitch diameter for the gear 254 is generally midway between crank portions 310 and 312. In the six-cylinder crankshaft of Figure 4, two crank portions 314 and 316 are located on one side of the gear 254 and two other crank portions 318 and 320 are located on the opposite side of the gear 254. In Figure 5, I have illustrated a four-cylinder engine crankshaft wherein the crankshaft gear 254 is located at the center of the crankshaft with two crank portions 322 and 324 located on one side of the gear 254 and two other crank portions 326 and 328 located on the opposite side of the gear 254. The end bearing supports 330 and 332 are located at the crankshaft ends. Bearing support 334 is located between the crank portions 324 and 322 and bearing portion 336 is located between the crank portions 328 and 326. The central bearing portion 338 is located generally on the transverse plane of the pitch diameter for the gear 254. Three of the four cylinders are illustrated only schematically in Figure 5 and are designated by reference characters 340, 342 and 344. A piston 278 located in the other cylinder is illustrated in Figure 5 and carries the same reference character used in Figure 3. Similarly, the same reference characters for designating the piston rod and the crankshaft bearing cap are used in Figure 3 and 5. Figure 6 is a line diagram that shows the clutches and the planetary elements of the gearing in schematic form. This is done to facilitate an understanding of the mode of operation of the gearing. In the schematic representation of Figure 6, the ring gear for planetary gear unit 150 as designated by the symbol R1 and the sun gear is designated by the symbol S1. Similarly, the ring gear and the sun gear for gear unit 176 are designated by the reference symbols R2 and S2, respectively. Clutches 102, 130 and 136 are designated respectively by the symbols OWC3, OWC1 and OWC2. Two output shafts are shown in Figure 6 but only one would be needed in the present construction. A second output shaft would be required if the transmission were to be adapted for driving torque at both the front and rear wheels. During forward drive operation, the clutch shown apart at 82 is disengaged. The turbine torque then is delivered directly to the gear 60 through member 78. The torque is transferred then from gear 60 through the overrunning clutch OWC3 to the torque input shaft 100 for the planetary gearing. Clutch C1 is applied as it is during operation and each of the first four forward driving ratios. Overrunning brake OWC4 192, which also carries designation OWC4, and the brake 198 provide a torque reaction for the sun gear S2, which serves as a reaction element during low speed ratio operation. To effect a ratio change to the second ratio from the low speed ratio, clutch C2 is applied together with clutch C1. Although clutch C1 remains applied, torque is not transferred through it since overrunning clutch OWC1 freewheels as torque is transferred directly from the shaft 100 through the engaged clutch C2 to the carrier for the planetary gear unit 150. This drives ring gear R2 for the gear unit 176. With sun gear S2 acting as a reaction point, output shaft 208 is driven with an increased ratio. Third speed ratio operation is achieved by engaging simultaneously clutches C1, C2 and the clutch C5, shown in part at 82 in Figure 1D. This causes turbine torque to be delivered to gear 68 which drives input gear 98 which is connected directly to the input shaft 100. Under these conditions, overrunning clutch OWC3 free wheels. A fourth speed ratio is achieved by engaging simultaneously the clutches C1, C2 and C3. This effects a 1:1 ratio in the gear units 150 and 176 as the gear 98 continues to serve as a torque input gear. Clutch C1 is released and the overdrive brake 128, which is designated by the symbol B1 in Figure 6, is applied. This anchors sun gear S1 for the planetary gear unit 150. The ring gear R1 then is overdriven, which overdrives the output shaft 208. Reverse drive is obtained by engaging simultaneously brake C4 (also identified as brake 166 in Figure 1F), the latter anchoring the carrier as sun gear S1 serves as a torque input element. Ring gear R1 then is driven in a reverse direction which drives the output shaft 208 in a reverse direction. The clutch and brake engagement and release pattern is shown in Figure 7. Reference may be made to it as well as to the schematic line diagram of Figure 6 to understand the operation of the planetary elements in each of the ratios. Figures 8A through 8D show a transaxle arrangement using the T-drive configuration wherein the front drive axis is situated between the hydrokinetic torque converter and the engine (not shown). This version of the design differs from the design described with reference to Figures 1A through 1G because the axis of the hydrokinetic torque converter and the planetary gearing is situated on the vehicle axis and intersects the axis of the crankshaft substantially at the mid-point of the crankshaft. Torque is transferred from the torque output element of the planetary gearing through a transfer drive including a drive chain. The driven element of the transfer drive is connected through a spline to an output pinion shaft located in spaced, parallel disposition with respect to the planetary gearing axis where it drivably engages the ring gear of the front axle differential assembly. In Figures 8C and 8D, numeral 350 generally designates a hydrokinetic torque converter having a bladed impeller 352, a bladed turbine 354 and a bladed stator 356. The impeller forms a part of an impeller assembly comprising a housing 358 which is connected to torque input shaft 360 by drive plate 362. Torque input shaft 360 corresponds to the torque input shaft for the T-drive version described with reference to Figures 1A through 1G. It is adapted to be driven by a right angle bevel gear which meshes with a crankshaft gear connected drivably to the crankshaft in a manner described in US Application No. 471,759. The gearing arrangement shown in Figures 8C and 8D is described in particular detail in U.S. Patent No. 4,938,097. Reference may be made to that patent for the purpose of supplementing this disclosure in order to assist in an understanding of the torque flow paths that are established by the planetary gearing shown in Figures 8C and 8D. The gearing of Figures 8C and 8D includes two simple planetary gear units 364 and 366. Torque is transferred from turbine 354 to torque input sleeve shaft 368, which is connected to clutch sleeve member 370, the connection being established by spline 372. Forward clutch 374 is connected through an overrunning coupling 376 to sun gear 378 for the first planetary gear unit 360. Ring gear 380 of the gear unit 360 is splined at 382 to the carrier 384 of gear unit 364. Planet pinions 386 are journalled on carrier 388 for gear unit 360. Pinions 386 mesh with sun gear 378 and ring gear 380. Clutch 374 is applied by clutch servo piston 390, which cooperates with the clutch sleeve member 370 to define a pressure chamber. When the pressure chamber is pressurized, the force on the piston 390 is transferred to pressure plate 392 through an extension 394 on the piston 390. The extension 394 passing through the clutch discs of coast clutch 396. Coast clutch 396 forms a bypass torque distribution path around the overrunning coupling 376. When coast braking is desired, clutch 396 is applied by coast clutch servo piston 398 located in an annular cylinder formed in the front clutch piston 390. Brake band 400 surrounds brake drum 402, which is anchored by the band during second speed ratio operation and fourth speed ratio operation. Brake drum 402 is connected to torque transfer member 404, which is connected to sun gear 406 of planetary gear unit 364. Planetary pinions 408 of gear unit 364 mesh with ring gear 410 and with sun gear 406. These are carried by planet carrier 384. Carrier 388 of gear unit 360 is connected to ring gear 410 and serves as a torque output element. Drive sprocket 412 is carried by torque output member 414, which is a sleeve that is splined to the carrier 388. During operation in low and in reverse, carrier 384 is adapted to be braked by a friction brake 416, applied by a fluid pressure-operated servo piston 420. An overrunning brake, generally indicated at 418, serves as a reaction torque distribution path in parallel with the friction torque distribution path of the brake 416. It serves to anchor the carrier 384 during low speed ratio operation. A direct drive clutch 422 establishes a connection between clutch sleeve member 370 and torque transfer member 424, which extends to and is connected with the ring gear 380 of gear unit 360. A reverse clutch 426 is adapted to connect the clutch sleeve member 370 with torque transfer member 404, which extends to the sun gear 406 of gear unit 364. Shaft 36 is connected to pump drive shaft 428, which extends through the planetary gearing and which is connected drivably to the pump gears of a positive displacement gear pump 430 located at the right-hand end of the transmission housing 432. The first driving speed ratio is obtained by engaging the forward clutch 374. Clutch 374 is engaged during operation in each of the four forward driving ratios. Turbine torque delivered from the turbine shaft 368 is transferred through the engaged clutch 374 and through overrunning coupling 376 to the sun gear 378. The carrier 384 is anchored by the overrunning brake 418. Thus, the carrier acts as a reaction point as it anchors the ring gear 380 of the gear unit 360. With the sun gear 378 acting as a torque input member, the carrier 388 and the sprocket 412 are driven with a reduced ratio. A ratio change to the second ratio from the first ratio is obtained by engaging brake band 400 which anchors the sun gear 406. The overrunning brake 418 then begins to free wheel as torque is delivered to the sprocket with an increased ratio. To obtain third speed ratio, both the direct clutch and the forward clutch are applied so that all of the elements of the gear unit are locked together for rotation in unison. The fourth ratio is an overdrive ratio which is achieved by anchoring sun gear 406 as brake band 400 is applied. The forward clutch 422 then delivers torque directly to the carrier 384, thus overspeeding the ring gear 410. Reverse drive is achieved by applying the low-and-reverse brake band 400 by releasing the forward clutch 374 and the direct clutch 422 and applying the reverse clutch 426. Turbine torque then is delivered directly to the clutch element 370 and through the engaged reverse clutch 426 to the torque transfer member 404, thus driving the sun gear 406. Brake 416 is applied so the carrier 384 acts as a reaction point as the ring gear 410 is driven in a reverse direction. Drive sprocket 412 is generally in transverse alignment with driven sprocket 432. A drive chain connects drivably the sprockets 412 and 432 a portion of the drive chain being shown at 434. Sprocket 432 is splined at 436 to output shaft 438. It is journalled by bearing 440 in a bearing opening 442 in the transmission housing 432. Shaft 438 is supported by spaced, tapered roller bearings 444 and 446. The inboard end of the shaft 438 carries bevel drive pinion 448 which meshes with differential ring gear 450 of the front axle and differential assembly generally indicated at 452. Assembly 452 functions in a manner similar to that of the front axle and differential assembly of the embodiment described with reference to Figures 1A through 1G. Ring gear 450 is carried by differential housing 454 which is mounted for rotation about the axis of front wheel drive shaft 456 extending to one traction wheel. The common axis of axle shaft 458 extends to the other front drive wheel. Differential housing 454 carries pinions 460 and 462 which mesh with side gears 464 and 466, which in turn drive the axle shafts. The displacement of the axis of shaft 438 with respect to the axis of the planetary gearing is sufficient to avoid interference between the torque converter and the torque flow path extending to the differential assembly 452 from the driving sprocket 412. Figure 9 is a schematic representation of the planetary gearing of Figure 8C. In Figure 9, direct clutch 422 carries a designation DC, reverse clutch 426 carries a designation RC, coast clutch 396 carries a designation CC, forward clutch carries a designation FC, the low-and-reverse brake 416 carries a designation LR, the overrunning clutch 418 carries a designation OWC1 and the overrunning clutch 376 carries a designation OWC2. Figure 8 shows a clutch and brake engagement and release pattern for the gear system schematically illustrated in Figure 10. It shows the brakes and the clutches that are applied or released to establish each of the forward driving ratios as well as a reverse ratio. The symbols used in Figure 8 correspond to the designations for the friction elements and the overriding clutches used in the schematic arrangement of Figure 10 so that the mode of operation indicated in Figure 8 can be correlated with the schematic diagram of Figure 10.

Claims: A T-drive vehicle powertrain assembly comprising an internal combustion engine (12) with a crankshaft, a drive gear (254) connected to and rotatable in unison with said crankshaft about the axis of the crankshaft at a location intermediate the crankshaft ends; a power transmission mechanism comprising a hydrokinetic torque converter (10) with an impeller (16) and a turbine (28) in fluid flow relationship and a gear system (14) having a power input shaft (100) mounted on an axis substantially perpendicular to said crankshaft axis, a power output shaft (208) of said transmission being adapted to be connected to vehicle traction wheels; a driving gear (252) connected drivably to said impeller (16) and meshing with said drive gear (254) whereby said engine and said transmission mechanism form a compact, unitary drive assembly, said power input shaft (100) being connected drivably to said turbine (28); said transmission having a power output shaft (208) extending in a substantially perpendicular direction with respect to the axis of said crankshaft; and a final transaxle gear drive (224) having differential gearing with axle drive gears (228,230), a geared connection between said power output shaft (208) and separate ones of said axle drive gears, each axle drive gear being adapted to be connected drivably to an axle shaft; the drive gear (254) being located at the centre of the crankshaft, and the final transaxle gear drive (224) being located between said engine (12) and said torque converter (10). A powertrain assembly as claimed in claim 1, wherein said driving gear (252) is mounted on an axis in spaced, parallel disposition with respect to the axis of said gear system (14); and further including a torque transfer drive means for establishing a torque flow path between said axes. A powertrain assembly as claimed in claim 2, wherein said torque transfer drive means comprises gears defining plural ratios and a controllable clutch means for selectively establishing each of said ratios. A powertrain assembly as claimed in claim 1, 2 or 3 wherein, said crankshaft has multiple crankshaft bearings including a bearing at each crankshaft end (292,294) and a bearing at a mid position (298) on said crankshaft whereby a mid-position nodal point is established thus increasing crankshaft torsional stiffness.

Inventors: CARRIERE DONALD L, CARRIERE, DONALD L.
Applicants: FORD FRANCE, FORD MOTOR CO, FORD WERKE AG, FORD FRANCE S. A., FORD MOTOR COMPANY, FORD MOTOR COMPANY LIMITED, FORD-WERKE AKTIENGESELLSCHAFT
Application year: 1992
Main classification: B60K17
IPC codes: F02B75, B60K17