Abstract:
A multiple speed power transmission includes an input, first and second input shafts, first and second clutches for releasably coupling the input to the input shafts, and a set of selectable power paths connecting the first and second input shafts to the output, the power paths having at least one power path that includes both the first and second input shafts.

Description:
BACKGROUND OF THE INVENTION 
     This invention relates to automatic transmissions having a layshaft kinematic arrangement, particularly to automatic transmissions having dual input clutches, but no torque converter. 
     Automatic transmissions for transmitting power between an input and an output, either over a continuously variable range of speed ratios or in discrete step changes among speed ratios, have associated with them several sources of parasitic losses, which adversely affect fuel economy. These losses are associated with a torque converter, open hydraulic friction clutches and brakes, hydraulic pump, and gear meshes. 
     To improve fuel economy in a motor vehicle having an automatic transmission, an automated shift manual (ASM) transmission can be used to eliminate or substantially reduce all of these parasitic losses except gear mesh losses. An ASM transmission generally performs gear ratio changes by first interrupting torque transmitted from the engine to the transmission input, preparing the transmission components associated with the next speed ratio, and then restoring torque at the input. A primary functional feature of ASM transmissions is the need to interrupt power transmitted from the engine to the transmission input shaft before or during each gear ratio change. 
     Dual clutch layshaft transmissions are essentially two ASM transmissions, one providing odd numbered gears and one providing even numbered gears. Shifts between odd numbered gears and even numbered gears can be accomplished without interrupting power flow. While operating in an odd numbered gear, couplers can be actuated to configure the transmission for the next even numbered gear. Dual clutch transmissions have parasitic losses only slightly higher than ASM transmissions. 
     When a motor vehicle is accelerated from rest, the mechanical power generated by the engine exceeds the power utilized by the vehicle. The transmission must dissipate the difference, generally as heat. Open torque converters are very efficient at converting the excess mechanical power into heat in the working fluid. Friction clutches, as used in ASM and dual clutch transmissions, are limited in the rate at which they can dissipate the excess power. The amount of energy that must be dissipated is determined by the torque level, the speed difference across the clutch, and the duration of the event. 
     The most effective way to limit the power that must be dissipated by the clutch is to provide additional torque multiplication in the gearbox. This has two benefits. First, it reduces the torque which the clutch must transmit. Second, it reduces the duration of the event because the gearbox input will become equal to the engine speed at a lower vehicle speed. The need for similar top gear ratios, which is dictated by cruising fuel economy, is unchanged, so the resulting gearbox must have substantially more total span. The difference between adjacent gear ratios is limited by the ability to make comfortable shifts. As a result, it is also necessary to increase the number of discrete gear ratios. 
     One reverse ratio has been considered sufficient, since speed is relatively low and fuel efficiency in reverse is not a significant concern. However, if the gear multiplication is high enough to satisfy clutch thermal considerations, it may be excessive for normal reverse driving, even at those relatively low speeds. Therefore, it is beneficial to provide a reverse ratio similar to the traditional reverse ratio in addition to one that has much more multiplication. 
     A way to increase the gear multiplication is to increase to ratio of the tooth counts for individual gear pairs. This would require increasing the distance between shafts due to limitations on how small the gears can be relative to the shaft diameter. Adding an additional forward and reverse ratio would ordinarily require at least four additional gears and an additional synchronizer sleeve. The resulting transmission would be much larger and likely would not fit into the package space available. 
     For example, using a dual clutch transmission in a truck requires a very high torque ratio to launch the truck from rest due to its heavy load especially on a grade. But it also requires low torque ratios for efficient highway cruising when the vehicle is lightly loaded and on level ground. Typically, layshaft transmissions accommodate these requirements by increasing the distance between the shafts, resulting in a transmission that is large and heavy. 
     SUMMARY OF THE INVENTION 
     A multiple speed power transmission that overcomes these difficulties includes an input, an output, first and second input shafts, and first and second clutches releasably coupling the input to the first and second input shafts, respectively. A first pinion is secured to the first input shaft, and a first gear secured to a layshaft is in continuous meshing engagement with the first pinion. A second gear is journalled on the layshaft, and a second pinion secured to the second input shaft is in continuous meshing engagement with the second pinion. A first coupler secured to the layshaft releasably couples the second gear to the layshaft. A first set of selectable power paths connects the first input shaft to the output, and a second set of selectable power paths connects the second input shaft to the output. 
     An advantage of this transmission is achieving the span with a smaller center distance compared to a conventional output reduction dual clutch arrangement. Also, a second reverse ratio and top gear ratio is available without requiring additional hardware. This arrangement allows the third forward gear and the second reverse gear to be activated concurrently, which is advantageous for rock cycling maneuvers. 
     The scope of applicability of the preferred embodiment will become apparent from the following detailed description, claims and drawings. It should be understood, that the description and specific examples, although indicating preferred embodiments of the invention, are given by way of illustration only. Various changes and modifications to the described embodiments and examples will become apparent to those skilled in the art. 
    
    
     
       DESCRIPTION OF THE DRAWINGS 
       These and other advantages will become readily apparent to those skilled in the art from the following detailed description of a preferred embodiment when considered in the light of the accompanying drawings in which: 
         FIG. 1  is a schematic diagram of an eight forward gear, two reverse gear transmission embodiment; 
         FIG. 2  is a chart containing a preferred number of teeth for each of the gears and pinions of the transmission of  FIG. 1 ; and 
         FIG. 3  is a chart containing the speed ratios between the input and output, and steps between the speed ratios for each of the forward and reverse gears of the transmission of  FIG. 1 . 
     
    
    
     DESCRIPTION OF THE PREFERRED EMBODIMENT 
     Referring to  FIG. 1 , a first embodiment of a transmission includes an input  10  driveably connected to a power source, such as an internal combustion engine or electric motor, and an output  11  for driving a load, such as the driven wheels of a motor vehicle, through a powertrain that may include a drive shaft, differential mechanism, and axle shafts. A first friction clutch  20 , which includes a clutch housing and clutch disc  22 , alternately driveably connects and disconnects a first input shaft  14  as clutch  20  is engaged and disengaged, respectively. A second friction clutch  16 , which includes a clutch housing and a clutch disc  18 , driveably connects and disconnects a second input shaft  12  as clutch  16  is engaged and disengaged, respectively. 
     Couplers  60 ,  62 ,  64 ,  66  are preferably synchronizers of the type used in automotive manual transmissions to connect a gear or pinion to a shaft, after synchronizing the speed of the shaft and that of the pinion or gear, and to disconnect the shaft and the associated pinion or gear. Such synchronizers include dog teeth on the gear or pinion, which engage teeth on a selector sleeve, which moves leftward and rightward from a neutral position to produce engagement. Alternatively, each coupler may be a dog clutch having teeth that are engaged with dog teeth on a gear or pinion without synchronizing the speed of the shaft with that of the pinion or gear. This invention may use couplers in any combination of synchronizers and dog clutches. Each coupler includes a hub secured to the shaft and the selector sleeve, which is supported on the hub such that it slides leftward and rightward into engagement with dog teeth on the adjacent gear or pinion. 
     In the case where a coupler is a synchronizer, it is provided with a conical surface, which engages mutually with a corresponding conical surface located on the gear or pinion. When the synchronizer is engaging with either of its adjacent gears, these conical surfaces are forced together into frictional contact, which synchronizes the speed of the gear to that of the shaft before the dog teeth engage. 
     Coupler  60  driveably connects input shaft  14  to pinions  30 ,  32 , and it disconnects those pinions from shaft  14 . Coupler  62  driveably connects layshaft  44  to pinion  42  and layshaft  36 , and it disconnects pinion  42  and layshaft  36  from layshaft  44 . Coupler  64  driveably connects layshaft  36  to gears  48 ,  52 , and it disconnects those gears from layshaft  36 . Coupler  66  driveably connects input shaft  12  to pinion  56 , gear  59 , and it disconnects those pinion  56  and gear  59  from input shaft  12 . A disc  70 , secured to layshaft  36 , carries a conical synchronizing surface and dog teeth, which complete a drive connection between layshafts  44 ,  36  through coupler  62 . 
     Input shaft  14  supports a pinion  24 , which is secured to the shaft and in continuous meshing engagement with a gear  26 , which is secured to an auxiliary layshaft  44 . Input shaft  14  also supports two pinions  30 ,  32 , which are journalled on shaft  14 . Pinion  30  is in meshing engagement with gear  34 , which is secured to layshaft  36 . Pinion  32  is in meshing engagement with gear  38 , which is secured to layshaft  36 . 
     Input shaft  12  supports a gear  40 , which is secured to shaft  12  and is in continuous meshing engagement with pinion  42 , which is journalled on layshaft  44 . 
     A pinion  46  is secured to input shaft  12  and is in continuous meshing engagement with a gear  48 , journalled on layshaft  36 . Similarly, pinion  50  is secured to input shaft  12  and is in continuous meshing engagement with an idler gear (not shown). The idler gear is in continuous meshing engagement with a gear  52 , journalled on layshaft  36 . 
     Gear  54  is secured to layshaft  36  and is in continuous meshing engagement with a pinion  56 , journalled on input shaft  12 . Similarly, output pinion  58  is secured to layshaft  36  and is in continuous meshing engagement with output gear  59 , which is secured to output  11 . 
     In first gear and low reverse (R 1 ), coupler  62  driveably connects pinion  42  and layshaft  44  through a power path, which driveably connects the first and second input shafts  14 ,  12 . The power path includes input shaft  14 , pinion  24 , gear  26 , auxiliary layshaft  44 , coupler  62 , pinion  42 , gear  40 , and input shaft  12 . 
     To accelerate the vehicle using the first forward gear, coupler  62  engages pinion  44 , coupler  64  engages gear  48 , and couplers  60 ,  66  are disengaged. Clutch  20  is engaged and clutch  16  is disengaged. The power path for the first speed ratio includes input  10 , clutch  20 , input shaft  14 , pinion  24 , gear  26 , auxiliary layshaft  44 , coupler  62 , pinion  42 , gear  40 , input shaft  12 , pinion  46 , gear  48 , coupler  64 , layshaft  36 , output pinion  58 , output gear  59 , and output  11 . Input  10  is driveably connected to input shaft  14  through clutch  20 . Layshaft  44  is driveably connected to pinion  42  through coupler  62 . Gear  48  is driveably connected to layshaft  36  through coupler  64 . The speed ratio for first gear is 5.248 when the gears and pinions have the number of teeth shown in  FIG. 2 . 
     To shift from first gear to second gear, coupler  64  remains engaged with pinion  48 , clutch  16  is progressively engaged, and clutch  20  is progressively disengaged. The power path for second gear includes input  10 , clutch  16 , input shaft  12 , pinion  46 , gear  48 , coupler  64 , layshaft  36 , output pinion  58 , output gear  59 , and output  11 . Input  10  is driveably connected to input shaft  12  and pinion  46  through clutch  16 . Pinion  46  drives gear  48 , which is driveably connected to layshaft  36  through coupler  64 . Pinion  58 , secured to layshaft  36 , drives output gear  59  and output  11 . The speed ratio for second gear is 4.234, when the gears and pinions have the number of teeth shown in  FIG. 2 . Coupler  62  may be disengaged after torque is transferred to the second gear power path. 
     To shift from the second gear to third gear, coupler  62  engages disc  70 , clutch  16  is progressively disengaged, and clutch  20  is progressively engaged. Following the 2-3 shift, coupler  64  is moved to the neutral position. The power path for third gear includes input  10 , clutch  20 , input shaft  14 , pinion  24 , gear  26 , layshaft  44 , coupler  62 , disc  70 , layshaft  36 , output pinion  58 , output gear  59 , and output  11 . Input  10  is driveably connected to input shaft  14  and pinion  24  through clutch  20 . Pinion  24  drives gear  26 , which is driveably connected to layshaft  36  through coupler  62 . Pinion  58 , secured to layshaft  36 , drives output gear  59  and output  11 . The speed ratio for third gear is 2.718, when the gears and pinions have the number of teeth shown in  FIG. 2 . 
     To shift from third gear to fourth gear, coupler  66  engages gear  59 , clutch  16  is progressively engaged, and clutch  20  is progressively disengaged. Following the 3-4 shift, coupler  66  is moved to the neutral position. The power path for fourth gear includes input  10 , clutch  16 , input shaft  12 , coupler  66 , pinion  56 , gear  54 , layshaft  36 , output pinion  58 , output gear  59 , and output  11 . Input  10  is driveably connected to input shaft  12  through clutch  16 . Coupler  66  driveably connects input shaft  12  to pinion  56 , which drives gear  54 , layshaft  36  and output pinion  58 . Output gear  59  and output  11  are driven by output pinion  58 . The speed ratio for fourth gear is 1.875, when the gears and pinions have the number of teeth shown in  FIG. 2 . 
     To shift from fourth gear to fifth gear, coupler  60  engages pinion  32 , clutch  20  is progressively engaged, and clutch  16  is progressively disengaged. Following the 4-3 shift, coupler  66  is moved to the neutral position. The power path for fifth gear includes input  10 , clutch  20 , input shaft  14 , coupler  60 , pinion  32 , gear  38 , layshaft  36 , output pinion  58 , output gear  59 , and output  11 . Input  10  is driveably connected to input shaft  14  through clutch  20 . Coupler  60  driveably connects pinion  32  and input shaft  14 . Pinion  32  drives gear  38 , layshaft  36  and output pinion  58 , which drives output gear  59  and output  11 . The speed ratio for fifth gear is 1.340, when the gears and pinions have the number of teeth shown in  FIG. 2 . 
     To shift from fifth gear to sixth gear, coupler  66  engages output gear  59 , clutch  16  is progressively engaged, and clutch  20  is progressively disengaged. Following the 5-6 shift, coupler  60  is moved to the neutral position. The power path for sixth gear includes input  10 , clutch  16 , input shaft  12 , coupler  66 , output gear  59 , and output  11 . Input  10  is driveably connected to input shaft  12  through clutch  16 . Coupler  66  driveably connects input shaft  12  to output gear  59  and output  11 . Sixth gear is a direct drive gear, whose speed ratio is 1.000. 
     To shift from sixth gear to seventh gear, coupler  60  engages pinion  30 , clutch  20  is progressively engaged, and clutch  16  is progressively disengaged. Following the 6-7 shift, coupler  66  is moved to the neutral position. The power path for seventh gear includes input  10 , clutch  20 , input shaft  14 , coupler  60 , pinion  30 , gear  34 , layshaft  36 , output pinion  58 , output gear  59 , and output  11 . Input  10  is driveably connected to input shaft  14  through clutch  20 . Coupler  60  driveably connects pinion  30  and input shaft  14 . Pinion  30  drives gear  34 , layshaft  36  and output pinion  58 , which drives output gear  59  and output  11 . The speed ratio for seventh gear is 0.770, when the gears and pinions have the number of teeth shown in  FIG. 2 . 
     To upshift from seventh gear to eighth gear, coupler  62  engages pinion  42 , clutch  16  is progressively engaged, and clutch  20  is progressively disengaged. The power path for eighth gear includes input  10 , clutch  16 , input shaft  12 , gear  40 , pinion  42 , coupler  62 , layshaft  44 , gear  26 , pinion  24 , input shaft  14 , coupler  60 , pinion  30 , gear  34 , layshaft  36 , output pinion  58 , output gear  59 , and output  11 . Input  10  is driveably connected to input shaft  12  and gear  40  through clutch  16 . Gear  40  drives pinion  42 , which is driveably connected to layshaft  44  through coupler  62 . Gear  26 , secured to layshaft  44 , drives pinion  24  and input shaft  14 . Coupler  60  driveably connects input shaft  14  and pinion  30 , which drives gear  34 , layshaft  36 , output pinion  58 , output gear  59 , and output  11 . The speed ratio for eight gear is 0.617, when the gears and pinions have the number of teeth shown in  FIG. 2 . 
     Each downshift is accomplished by reversing the steps of the corresponding upshift. 
     To accelerate the vehicle in a low reverse gear (R 1 ), coupler  62  engages pinion  42 , coupler  64  engages gear  52 , couplers  60 ,  66  are disengaged, and clutch  20  is engaged. The power path for low reverse gear includes input  10 , clutch  20 , input shaft  14 , pinion  24 , gear  26 , layshaft  44 , coupler  62 , pinion  42 , gear  40 , input shaft  12 , pinion  50 , an idler gear (not shown), gear  52 , coupler  64 , layshaft  36 , output pinion  58 , output gear  59 , and output  11 . Input  10  is driveably connected to shaft  14  through clutch  20 . Shaft  14  is drives pinion  24 , which drives gear  26 . Coupler  62  driveably connects pinion  42  and layshaft  44 . Pinion  42  drives gear  40 , input shaft  12  and pinion  50 , which drives a reverse idler (not shown) and gear  52 . Coupler  64  driveably connects gear  52  to layshaft  36 , which drives output pinion  58  output gear  59  and output  11 . The speed ratio for the first reverse gear is −4.897, when the gears and pinions have the number of teeth shown in  FIG. 2 . 
     To shift from the first reverse gear to a second reverse gear, coupler  64  remains engaged to gear  52 , clutch  16  is progressively engaged, and clutch  20  is progressively disengaged. The power path for the second reverse gear includes input  10 , clutch  16 , input shaft  12 , pinion  50 , gear  52 , coupler  64 , layshaft  36 , output pinion  58 , output gear  59 , and output  11 . Input  10  is driveably connected to shaft  12  through clutch  16 . Reverse pinion  50  drives its reverse idler (not shown) and reverse gear  52 . Coupler  64  driveably connects gear  52 , layshaft  36  and output pinion  58 , which drives output gear  59  and output  11 . The speed ratio for the second reverse gear is −3.924, when the gears and pinions have the number of teeth shown in  FIG. 2 . 
     In accordance with the provisions of the patent statutes, the preferred embodiment has been described. However, it should be noted that the alternate embodiments can be practiced otherwise than as specifically illustrated and described.