Patent Document

RELATED APPLICATIONS 
     This application claims priority to U.S. Provisional Patent Application Ser. No. 61/603,990, filed Feb. 28, 2012, which is hereby incorporated by reference in its entirety. 
    
    
     FIELD OF THE DISCLOSURE 
     The present disclosure relates to a gearing arrangement for a transmission, and in particular to a planetary gearing arrangement for an automotive transmission having a Polak gearing arrangement. 
     BACKGROUND 
     A conventional transmission for a powered vehicle can include a gearbox, i.e., gears, synchronizers, dog clutches, clutch plates and reaction plates, a number of automatically selectable gears, planetary gear sets, hubs, pistons, shafts, and other housings. The clutches can be stationary brakes/clutches or rotating clutches. The transmission can have an internal shaft which rotates one or more clutches or shafts in the transmission. 
     The number of clutches and gear schemes can be used to achieve different gear ratios. Each gear ratio can define a range in which the transmission operates. A higher numerical gear ratio, for example, can be achieved at a lower transmission output speed. This can be important if a vehicle is heavily loaded or requires higher torque to ascend an elevation, for instance. Alternatively, a lower numerical gear ratio can be achieved at a higher transmission output speed, particularly when a vehicle is travelling at higher speeds on a highway. The lower numerical gear ratio can help increase fuel efficiency at these speeds. 
     To achieve different gear ratios, gear schemes are configured in which brakes, clutches, shafts, and gearsets are packaged in a transmission. The gear scheme can include one or more planetary gearsets. A planetary gearset can include a sun gear, a ring gear, and a carrier. One or more of the sun gear, ring gear, and carrier can be input or output of the planetary gearset. The manner in which torque is transmitted through the clutches and gearsets determines the different gear ratios for a given transmission. 
     In some arrangements, however, the reverse gear ratio may be numerically too large to obtain a required reverse speed. In addition, the gear step between the first and second forward gears may be too large to permit the 1-2 and 2-1 shifts to be made with the torque converter clutch applied. Making these shifts with the torque converter clutch applied can improve transmission efficiency, fuel economy, and reduce transmission heat generation. Gear step is defined in the present disclosure as the ratio of two gear ratios. 
     There is a need for an automatic transmission having a gearing arrangement for automatically selecting gears and which the gearbox is configured for an available fast reverse and a close gear step 1-2 shift. 
     SUMMARY 
     In an exemplary embodiment of the present disclosure, an automatic transmission is provided having an input adapted to couple to a torque-generating mechanism and an output coupled to the input. The transmission also includes a first rotating torque-transferring mechanism disposed along a first torque path and coupled to the input. A second rotating torque-transferring mechanism is disposed along a second torque path and is coupled to the input independent of the first torque-transferring mechanism. The transmission includes a plurality of stationary torque-transferring mechanisms, each of which is disposed between the input and output. The transmission includes a first planetary gearset, a second planetary gearset, a third planetary gearset, and a fourth planetary gearset, where each gearset includes a sun gear, a ring gear, and a carrier assembly. Moreover, the ring gear of the third planetary gearset is coupled to the carrier assembly of the second planetary gearset and the carrier assembly of the fourth planetary gearset. 
     In one aspect of this embodiment, the sun gear of the first planetary gearset is coupled to the input along a third torque path and the ring gear of the first planetary gearset is coupled to the first stationary torque-transferring mechanism. In another aspect, the ring gear of the fourth planetary gearset is coupled to the fourth stationary torque-transferring mechanism and the carrier assembly of the fourth planetary gearset. In a different aspect, the first rotating torque-transferring mechanism, second rotating torque-transferring mechanism, first stationary torque-transferring mechanism, second stationary torque-transferring mechanism, third stationary torque-transferring mechanism, and fourth stationary torque-transferring mechanism are engaged in combinations of at least two to establish at least seven forward speed ratios and two reverse speed ratios between the input and the output. Here, the gear step between the first forward speed ratio and second forward speed ratio can be about 1.5 or less. 
     In a related aspect, the first torque path is at least partially defined by the input, the first rotating torque-transferring mechanism, the sun gear of the second planetary gearset, and the sun gear of the third planetary gearset. The second torque path is at least partially defined by the input, the second rotating torque-transferring mechanism, and the carrier assembly of the second planetary gearset. In a further aspect, the carrier assembly of the first planetary gearset is coupled to the ring gear of the second planetary gearset; the carrier assembly of the second planetary gearset is coupled to the ring gear of the third planetary gearset; and the carrier assembly of the third planetary gearset is coupled to the sun gear of the fourth planetary gearset and the output. In yet a further aspect, the ring gear of the third planetary gearset is coupled to at least three carrier assemblies and one stationary torque-transferring mechanism. Moreover, a third torque path is at least partially defined by the input, the sun gear of the first planetary gearset, and the carrier assembly of the first planetary gearset. 
     In another embodiment, a gear scheme of an automatic transmission is provided. The gear scheme includes a transmission input and a transmission output, the input adapted to be coupled to a torque-generating mechanism. The scheme also includes a first rotating torque-transferring mechanism and a second rotating torque-transferring mechanism, each being independently coupled to the transmission input. A first planetary gearset is coupled to the first rotating torque-transferring mechanism, where the first planetary gearset is coupled to a first stationary torque-transferring mechanism. A second planetary gearset is coupled to the first planetary gearset and the first rotating torque-transferring mechanism, where the second planetary gearset is coupled to a second stationary torque-transferring mechanism. A third planetary gearset is coupled to the second rotating torque-transferring mechanism and the transmission output, where the third planetary gearset is coupled to a third stationary torque-transferring mechanism. In addition, a fourth planetary gearset is coupled to the third planetary gearset and a fourth stationary torque-transferring mechanism. The fourth planetary gearset includes a sun gear, a ring gear, and a carrier assembly, where the ring gear is coupled to the fourth stationary torque-transferring mechanism. 
     In one aspect of this embodiment, the third planetary gearset comprises a sun gear, a ring gear, and a carrier assembly, the ring gear of the third planetary gearset being coupled to the carrier assembly of the fourth planetary gearset. In a second aspect, the first planetary gearset comprises a sun gear, a ring gear, and a carrier assembly, the sun gear of the first planetary gearset being coupled to the transmission input and the ring gear of the first planetary gearset is coupled to the first stationary torque-transferring mechanism. Here, a torque path is defined between the transmission input and transmission output, where the sun gear and carrier assembly of the first planetary gearset are disposed along the torque path. 
     In another aspect, a first torque path is defined between the input and the output of the transmission, where the first rotating torque-transferring mechanism, an input of the second planetary gearset, an input of the third planetary gearset, and an output of the third planetary gearset are disposed along the first torque path. In a different aspect, a second torque path is defined between the input and the output of the transmission, where the second rotating torque-transferring mechanism, a ring gear of one of the planetary gearsets, and at least two carrier assemblies of the planetary gearsets are disposed along the second torque path. 
     In a related embodiment, the first rotating torque-transferring mechanism, second rotating torque-transferring mechanism, first stationary torque-transferring mechanism, second stationary torque-transferring mechanism, third stationary torque-transferring mechanism, and fourth stationary torque-transferring mechanism are engaged in combinations of at least two to establish at least seven forward speed ratios and two reverse speed ratios between the transmission input and the transmission output. Here, the gear step between the first forward speed ratio and second forward speed ratio can be about 1.5 or less. Moreover, in another related aspect, the third planetary gearset includes a ring gear being coupled to at least three carrier assemblies and the third stationary torque-transferring mechanism. In a similar aspect, the first planetary gearset includes a carrier assembly being coupled to the second planetary gearset, the second planetary gearset includes a carrier assembly being coupled to the third planetary gearset, and the third planetary gearset includes a carrier assembly being coupled to the sun gear of the fourth planetary gearset and the transmission output. 
    
    
     
       BRIEF DESCRIPTION OF THE DRAWINGS 
       The above-mentioned aspects of the present disclosure and the manner of obtaining them will become more apparent and the disclosure itself will be better understood by reference to the following description of the embodiments of the disclosure, taken in conjunction with the accompanying drawings, wherein: 
         FIG. 1  is an exemplary block diagram and schematic view of one illustrative embodiment of a powered vehicular system; 
         FIG. 2  is a first gearing scheme for a multi-speed automatic transmission; and 
         FIG. 3  is a second gearing scheme for an improved multi-speed automatic transmission with a fast reverse speed and close gear step 1-2 shift capability. 
     
    
    
     Corresponding reference numerals are used to indicate corresponding parts throughout the several views. 
     DETAILED DESCRIPTION 
     The embodiments of the present disclosure described below are not intended to be exhaustive or to limit the disclosure to the precise forms disclosed in the following detailed description. Rather, the embodiments are chosen and described so that others skilled in the art may appreciate and understand the principles and practices of the present disclosure. 
     Referring now to  FIG. 1 , a block diagram and schematic view of one illustrative embodiment of a vehicular system  100  having a drive unit  102  and transmission  118  is shown. In the illustrated embodiment, the drive unit  102  may include an internal combustion engine, diesel engine, electric motor, or other power-generating device. The drive unit  102  is configured to rotatably drive an output shaft  104  that is coupled to an input or pump shaft  106  of a conventional torque converter  108 . The input or pump shaft  106  is coupled to an impeller or pump  110  that is rotatably driven by the output shaft  104  of the drive unit  102 . The torque converter  108  further includes a turbine  112  that is coupled to a turbine shaft  114 , and the turbine shaft  114  is coupled to, or integral with, a rotatable input shaft  124  of the transmission  118 . The transmission  118  can also include an internal pump  120  for building pressure within different flow circuits (e.g., main circuit, lube circuit, etc.) of the transmission  118 . The pump  120  can be driven by a shaft  116  that is coupled to the output shaft  104  of the drive unit  102 . In this arrangement, the drive unit  102  can deliver torque to the shaft  116  for driving the pump  120  and building pressure within the different circuits of the transmission  118 . 
     The transmission  118  can include a planetary gear system  122  having a number of automatically selected gears. An output shaft  126  of the transmission  118  is coupled to or integral with, and rotatably drives, a propeller shaft  128  that is coupled to a conventional universal joint  130 . The universal joint  130  is coupled to, and rotatably drives, an axle  132  having wheels  134 A and  134 B mounted thereto at each end. The output shaft  126  of the transmission  118  drives the wheels  134 A and  134 B in a conventional manner via the propeller shaft  128 , universal joint  130  and axle  132 . 
     A conventional lockup clutch  136  is connected between the pump  110  and the turbine  112  of the torque converter  108 . The operation of the torque converter  108  is conventional in that the torque converter  108  is operable in a so-called “torque converter” mode during certain operating conditions such as vehicle launch, low speed and certain gear shifting conditions. In the torque converter mode, the lockup clutch  136  is disengaged and the pump  110  rotates at the rotational speed of the drive unit output shaft  104  while the turbine  112  is rotatably actuated by the pump  110  through a fluid (not shown) interposed between the pump  110  and the turbine  112 . In this operational mode, torque multiplication occurs through the fluid coupling such that the turbine shaft  114  is exposed to drive more torque than is being supplied by the drive unit  102 , as is known in the art. The torque converter  108  is alternatively operable in a so-called “lockup” mode during other operating conditions, such as when certain gears of the planetary gear system  122  of the transmission  118  are engaged. In the lockup mode, the lockup clutch  136  is engaged and the pump  110  is thereby secured directly to the turbine  112  so that the drive unit output shaft  104  is directly coupled to the input shaft  124  of the transmission  118 , as is also known in the art. 
     The transmission  118  further includes an electro-hydraulic system  138  that is fluidly coupled to the planetary gear system  122  via a number, J, of fluid paths,  140   1 - 140   J , where J may be any positive integer. The electro-hydraulic system  138  is responsive to control signals to selectively cause fluid to flow through one or more of the fluid paths,  140   1 - 140   J , to thereby control operation, i.e., engagement and disengagement, of a plurality of corresponding friction devices in the planetary gear system  122 . The plurality of friction devices may include, but are not limited to, one or more conventional brake devices, one or more torque transmitting devices, and the like. Generally, the operation, i.e., engagement and disengagement, of the plurality of friction devices is controlled by selectively controlling the friction applied by each of the plurality of friction devices, such as by controlling fluid pressure to each of the friction devices. In one example embodiment, which is not intended to be limiting in any way, the plurality of friction devices include a plurality of brake and torque transmitting devices in the form of conventional clutches that may each be controllably engaged and disengaged via fluid pressure supplied by the electro-hydraulic system  138 . In any case, changing or shifting between the various gears of the transmission  118  is accomplished in a conventional manner by selectively controlling the plurality of friction devices via control of fluid pressure within the number of fluid paths  140   1 - 140   J . 
     The system  100  further includes a transmission control circuit  142  that can include a memory unit  144 . The transmission control circuit  142  is illustratively microprocessor-based, and the memory unit  144  generally includes instructions stored therein that are executable by the transmission control circuit  142  to control operation of the torque converter  108  and operation of the transmission  118 , i.e., shifting between the various gears of the planetary gear system  122 . It will be understood, however, that this disclosure contemplates other embodiments in which the transmission control circuit  142  is not microprocessor-based, but is configured to control operation of the torque converter  108  and/or transmission  118  based on one or more sets of hardwired instructions and/or software instructions stored in the memory unit  144 . 
     In the system  100  illustrated in  FIG. 1 , the torque converter  108  and the transmission  118  include a number of sensors configured to produce sensor signals that are indicative of one or more operating states of the torque converter  108  and transmission  118 , respectively. For example, the torque converter  108  illustratively includes a conventional speed sensor  146  that is positioned and configured to produce a speed signal corresponding to the rotational speed of the pump shaft  106 , which is the same rotational speed of the output shaft  104  of the drive unit  102 . The speed sensor  146  is electrically connected to a pump speed input, PS, of the transmission control circuit  142  via a signal path  152 , and the transmission control circuit  142  is operable to process the speed signal produced by the speed sensor  146  in a conventional manner to determine the rotational speed of the turbine shaft  106 /drive unit output shaft  104 . 
     The transmission  118  illustratively includes another conventional speed sensor  148  that is positioned and configured to produce a speed signal corresponding to the rotational speed of the transmission input shaft  124 , which is the same rotational speed as the turbine shaft  114 . The input shaft  124  of the transmission  118  is directly coupled to, or integral with, the turbine shaft  114 , and the speed sensor  148  may alternatively be positioned and configured to produce a speed signal corresponding to the rotational speed of the turbine shaft  114 . In any case, the speed sensor  148  is electrically connected to a transmission input shaft speed input, TIS, of the transmission control circuit  142  via a signal path  154 , and the transmission control circuit  142  is operable to process the speed signal produced by the speed sensor  148  in a conventional manner to determine the rotational speed of the turbine shaft  114 /transmission input shaft  124 . 
     The transmission  118  further includes yet another speed sensor  150  that is positioned and configured to produce a speed signal corresponding to the rotational speed of the output shaft  126  of the transmission  118 . The speed sensor  150  may be conventional, and is electrically connected to a transmission output shaft speed input, TOS, of the transmission control circuit  142  via a signal path  156 . The transmission control circuit  142  is configured to process the speed signal produced by the speed sensor  150  in a conventional manner to determine the rotational speed of the transmission output shaft  126 . 
     In the illustrated embodiment, the transmission  118  further includes one or more actuators configured to control various operations within the transmission  118 . For example, the electro-hydraulic system  138  described herein illustratively includes a number of actuators, e.g., conventional solenoids or other conventional actuators, that are electrically connected to a number, J, of control outputs, CP 1 -CP J , of the transmission control circuit  142  via a corresponding number of signal paths  72   1 - 72   J , where J may be any positive integer as described above. The actuators within the electro-hydraulic system  138  are each responsive to a corresponding one of the control signals, CP 1 -CP J , produced by the transmission control circuit  142  on one of the corresponding signal paths  72   1 - 72   J  to control the friction applied by each of the plurality of friction devices by controlling the pressure of fluid within one or more corresponding fluid passageway  140   1 - 140   J , and thus control the operation, i.e., engaging and disengaging, of one or more corresponding friction devices, based on information provided by the various speed sensors  146 ,  148 , and/or  150 . The friction devices of the planetary gear system  122  are illustratively controlled by hydraulic fluid which is distributed by the electro-hydraulic system in a conventional manner. For example, the electro-hydraulic system  138  illustratively includes a conventional hydraulic positive displacement pump (not shown) which distributes fluid to the one or more friction devices via control of the one or more actuators within the electro-hydraulic system  138 . In this embodiment, the control signals, CP 1 -CP J , are illustratively analog friction device pressure commands to which the one or more actuators are responsive to control the hydraulic pressure to the one or more frictions devices. It will be understood, however, that the friction applied by each of the plurality of friction devices may alternatively be controlled in accordance with other conventional friction device control structures and techniques, and such other conventional friction device control structures and techniques are contemplated by this disclosure. In any case, however, the analog operation of each of the friction devices is controlled by the control circuit  142  in accordance with instructions stored in the memory unit  144 . 
     In the illustrated embodiment, the system  100  further includes a drive unit control circuit  160  having an input/output port (I/O) that is electrically coupled to the drive unit  102  via a number, K, of signal paths  162 , wherein K may be any positive integer. The drive unit control circuit  160  may be conventional, and is operable to control and manage the overall operation of the drive unit  102 . The drive unit control circuit  160  further includes a communication port, COM, which is electrically connected to a similar communication port, COM, of the transmission control circuit  142  via a number, L, of signal paths  164 , wherein L may be any positive integer. The one or more signal paths  164  are typically referred to collectively as a data link. Generally, the drive unit control circuit  160  and the transmission control circuit  142  are operable to share information via the one or more signal paths  164  in a conventional manner. In one embodiment, for example, the drive unit control circuit  160  and transmission control circuit  142  are operable to share information via the one or more signal paths  164  in the form of one or more messages in accordance with a society of automotive engineers (SAE) J-1939 communications protocol, although this disclosure contemplates other embodiments in which the drive unit control circuit  160  and the transmission control circuit  142  are operable to share information via the one or more signal paths  164  in accordance with one or more other conventional communication protocols. 
     With reference to  FIG. 2 , an exemplary gearing scheme  200  is provided for transferring torque from an input  202  of the transmission to an output  204  thereof. The input  202  and output  204  can be disposed along the same centerline as shown in  FIG. 2 . Moreover, the gearing scheme  200  further includes a plurality of clutches. The plurality of clutches can include a pair of rotating clutches, i.e., C 1  and C 2 , and three stationary clutches or brakes, i.e., C 3 , C 4 , and C 5 . Each of the clutches or brakes can include one or more plates. The plates can include friction material and thus comprise friction plates, whereas other plates can be reaction plates. 
     The gearing scheme  200  can also include a plurality of planetary gearsets. For example, in  FIG. 2 , the scheme  200  includes a first planetary gearset  206 , a second planetary gearset  208  and a third planetary gearset  210 . For purposes of this disclosure, the first planetary gearset can be referred to as a P 1  planetary gearset. Likewise, the second and third planetary gearsets can be referred to as P 2  and P 3 , respectively. Each planetary gearset can include a sun gear, a ring gear, and a carrier. For instance, the P 1  planetary gearset  206  includes a P 1  sun gear  214 , a P 1  carrier  216 , and a P 1  ring gear  218 . The P 2  planetary gearset  208  includes a P 2  sun gear  220 , a P 2  carrier  222 , and a P 2  ring gear  224 . Similarly, the P 3  planetary gearset  210  includes a P 3  sun gear  226 , a P 3  carrier  228 , and a P 3  ring gear  230 . 
     In  FIG. 2 , the lines connecting the different components can refer or indicate paths through which torque can be transferred. In addition, where the lines are broken and horizontal lines are shown (e.g., forming an equal sign (“=”)), these locations can refer to gears meshing (or, alternative embodiments, being splined) to one another. For instance, the P 1  sun gear  214  is shown meshing with pinion gears coupled to the P 1  carrier  216  and the pinion gears of the P 1  carrier  216  are meshing with the P 1  ring gear  218 . When the C 1  or C 2  clutch is unapplied, torque does not pass through the clutches. Similarly, when the C 3  clutch (or brake) is applied, for example, the P 1  ring gear  218  is held and cannot rotate. From this point forward, the C 1  and C 2  clutches will be referred to as “clutches” whereas the C 3 -C 5  clutches will be referred to as “brakes”, but it should be understood that these components may be different in other embodiments. 
     In the present disclosure, the gearing scheme  200  is such that two clutches (or two brakes or one clutch and one brake) are applied to achieve a particular range or gear ratio. In other gearing scheme embodiments, however, a range may be achieved by applying any combination of clutches or brakes (e.g., one clutch, three clutches, four clutches, etc.). 
     In one particular embodiment, a gearing scheme can be arranged such that the following gear ratio ranges are achieved for different ranges: 
     
       
         
               
               
               
               
             
           
               
                   
                   
               
               
                   
                   
                   
                 Applied 
               
               
                   
                 Gear 
                   
                 Clutches or 
               
               
                   
                 Range 
                 Gear Ratio 
                 Brakes 
               
               
                   
                   
               
             
             
               
                   
                 F1 
                 2.5-6.0 
                 C1 &amp; C5 
               
               
                   
                 F2 
                 1.5-4.5 
                 C1 &amp; C4 
               
               
                   
                 F3 
                 1.1-3.0 
                 C1 &amp; C3 
               
               
                   
                 F4 
                 0.9-1.1 
                 C1 &amp; C2 
               
               
                   
                 F5 
                 0.25-0.99 
                 C2 &amp; C3 
               
               
                   
                 F6 
                 0.25-0.99 
                 C2 &amp; C4 
               
               
                   
                 Reverse 
                 (−3.0)-(−7.0) 
                 C3 &amp; C5 
               
               
                   
                   
               
             
          
         
       
     
     “F 1 ” refers to a first forward range, “F 2 ” refers to a second forward range, etc. The combination of gear ratios, ranges, and applied clutches and brakes is exemplary and non-limiting. Other embodiments are possible in which additional or fewer ranges are possible. The gear ratio can also be adjusted as desired to accommodate a close ratio or wide ratio transmission. As is known, the difference between a close ratio and wide ratio transmission is the number of gear teeth on the various parts of the planetary gearsets. 
     In the gearing scheme  200 , the reverse gear ratio may be too slow to meet the needs of a customer requiring a faster reverse ratio. In addition, the large F 1  to F 2  gear step prevents the shift from being made with the torque converter clutch applied, which can result in lower transmission efficiency, lower fuel economy, and higher transmission heat generation. In the present disclosure, however, an improved gearing scheme  300  can overcome such limitations so that the transmission can provide a reverse ratio that allows a vehicle to move faster in reverse than conventional reverse ranges in multi-ratio transmissions. Further, the F 1  to F 2  shift can be made with the torque converter clutch applied. 
     In the first forward range, F 1 , of gear scheme  200 , the torque path through the transmission is such that the C 1  clutch and C 5  brake are applied. When the C 5  brake is applied, it prevents the P 3  ring gear  230  from rotating. With the P 3  ring gear  230  stopped, the P 3  sun gear  226  is an input of torque and the P 3  carrier  228  is the output. Here, the P 3  sun gear  226  and P 3  carrier  228  rotate in the same direction, but the P 3  carrier  228  rotates at a slower speed. 
     Referring to  FIG. 2 , the reverse range, R, can be achieved by applying the C 3  and C 5  brakes. Both of these brakes are stationary and hold ring gears from rotating. Specifically, the C 3  brake holds the P 1  ring gear  218  and the C 5  brake holds the P 3  ring gear  230 . Since neither the C 1  nor C 2  clutches are applied, torque is transferred from the transmission input  202  to the first planetary gearset  206  via the P 1  sun gear  214 . Since the P 1  ring gear  218  is held, the P 1  carrier  216  rotates in the same direction as the P 1  sun gear  214 , albeit at a slower speed, and transfers torque to the P 2  planetary gearset  208  via the P 2  ring gear  224 . 
     Since the C 5  brake holds the P 3  ring gear  230 , the P 2  carrier  222  cannot rotate. As such, the P 2  sun gear  220  is the output of the P 2  planetary gearset  208 . Since the P 2  carrier  222  is held, the P 2  ring gear  224  and P 2  sun gear  220  rotate in opposite directions relative to each other, but with the P 2  sun gear  220  rotating at a higher speed than the P 2  ring gear  224 . With the P 2  sun gear  220  now rotating in an opposite direction from the transmission input, the P 2  sun gear  220  also drives the P 3  sun gear  226  to rotate in the opposite direction of the transmission input  202 . The P 3  sun gear  226  continues to drive the P 3  carrier  228  in the opposite direction of the transmission input  202 , albeit at a slower speed, and thus the transmission output  204  (which is driven by the P 3  carrier  228 ) rotates in the reverse direction (i.e., opposite from the direction of which the transmission input  202  rotates). 
     Referring to  FIG. 3 , another exemplary embodiment of a gear scheme  300  is shown. The gear scheme  300  in  FIG. 3  is similar to the gear scheme  200  of  FIG. 2 , except for the addition of a fourth planetary gearset  312 . Thus, the inputs and outputs of the P 1  carrier, P 2  carrier, and P 3  carrier in the gear scheme  300  are the same as the previously described gear scheme  200 . 
     As shown, the gear scheme  300  can include a transmission input  302 , a transmission output  302 , two rotating clutches, C 1  and C 2 , and four stationary brakes, i.e., C 3 , C 4 , C 5 , and C 6 . The input  302  and output  304  can be disposed along the same transmission centerline as shown in  FIG. 3 . Alternatively, the input  302  and output  304  can be disposed along different centerlines. The gear scheme  300  also can include a first planetary gearset  306 , a second planetary gearset  308 , a third planetary gearset  310 , and a fourth planetary gearset  312 . The first planetary gearset  306 , referred to herein as the P 1  planetary gearset, can include a P 1  sun gear  314 , a P 1  carrier  316 , and a P 1  ring gear  318 . Likewise, the second planetary gearset  308 , or P 2  planetary gearset, can include a P 2  sun gear  320 , a P 2  carrier  322 , and a P 2  ring gear  324 . The third planetary gearset  310 , or P 3  planetary gearset, can include a P 3  sun gear  326 , a P 3  carrier  328 , and a P 3  ring gear  330 . Similarly, the fourth planetary gearset  312 , or P 4  planetary gearset, can include a P 4  sun gear  332 , a P 4  carrier  334 , and a P 4  ring gear  336 . 
     The gear scheme  300  can include similar ranges, gear ratios, and applied/unapplied clutches as described above with respect to gear scheme  200 . Thus, in a first forward range or speed ratio, F 1 , the C 1  clutch and C 5  brake can be applied. In this embodiment, the P 3  ring gear  330  is held fixed by the C 5  brake. Torque therefore passes through the P 3  planetary gearset  310  via the P 3  sun gear  326  and it outputs through the P 3  carrier  328 . The P 3  sun gear  326  and P 3  carrier  328  rotate in the same direction, but the P 3  carrier  328  rotates at a slower speed multiplying torque to the transmission output  304  by the same ratio as gear scheme  200 . 
     In reverse, however, the gear scheme  300  differs over the previous gear scheme  200 . The addition of the fourth planetary gearset  312  in  FIG. 3  results in two available reverse gear or speed ratios. R 1 , or a first reverse gear or speed ratio of gear scheme  300 , has a similar torque path and gear ratio as the single reverse ratio in the gear scheme  200  of  FIG. 2 . R 2 , or a second reverse gear or speed ratio of gear scheme  300 , has a “faster” reverse gear or speed ratio than the single reverse gear or speed ratio of gear scheme  200 . For purposes of this disclosure, the “faster” reverse gear or speed ratio (R 2 ) in scheme  300  means the numerical gear or speed ratio is smaller than the single reverse gear or speed ratio of scheme  200 . In addition, a reverse range can refer to a gear ratio or speed ratio for purposes of this disclosure. 
     The second reverse ratio, R 2 , of scheme  300  can be achieved by applying the C 3  and C 6  brakes. Both of these brakes can be stationary brakes (i.e., not rotating) and hold corresponding ring gears from rotating. Specifically, the C 3  brake holds the P 1  ring gear  318  and the C 6  brake holds the P 4  ring gear  336 . Since neither the C 1  nor C 2  clutches are applied, torque is transferred from the transmission input  302  to the first planetary gearset  306  via the P 1  sun gear  314 . Since the P 1  ring gear  318  is held, the P 1  carrier  316  rotates in the same direction as the P 1  sun gear  314 , albeit at a slower speed, and transfers torque to the P 2  planetary gearset  308  via the P 2  ring gear  324 . 
     With the transmission output  304  rotating in the reverse direction and the P 4  ring gear  336  held, the P 4  carrier  334 , P 3  ring gear  330 , and P 2  carrier  322  all rotate in the reverse direction as well. Now, with the P 2  ring gear  324  rotating forward and the P 2  carrier  322  rotating in reverse, the P 2  sun gear  320  and the P 3  sun gear  326  rotate at a high speed in reverse. Since the P 3  sun gear  326  and P 3  ring gear  330  are both rotating in the reverse direction, the P 3  carrier  328  and the transmission output  304  rotate at a relatively high speed in the second reverse ratio, R 2 , of gear scheme  300  compared to the single reverse ratio of gear scheme  200  and the first reverse ratio, R 1 , of gear scheme  300 . In one non-limiting embodiment, R 2  of gear scheme  300  can have a gear ratio of about −1.95 while the single reverse ratio in gear scheme  200  and first reverse ratio, R 1 , of gear scheme  300  have gear ratios of about −4.80. The second reverse ratio, R 2 , of gear scheme  300  can therefore provide a faster vehicle speed in reverse for those vocations requiring it. 
     Gear scheme  300  can also provide an additional forward gear ratio that fits between the F 1  and F 2  gear ratios of gear scheme  200 . This allows for seven forward speeds to be available with gear scheme  300  and can reduce the large gear step typically associated with the F 1  to F 2  shift of gear scheme  200 . The F 1  torque path and gear ratio of gear scheme  300  are similar to that of the F 1  torque path in gear scheme  200 . The F 3  torque path and gear ratio of gear scheme  300  are similar to that as F 2  of gear scheme  200 , and likewise the pattern for the remaining higher forward ranges of gear scheme  300  are similar to those of the gear scheme  200 . However, the F 2  torque path and gear ratio of gear scheme  300  are different from that of the gear scheme  200 . 
     The second forward range, F 2 , of gear scheme  300  can be achieved by applying the C 1  clutch and C 6  brake. In  FIG. 3 , the P 4  ring gear  336  is held by the C 6  brake. Input into the P 3  sun gear  326  of the P 3  planetary gearset  310  is via the applied C 1  clutch. The P 3  sun gear  326  can transfer torque to the P 3  carrier  328  which drives the transmission output  304 . The P 3  sun gear  326  and P 3  carrier  328  rotate in the forward direction and further drive the P 4  sun gear  332 . With the P 4  ring gear  336  being held, the P 4  sun gear  332  drives the P 4  carrier  334 . The P 4  carrier  334  rotates in the forward direction and further drives the P 3  ring gear  330 . Since the P 3  ring gear  330  is rotating forward rather than being fixed as in F 1 , it combines with the forward rotation of the P 3  sun gear  326  to drive the P 3  carrier  328  and transmission output  304  at a faster rotational speed in the forward direction than that of F 1 , but slower than that of F 3  of gear scheme  300 . 
     In one non-limiting aspect of this embodiment, the F 1  and F 2  forward ranges in gear scheme  200  can have gear ratios of approximately 3.51 and 1.91 with a gear step of 1.84 therebetween. In gear scheme  300 , however, forward ranges F 1 , F 2 , and F 3  can have respective gear ratios of 3.51, 2.56, and 1.91 with gear steps of 1.37 and 1.34. As a result, a gear step at about or below 1.5 formed between forward ranges F 1  and F 2  of gear scheme  300  can allow the shift from F 1  to F 2  to be made with the torque converter clutch applied, thereby improving transmission efficiency and fuel economy while reducing transmission heat generation. 
     As for the other forward ranges in gear scheme  300 , the torque flow path between the transmission input  302  and transmission output  304  is similar to the other forward ranges in gear scheme  200 . For instance, in a third forward range F 3  of gear scheme  300 , the C 4  brake holds the P 2  ring gear  324  and the C 1  clutch is applied. Torque therefore passes through the P 2  planetary gearset  308  via the P 2  sun gear  320  and is output through the P 2  carrier  322 . The P 2  carrier  322  rotates in the same direction as the P 2  sun gear  320  and is coupled to the P 3  ring gear  330 . Torque also passes through the P 3  planetary gearset  310  as an input via the P 3  sun gear  326 . The P 3  carrier  328  outputs torque to the transmission output  304  as shown. 
     In a fourth forward range F 4  of gear scheme  300 , the C 3  brake can hold the first ring gear  318  and the C 1  clutch is applied. Torque can therefore enter the P 1  planetary gearset  306  via the P 1  sun gear  314  and output via the P 1  carrier  316 . The P 1  carrier  316  is coupled to the P 2  ring gear  324 , and as a result torque passes through the P 2  planetary gearset  308  through the P 2  ring gear  324  (via the P 1  carrier  316 ) and the P 2  sun gear  320 . The P 2  carrier  322  is the output of the P 2  planetary gearset  308  and transfers torque to the P 3  ring gear  330 . Thus, torque passes through the P 3  planetary gearset  310  via the P 3  ring gear  330  (via the P 2  carrier) and P 3  sun gear  326 . The P 3  carrier  328  outputs torque to the transmission output  304 . 
     In a fifth forward range F 5  of gear scheme  300 , the C 1  and C 2  clutches are applied, but none of the stationary brakes are applied. Thus, torque from the transmission input  302  enters the P 3  planetary gearset  310  via the P 3  sun gear  326  and the P 3  ring gear  330  (i.e., via the P 2  carrier  322 ). With the P 3  sun gear  326  and P 3  ring gear  330  rotating at input speed, the P 3  carrier  328  and transmission output  304  rotate at input speed thereby resulting in a gear ratio of approximately 1.0. 
     In a sixth forward range F 6  of gear scheme  300 , the C 2  clutch and C 3  brake are applied. Unlike the previously described embodiments, the C 1  clutch is not applied. The C 3  brake holds the P 1  ring gear  318 . Torque therefore passes through the P 1  planetary gearset  306  via the P 1  sun gear  314  and is output through the P 1  carrier  316 . Torque passes through the P 2  planetary gearset  308  through the P 2  ring gear  324  (via the P 1  carrier  316 ), P 2  carrier  322 , and the P 3  ring gear  330  (via the applied C 2  clutch). This overdrives the P 2  sun gear  320  and the P 3  sun gear  326 . With the P 3  sun gear  326  being overdriven and the P 3  ring gear rotating at input speed, the P 3  carrier  328  and transmission output  304  are overdriven. In this disclosure, the “overdriven” condition refers to the transmission output  304  rotating at a higher speed than the transmission input  302 . 
     In a seventh forward range F 7  of gear scheme  300 , the C 2  clutch and C 4  brake are applied or held. In this range, the C 4  brake holds the P 2  ring gear  324 . Torque is input to the P 2  planetary gearset  308  directly from the transmission input  302  to the P 2  carrier  322  and P 3  ring gear  330 . This condition overdrives the P 2  sun gear  320  and P 3  sun gear  326  to a larger degree than the sixth forward range F 6  of gear scheme  300 . With the P 3  sun gear  326  being overdriven to a greater degree and the P 3  ring gear  330  rotating at input speed, the P 3  carrier  328  and transmission output  304  are overdriven to a larger degree than in the sixth forward range F 6  of gear scheme  300 . 
     In the above-described embodiments, the torque flow paths can differ depending on the gear scheme and which clutches or brakes are applied/unapplied. In one embodiment, the gear ratio in the first forward range F 1  is greater than the gear ratios for the second (F 2 ), third (F 3 ), fourth (F 4 ), fifth (F 5 ), sixth (F 6 ), and seventh forward ranges (F 7 ). The gear ratio in the seventh forward range F 7  is less than the gear ratios in the first, second, third, fourth, fifth, and sixth forward ranges. With the ability to achieve a second range in reverse and an additional forward range, thereby providing smaller gear steps between ranges, the gear scheme  300  can achieve efficiency and fuel economy advantages over many conventional gearing arrangements. 
     While exemplary embodiments incorporating the principles of the present disclosure have been disclosed hereinabove, the present disclosure is not limited to the disclosed embodiments. Instead, this application is intended to cover any variations, uses, or adaptations of the disclosure using its general principles. Further, this application is intended to cover such departures from the present disclosure as come within known or customary practice in the art to which this disclosure pertains and which fall within the limits of the appended claims.

Technology Category: 2