Patent Publication Number: US-2009223317-A1

Title: Gearshift Interlock

Description:
CROSS-REFERENCE TO RELATED APPLICATIONS 
     This application claims the benefit of U.S. Provisional Application No. 60/816,779, filed Jun. 27, 2006. 
    
    
     FIELD OF THE INVENTION 
     The field of the present invention is that of gearshift interlocks and automotive transmissions which utilize gearshift interlocks. 
     BACKGROUND OF THE INVENTION 
     Generally speaking, land vehicles require a powertrain consisting of three basic components. These components include power plant (such as an internal combustion engine), a power transmission, and wheels. The power transmission component is typically referred to simply as the “transmission.” Engine torque and speed are converted in the transmission in accordance with the tractive-power demand of the vehicle. Presently, there are two typical transmissions widely available for use in conventional motor vehicles. The first and oldest type is the manually operated transmission. These transmissions include a foot-operated start-up or launch clutch that engages and disengages the driveline with the power plant and a gearshift lever to selectively change the gear ratios within the transmission. When driving a vehicle having a manual transmission, the driver must coordinate the operation of the dutch pedal, the gearshift lever, and the accelerator pedal to achieve a smooth and efficient shift from one gear to the next. The structure of a manual transmission is simple and robust and provides good fuel economy by having a direct power connection from the engine to the final drive wheels of the vehicle. Additionally, since the operator is given complete control over the timing of the shifts, the operator is able to dynamically adjust the shifting process so that the vehicle can be driven most efficiently. One disadvantage of the manual transmission is that there is an interruption in the drive connection during gear shifting. This results in losses in efficiency. In addition, there is a great deal of physical interaction required on the part of the operator to shift gears in a vehicle that employs a manual transmission. 
     The second, and newer choice for the transmission of power in a conventional motor vehicle is an automatic transmission. Automatic transmissions offer ease of operation. The driver of a vehicle having an automatic transmission is not required to use both hands, one for the steering wheel and one for the gearshift, and both feet, one for the clutch and one for the accelerator and brake pedal in order to safely operate the vehicle. In addition, an automatic transmission provides greater convenience in stop and go situations, because the driver is not concerned about continuously shifting gears to adjust to the ever-changing speed of traffic. Although conventional automatic transmissions avoid an interruption in the drive connection during gear shifting, they suffer from the disadvantage of reduced efficiency because of the need for hydrokinetic devices, such as torque converters, interposed between the output of the engine and the input of the transmission for transferring kinetic energy therebetween. In addition, automatic transmissions are typically more mechanically complex and therefore more expensive than manual transmissions. 
     For example, torque converters typically include impeller assemblies that are operatively connected for rotation with the torque input from an internal combustion engine, a turbine assembly that is fluidly connected in driven relationship with the impeller assembly and a stator or reactor assembly. These assemblies together form a substantially toroidal flow passage for kinetic fluid in the torque converter. Each assembly includes a plurality of blades or vanes that act to convert mechanical energy to hydrokinetic energy and back to mechanical energy. The stator assembly of a conventional torque converter is locked against rotation in one direction but is free to spin about an axis in the direction of rotation of the impeller assembly and turbine assembly. When the stator assembly is locked against rotation, the torque is multiplied by the torque converter. During torque multiplication, the output torque is greater than the input torque for the torque converter. However, when there is no torque multiplication, the torque converter becomes a fluid coupling. Fluid couplings have inherent slip. Torque converter slip exists when the speed ratio is less than 1.0 (RPM input&gt;than RPM output of the torque converter). The inherent slip reduces the efficiency of the torque converter. 
     While torque converters provide a smooth coupling between the engine and the transmission, the slippage of the torque converter results in a parasitic loss, thereby decreasing the efficiency of the entire powertrain. Further, the torque converter itself requires pressurized hydraulic fluid in addition to any pressurized fluid requirements for the actuation of the gear shifting operations. This means that an automatic transmission must have a large capacity pump to provide the necessary hydraulic pressure for both converter engagement and shift changes. The power required to drive the pump and pressurize the fluid introduces additional parasitic losses of efficiency in the automatic transmission. 
     In an ongoing attempt to provide a vehicle transmission that has the advantages of both types of transmissions with fewer of the drawbacks, combinations of the traditional “manual” and “automatic” transmissions have evolved. Most recently, “automated” variants of conventional manual transmissions have been developed which shift automatically without any input from the vehicle operator. Such automated manual transmissions typically include a plurality of power-operated actuators that are controlled by a transmission controller or some type of electronic control unit (ECU) to automatically shift synchronized clutches that control the engagement of meshed gear wheels traditionally found in manual transmissions. The design variants have included either electrically or hydraulically powered actuators to affect the gear changes. However, even with the inherent improvements of these newer automated transmissions, they still have the disadvantage of a power interruption in the drive connection between the input shaft and the output shaft during sequential gear shifting. Power interrupted shifting results in a harsh shift feel that is generally considered to be unacceptable when compared to smooth shift feel associated with most conventional automatic transmissions. 
     To overcome this problem, other automated manual type transmissions have been developed that can be power-shifted to permit gearshifts to be made under load. Examples of such power-shifted automated manual transmissions are shown in U.S. Pat. No. 5,711,409 issued on Jan. 27, 1998 to Murata for a Twin-Clutch Type Transmission, and U.S. Pat. No. 5,966,989 issued on Apr. 4, 2000 to Reed, Jr. et al for an Electro-mechanical Automatic Transmission having Dual Input Shafts. These particular types of automated manual transmissions have two clutches and are generally referred to simply as dual, or twin, clutch transmissions. The dual clutch structure is most often coaxially and cooperatively configured so as to derive power input from a single engine flywheel arrangement. However, some designs have a dual clutch assembly that is coaxial but with the clutches located on opposite sides of the transmission&#39;s body and having different input sources. Regardless, the layout is the equivalent of having two transmissions in one housing, namely one power transmission assembly on each of two input shafts concomitantly driving one output shaft. Each transmission can be shifted and clutched independently. In this manner, uninterrupted power upshifting and downshifting between gears, along with the high mechanical efficiency of a manual transmission is available in an automatic transmission form. Thus, significant increases in fuel economy and vehicle performance may be achieved through the effective use of certain automated manual transmissions. 
     The dual clutch transmission structure may include two dry disc clutches each with their own clutch actuator to control the engagement and disengagement of the two-clutch discs independently. While the clutch actuators may be of the electromechanical type, since a lubrication system within the transmission requires a pump, some dual clutch transmissions utilize hydraulic shifting and clutch control. These pumps are most often gerotor types, and are much smaller than those used in conventional automatic transmissions because they typically do not have to supply a torque converter. Thus, any parasitic losses are kept small. Shifts are accomplished by engaging the desired gear prior to a shift event and subsequently engaging the corresponding clutch. With two clutches and two inputs shafts, at certain times, the dual clutch transmission may be in two different gear ratios at once, but only one clutch will be engaged and transmitting power at any given moment. To shift to the next higher gear, first the desired gears on the input shaft of the non-driven clutch assembly are engaged, then the driven clutch is released and the non-driven clutch is engaged. 
     This requires that the dual clutch transmission be configured to have the forward gear ratios alternatingly arranged on their respective input shafts. In other words, to perform up-shifts from first to second gear, the first and second gears must be on different input shafts. Therefore, the odd gears will be associated with one input shaft and the even gears will be associated with the other input shaft. In view of this convention, the input shafts are generally referred to as the odd and even shafts. Typically, the input shafts transfer the applied torque to a single counter shaft, which includes mating gears to the input shaft gears. The mating gears of the counter shaft are in constant mesh with the gears on the input shafts. The counter shaft also includes an output gear that is meshingly engaged to a gear on the output shaft. Thus, the input torque from the engine is transferred from one of the clutches to an input shaft, through a gear set to the counter shaft and from the counter shaft to the output shaft. 
     Gear engagement in a dual clutch transmission is similar to that in a conventional manual transmission. One of the gears in each of the gear sets is disposed on its respective shaft in such a manner so that it can freewheel about the shaft. A synchronizer is also disposed on the shaft next to the freewheeling gear so that the synchronizer can selectively engage the gear to the shaft. To automate the transmission, the mechanical selection of each of the gear sets is typically performed by some type of actuator that moves the synchronizers. A reverse gear set includes a gear on one of the input shafts, a gear on the counter shaft, and an intermediate gear mounted on a separate counter shaft meshingly disposed between the two so that reverse movement of the output shaft may be achieved. 
     In the above noted transmission, synchronizer mechanisms for the 1-3 gear combination, 2-R gear combination and 4-6 gear combination are often associated with one another. It is desirable to provide an interlock for the synchronizer mechanisms to prevent simultaneous engagement of associated gears. 
     SUMMARY OF THE INVENTION 
     To meet the aforementioned and other manifold desires, a revelation of the present invention is brought forth. In a preferred embodiment, the present invention provides a gearshift interlock including a first shift block operatively associated with a first synchronized gear. The first shift block is movable between neutral and actuated positions and has a detent. A second shift block is provided operatively associated with a second synchronized gear. The second shift block is movable between neutral and actuated positions and has a detent. A lockout member is provided wherein movement of one of the shift blocks from the neutral position toward the actuated position causes that shift block to urge the lockout member to engage the other shift block detent preventing movement of the other shift block. 
     Other features of the invention will become more apparent to those skilled in the art as the invention is further revealed in the accompanying drawings and detailed description of the invention. 
    
    
     
       BRIEF DESCRIPTION OF THE DRAWINGS 
         FIG. 1  is a schematic view of an inventive preferred embodiment dual clutch transmission utilizing a gearshift interlock of the present invention. 
         FIG. 2  is a partial perspective view of a shift fork connection with a shift block of a gearshift interlock of the present invention. 
         FIG. 3  is a side schematic view of a gearshift interlock of the present invention. 
         FIGS. 4A-4C  are schematic front views illustrating operation of the gearshift interlock shown in  FIG. 2 . 
         FIGS. 5A-5C  are schematic top views illustrating operation of the gearshift interlock shown in  FIG. 2 . 
         FIGS. 6-10  are views similar to  FIG. 4B  of alternate preferred embodiments gearshift interlocks of the present invention. 
         FIGS. 11 and 11A  are front and side elevation views of a shift fork shown in  FIG. 2 . 
     
    
    
     DETAILED DESCRIPTION OF THE INVENTION 
     A representative dual clutch transmission that may be used with a gearshift interlock of the present invention is generally indicated at  10  in the schematic illustrated in  FIG. 1 . Specifically, as shown in  FIG. 1 , the dual clutch transmission  10  includes a dual, coaxial clutch arrangement including clutch mechanisms  32  and  34 , a first input shaft, generally indicated at  14 , a second input shaft, generally indicated at  16 , that is coaxial to the first, a counter shaft, generally indicated at  18 , an output shaft  20 , a reverse counter shaft  22 , a plurality of synchronizers, generally indicated at  24 , and a plurality of shift actuators generally (not shown). 
     The dual clutch transmission  10  forms a portion of a vehicle powertrain and is responsible for taking a torque input from a prime mover, such as an internal combustion engine, and transmitting the torque through selectable gear ratios to the vehicle drive wheels. The dual clutch transmission  10  operatively routes the applied torque from the engine through the dual, coaxial clutch arrangement  7  to either the first input shaft  14  or the second input shaft  16 . The input shafts  14  and  16  include a first series of gears, which are in constant mesh with a second series of gears disposed on the counter shaft  18 . Each one of the first series of gears interacts with one of the second series of gears to provide the different gear ratio sets used for transferring torque. The counter shaft  18  also includes a first output gear that is in constant mesh with a second output gear disposed on the output shaft  20 . The plurality of synchronizers  24  are disposed on the two input shafts  14 ,  16  and on the counter shaft  18  and are operatively controlled by the plurality of shift actuators to selectively engage one of the gear ratio sets. Thus, torque is transferred from the engine to the dual, coaxial clutch arrangement  7 , to one of the input shafts  14  or  16 , to the counter shaft  18  through one of the gear ratio sets, and to the output shaft  20 . The output shaft  20  further provides the output torque to the remainder of the powertrain. Additionally, the reverse counter shaft  22  includes an intermediate gear that is disposed between one of the first series of gears and one of the second series of gears, which allows for a reverse rotation of the counter shaft  18  and the output shaft  20 . Each of these components will be discussed in greater detail below. 
     Specifically, the dual, coaxial clutch arrangement  7  includes a first clutch mechanism  32  and a second clutch mechanism  34 . The first clutch mechanism  32  is, in part, physically connected to a portion of the engine flywheel (not shown) and is, in part, physically attached to the first input shaft  14 , such that the first clutch mechanism  32  can operatively and selectively engage or disengage the first input shaft  14  to and from the flywheel. Similarly, the second clutch mechanism  34  is, in part, physically connected to a portion of the flywheel and is, in part, physically attached to the second input shaft  16 , such that the second clutch mechanism  34  can operatively and selectively engage or disengage the second input shaft  16  to and from the flywheel. As can be seen from  FIG. 1 , the first and second clutch mechanisms  32 ,  34  are coaxial and axially spaced from one another such that the clutch housing of the first clutch mechanism  32  is in front of the clutch housing of the second clutch mechanism  34 . The first and second input shafts  14 ,  16  are also coaxial and co-centric such that the second input shaft  16  is hollow having an inside diameter sufficient to allow the first input shaft  14  to pass through and be partially supported by the second input shaft  16 . The first input shaft  14  includes a first input gear  38  and a third input gear  42 . The first input shaft  14  is longer in length than the second input shaft  16  so that the first input gear  38  and a third input gear  42  are disposed on the portion of the first input shaft  14  that extends beyond the second input shaft  16 . The second input shaft  16  includes a second input gear  40 , a fourth input gear  44 , a sixth input gear  46 , and a reverse input gear  48 . As shown in  FIG. 1 , the second input gear  40  and the reverse input gear  48  are fixedly supported on the second input shaft  16  and the fourth input gear  44  and sixth input gear  46  are rotatably supported about the second input shaft  16  upon bearing assemblies  50  so that their rotation is unrestrained unless the accompanying synchronizer is engaged, as will be discussed in greater detail below. 
     The counter shaft  18  is a single, one-piece shaft that includes the opposing, or counter, gears to those on the inputs shafts  14 ,  16 . As shown in  FIG. 1 , the counter shaft  18  includes a first counter gear  52 , a second counter gear  54 , a third counter gear  56 , a fourth counter gear  58 , a sixth counter gear  60 , and a reverse counter gear  62 . The counter shaft  18  fixedly retains the fourth counter gear  58  and sixth counter gear  60 , while first, second, third, and reverse counter gears  52 ,  54 ,  56 ,  62  are supported about the counter shaft  18  by bearing assemblies  50  so that their rotation is unrestrained unless the accompanying synchronizer is engaged as will be discussed in greater detail below. The counter shaft  18  also fixedly retains a first drive gear  64  that meshingly engages the corresponding second driven gear  66  on the output shaft  20 . The second driven gear  66  is fixedly mounted on the output shaft  20 . The output shaft  20  extends outward from the transmission  10  to provide an attachment for the remainder of the powertrain. 
     The reverse counter shaft  22  is a relatively short shaft having a single reverse intermediate gear  72  that is disposed between, and meshingly engaged with, the reverse input gear  48  on the second input shaft  16  and the reverse counter gear  62  on the counter shaft  18 . Thus, when the reverse gears  48 ,  62 , and  72  are engaged, the reverse intermediate gear  72  on the reverse counter shaft  22  causes the counter shaft  18  to turn in the opposite rotational direction from the forward gears thereby providing a reverse rotation of the output shaft  20 . It should be appreciated that all of the shafts of the dual clutch transmission  10  are disposed and rotationally secured within the transmission  10  by some manner of bearing assembly such as roller bearings, for example, shown at  68  in  FIG. 1 . 
     The engagement and disengagement of the various forward and reverse gears is accomplished by the actuation of the synchronizers  24  within the transmission. As shown in  FIG. 1  in this example of a dual clutch transmission  10 , there are four synchronizers  74 ,  76 ,  78 , and  80  utilized to shift through the six forward gears and reverse. It should be appreciated that there are a variety of known types of synchronizers that are capable of engaging a gear to a shaft and that the particular type employed for the purposes of this discussion is beyond the scope of the present invention. Generally speaking, any type of synchronizer that is movable by a shift fork or like device may be employed. As shown in the representative example of  FIG. 1 , the synchronizers (with the exception of synchronizer  76 ) are dual actuated synchronizers, such that they selectively engage one of two separate gears to the same respective shaft. Specifically with reference to the example illustrated in  FIG. 1 , synchronizer  78  can engage the first counter gear  52  on the counter shaft  18  or engage the third counter gear  56 . Synchronizer  80  can engage the reverse counter gear  62  or engage the second counter gear  54 . Likewise, synchronizer  74  can engage the fourth input gear  44  or engage the sixth input gear  46 . Single acting synchronizer  76  can selectively connect the end of the first input shaft  14  to the output shaft  20  thereby providing a direct 1:1 (one to one) drive ratio for fifth gear. It should be appreciated that this example of the dual clutch transmission is representative and that other gear set, synchronizer, and shift actuator arrangements are possible within the dual clutch transmission  10  as long as the even and odd gear sets are disposed on opposite input shafts. 
     To actuate the synchronizers  74 ,  76 ,  78 , and  80 , this representative example of a dual clutch transmission  10  utilizes hydraulically driven shift actuators with attached shift forks. The dual actuated synchronizers  78 ,  74  and  80  all incorporate a gearshift interlock  70  (only the gearshift interlock for the synchronizer  78  is shown for clarity of illustration) of the present invention to prevent inadvertent simultaneous multiple gear engagement. 
     Referring to  FIGS. 2-5C ,  11 , and  11 A the gearshift interlock  70  arrangement of the present invention includes a first shift block  102 . The first shift block  102  is operatively associated with a first synchronized gear  56 . The first shift block  102  has a cut out  103  formed to accept a foot  105  of a shift fork  107 . 
     The shift block  102  is linearly slideably mounted in a housing  110  having a closed end  113  and an open end  111 . Adjacent the open end  111  is a blind flange  118 . The first shift block  102  is sealed within a first control volume  106  along a first extreme end, and a second control volume  108  along a second extreme end. The first shift block  102  has a neutral position  115  as shown in  FIGS. 4B and 5B . To hydraulically move the shift block  102  to a fully actuated position  125 , the control volume  106  is pressurized (via an inlet/outlet line  109 ) and or the controlled volume  108  is depressurized (via an inlet/outlet line  101 ). The shift block  102  will move in a direction of arrow  122  to the position  125 . To return the shift block  102  to the neutral position  115 , the control volume  108  is pressurized and or the control volume  106  is depressurized. 
     The shift block  102  has an integrally formed conical detent  114 . The detent  114  faces a generally adjacent second shift block  116 . The second shift block  116  is operatively associated with a second synchronized gear  52  (via a shift fork, not shown) that is a mirror image of the shift fork  107 . The shift forks have axially and laterally offset collars  117  allowing a centerline  127  of the collars to be axially aligned with each other. The second shift block  116  is typically a mirror image the first shift block  102  and as shown in  FIGS. 4B and 5B  shares a common neutral position  115 . The second shift block  116  is hydraulically moved along a path  119  that is parallel to a path  120  of travel for the first shift block  102 . Actuation of the second shift block  116  causes the second shift block  116  to move in a direction of arrow  123  opposite of that of arrow  122  to a position  123 . The second shift block  116  is sealed along its extreme ends in control volumes  124  and  126 . 
     Positioned between the shift blocks  102  and  116  in a concave seat  128  is a spherical lockout member or ball  130 . When the shift blocks  102  and  116  are in the neutral positions as shown in  FIGS. 4B and 5B  the lockout ball  130  is positioned generally within both of the detents  114  (with a slight amount of clearance  134  with both detects  114 ). When the shift block  102  is moved in the direction of arrow  122  during activation, a ride out surface  132  of the shift block  102  urges the lockout ball  130  fully into the detent  114  of the second shift block  116 . With the lockout member  130  fully engaged within the detent  114  of the second shift block  116 , the second shift block  116  is locked out from movement ( FIGS. 4A and 5A ). Consequently, the gear  52  associated with the second shift block  116  cannot be engaged. When the first shift block  102  is returned to its neutral position  115  shown in  FIGS. 4B and 5B , the locking ball slight clearance  134  with the detents  114  is restored. From the neutral position  115  the second shift block  116  and its associated gear can be engaged causing the lockout ball to  130  fully engage with the detent  114  of the first shift block  102  and the first block  102  and its associated gear is blocked from engagement ( FIGS. 4C and 5C ). 
     Referring to  FIG. 6 , an alternate preferred embodiment of gearshift interlock  147  is shown. The shift blocks  148  and  150  are almost identical to those of aforedescribed. Spherical balls  151  provide a multiple-piece lockout member. The lockout balls  151  are positioned in a generally flat seat  152 . 
     Referring to  FIG. 7 , an alternate preferred embodiment gearshift interlock  167  is provided having a generally concave seat  168  and arcuate lockout member  170 . Instead of the translational movement of the lockout ball, lockout member  170  is urged into arcuate movement upon activation of one of the shift blocks  174 ,  175 . 
     Referring to  FIG. 8 , an alternate embodiment gearshift interlock  187  is provided having convex bent elongated pendulum lockout member  190 . The pendulum  190  is positioned on a concave seat  192  and pivots about a pivot point  194 . The detent faces  195  and  196  of shift blocks  197  and  198  are parallel facing instead of the cross facing as with the detents  114  of  FIGS. 4A-5C . 
     Referring to  FIG. 9  an alternate embodiment gearshift interlock  207  is provided with a bent elongated lockout member pendulum  210  and a pivot point  212 . The pivot point  212  is connected with a stem  216  that extends through an aperture  214  in the pendulum  210 . The stem  216  has a threaded portion  218  that is threaded within a bore of the housing  110 . The pivot point  212  can be adjusted axially to insure that the pendulum properly engages with the detent faces  219  and  220  of the of the shift blocks  222  and  224 . 
     Referring to  FIG. 10 , an alternate embodiment gearshift interlock  227  is provided. The gearshift interlock  227  has a straight pendulum  228 . The gearshift interlock  227  has angled outward facing detent faces  232 ,  233  on shift blocks  234  and  236 . 
     While preferred embodiments of the present invention have been disclosed, it is to be understood it has been described by way of example only, and various modifications can be made without departing from the spirit and scope of the invention as it is encompassed in the following claims.