Patent Publication Number: US-11391330-B2

Title: High performance synchronous transmission

Description:
CROSS-REFERENCE TO RELATED APPLICATIONS 
     This application claims priority to PCT application no. PCT/EP2017/081062, filed Nov. 30, 2017, which claims priority to Italian application no. 102016000122052, filed Dec. 1, 2016, the contents of which are incorporated by reference in their entireties. 
     The present invention relates to a high performance transmission, in particular to be used aboard a motorcycle, such as a scooter, as element for transmitting the motion generated by a motor to a driving wheel, in particular the rear wheel of the motorcycle. 
     In the latest generation scooters, the most commonly used transmission is of CVT (Continuosly variable transmission) type, known as continuous transmission or continuous variator. 
     It has the advantage of either providing a continuous traction and not requiring the manual actuation of the different ratios. However, by using elements of sliding type, such transmission is characterized by a low performance, above all in the transitory procedures when the hysteresis effect of the transmission belt is maximum. 
     This pushes downwards the general performance of the vehicle and increases the consumption thereof. 
     On the other hand, a much felt need in the field is that of limiting the consumption as much as possible, however maintaining, whenever requested by the market, the comfort level thereto the users got used by the CVT gearbox. 
     The object underlying the present invention is then to increase considerably the overall performance of the transmission in two-wheel vehicles for urban transport. 
     However, in the design of a transmission for scooters and the like, there is a base constraint which consists in that the crankshaft, which receives the motion from the piston in the cylinder, and the hub shaft, which transmits at the end of the transmission kinematic chain the motion to the rear wheel, are parallel therebetween and they are arranged at a distance depending from the engine position. 
     If, with a transmission of CVT type, these two shafts are substantially connected by a belt extending between two pulleys kinematically connected to such shafts by filling-in the distance therebetween, this scheme is not easy to be applied in case of a synchronous transmission, which uses a plurality of toothed wheels engaged therebetween with different transmission ratios, but with the constraint that they cannot be placed all side-by-side. 
     Moreover, another inherent difficulty in a synchronous transmission is the need for having an automatic gearbox, according to the operating condition of the vehicle. In fact, it is necessary to implement increasing or decreasing shifts without producing wrenches, jerks and abrupt slowing-down, with maximum running graduality and mildness. 
     The solution idea, to the problem of providing a transmission of the above-mentioned type, consists in optimizing the performance of the transmission itself, which could use a synchronous belt between two toothed pulleys or in case another synchronous system, for example a pinion-chain-toothed wheel system, and however a high performance system, for transmitting the motion between crankshaft and hub shaft, with fixed transmission ratio, instead of the CVT belt with variable transmission ratio, and a mechanical gearbox which, with a predefined number of ratios, replaces the ratio variations obtained by the pulleys of a CVT. 
     In particular, this new type of transmission has the problem of reducing the overall dimensions, however while keeping the number of gearwheels necessary to actuate a four-or-more-speed gearbox. 
     Therefore, the above-mentioned problem is solved by a high performance synchronous transmission as specified above as defined in the enclosed claim  1 . 
     The main advantage of the high-performance transmission according to the present invention lies in the fact that the side overall dimensions of the transmission is kept to the minimum thanks to the compactness allowed by the transmission structure. 
    
    
     
       The present invention will be described hereinafter according to some preferred embodiments thereof, provided by way of example and not with limitative purposes by referring to the enclosed drawings, wherein: 
         FIG. 1  shows a side elevational view of a scooter incorporating the transmission according to the present invention; 
         FIG. 2  shows a perspective view of the transmission of  FIG. 1 , closed in its container, and of the related engine block; 
         FIG. 3  shows a front view of an embodiment example of high performance synchronous transmission according to the present invention, without outer casing; 
         FIG. 4  shows a perspective top view and a view in flat longitudinal section of the transmission of  FIG. 3 ; 
         FIG. 5  shows a top plan view of the transmission of  FIG. 3 ; 
         FIG. 6  shows a rear perspective view of the transmission of  FIG. 3 ; 
         FIG. 7  shows a perspective section view of a first detail of the transmission of  FIG. 3 ; 
         FIG. 8  shows a perspective view of a second detail of the transmission of  FIG. 3 ; 
         FIG. 9  shows a perspective view of some components of the first detail of  FIG. 7 ; 
         FIG. 10  shows a front and perspective view in partial section of a third detail of the transmission of  FIG. 3 ; 
         FIG. 10A  shows a connecting scheme to combine the details of the previous figures; 
         FIG. 11  shows a top perspective partial view and a flat longitudinal section view of the transmission of  FIG. 3 , that is of the right side thereof; 
         FIG. 12  shows a perspective section view of a fourth detail of the transmission of  FIG. 3 ; 
         FIGS. 13A and 13B  show a perspective view and another perspective view in partial section of a fifth detail of the transmission of  FIG. 3 , respectively; 
         FIG. 14A  shows several schemes for actuating the transmission of  FIG. 3 , according to some variants thereof; 
         FIG. 14B  shows an operation diagram describing the behaviour of some portions of the fifth detail of  FIGS. 13A and 13B ; 
         FIGS. 14C and 14D  show a perspective view and a side view of a component of the fifth detail of  FIGS. 13A and 13B  otherwise not visible in such figures, respectively; 
         FIG. 15  shows a top perspective partial view and in flat longitudinal section of the transmission of  FIG. 3 , that is of the left side thereof; 
         FIG. 16  shows a perspective view of a sixth detail of the transmission of  FIG. 3 ; 
         FIG. 17  shows a first perspective section view of the sixth detail of  FIG. 16 ; 
         FIG. 18  shows a second perspective section view of the sixth detail of  FIG. 16 ; 
         FIG. 19  shows a third perspective section view of the sixth detail of  FIG. 16 ; 
         FIGS. 20A, 20B and 20C  show respective perspective views of some components of the sixth detail of FIG.  16 , in particular  FIGS. 20A and 20B  show respective sides of the same component; 
         FIG. 21  shows a schematic view of an additional component of the sixth detail of  FIG. 16 ; 
         FIG. 22  illustrates the operation of said sixth detail of  FIG. 16  in relation to the component of  FIG. 21 ; 
         FIG. 23  shows a perspective view of a seventh detail of the transmission of  FIG. 3 ; and 
         FIG. 24  shows a section of the seventh detail of  FIG. 20 . 
     
    
    
     By referring to  FIGS. 1 and 2 , a motorcycle, and in particular a scooter, is designated as a whole with  100 . The invention relates to the field of the vehicles with a saddle, or straddle vehicles which are driven astride, generally, with two, three or four wheels, with particular reference to the scooters having a propulsion unit arranged in a position under a saddle  101 , inside a chassis  102 , which herein is represented laterally, extending from a front wheel  103 , controlled by handlebar  104  to a driving rear wheel  105 . 
     The propulsion unit  106  ( FIG. 2 ) or, briefly, the engine is of the type having one or more cylinders arranged in a position approximately tilted on the median plane of the vehicle corresponding to the rotation plane of the two wheels during the forward rectilinear running. 
     The engine  106  has an engine block  107  in one single piece which receives, in the present embodiment example, a cylinder  108  and a related (not shown) piston. 
     The piston acting in said cylinder  108  is connected to a crankshaft  2  positioned transversally and perpendicularly to said median plane. On a represented side ( FIGS. 1 and 2 ) of the scooter  100  a transmission device  1  or, more briefly, a transmission of the motion from the crankshaft to the hub of the rear wheel  105  is provided. 
     The herein described transmission is of the synchronous or nearly synchronous type, and it uses a pair of pulleys kinematically connected by a annular belt, preferably a toothed belt on toothed pulleys or a high performance belt, for example of the Stretch Fit® type or the like. 
     It is to be meant that what described hereinafter can be applied wholly or partially even to an equivalent synchronous transmission of other type, for example a pinion-chain-toothed wheel transmission. 
     By referring to the present example, the transmission  1  has a container  109  which receives inside the transmission elements which will be described hereinafter with greater details. The container  109  is connected to the engine block  107  by creating a tunnel-like casing containing the crankshaft  2  and all transmission elements connected thereto. 
     Moreover, the container  109 , on the exposed side of the motorcycle  100 , is closed by a cover  110  of the transmission  1 , substantially extending from the engine  106  to the hub shaft  75  of the driving wheel  105 . The cover  110  is fastened to the container  109  by means of suitable bolts  111 . Openings, slits, air intakes for accessing and/or cooling down the transmission elements through said cover  110  could be provided. 
     The cover  110  is rested upon a fastening edge  112  of the container  109 , equipped with fastening seats  113  for said bolts  111  and additional seats of front connection  114 , with a hinge connection of axis A to allow the engine block  107  and the transmission  1  to oscillate, and rear connection  115 , connected to a rear suspension  116 , for connecting the casing  109  and the whole transmission  1  to the frame of the vehicle  100 . 
     Such transmission is of the several-speed type and of synchronous type  1 , and it is arranged for connecting the crankshaft  2 , which receives the motion from the movement of one or more pistons, to the hub shaft  75 , by considering that these two shafts are parallel therebetween and placed at a prefixed distance. The hub shaft  75 , at one distal end thereof, is equipped with a pinion  76  connecting to the rear wheel  105 . 
     They are both perpendicular to the median plane of the vehicle, defined by the rotation plane of the front and rear wheels. It is further to be meant that the use of this type of transmission is not limited to the herein represented two-wheel scooter, but it can be extended to a scooter with a pair of front wheels or to a scooter with four wheels. 
     By referring to  FIG. 3  and to the subsequent figures, the transmission is herein designated as a whole with  1  and it comprises a crankshaft  2  comprising a crank  3  thereto a connecting rod  4  is connected which receives the motion by a not represented piston; however, it is to be meant that such transmission could be applied even to several-cylinder engines. 
     The crankshaft  2  extends from both sides of the crank: in the direction opposite to the transmission the crankshaft will be connected, by way of example, to an electric engine-generator, in case but not exclusively for an operation of hybrid type, and to a cooling valve. 
     In the direction of the transmission, the crankshaft comprises a starting centrifugal clutch  5 , which is useful to manage the starting of the vehicle from standstill. 
     In fact, the rotation of the crankshaft  2  puts in rotation the small hub  6  of the shaft and the mass-bearing plate  7  connected thereto, which drugs into rotation two clutch masses  8  ( FIG. 8 ) which tend to move away therefrom due to the effect of the centrifugal force acting thereon, in opposition to clutch springs  9 . 
     Once reached the defined rotation regime, the masses  8 , through the friction material placed on the outer periphery thereof, transmit the motion to a first clutch housing  10 , which is stiffly keyed on a bushing  11  assembled on clutch bearings  12  in order to guarantee the rotation between crankshaft  2  and housing  10  when the clutch is in the idle position. 
     Moreover, a driving pulley  13  is arranged on the bushing  11 , this pulley surrounding the distal end thereof and which is stiffly keyed thereto. The driving pulley  13  ( FIG. 7 ) has a mobile coupling element  14 , shaped like a crown and inserted inside thereof, that is between the pulley  13  and the bushing  11 , and capable of sliding with respect to the distal end of the shaft  2  thereabout it is slidingly put, so as to be free to translate in opposition to a pre-loaded spring  15  arranged between the driving pulley  13  and the distal end of the mobile crown  14 . 
     On the distal end of the crankshaft  2  a fixed coupling element  77  is instead provided, integral thereto, with respect thereto said mobile coupling element  14  slides. 
     The two mobile and fixed coupling elements, which can be slidingly and axially controlled to engage and disengage, so as to connect directly the driving pulley  13  to the crankshaft  2  by excluding the centrifugal clutch  5 , constitute a coupling between the crankshaft  2  and the driving pulley, with the function of excluding the centrifugal clutch of the present transmission. 
     The fixed coupling element  77  has a first axial toothing  78  projecting outwardly and radially; the mobile coupling element  14  has both a second axial toothing  79  projecting inside thereof and radially, and it is intended for coupling with said first axial toothing  77 , and a third axial toothing  21  projecting outwardly and radially ( FIG. 9 ). 
     Under axial toothing a toothing is meant whose teeth extend according to an axial direction on the belonging element, arranged for coupling by sliding in an axial direction with the teeth of a complementary axial toothing. 
     The driving pulley  13 , in turn, has a fourth axial toothing  22  which is intended to couple with said third axial toothing of the mobile coupling element  14 . It has a cup shaped element  16 , rested on a peripheral edge thereof and so as to project outside the driving pulley  13 , coaxial thereto, and capable of pushing, if forced in this direction, on the mobile coupling element  14  thus by making the second axial toothing  79  and the third axial toothing  21  to slide with respect to the first axial toothing  78  and to the fourth axial toothing  22 , respectively. 
     The cup shaped element  16  acts as first actuation button  16  or engagement button. 
     The configuration is so that, when a pressure is not exerted on the first actuation button  16 , the mobile coupling element  14  translates coaxially to the crankshaft  2  by moving away from the bushing  11 , pushed by the force of the preloaded spring  15 . Such translation determines the coupling of the first toothing  78  of the fixed coupling element  77 , integral to the crankshaft  2 , and of the second toothing  79  of the mobile coupling element  14 , whereas the third axial toothing of the mobile coupling element  14  and the fourth axial toothing  22  of the driving pulley  13  are always engaged therebetween, but they allow to the mobile coupling element  14  to slide with respect to the driving pulley  13  by implementing a prismatic coupling. 
     Therefore, by leaving free the first actuation button  16  a direct mechanical connection between the crankshaft  2 , the fixed coupling element  77 , the mobile coupling element  14  (due to the first and the second axial toothing  78 ,  79 ) and the driving pulley  13  (due to the third and the fourth toothing  21 ,  22 ) is determined and, under such operating condition, the driving pulley  13  is dragged into rotation by the crankshaft  2  whatever the rotation regime of the latter may be, that is whatever the operating state of the centrifugal clutch  5  may be. 
     In this operating state, even the bushing  11  is dragged into rotation by the driving pulley  13 , even if it does not receive the motion by the centrifugal clutch  5 : it can rotate freely on the bearings  12  even if the rotation regime thereof to say the truth is equal to that of the crankshaft  2 , but wrenches in the transition phase are avoided. On the contrary, if the centrifugal clutch  5  is engaged, the rotation regime thereof is equal to that of the crankshaft  2  and of the driving pulley  13 . 
     This operating state, as it will clearly appear here below in the present description, corresponds to the second, third and fourth speed, that is to any higher speed than the first speed, wherein one wishes that the driving pulley  13  transmits the motion to the driving wheel  105 , regardless the rotation regime of the crankshaft  2 , then even below the threshold which would determine the disengagement of the centrifugal clutch  5 . 
     On the contrary, if the first actuation button  16  is pressed, the mobile coupling element  14  is pushed in the direction of the bushing  11  in opposition to the action of the preloaded spring  15 , by disengaging the first and the second axial toothing  78 ,  79  and then by disengaging the driving pulley  13  from the crankshaft  2 . In this state, the mobile coupling element  14  can receive and transmit the motion given to it by the bushing  11  through the centrifugal clutch  5 . In fact, the bushing  11  is released from the crankshaft  2  thanks to the bearings  12 . 
     This state corresponds to the first speed or the idle state, the centrifugal clutch  5  determining the passage from one to the other one and viceversa depending upon the rotation regime of the crankshaft  2 . 
     Therefore, by summing up, in the first speed the centrifugal clutch operates normally, by allowing the motion transmission and the starting of the motorcycle  100  above a predetermined rotation regime of the crankshaft  2 , wherein the centrifugal clutch  5  causes its own engagement. 
     In the second speed, and in the subsequent speeds the centrifugal clutch  5  actually is excluded from the kinematic chain, as the motion is transmitted by the crankshaft  2  directly to the driving pulley  13 , regardless the rotation regime of the crankshaft  2 , therefore even below said threshold wherein the centrifugal clutch  5  is not engaged. 
       FIG. 7  represents this second state, with a gap between distal end of the bushing  11  and proximal end of the mobile coupling element  14 . 
     The driving pulley  13  instead is useful to transmit the motion from the crankshaft  2  to the axis of a driven pulley  17  which constitutes the input of the real gearbox. 
     The two driving  13  and driven  17  pulleys are toothed and they are connected by a synchronous belt  18  with a fixed transmission ratio. The side containments of the belt  18  are mounted on this driving pulley  13  in order to optimize the transmission performance ( FIG. 10 ). 
     To this regard, a control lever  20  is provided to exert a pressure on the first actuation button  16 , that is on the engagement button. 
     Therefore, upon starting in the first speed, the control lever  20  is active, it pushes on the first actuation button  16  and thus on the crown-like mobile coupling element  14 , so as to disengage between them the first and the second toothing  78 ,  79 . 
     From the second speed on, the lever  20  moves away from the first actuation button  16  and it does not exert any pressure, in a not intervening position. 
     This implements the possibility of running at lower engine rotation regimes than those upon connecting the clutch  5 , an impossible procedure on all systems with automatic centrifugal clutch, including CVT systems. 
     The lever  20 , on this regard, has a pressing end  24  and it is oscillating with respect to a fulcrum  25  integral to a fixed portion of the transmission, thus to the container  109  ( FIG. 4 ). The way in which the control lever  20  is actuated will be described hereinafter. 
     As previously explained, the annular belt  18  which is wound on the driving pulley  13  implements a synchronous connection, as it is toothed, and it requires the presence of a fixed tensioner  30 , arranged in the lower branch of the belt  18  ( FIG. 10A ), arranged outside the ring formed by the belt  18  and pressing towards the inside of the ring itself. 
     This belt  18  is required for transmitting the motion from the axis of the crankshaft  2  to the axis of the gearbox input, placed in the area of the rear wheel  105 . 
     The transmission ratio is fixed and the tensioner  30  has to keep a constant load under all use conditions. 
     As already highlighted, it is to be noted that it is not strictly necessary that the belt  18  has to be toothed, as there are so-called high transmission performance belts, that is substantially synchronous or almost synchronous, with or without tensioning device  30 . 
     The latter ( FIG. 10 ) has an eccentricity in the central fastening thereof: the tensioning device  30  has a fixed pin  31 , integral to a fixed portion of the transmission, whereon a circular and eccentric supporting element  32  is assembled, which forms a circular periphery whereon a tensioner bearing is assembled, whereon, in turn, a pressing wheel  34  is assembled, positioned so as to exert a pressure between the smooth outer periphery thereof  35  on the tooth outer face  36  of the belt  18 . 
     The fixed pin  31  is arranged eccentric with respect to the supporting element  32 , so that, by rotating the latter during assembly, it is possible to move the wheel  34  by loading the belt  18 . 
     The fixed pin  31  is of the screw type and, once tightened, it locks the supporting element  32  in the wished operating position thereof. 
     If released, the fixed pin  31  allows again the rotation of the supporting element  32  thus by moving away the pressing wheel  34  from the belt  18 , by making easy, for example, the replacement thereof at the end of its life cycle. Then, it is sufficient to reposition the eccentric supporting element  32  in the maximum tensioning position thereof. 
     The driven pulley  17  is a toothed pulley too, or of other type depending upon the selected belt. It transmits the motion from the belt  18  to an input clutch  40  ( FIGS. 11 and 12 ) which performs materially the speed shifts. 
     The input clutch  40  is a clutch of the disc type and it comprises a second clutch housing  41  connected to the driven pulley  17 . The input clutch  40  transmits the motion to a primary shaft of the gearbox  51  the distal end thereof, facing towards the cover  110  of the transmission  1 , is connected to a clutch hub  42 . 
     The second clutch housing  41 , too, is assembled on the primary shaft of the gearbox  51  by means of a pair of first clutch bearings  37 , thanks thereto the rotation of the shaft  51  does not affect the clutch housing and vice versa. 
     Inside the housing  41  two clutch discs of the input clutch  40  are included: a more external first clutch disc  38  is connected to the housing  41 , whereas a second clutch disc  39  is faced thereto more internally. It is connected and integral to an inner disc-pushing element  44 , which surrounds and includes the clutch hub  42  thereto it is connected. The inner disc-pushing element acts axially on the clutch discs  38 ,  39  by opening and closing them. 
     A clutch cover  26  is connected to the first clutch disc  38 , which cover encloses the space included in the second clutch housing  41  and supports the disc-pushing elements which will be described hereinafter. 
     To this regard, a clutch spring  46  is positioned between the clutch hub  42  and an inner disc-pushing element  45  covering and towering above the clutch hub  42 . At the distal end of the primary shaft  51 , thus at the rotation centre thereof, the outer disc-pushing element  45  comprises a second actuation button  48  assembled on a second clutch bearing  49  which releases it from the rotation of the outer disc-pushing element  45 . 
     Onto the second actuation button  48  a pressure can be exerted which determines the detachment of the input clutch  40 . 
     The clutch discs  38 ,  39 , usually, are closed due to the effect of the load of the clutch spring  46 . The motion is then transmitted by the driven pulley  17  to the housing  41  and to the discs  38 ,  39 , and therefrom to the two disc-pushing elements  44 ,  45  and to the clutch hub, and then to the primary shaft  51 . 
     When a pressure is exerted on the second actuation button  48 , this pushes towards the distal end of the primary shaft  51  the outer disc-pushing element  45 : through the inner disc-pushing element the second clutch disc  39  is moved away from the first clutch disc  38 , thus by interrupting the kinematic continuity between second clutch housing  41  and clutch hub  42 . 
     The pressure on the actuating button is obtained by means of a clutch lever  47  the clutch fulcrum  27  thereof is connected to a fixed portion of the transmission  1 , that is to the container  109  or to the transmission cover  110 , analogously to what described for the control lever  20 . 
     The clutch lever  47  exerts a pressure through a pressing operating end  28 , pressure which opposes to the load of the clutch spring  46  which defines the dragging load of the clutch  40 . 
     The actuation of the clutch lever  47  will be described in greater detail here below in the present description. 
     By referring to  FIGS. 23 and 24 , on the input clutch  40 , between it and the clutch lever  47 , an adjustment of the clearance between the pressing operating end thereof  28  and the second actuation button  48  connected to the outer disc-pushing element  45  is provided. 
     Such adjustment is obtained by means of a clearance-adjusting element  90  which allows to adjust an assembly clearance defined and in case adjustable in time for maintenance. Such adjustment allows to set to zero the assembly clearances, due to the tolerances and in time the possible wears which could modify the timing between actuator and clutch itself. Once adjusted this intervention point, it follows that the actions of the device actuating the speeds, which will be described hereinafter, will be always synchronized and phased in the same way and with the margins provided by the tolerance on the actuation clearance. 
     The clearance-adjusting element  90  provides a locking nut  29 , assembled on said operating end  28  at a threaded hole  43 , which is useful to assemble an adjusting screw  92  inserted in said nut  29  and in said hole  43 . 
     The axial position of the adjusting screw  92  can be maneuvered simply by acting, with a suitable wrench, on its head  93 , so as to adjust the incidence of the operating end  28  by regulating the wished clearance. 
     In fact, by variating the axial position of the adjusting screw  92  the resting terminal thereof  94 , interfering with the second actuation button  48  assembled on the second clutch bearing  49  ( FIG. 24 ), is translated. 
     The input clutch  40  is arranged for driving a mechanical gear transmission  50 , the number of ratios thereof is not constraining. In the scheme which will be described hereinafter four ratios are provided. 
     The used gearbox scheme provides a primary axis and two secondary axes, and a final hub shaft, that is the wheel axis. This scheme can be the most suitable one for the type applying to a scooter, due to compactness in axial direction and versatility in managing ratios. 
     The gearbox  50  then comprises: a primary shaft  51 , already mentioned with reference to the input clutch  40  which transmits to it the motion, with an input gearwheel  60  which is connected to the clutch hub  42 ; a first secondary shaft  52  which is assigned to the first and third speed by means of a first running toothed wheel  61  and a third running toothed wheel  63  with different diameters, with a respective first output gearwheel  71  for engaging with the hub shaft  75  connected to the rear driving wheel  105 ; a second secondary shaft  53 , intended to the second and fourth speed with a second running toothed wheel  62  and a fourth running toothed wheel  64 , with output gearwheel  72  for engaging with the hub shaft  75  connected to the rear driving wheel  105 ; and at last the already mentioned hub shaft  75  which supports an output toothed wheel  73  with big diameter, so as to implement an additional reduction in the transmission ratio at the hub shaft  75 . 
     The above-mentioned toothed wheels  61 ,  62 ,  63  and  64  of the first, second, third and fourth speed, respectively, are assembled freely on the respective secondary shaft  52 ,  53  so that they can rotate with respect thereto, by remaining in a fixed and predetermined axial position, and they are respectively engaged with a first running pinion  54 , a second running pinion  55 , a third running pinion  56  and a fourth running pinion  57  arranged fixed and integral to the primary shaft  51 , for transmitting the first speed (first toothed wheel  61  of the first secondary shaft  52  and first pinion  54  of the primary shaft  51 ), of the second speed (second toothed wheel  62  of the second secondary shaft  53  and second pinion of the primary shaft  51 ), of the third speed (third toothed wheel  63  of the first secondary shaft  52  and third pinion  56  of the primary shaft  51 ) and of the fourth speed (fourth toothed wheel  64  of the second secondary shaft  53  and fourth pinion  57  of the primary shaft  51 ), with a transmission ratio decreasing from the first to the fourth speed due to the different diameters of the respective toothed wheels  61 ,  62 ,  63  and  64  of the two secondary shafts  52 ,  53  and of the pinions  54 ,  55 ,  56 , and  57  of the primary shaft  51  ( FIG. 13B ). 
     It is to be meant that, when they are not engaged, the toothed wheels  61 ,  62 ,  63  and  64  rotate dragged by the pinions  54 ,  55 ,  56 , and  57  without transmitting the motion to their own secondary shafts  52 ,  53 . 
     To this regard, respective first sliding coupling  65  and second sliding coupling  66  act on each secondary shaft  52 ,  53 , the couplings being controlled in axial translation with respect to the secondary shafts  52 ,  53  by a corresponding first coupling fork  67  and second coupling fork  68 . 
     The sliding couplings  65 ,  66  are wheels which have, on their own respective inner crown placed around the respective secondary shaft  52 ,  53 , a first spline coupling  131  and a second spline coupling  132  ( FIG. 14C ), respectively, in engagement with corresponding splines formed on the respective secondary shaft  52 ,  53 . It is to be meant that said sliding couplings  65 ,  66  are free to rotate with respect to their coupling forks  67 ,  68 . 
     The coupling forks  67 ,  68  are equipped with a cam transferring end  69  which are maneuvered by means of a desmodromic drum  70  having a cylindrical surface  79  whereon one single desmodromic track  19  is formed. 
     The first sliding coupling  65  has first coupling pins  133  and second coupling pins  134  on the opposite sides thereof, projecting in axial direction respectively in the direction of the first toothed wheel  61  and the third toothed wheel  63 . 
     Analogously, the second sliding coupling  66  has third coupling pins  135  and fourth coupling pins  136  on the opposite sides thereof, projecting in axial direction respectively in the direction of the second toothed wheel  62  and of the fourth toothed wheel  64 . 
     With the axial sliding motion of the respective sliding couplings  65 ,  66  the pins  133 ,  134 ,  135  and  136  are intended to engage in the toothed wheels  61 ,  62 ,  63  and  64  thereto they are facing, the latter wheels having first coupling recesses  137 , second coupling recesses  138 , third coupling recesses  139  and fourth coupling recesses  140 , respectively. 
     According to the herein described operation principle, the cam follower ends  69  of the coupling forks  67 ,  68  are constrained to follow the path defined by the track  19  implemented in the desmodromic drum  70 , during the rotation thereof. 
     The actuation of the desmodromic drum  70 , which rotates by an angular amount varying depending upon the speed to be selected, leads to a translation in axial direction of the forks  67 ,  68 . 
     Each one of the two forks  67 ,  68  is connected to a selector element  65 ,  66 , one per each secondary shaft of the gearbox which in turn is keyed to its own shaft by means of a grooved profile  131 ,  132 . The adoption of a coupling with grooved profile allows to transmit the rotary motion and at the same time it allows the translation in axial direction of the selector element. 
     Each selector element on each face is equipped with projections, in particular four, suitably shaped to insert in corresponding recesses, suitably implemented on the toothed wheels assembled on the two secondary shafts of the gearbox, divided as follows: I and III speed on one shaft, II and IV on another one. 
     Each time, depending upon the selected speed, the selector element will move on one side or on another one. Upon each speed shift, both selector elements will move by engaging or disengaging the responsible gearwheel. 
     For example, in the speed passage from the I st  to the II nd  ratio, the selector element  65  placed on the first one of the two secondary shafts of the gearbox will move from the engagement position to the neutral one, at the same time the selector element  66  assembled on the second secondary shaft of the gearbox will move from the neutral position to the engagement position, by keying the toothed wheel  62  related to the II nd  speed to its own secondary shaft, that is the projections of the selector element will enter the recesses implemented on the toothed wheel of the II nd  speed. 
     Since, as said, the actuation of the selectors is contemporary and specular, it is possible to implement a desmodromic drum equipped with one single track capable of actuating all four speeds. All this to the advantage of the simplicity of the layout of the solution and of the implementation inexpensiveness. 
     It is to be noted that the coupling forks  67 ,  68  are identical between them and with symmetrical sides, they are rotated one with respect the other one by 180°, with a greater construction simplicity. Even the sliding couplings  65 ,  66  are equal therebetween. 
     The profile of the desmodromic track  19  is represented in  FIG. 14B : S 1  designates the representation of the track  19  from the point of view of the first coupling fork  67  acting on the first secondary shaft  52 , and S 2  designates the representation of the track  19  from the point of view of the second coupling fork  68  acting on the second secondary shaft  53 . 
     C 1  and C 2 , respectively, designate the cams&#39; profiles which will be described hereinafter with greater details, controlling the clutch lever  47  and the control lever  20 , respectively. 
       1   a ,  2   a ,  3   a  and  4   a  designate the speed engagement from the first one to the fourth one, F designates an idle status, wherein the transmission of the motion from the driven pulley  17  to the primary shaft  51  does not occur, through the synchronizing device  40  ( FIG. 14B ). 
     The track paths S 1  and S 2  are formed, in the present embodiment example, from one single peripheral track  19  which is divided into fourth tracts, each one with a width of 90°. 
     It then comprises two central opposite tracts, which follow a neutral periphery, and two opposite tracts staggered therebetween and with respect to the two central tracts, still with a peripheral course. Such tracts are connected therebetween by respective ramps. 
     In particular, each ramp comprises an ascending tract, a linear tract extending from a maximum point of the ascending tract, and a descending tract extending from the linear tract, wherein the ascending tract, the linear tract and the descending tract define a substantially trapezoidal profile. 
     From the tracks S 1  and S 2  the translations of the sliding couplings  65 ,  66  with respect to the respective secondary shafts  52 ,  53 , determining the gearwheel engagement, are deduced. The engagement of each speed is alternated by an idle status. 
     By referring to  FIG. 14B , the first sliding coupling  65  and the respective first coupling fork  67  are translated axially when the corresponding cam follower end  69  moves in the staggered tracts of the track  19  which moves it in the direction of the first and the third toothed wheel  61 ,  63 . On the contrary, when this cam follower end  69  is in the central tracts the first secondary shaft  52  does not transmit the motion. 
     Analogously, when the second sliding coupling  66  and the respective second coupling fork  68  are translated axially, the corresponding cam follower end  69  moves in the staggered tracts of the track  19  which moves it in the direction of the second and the fourth toothed wheel  62 ,  64 . When this cam follower end  69 , instead, is in the central tracts, the second secondary shaft  53  does not transmit the motion. 
     In this example, the cam follower ends  69  of the coupling forks  67 ,  68  are spaced apart by an arc of 90° on the desmodromic drum. 
     It is to be noted that the hub shaft  75 , the two secondary shafts  52 ,  53  and the primary shaft  51  have axes parallel therebetween, grouped at the rear wheel  105 . 
     Even the rotation axis of the desmodromic drum  70  is parallel to the axes of the previously mentioned shafts. 
     As it will be described in more details hereinafter it is actuated by an actuator  80  which will be described hereinafter. 
     The used gearbox scheme provides some possible variants which will be described by referring to figure (see  FIG. 14A ). 
     Scheme A: four ratios with constant delta revolution ratio scale. It is the simplest and most compact solution: it provides two pairs of identical gearwheels between first secondary shaft and second secondary shaft, with two sliding couplings and coupling forks identical therebetween and one single track for defining the speeds on the desmodromic drum. 
     Scheme B: it is the solution shown in relation to the herein described embodiment example, it provides four ratios with progressive delta revolution ratio scale. The solution provides an identical pair of (first and second) gearwheels between first secondary shaft and second secondary shaft; two sliding couplings and two coupling forks identical therebetween and one single track for defining the speeds on the desmodromic drum. 
     Scheme C: solution with four ratios with constant delta revolutions and double clutch: this possible variant provides the use of a double clutch for shifting speed, which can be useful for passing between speeds without torque holes from one gearwheel to the other one. It provides due sliding couplings and two coupling forks identical therebetween and two distinct tracks implemented on the cylindrical surface of one single desmodromic drum. 
     Schema D: solution with six progressive delta revolution ratios. This variant provides the adoption of six speeds. The same scheme can be proposed with constant or progressive delta revolution ratios. 
     As it clearly results from the scheme of actuations, due to the effect of the previously described geometry the two tracks of the secondary shafts (S 1  and S 2 ) result to be identical but staggered by 90°, this thanks to the used gearbox scheme. Therefore, by positioning the two coupling forks  67 ,  68  on a desmodromic track  19  staggered by 90° of the desmodromic drum  70 , the possibility is obtained of having the two coupling forks  67 ,  68  equal therebetween and one single track on the desmodromic drum  70 , with a higher constructive convenience. 
     The electromechanical actuator  80  has the purpose of defining, for each gearwheel shifting procedure, the opening of the rear clutch by means of the dedicated clutch lever  47 , the movement of the two coupling forks  67 ,  68 , by disengaging the ongoing gearwheel and by engaging the subsequent or the previous one, the re-closing of the clutch  40 . Moreover, the actuator  80  is arranged to actuate the control lever  20  of the front centrifugal clutch  5  in the first speed. In this way, by using one single rotating electric engine, all these procedures are synchronized. 
     The electromechanical actuator  80  comprises a rotating electric motor  81 , suitably fed by means of a control unit in order to make the motor axis to rotate according to both rotation directions. It is to be noted that the rotation axis of the electric motor is perpendicular to the axes of the primary  51 , secondary  52  and  53  and hub  75  shafts. 
     On the rotative output of the electric motor  81  a pair of gearwheels  82 ,  83  is provided for reducing the transmission ratio outgoing from the motor, gearwheels with parallel axes, controlling a first actuator shaft  84  by means of an engagement of irreversible type allowing a greater precision and less influence of the clearances. The opposite ends of the actuator shaft  84  are supported by first actuator bearings  95 . The axis of the first actuator shaft  84 , too, is perpendicular to the axes of the primary  51 , secondary  52  and  53  and hub  75  shafts, and this allows a reduction in the overall dimensions. 
     The first actuator shaft  84  is engaged to an actuator pinion  96  which controls, with a suitable reduction ratio, a second actuator shaft  85  perpendicular to the previous one and then it is parallel to the axes of the primary  51 , secondary  52  and  53  and hub  75  shafts. 
     It extends on both sides of the actuator pinion  96  to control both the previously described desmodromic drum  70  and a cam set which actuates the clutch  47  and control  20  levers, with a pair of second actuator bearings  97  arranged on the side of the cam set. 
     The desmodromic drum  70  is on the side of the transmission  1  corresponding to the internal combustion engine and to the rear wheel; said cam system, together with said levers  20 ,  47 , is on the side of the transmission  1  covered by the cover  110 , wherein there is the synchronizer device  40  too. 
     The desmodromic drum  70  is controlled by a first actuator toothed wheel  98  which is keyed directly on the second actuator shaft  85 ; it is engaged to a second actuator toothed wheel  58  positioned between the actuator  80  and the gearbox  50 , which controls directly in rotation a third actuator shaft  59  fastened to the base of the desmodromic drum  80  which thus is suitably rotated. 
     In the present example, the transmission ratio between the second and the third actuator shaft  85 ,  59  is 1:1, thus a rotation angle of 90° of the desmodromic drum  70 , and then then a gearwheel shifting ( FIG. 14B ), corresponds to a rotation angle of the first (or the second) actuator wheel  98  of 90°. This in case of a four-speed gearbox. 
     Therefore, the engagement of a precise gearwheel corresponds to each position of the first actuator toothed wheel  98  staggered by 90°. To this regard, then, it is possible to provide a feedback signal indicating the engaged gearwheel, determined by the rotation of the actuator  80 . 
     Therefore, the first actuator toothed wheel  98  comprises a plurality of magnets  119  N and S, in particular four magnets (two per polarity) arranged alternated and spaced apart on a single periphery of an arc of 90°. 
     It is to be meant that, in a three-speed solution, three magnets could be sufficient. The magnets  119  N and S are arranged on the side of the wheel  98  wherein it is connected to the second actuator shaft  85 . 
     On this side, inside an actuator casing  99  extending for the whole extension of the second actuator shaft  85 , there is a detection card  120  which comprises a pair of Hall sensors  121 , arranged on a periphery corresponding to that of the magnets  119 , and separated by an arc of 90°. 
     The actuator casing  99  ( FIG. 18 ) is integral to the container of the transmission  109 , as well as the card  120 , which is connected to a control unit, which receives said feedback signal, by means of a connector  122  ( FIG. 22 ). The card  120  comprises even other chips which carry out other functions assigned thereto. 
     The Hall sensors  121  are capable of detecting the polarities of the magnets of the first actuator toothed wheel  85 , as each one thereof produces a peak signal with different polarity according to the polarity of the magnet  119  which passes nearby. By translating the signal which corresponds to N with  0  and the signal which corresponds to S with  1  (or viceversa), the pair of sensors  121  provides a binary signal according to the table of  FIG. 22  (N-N; N-S; S-S, S-N) which as a whole can assume four distinct values, each one thereof will correspond to a speed. 
     In this way, which is wholly passive and it depends only upon the rotation of the cam actuator of a synchronizer, it is possible generating a signal representing the engaged gearwheel, which could be used for any purpose, in particular it could provide an indication of the really engaged gearwheel to one or more control units. 
     On the other side of the actuator pinion  84 , on the end of the second actuator shaft  85  there is a cam set  86  which comprises a first cam  87  formed on the periphery of a first cam disc  88  arranged adjacent to the actuator pinion  84 . 
     The first cam  87  is fixed bot radially and axially, that is it is immobile with respect to the actuator casing  99  thereto it is constrained. 
     The cam profile of such first cam  87  has four peaks and four valleys, each one spaced apart by 90°, the valleys corresponding to a respective speed from the first to the fourth, thereto the magnets  119  of the first actuator toothed wheel  98  will equally correspond; the peaks, as it will be seen shortly, instead, correspond to the idle positions F ( FIG. 14B ). 
     The cam set  86  further comprises a cam follower  89  having a cam follower bushing  123  put on the top of the second actuator shaft  85 , which is constrained thereto from the rotational point of view, but it is axially mobile therealong thanks to one or more not represented axial ribs, which form a prismatic pair. 
     The cam follower  89  further comprises a second cam disc  126  having two opposite faces, one faced towards the first cam  87  and having a cam profile analogous to the one of the first cam  87 , but it is specular, that is it has four peaks and four valleys each one spaced apart by 90°, the valleys corresponding to a respective speed from the first one to the fourth one, thereto the magnets  119  of the first actuator toothed wheel  98  will equally correspond; the peaks, as it will be seen shortly, instead correspond to the idle positions F ( FIG. 14B ). 
     Therefore, the cam follower  89  moves away from the actuator pinion  84  when it is necessary to obtain an idle status F by pressing at the same time the second actuation button  48  of the input clutch  40 , to interrupt the transmission of the motion from the driven pulley  17  to the primary shaft  51 , and by acting on the sliding couplings  65 ,  66  of the gearbox  50  through the rotation of the desmodromic drum  70 . 
     Analogously, the cam follower  89  approaches to the actuator pinion  84  when any gearwheel is engaged, that is when the second actuation button  48  has not to be pressed to allow the transmission of the motion from the driven pulley  17  to the primary shaft  51 . 
     In order to obtain this approaching or return it will be necessary to arrange a return mechanism of conventional type, for example on the clutch lever  47 . 
     In order to obtain said pressure on the second actuation button  48 , on the face of the second cam disc  126  opposite to the profile of the cam follower it has an actuating projection  127  which represents an extension of the second actuator shaft  85  but which is mobile with alternated motion in response to the interaction between first cam  87  and cam follower  89 . 
     The actuating projection  127  acts directly on an actuation end  125  of the clutch lever  47 , opposite to the second pressing end  28 , by obtaining the oscillation of the clutch lever  47  upon each gearwheel shifting, to determine the idle status F and the gearwheel shifting driven through the desmodromic drum  70 . 
     This oscillation is represented by the track C 1  in  FIG. 14B . 
     Moreover, on the face of the second cam disc  126  opposite to the profile of the cam follower the second cam disc  126  has an additional cam profile determining a second cam  128  on such face. Such profile has a projection corresponding to the engagement of the first speed and it acts on an actuation end  124  of the control lever  20  opposite to the first pressing end  24 , by obtaining the oscillation of the control lever  20  which then could act on the first actuation button  16 , acting as actuation button too, which allows the effective detachment of the centrifugal clutch  5  on the crankshaft  2 , but only in the first speed, as previously described. 
     This oscillation is represented by the track C 2  in  FIG. 14B . 
     To the above-described synchronous transmission a person skilled in the art, in order to satisfy additional and contingent needs, could introduce several additional modifications and variants, all however comprised within the protection scope of the present invention, as defined by the enclosed claims.