Abstract:
A strut-type synchronizer for coupling a gear to a rotating shaft includes a hub rotating with the shaft, a clutch ring rotatable with the gear, and a friction ring disposed between the hub and the clutch ring. The hub supports several struts for relative axial movement, as urged by an encircling splined sleeve through a detent, into engagement with the friction ring. The friction ring, which does not itself directly engage the sleeve, has a friction surface that is thus urged by the struts into engagement with a complementary friction surface of the clutch ring, whereupon relative rotation between the friction ring and the clutch ring ceases, the detent is overcome, and the sleeve&#39;s spline teeth engage complementary spline teeth on the clutch ring to fully engage the synchronizer. The friction surfaces include two axially-spaced frustoconical portions, by which an increased cone torque is achieved during synchronization.

Description:
FIELD OF THE INVENTION 
     The invention relates generally to parallel-shaft transmissions in which a synchronizer is employed to match the rotational speeds of a driving hub and an adjacent driven gear and subsequently mechanically couple the gear to the hub using an axially-slidable synchronizer sleeve. 
     BACKGROUND OF THE INVENTION 
     Typical strut-type synchronizers for manual automotive transmissions include a “shifting sleeve carrier” or “hub” that rotates with a given shaft either adjacent to a single gear, or in between a pair of gears, that are respectively mounted on the shaft for free-rotation relative to the shaft. The hub includes a set of peripheral spline teeth and several peripheral axial slots or pockets, each pocket supporting a respective strut for axial movement relative to the pocket. An annular shifting sleeve, encircling and in splined engagement with the hub, engages each strut through a spring-loaded detent that typically includes a circumferential detent groove defined on the inner periphery of the sleeve. 
     The detent is generally designed to maximally transmit, to the strut, a relatively small fraction of an axial shifting load that is applied to the sleeve through a suitable combination of detent spring rate, the “ramp angle” defined by the detent groove&#39;s walls or “ramps,” as measured relative to a reference line that is parallel to the rotational axis of the shaft, and the coefficient of friction between the detent and the sleeve (and, in the case of a ball-type detent, the coefficient of friction between the detent ball and the struts&#39; ball-receiving passage). The maximum axial detent load on the strut is generally known as the synchronizer&#39;s “breakthrough load” or “BTL.” Through known approaches, transmissions intended to provide relatively “smooth” shifts typically employ synchronizer detents that achieve breakthrough loads in the range of about 40 N to about 50 N, or perhaps as little as about 5 percent of the typical 400-800 N applied axial shifting load (often expressed as a ratio of the applied axial shifting load F to the breakthrough load BTL, such known ratios are typically well in excess of 10:1, and perhaps even 20:1 or more), using ramp angles in the range of about 30 to 40 degrees. Even when “firm” shift points are desired, known approaches maximally limit the breakthrough load to no more than about 80-100 N, or roughly fifteen percent of the applied load (i.e, an F-to-BTL ratio of at least about 6.6:1), using maximum ramp angles of up to about 45 degrees. 
     To rotationally couple one gear to the shaft, a fork applies the axial shifting load to the sleeve to thereby move the sleeve axially relative to the hub toward the gear. The sleeve then operates through the detent to apply a small fraction of the applied axial shifting load F (up to the detent&#39;s breakthrough load BTL) to each strut, thereby moving each strut axially toward the gear. The struts, in turn, axially engage a respective baulking or blocking ring, disposed between the hub and the gear and rotatable with the hub, toward a clutch ring that is itself rotationally coupled to the gear. Ultimately, the struts cooperatively urge a conical friction surface on the blocking ring into engagement with a complementary friction surface or “cone” defined on the clutch ring. The resulting frictional engagement between the blocking ring and the clutch ring generates a cone torque that rotationally accelerates or decelerates the clutch ring and its coupled gear relative to the blocking ring and, hence, reduce the rotational speed differential between the hub and the gear. 
     Ultimately, the blocking ring frictionally bears against the clutch ring sufficiently to stop relative rotation between the sleeve and the gear. The continued application of the axial shifting load to the sleeve then overcomes the detent, and the sleeve moves farther toward the gear as the unloaded strut likewise ceases to axially bear against the blocking ring, with leading-edge chamfers on the sleeve&#39;s spline teeth engaging opposing entrance-edge chamfers on a further set of peripheral teeth defined on the blocking ring. After the mating chamfers of the sleeve and the clutch ring cooperate to “clock” the sleeve&#39;s spline teeth into angular registration with the blocking ring&#39;s peripheral teeth, the sleeve moves even farther and its leading-edge chamfers engage opposing entrance-edge chamfers on a set of peripheral teeth defined on the clutch ring. After this second set of mating chamfers cooperate to clock the sleeve&#39;s spline into angular registration with the clutch ring&#39;s peripheral teeth, the sleeve moves into “full engagement” with the clutch ring to thereby rotationally “lock up” the selected gear with the shaft. 
     The foregoing known strut-type synchronizers typical achieve lockup synchronization times in the range of perhaps about 200-300 msec for a typical manual transmission, from the initial movement of the sleeve from its neutral position about the hub, to the point at which the opposed chamfers of the sleeve spline and the clutch ring first begin to mesh (it will be appreciated that an additional time of perhaps about 100 msec is still required to axially displace the sleeve into “full engagement” with the clutch ring, as described above). 
     In another known strut-type synchronizer, as disclosed in U.S. Pat. No. 5,085,303, a pair of radially-nested friction surfaces or “cones,” defined on either side of a “middle cone ring” rotating with the gear and respectively engaging complementary friction surfaces on the blocking ring and an “inner cone ring” rotating with the synchronizer&#39;s mainshaft, cooperate in response to the axial movement of the sleeve and struts to generate a relatively-increased cone torque, even when using a relatively-reduced applied axial shifting load or a relatively-reduced detent breakthrough load. Unfortunately, such dual cone synchronizers necessarily feature both an increased parts count, including discrete parts for the blocking ring, the two cone rings, and the clutching ring, and a relatively-increased overall axial synchronizer dimension, in order to achieve the relative increase in generated cone torque. Further, such “dual-cone” synchronizers are susceptible to torque-generation losses as the multiple cone rings link together, and due to tolerance stack-up, such that the best estimate for the resulting cone torque is the square root of the sum of the squares of the cone torque generated due to engagement of each individual cone. 
     SUMMARY OF THE INVENTION 
     In accordance with the invention, a strut-type synchronizer for rotatably coupling a gear to a shaft includes an annular hub rotatably coupled to the shaft, and an annular clutch ring that is axially-spaced from the hub and rotationally coupled to the gear. The hub includes a peripheral surface in which both a set of external spline teeth and several circumferentially-spaced pockets are defined. The clutch ring likewise includes a peripheral surface on which a set of external spline teeth are defined, as well as a conical radially-outer friction surface projecting toward a first axial face of the hub. The synchronizer further includes an annular friction ring encircling the shaft and disposed between the hub and the clutch ring, wherein friction ring includes an axial face in opposition to the hub&#39;s axial face, and a radially-inner friction surface complementary to and in general opposition with the conical friction surface of the clutch ring. Several struts, each disposed in a respective pocket of the hub so as to be axially movable relative to the hub into engagement with the friction ring&#39;s axial face, rotationally couple the friction ring to the hub. 
     The synchronizer also includes an annular sleeve encircling the hub, wherein the sleeve includes a set of internal spline teeth engaging the external spline teeth of the hub, such that the sleeve rotates with the hub, and further adapted to engage the external spline teeth of the clutch ring when the sleeve is shifted axially relative to the hub, for example, under the control of a fork that engages suitable surface features on a radially outer surface of the sleeve, to thereby couple the hub to the clutch ring. The synchronizer further includes a detent operative to couple the sleeve to each of the struts for axial movement relative to the hub upon application of an axial shifting load to the sleeve. 
     When the applied axial shifting load exceeds the detent&#39;s breakthrough load, the sleeve decouples from the strut and continues to move axially until the sleeve&#39;s internal spline teeth engage the external spline teeth of the clutch ring. While the invention contemplates use of any suitable detent configuration, by way of example only, in an exemplary embodiment, a radial passage is defined in the strut, and a detent ball is disposed in the strut passage so as to be positioned proximate to a circumferential detent groove defined in the radially inner surface of the sleeve. A detent spring is disposed in each pocket of the hub such that a radially-inner end of the spring is supported by a radial surface of the hub, and a radially outer end of the spring extends through the radial passage defined in the strut to bias the detent ball into engagement with the sleeve&#39;s detent groove. 
     In accordance with an aspect of the invention, the detent parameters, such as the detent spring rate, the coefficient of friction between the detent ball and the sleeve, and the ramp angle defined by the detent groove&#39;s ramps, are selected to achieve a breakthrough load significantly greater than about 100 N. Preferably, the detent&#39;s breakthrough load is greater than about 125 N, i.e., the detent achieves an F-to-BTL ratio of no greater than about 4.8:1, for example, when using an average applied axial shifting load F of about 600 N. Most preferably, the detent&#39;s breakthrough load is greater than about 150 N, i.e., the detent achieves an F-to-BTL ratio of no greater than about 4:1 when using an average applied axial shifting load F of about 600 N. Indeed, by way of further example only, in an exemplary embodiment, the detent achieves a breakthrough load BTL of at least about 200 N in response to an average applied axial shifting load of about 600 N, or an F-to-BTL ratio of roughly 3:1. 
     Given the known practical constraints on detent spring selection, as well as the desirability of maintaining a relatively short axial distance between the point at which the detent is first loaded and the point at which the mating chamfers engage, the exemplary embodiment advantageously employs relatively-higher ramp angles of at least 45 degrees and, preferably, a ramp angle greater than about 50 degrees, as measured relative to a tangential reference plane, to increase the synchronizer&#39;s breakthrough load and, hence, effectively “slow-down” its nominal synchronization time, without having to greatly increase the detent&#39;s nominal spring rate. 
     In accordance with another aspect of the invention, the internal spline teeth of the sleeve have a minimum crest radius, and the friction ring has a maximum radial dimension that is less than the minimum crest radius of the internal spline teeth of the sleeve. In this manner, the internal spline teeth of the sleeve are radially spaced from the friction ring when the sleeve encircles the friction ring. It will be appreciated that, in contrast with known blocking rings which operate to slow the axial movement of the sleeve as the sleeve-to-blocking ring engagement generates the required index torque to “clock” the blocking ring&#39;s external spline teeth into registration with those of the clutch ring, the absence of direct contact between the sleeve and the friction ring ensures relatively rapid axial movement of the sleeve once the detent&#39;s breakthrough load has been overcome. 
     In accordance with another aspect of the invention, the conical friction surface of the clutch ring preferably defines a pair of axially-spaced frustoconical portions that are each rotatable with the gear about the shaft axis, with the “smaller” cone section projecting closer to the hub&#39;s axial face, and having a lesser maximum diameter than the minimum diameter of the “larger” cone, to thereby define a synchronizer with a “stepped” cone when viewed in longitudinal cross-section. The friction ring likewise features a “stepped” complementary friction surface, for engagement with the stepped cone sections on the clutch ring. The total cone torque generated by the stepped cone configuration of the exemplary embodiment advantageously equals the sum of the cone torques generated by each cone&#39;s individual engagement with its complementary friction surface of the friction ring. In this manner, the stepped cone configuration of the invention advantageously provides a significant increase in generated cone torque, without increasing the synchronizer&#39;s parts count. Indeed, in one embodiment, a stepped cone synchronizer in accordance with the invention features a stepped-cone clutch ring that is integrally formed with its respective gear, thereby further beneficially reducing synchronizer parts count and tolerance stack-up, as well as a further-relatively-reduced overall axial synchronizer dimension. 
     Preferably, the annular transition on the clutch ring between the smaller and larger frustoconical portions, as well as the complementary annular transition defined on the friction ring, is slightly canted relative to a reference plane orthogonal to the shaft&#39;s rotational axis, to thereby reduce the likelihood of a collection of oil at these transition sections upon engagement of the friction ring with the clutch ring. While the invention contemplates any suitable angle of inclination of these transitions, in a preferred embodiment, the angle of inclination is preferably greater than about 1 degrees and less than about 4 degrees relative to the orthogonal reference plane. Most preferably, the angle of inclination of each transition is about 2 degrees relative to the orthogonal reference plane. 
     Other objects, features, and advantages of the present invention will be readily appreciated upon a review of the subsequent description of the preferred embodiment and the appended claims, taken in conjunction with the accompanying Drawings. 
    
    
     
       BRIEF DESCRIPTION OF THE DRAWINGS 
         FIG. 1  is a partial longitudinal section of a parallel-shaft automotive transmission incorporating an exemplary stepped-cone synchronizer featuring a high breakthrough load in accordance with several aspects of the invention; 
         FIG. 2  is a partial transverse section of the synchronizer of  FIG. 1 , taken along Line  2 - 2  thereof; 
         FIG. 3  is a partial exploded view of the synchronizer of  FIG. 1 , showing only the mainshaft, the left gear and left clutch ring, the left friction ring, and the hub-strut-sleeve assembly for clarity of illustration; 
         FIG. 4  is an enlarged partal section of the synchronizer of  FIG. 1  showing the projecting stepped-cone of the clutch ring and complementary friction surfaces of the friction ring, as well as the high-breakthrough-load detent; 
         FIG. 5  shows several plots of detent ramp angle θ to the developed breakthrough load BTL, for three different nominal values for the detent spring force; 
         FIG. 6  shows several plots of detent ramp angle θ to the ratio of the applied axial shifting load F to the developed breakthrough load BTL, for the same three different nominal values for the detent spring force; and 
         FIG. 7  shows a partial longitudinal section of an alternative synchronizer in accordance with the invention, in which the clutch ring associated with one of the gears is integrally formed with the gear. 
     
    
    
     DETAILED DESCRIPTION OF THE INVENTION 
     Referring to  FIGS. 1 and 2 , an exemplary synchronizer  10  according to the invention, and incorporated within a parallel-shaft automotive transmission  12 , is shown as being disposed between left and right transmission gears  14 , 16  that are respectively mounted as with suitable needle bearings for free rotation about a mainshaft  18 . The synchronizer  10  includes a hub  20  splined to the mainshaft  18  so as to rotate with the mainshaft  18 . As best seen in  FIG. 2 , the outer periphery  22  of the hub  20  defines a set of external spline teeth  24  and a plurality of circumferentially-spaced pockets  26  (only one of which is illustrated in the sectional view that is  FIG. 2 ). Each pocket  26 , which includes a radially-inner base surface  28 , is adapted to support a respective strut  30  for relative axial movement, as further described below. 
     Referring to  FIG. 1 , an annular left clutch ring  32  is welded to the left gear  14 , for rotation with the left gear  14  about the axis  34  of the mainshaft  18  opposite a left axial face  36  of the hub  20 . Similarly, an annular right clutch ring  38  is welded to the right gear  16 , for rotation with the right gear  16  about the mainshaft axis  34  opposite a right axial face  40  of the hub  20 . 
     As seen in  FIG. 1  and the partial exploded view shown in  FIG. 3 , each clutch ring  32 , 38  includes a peripheral surface defining another set of external spline teeth  42 , and a radially-outer frustoconical friction surface (“external friction surface  44 ”) projecting axially toward the left and right face  36 , 40  of the hub  20 , respectively. A pair of friction rings  46 , 48  encircle the mainshaft  18  between the hub  20  and the left clutch ring  32 , and between the hub  20  and the right clutch ring  38 , respectively. Each friction ring  46 , 48  includes an axial face  50 , 52  opposing the respective left and right faces  36 , 40  of the hub  20 . Each friction ring  46 , 48  further includes a radially-inner frustoconical friction surface (“internal friction surface  54 ”) complementary to the respective external friction surfaces  44  of the left and right clutch rings  32 , 38 . 
     As best seen in  FIGS. 3 and 4  in the context of the left clutch ring  32  and left friction ring  46 , in accordance with an aspect of the invention, the external friction surface  44  of each clutch ring  32 , 38  beneficially includes a first frustoconical portion  56  having a maximum radius R 1max , and a second frustoconical portion  58 , axially spaced from the first frustoconical portion  56 , that has a minimum radius R 2min  greater than the maximum radius R 1max  of the first frustoconical portion  56 . The first and second frustoconical portions  56 , 58  of each clutch ring&#39;s external friction surface  44  are themselves separated by a frustoconical transition  60 , to thereby define a “stepped cone” when viewed in longitudinal cross-section, as seen in  FIGS. 1 ,  3 , and  4 . The first and second frustoconical portions  56 , 58  of each clutch ring&#39;s external friction surface  44  are provided with the same nominal angle of inclination, or “cone angle α,” typically ranging between about 6 degrees and about 8.5 degrees relative to the axis  34  of the mainshaft  18 . The frustoconical transition preferably has an angle of inclination β of at least about 1 degrees relative to a radial reference plane, i.e., a reference plane  62  that is normal to the mainshaft axis. Preferably, the angle of inclination of the frustoconical transition  60  is no greater than about 4 degrees. Most preferably, the angle of inclination of the frustoconical transition  60  is about 2 degrees. 
     And, as best seen in  FIG. 3 , the complementary internal friction surface  54  of each friction ring  46 , 48  likewise includes two axially-spaced frustoconical portions  64 , 66  separated by a frustoconical transition  68 . Preferably, the angle of inclination of the friction ring&#39;s complementary frustoconical transition  68  is roughly equal to that of the clutch ring&#39;s frustoconical transition  60 , to further reduce the likelihood of collecting lubricating oil adjacent to the friction ring&#39;s frustoconical transition  68  upon engagement of the left friction ring  46  with the left clutch ring  32 . 
     Referring again to  FIGS. 1 and 2 , the synchronizer  10  further includes a sleeve  70  encircling the hub  20  and axially movable relative to the hub  20  in response to an axial shifting load applied, for example, to a peripheral surface feature  72  of the sleeve  70  as by a suitable fork (not shown). The sleeve  70  includes a set of internal spline teeth  74  adapted to matingly engage the external spline teeth  24 , 42  of the hub  20  and of the clutch rings  32 , 38 , respectively. In accordance with an aspect of the invention, each friction ring  46 , 48  has a maximum radial dimension R FRmax  (as seen in  FIG. 1 ), and the sleeve&#39;s internal spline teeth  74  have a minimum crest radius R Smin  (as seen in  FIG. 2 ) that is greater than the maximum radial dimension R FRmax  of the friction rings  46 , 48 . In this manner, the sleeve&#39;s internal spline teeth  74  remain radially spaced from each friction ring  46 , 48  when the sleeve  70  encircles the friction ring  46 , 48 , thereby preventing any direct engagement between the sleeve  70  and the friction rings  46 , 48 . 
     While the invention contemplates use of any suitable materials for the friction rings  46 , 48 , including a sintered bronze material, a further advantage of preventing any direct engagement between the sleeve  70  and the friction rings  46 , 48 , i.e., the absence of any external spline teeth on either of the friction rings  46 , 48 , is that the invention contemplates a broader range of materials selection and manufacture for the friction rings  46 , 48 , including the bonding of a suitable friction material to a forged, cast, or powdered metal annular substrate. 
     As best seen in  FIGS. 1 ,  2 , and  4 , each strut  30  supported within its corresponding pocket  26  of the hub  20  includes a pair of end surfaces  76 , 78  adapted to engage the respective opposed axial face  50 , 52  of one of the friction rings  46 , 48 , whereby each friction ring  46 , 48  is rotatable with the hub  20 . A detent  80  operates to couple each strut  30  to the sleeve  70  for axial movement relative to the hub  20 . 
     While the detent  80  may be of any suitable construction, in the exemplary synchronizer  10 , a radial passage  82  is defined in each strut  30 , and a detent ball  84  is partially captured within the strut passage  82  so as to be positioned proximate to a circumferential detent groove  86  defined in a radially-inner surface of the sleeve  70 . The detent groove  86  defines a pair of opposed ramps  88  that are generally disposed at a ramp angle θ relative to the axis  34  of the mainshaft  18 . A detent spring  90  is disposed in each pocket  26  of the hub  20 , such that a radially-inner end of the spring  90  is supported by the base surface  28  of the hub  20  (as best seen in  FIGS. 1 and 2 ), and a radially outer end of the spring  90  extends through the strut&#39;s radial passage  82  to bias the detent ball  84  into engagement with the sleeve&#39;s detent groove  86  (as best seen in  FIG. 4 ). 
     Because the exemplary synchronizer  10  does not employ a blocking ring to otherwise slow the axial movement of the sleeve  70  towards one of the clutching rings  32 , 38  once the detent&#39;s breakthrough load BTL has been overcome, the parameters of the detent  80 , such as the detent spring rate, the coefficient of friction between the detent ball  84  and the sleeve  70 , and the ramp angle θ, are selected to achieve a breakthrough load significantly greater than about 100 N, in order to increase the time period required to initially overcome the detent after synchronization has commenced. Preferably, the detent parameters provide a breakthrough load greater than about 125 N. Most preferably, the detent  80  is designed to provide a breakthrough load greater than about 150 N. 
     Because of practical limitations on increasing the detent spring force above perhaps 40 N, including packaging constraints (within the strut&#39;s radial passage), manufacturability, and perhaps even a potential spring rate variability due to tolerance stack-ups, as well as the relatively fixed coefficient of friction between the detent ball  84  and the ramp  88 , the detent ramp angle θ and detent spring rates are preferably selected so as to operationally place the detent  80  within Region A of  FIG. 5  (wherein a relatively-lower spring rate of about 20 N is shown in phantom line, an intermediate spring rate of about 30 N is shown in broken line, and a relatively higher spring rate of about 40 N is shown in solid line). Thus, a breakthrough load BTL greater than about 150 N is conveniently provided through use of a detent groove ramp angle θ greater than about 45 degrees (to thereby lie in Region A of  FIG. 5 ). Most preferably, the detent spring rate is selected to utilize a ramp angle θ that is at least about 48 degrees, but no greater than about 57 degrees, i.e., to operationally place the detent  80  within Region B of  FIG. 5 . By way of comparison only, known synchronizers with blocking rings typically utilize detent ramp angles significantly less than 100 N, to thereby operationally place the detents of such known synchronizers with blocking rings within Region C of  FIG. 5 . 
     As a further preferred criterion, for a synchronizer  10  to be operated with an average applied shifting load F, the ratio of the average applied shifting load F to the detent&#39;s breakthrough load BTL is preferably less than about 4.8:1 and, most preferably, is less than about 4:1. Shown graphically in  FIG. 6 , the detent ramp angle θ and detent spring rate are preferably selected so as to operationally place the detent  80  within Region D of  FIG. 6  (wherein a relatively-lower spring rate of about 20 N is shown in phantom line, an intermediate spring rate of about 30 N is shown in broken line, and a relatively higher spring rate of about 40 N is shown in solid line), and most preferably to operationally place the detent  80  within Region E of  FIG. 6 . Once again, by way of comparison only, known synchronizers with blocking rings typically utilize detent ramp angles and spring rates which operationally place such detents within Region F of  FIG. 6 . Thus, in a constructed embodiment of the exemplary synchronizer  10 , where an average applied shifting load of about 600 N is generated using an electromagnetic actuator (not shown) in response to a shift demand signal, a desired breakthrough load of about 200 N, i.e., an F-to-BTL ratio of about 3:1, is conveniently achieved using a detent spring  90  having a spring rate of 40 N and a detent ramp angle θ of about 53 degrees. 
     Once the detent  80  is overcome by the applied axial shifting load F, the sleeve  70  moves axially towards one of the clutch rings  32 , 38  until leading-edge chamfers  92  defined on the sleeve&#39;s internal spline teeth  74  (as best seen in  FIG. 4 ) matingly engage corresponding leading-edge chamfers  94  defined on the external spline teeth  42  of one of the clutch rings  32 , 38 . The leading-edge chamfers  92  on the sleeve&#39;s spline teeth  74  are preferably disposed at a slightly greater angle than the corresponding leading-edge chamfers  94  on the clutch ring  32 , 38 , for example, perhaps 2-3 degrees greater, to facilitate engagement. It will be appreciated that the mass of the sleeve  70  and the distance between the respective chamfers  92 , 94  of the sleeve  70  and the clutch rings  32 , 38  upon detent release are also preferably chosen to thereby achieve a desired time, from initial detent loading to full engagement of the sleeve  70  with the clutch ring  32 , 38 , that is generally in the range of about 180 msec, for “firm” or “power” shifts, to about 300 msec, for “smooth” shifts, based upon the anticipated axial shifting load F that will be applied to the sleeve  70 . It will be appreciated that, by eliminating any sleeve engagement with either of the friction rings  46 , 48 , the invention advantageously allows for a substantial reduction in the time required to achieve full engagement once the detent  80  is overcome, for example, where “firm” shifts are desired, as when operating the transmission using electronically-controlled electromagnetic actuators operating in a “sport” program mode. 
       FIG. 7  shows an alternate embodiment  110  of a transmission synchronizer in accordance with several aspects of the invention, wherein the features of the right clutch ring, including its stepped friction surfaces  144 , are integrally formed with the left gear  116 . It will be appreciated that, where the gear teeth defined on a given gear are significantly taller than the clutch ring&#39;s external spline teeth  142 , as often is the case for numerically-higher-ratio gears  116 , the features of the clutch ring including its projecting external friction surface  144 , are advantageously formed directly on the gear  116  to thereby reduce parts count and potential tolerance stack-ups. 
     Thus, the invention advantageously achieves a synchronizer of reduced overall axial dimension, capable of providing a wider range of synchronization times, with a reduced parts count and a reduced tolerance stack-up. 
     While the above description constitutes the preferred embodiments, it will be appreciated that the invention is susceptible to modification, variation and change without departing from the proper scope and fair meaning of the subjoined claims. For example, while the disclosed embodiments both advantageously feature stepped cones on the clutch ring and the elimination of the index torque component of known strut-type synchronizers employing blocking rings, because the stepped cones achieve a greater cone torque in response to a given applied axial force, it will be appreciated that an advantage can yet be obtained in a conventional strut-type synchronizer when incorporating the stepped friction surfaces of the exemplary synchronizer  10 , for example, by permitting use of a reduced detent ramp angle or a detent spring with a reduced spring rate. Further, while the external and internal friction surfaces  44 , 54  of the two clutch rings  32 , 38  and friction rings  46 , 48  of the exemplary synchronizer  10  have an identical configuration, it will be appreciated that the invention contemplates use of different friction surface configurations on either side of the hub, as desired, such as a single frustoconical friction surface on one clutch ring/friction ring pair and a stepped cone configuration on the other clutch ring/friction ring pair.