WO2010034223A1 - 导向式牙嵌超越离合器 - Google Patents

导向式牙嵌超越离合器 Download PDF

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Publication number
WO2010034223A1
WO2010034223A1 PCT/CN2009/073774 CN2009073774W WO2010034223A1 WO 2010034223 A1 WO2010034223 A1 WO 2010034223A1 CN 2009073774 W CN2009073774 W CN 2009073774W WO 2010034223 A1 WO2010034223 A1 WO 2010034223A1
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WO
WIPO (PCT)
Prior art keywords
ring
overrunning clutch
fixed
introduction
rotation
Prior art date
Application number
PCT/CN2009/073774
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English (en)
French (fr)
Inventor
洪涛
Original Assignee
Hong Tao
Priority date (The priority date is an assumption and is not a legal conclusion. Google has not performed a legal analysis and makes no representation as to the accuracy of the date listed.)
Filing date
Publication date
Application filed by Hong Tao filed Critical Hong Tao
Publication of WO2010034223A1 publication Critical patent/WO2010034223A1/zh

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Classifications

    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F16—ENGINEERING ELEMENTS AND UNITS; GENERAL MEASURES FOR PRODUCING AND MAINTAINING EFFECTIVE FUNCTIONING OF MACHINES OR INSTALLATIONS; THERMAL INSULATION IN GENERAL
    • F16D—COUPLINGS FOR TRANSMITTING ROTATION; CLUTCHES; BRAKES
    • F16D41/00—Freewheels or freewheel clutches
    • F16D41/18—Freewheels or freewheel clutches with non-hinged detent
    • F16D41/185—Freewheels or freewheel clutches with non-hinged detent the engaging movement having an axial component
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F16—ENGINEERING ELEMENTS AND UNITS; GENERAL MEASURES FOR PRODUCING AND MAINTAINING EFFECTIVE FUNCTIONING OF MACHINES OR INSTALLATIONS; THERMAL INSULATION IN GENERAL
    • F16D—COUPLINGS FOR TRANSMITTING ROTATION; CLUTCHES; BRAKES
    • F16D41/00—Freewheels or freewheel clutches
    • F16D41/02—Freewheels or freewheel clutches disengaged by contact of a part of or on the freewheel or freewheel clutch with a stationarily-mounted member
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F16—ENGINEERING ELEMENTS AND UNITS; GENERAL MEASURES FOR PRODUCING AND MAINTAINING EFFECTIVE FUNCTIONING OF MACHINES OR INSTALLATIONS; THERMAL INSULATION IN GENERAL
    • F16D—COUPLINGS FOR TRANSMITTING ROTATION; CLUTCHES; BRAKES
    • F16D41/00—Freewheels or freewheel clutches
    • F16D41/22—Freewheels or freewheel clutches with clutching ring or disc axially shifted as a result of lost motion between actuating members

Definitions

  • the present invention relates to a clutch device in the field of mechanical transmission, and other transmission devices such as a dental differential including such a clutch device, and more particularly, but not exclusively, to a transmission torque and a rotation override function.
  • the jaw-mounted overrunning clutch Background technique
  • the friction type overrunning clutch has the disadvantages of low bearing capacity, poor reliability, low transmission efficiency, difficult processing and assembly, high cost, easy wear and small application range ("The development status of the overrunning clutch” And Trends, Zhang Jizheng, etc., The 3rd International Conference on Mechanical and Technological History of China and Japan, Kunming, 2002, 398 ⁇ 403). Embedded, although it has the potential to overcome the above-mentioned deficiencies, but because of the collision, collision noise and industry-specific knowledge beyond the rotation, in addition to the SSS (Synchro-Self-Shifting) synchronous clutch, this type of overrunning clutch is almost not obtained. Due attention or application, especially for the jaw type, the bearing capacity is huge, the radial dimension is relatively small, and the advantages of no relative slip after joining/fitting are not substantially utilized.
  • the main working principle of the SSS synchronizing clutch is that it relies on the action of the one-way pawl and the relative rotation of the helical spline tooth pair between the moving ring and the drive shaft, so that the force transmitting ring is both circumferentially oriented. Axial engagement/fitting and separation are achieved in relative motion (this is called so-called synchronous self-movement). Among them, the helical spline tooth pair has the dual functions of transmitting torque and turning guidance. With the friction-type overrunning clutch, the SSS Synchronous Clutch combines the advantages of high speed and high torque (maximum torque is equivalent to one million Nm of the friction overrunning clutch).
  • the manufacturing and assembly precision and process requirements are too high, the coaxiality requirements are too strict, the cost is too high, the moving ring axial movement amount is large, the axial structure, the radial size and the exclusive space of the core structure are large, and it is difficult to
  • the miniaturization has led to its narrow application range (mainly used in a few drive shaft systems such as ships and large generator sets), and its versatility is poor.
  • the spiral rising angle characteristic of the spline teeth causes the spline tooth surface frictional strength to be too high, and the effective utilization rate of the mechanical/mechanical potential of the material is significantly lower than 100%, thereby reducing the bearing capacity.
  • due to the structure and working principle its carrying capacity is not large enough, and it is impossible to realize the bidirectional bearing and the artificial controllability of the working condition. In practice, the jointing/mating or separation reliability is not high enough. Summary of the invention
  • the present invention is directed to solving the above problems.
  • the object of the present invention is to provide a fitting and a separation which are all completed by a rotating guiding mechanism, the separating and fitting trajectories are relatively fixed, the driving force of the force transmitting tooth layout and the guiding mechanism is diverse, and there is no sliding wear between the tooth tips and the bearing capacity.
  • the larger guided jaw overrunning clutch optionally has one-way, two-way overrunning clutch conditions, as well as controlled glider, clutch or coupling conditions.
  • a guided jaw overrunning clutch of the present invention includes an axially-fitted force-transmitting engagement mechanism for transmitting torque, having a retaining ring that rotates about the same axis and respectively forms a force-transmitting tooth and
  • the moving ring, the moving ring and the second rotating member outside the clutch constitute an external transmission mechanism capable of both axial movement and torque transmission; at least one rotary guiding type deriving mechanism that starts to overturn between the second rotating member and the fixed ring When the moving ring is axially moved away from the fixing ring to release the axial fitting of the force transmitting fitting mechanism; at least one rotational guiding type guiding mechanism that causes the movement to start when the second rotating member and the fixed ring start to reversely rotate The ring is axially moved toward the retaining ring to restore the axial fitting of the force transmitting fitting mechanism; at least one fixed point actuating mechanism is used in the anti-overturning process described above, when the moving ring is at a specific circumferential position relative to the
  • the angle between the engaging conical surface and the axis of rotation is greater than or equal to zero. And less than or equal to 180°, at least one of the guiding mechanism and the guiding mechanism activates the motive force by the relative rotation between the moving ring and the fixing ring, and the inlet margin when the force transmitting fitting mechanism is fitted is greater than zero.
  • Failure Refers to the meaning that the mechanism or part does not function properly due to human or objective reasons, and loses its basic function.
  • the pawl and the ratchet are axially misaligned with each other in a manner that destroys the basic relationship of the member, or the pawl is forcibly restrained in the separated position by the basic motion of the breaking member, and the ratchet mechanism will be The person loses the possibility of meshing and fails.
  • Rotating guide mechanism A mechanism that is driven by a relative rotation of the circumference to generate/get axial relative movement.
  • the utility model relates to a coaxial screw mechanism with a strict spiral angle and a different-axis external mesh helical gear mechanism, and a radial pin groove mechanism, an end face ratchet mechanism, an end face fitting mechanism and a cylindrical cam mechanism, etc. .
  • the introduction mechanism and the derivation mechanism are both rotation guide mechanisms.
  • Axial fitting The fitting or separation of both fittings in the fitting mechanism is accompanied by a fitting mode in which both axial directions are relatively moved; the fitting teeth can be formed on both end faces, tapered faces or cylindrical faces.
  • the rotational driving guide mechanism and the actuating motive force of the introduction mechanism exhaust all possibilities, which may be relative rotation between the moving ring and the second rotating member, or may be a moving ring and Relative rotation between the retaining rings.
  • the force transmitting teeth of the force-transmitting fitting mechanism are no longer uniquely cylindrically distributed, enriching the layout form of the transmission teeth, and actuating the selection mechanism, the commutation and the reversing drive mechanism make it possible to control the working conditions and is simple. Fast.
  • Figure 1 is a simplified axial cross-sectional view of a one-way overrunning clutch of package one in accordance with the present invention.
  • Figure 2 is a schematic view of the moving ring of Figure 1, (a) is an axial half-sectional view of the bottom view, and (b) is a front view.
  • 3 is a partial series development view of the radial projection of the three-dimensional force-engaging mechanism, the derivation mechanism, and the introduction mechanism of FIG.
  • FIG. 4 is a simplified axial cross-sectional view of a one-way overrunning clutch of package type two in accordance with the present invention.
  • Figure 5 is a schematic view of the pawl of Figure 4, (a) is a bottom view, (b) is a front view, and (c) is a plan view.
  • Figure 6 is an enlarged schematic view showing a partial structure of the pawl blocking ring of Figure 4;
  • Figure 7 is a schematic view of an alternative pawl of Figure 4, (a) being a bottom view, (b) being a front view, and (c) being a top view.
  • Figure 8 is an axial cross-sectional view of a two-way overrunning clutch in the form of a shaft-shaft transmission in accordance with the present invention.
  • Figure 9 is a schematic view of the moving ring of Figure 8, (a) is an axial half-section of the right side view, and (b) is a front view.
  • Figure 10 is a schematic view of the retaining ring of Figure 8, (a) is an axial half-sectional view of the front view, and (b) is an enlarged schematic view of the partial structure of the H-direction in (a).
  • Figure 11 is a schematic view of the pawl of Figure 8, (a) is a bottom view, (b) is a front view, and (c) is a plan view.
  • Figure 12 is a schematic view of the lead-out ring of Figure 8, (a) is an axial cross-sectional view of a simplified front view, and (b) is an enlarged schematic view of a partial structure of the left view.
  • Figure 13 is a schematic view of the introduction ring of Figure 8, (a) is an axial half cross-sectional view of the bottom view, and (b) is a front view.
  • Figure 14 is an enlarged schematic view showing a partial structure of the pawl blocking ring of Figure 8.
  • Figure 15 is a schematic view of the swinging rod of Figure 8, (a) being a front view and (b) being a plan view.
  • Figure 16 (a) is an enlarged schematic view showing a partial structure of the drive ring of Figure 8, and (b) is an alternative structure thereof.
  • Figure 17 is a partial series development view of the radial projection of the tooth profile of the force-transmitting fitting mechanism (1), the deriving mechanism (2), and the introduction mechanism (2) of Figure 8 on the same outer cylindrical surface, and a partial unfolded view of the corresponding imaginary axial projection profile of the fixed-point actuation mechanism (3), wherein (a) corresponds to the force transmission condition, (; b) corresponds to the overrun condition, and (C) corresponds to the fitted reset Process, the arrows in the figure indicate the direction of rotation of the moving ring relative to the fixed ring.
  • FIG. 18 is a reversing principle diagram showing the relationship between the tooth profiles of the force transmitting fitting mechanism and the deriving mechanism when the half tooth out-of-way reversing scheme is adopted in the two-way overrunning clutch, (a) the translating device corresponding to the direction one (b) corresponds to the overrun condition of direction two, and the arrow in the figure indicates the direction of the overrun.
  • Fig. 19 is a schematic view of the swing lever of Fig. 20, (a) being a front view and (b) being a plan view.
  • Figure 20 is an axial cross-sectional view of a two-way overrunning clutch of package type two in accordance with the present invention.
  • Figure 21 is a schematic view of the moving ring of Figure 20, (a) is a simplified axial sectional view of the right side view, and (b) is a simplified front view.
  • Figure 22 is a schematic view of the retaining ring of Figure 20, (a) is an axial half-sectional view of the front view, (b) is an enlarged schematic view of the local structure in the H direction in (a), and (c) is the TT cross section in (b) Partial enlarged view.
  • Figure 23 is a partial series development view of the radial projection of the tooth profile of the force-transmitting fitting mechanism (1), the deriving mechanism (2), and the introduction mechanism (2) of Figure 20 on the same outer cylindrical surface, and a partial development view (3) of the corresponding imaginary axial projection profile of the fixed-point actuation mechanism at the axial section of the pawl, wherein (a) corresponds to the force transmission condition, (b) corresponds to the overrun condition, (c Corresponding to the fitting reset process, (d) corresponds to a schematic diagram of the force transmission condition after the optional introduction ring is inserted; the arrow in the figure indicates the direction of rotation of the moving ring relative to the fixed ring.
  • Figure 24 is a schematic view of the optional member introduction ring of Figure 20, (a) is a front view, and (b) is an axial half cross-sectional view of the left side view. detailed description
  • Embodiment 1 Wheel-shaft-driven one-way overrunning clutch with package form one C1
  • the retaining ring 50 of the overrunning clutch C1 and the support shell 230 are integrally fastened by a screw 218, and the two are axially fixed to the second rotating member 208 by means of two bearings 224, so as to transmit the force fitting mechanism M1,
  • the lead-out mechanism M2, the introduction mechanism M3 and the fixed-point actuation mechanism M4 are packaged and can be rotated about the axis 200 individually or integrally.
  • the force transmitting fitting mechanism M1 includes a fixed ring 50 and a moving ring 70 which are axially oppositely fitted, as shown in FIGS. 1 to 3.
  • the circumferentially evenly distributed force transmitting teeth 72 of the moving ring 70 preferably have a force transmitted parallel to the axis 200.
  • the flank 74 the other flank of which is an associated leading flank 104 that is inclined relative to the axis 200.
  • the root of the force transmitting tooth side 74 can also be retracted or convex (but not beneficial for fabrication, torque transfer and meshing stability).
  • the structure and the number of teeth of the fitting end face on the retaining ring 50 are exactly equal to the moving ring 70, and the force transmitting teeth 52 correspondingly have a force transmitting flank 54 and a leading flank 94. That is, the force transmitting teeth 52 are conformed with the lead-out teeth 92, the force transmitting teeth 72 are associated with the auxiliary lead-out teeth 102, and each of the circumferentially-occupied tooth bodies corresponds to a different circumferential direction, and the cross-section is generally serrated. . Therefore, both of the fitting members constitute the one-way force-extracting mechanism M1 and also constitute the one-way outlet mechanism M2, and correspond to different circumferential directions. This is most clearly shown in Figure 3. The width of the slot of the fixed ring 50 and the moving ring 70 is significantly larger than the width of the tip of the other end, and the difference between the two is that the inlet margin is significantly greater than zero.
  • the concept of engaging the conical surface is similar to the pitch cone in the straight bevel gear.
  • the introduction mechanism ⁇ 3 includes a pin-type pawl 110 and a helical-type auxiliary introduction tooth 162 which are engageable with each other, and the former is fitted to the fixed ring in a circumferentially and axially fixed manner together with the return spring 114.
  • the latter is circumferentially evenly distributed on the outer cylindrical surface of the sub-introduction ring 160, which is integrally formed with the splined base ring 76 of the moving ring 70.
  • the angle of elevation of the flank 164 of the associated lead-in tooth 162 ensures that its corresponding friction pair is not self-locking and has the same helical direction as the associated leading flank 104 (for example, both left-handed), but the circumferential orientation of the sides of the two teeth is just right. Reciprocal.
  • the fixed point actuation mechanism ⁇ 4 includes a pawl 110, a return spring 114, and a ratchet 132 integrally formed with the auxiliary lead-in tooth 162 having a number of teeth equal to the number of teeth of the force transmitting tooth 72, the engaging surface 134 being the auxiliary leading flank 164. That is, the fixed point actuation mechanism ⁇ 4 and the introduction mechanism ⁇ 3 are mixed into a space guiding ratchet mechanism, see Figs. 1 to 3. And critically, when the force-carrying engagement mechanism M1 is fully engaged, the fixed-point actuation mechanism M4 is still engaged to prevent the non-transit separation of the former.
  • the overrunning clutch C1 can be in the second shaft of the latter shaft hole and Torque is transmitted between the gears (both not shown) that are coupled to the stationary ring 50 by the teeth 202.
  • the torque transmitted to the retaining ring 50 via the teeth 202 is transmitted to the moving ring 70 via the force transmitting mechanism M1 and ultimately transmitted to the second rotating member 208 by the spline pair, or vice versa, see FIG. a).
  • the pawl 110 and the ratchet 132 of the fixed-point actuation mechanism M4 are always in meshing state and have a reliable axial locking action, so that the moving ring 70 cannot be axially separated from the fixed ring 50 unless it is rotated beyond Time. Therefore, the force transmission conditions are stable and reliable.
  • the circumferential compressive stress and bending moment on the force transmitting teeth 52, 72 and the spline teeth 78 can be 100% used for torque transmission, that is, at the surface compressive stress and the bending strength. In terms of the effective utilization of the mechanical potential of the material, it can reach 100%, which is significantly higher than the level of about 70 to 90% of the prior art spiral spline pair.
  • the amount of axial movement of the moving ring 70 is independent of the tooth length of the end face type force transmitting teeth 52, 72, and does not directly limit the length of the force transmitting teeth, that is, the torque transmitting capability, as in the SSS synchronous clutch.
  • this embodiment has the advantages of greater load carrying capacity and higher impact resistance, smaller radial and axial dimensions, and correspondingly higher operating speeds and wider application range. It can be used in large transmission applications, and can be applied to small transmission applications with better versatility.
  • the specific parameters of the Synchronous Clutch 55T of the UK SSS Gears Limited which is representative of the prior art and products, are: working torque 5,000 Nm, destructive torque 15,000 Nm, transmission tooth maximum joint outer diameter 155 mm, exclusive shaft
  • the length is 207.5 mm (the displacement of the moving ring should be no less than 13 mm); and the 3 H JI-164 introduced in the book "Automobile Axle Design” (Liu Weixin, Tsinghua University Press, April 2004, p273 ⁇ 277)
  • the jaw-type freewheel differential has a calculated torque of 15,680 Nm for the one-side overrunning clutch when the maximum joint outer diameter of the transmission tooth is 155 mm, and the 18 transmission teeth are still cross-section.
  • the inverted trapezoidal two-way tooth has an axial common length of less than 60 mm and a moving ring displacement of only 5.5 mm. If the sawtooth one-way force transmission tooth of the embodiment is replaced, the bearing capacity is about 4 to 5 times. That is to say, a simple analogy can be seen that the carrying capacity of the present embodiment has an equivalent load-carrying outer diameter of 62,720 to 78,400 Nm, which is about 12 to 15 times that of 55T.
  • the present embodiment is the same as the number of ratchet teeth of the prior art such as the product 55T, then the two sides (the statistical or probability average) circumferential angles before the fixed-point actuation mechanism M4 is actuated are the same, and thus, The meshing impact strength of the two sides will only depend on the angle of rotation required for the introduction process. Needless to say, in the case of having the same guiding helix angle, the rotation angle of this embodiment is doubled smaller than that of the prior art.
  • the amount of movement of the moving ring 70 has been reduced by a factor of two, and the second is the embodiment.
  • the radius of the introduction mechanism M3 is significantly larger, and the larger radius means that the circumferential angle corresponding to the same circumferential distance is smaller (the introduction mechanism M3 of the embodiment can be arranged to the outer cylindrical surface of the moving ring 70, see The following instructions). Therefore, in this embodiment, the total circumferential angle of the anti-override rotation required for completing the fitting reset, that is, the slip angle, the difference between the rotational speeds of the two moments of the engagement force M1 and the meshing impact strength are significant or at least Significantly smaller than the prior art.
  • this embodiment does not have a separate asymmetrical rotating member, and has no frictional resistance proportional to the centrifugal force. Therefore, a relatively higher rotational speed and a higher torque are achieved, and the transmitted power is relatively greatly increased. While gaining these performance advantages, the relatively more mature and lower cost process structure, such as zigzag force transmission teeth and straight tooth splines, is simpler to manufacture and assemble, and the cost is lower. Moreover, referring to FIG.
  • the axial fitting of the force transmitting fitting mechanism M1 adopts a wedge pattern having a large inlet margin starting from the crest, instead of a small or zero inlet starting from the flank
  • the equal width mode of the margin; the derivation mechanism M2 as the control portion and the introduction mechanism M3, the force transmitting fitting mechanism M1 as the force transmitting portion, and the spline tooth pair (ie, the external transmission mechanism) are independent, and neither Affected by the wear of other mechanisms; Due to the introduction and derivation process, the fixed-point actuation mechanism M4 is always in an effective state, and can be engaged with the force-transmitting engagement mechanism M1 at the same time without the possibility of withstanding the torque, and the components do not need to be As in the prior art, the axial misalignment is repeated to ensure the reciprocal state of the above-mentioned meshing state; therefore, this embodiment has higher operational reliability, longer working life, and lower use and maintenance requirements than the prior art. Moreover, the reliability of the axial fitting of the force-car
  • the fixed point actuation mechanism M4 in this embodiment is essentially a guide ratchet mechanism.
  • the axial force component exists on the pawl 110. Therefore, as long as the pawl 110 is unidirectionally defined by an axial support member 136 as shown in Fig. 8, the above-described impact of the impact is reduced, especially in the case where the support member 136 itself has axial elasticity or is interposed with elastic members therebetween. in the case of. Therefore, compared with the prior art, the fixed-point actuation mechanism M4 of the present embodiment is not only subjected to less impact, but also has higher impact resistance and a more significant possibility of rigidity damage. Significantly improved overall reliability and longevity with key points.
  • the pawl and the ratchet of the guiding ratchet mechanism can be not only radially indexed, but also arranged or repeatedly arranged at the C portion of the moving ring 70 and the D portion of the supporting shell 230 to be adapted separately or simultaneously.
  • the working speed is high and low, and the pawl or ratchet can still be fixed to the circumferential direction of the fixing ring 50 by the screw 218.
  • a clear override such as the drive shaft of a low-speed high-speed split starter or the high-speed engagement of a low-speed split twin-engine helicopter, it is easy to achieve a no-load torque overrun by centrifugal force. .
  • the return spring 114 in the guide ratchet mechanism is not required and its function can be fully provided by its own gravity or centrifugal force by means of the circumferential distribution density of the pawl 110 (e.g., greater than four). It is also easy to understand that the present invention does not limit the structural form of the pawl 110.
  • the cylindrical shape appears in FIG. 1 for convenience of description, and it can have any other structural form, such as a figure.
  • the fixed-point actuation mechanism M4 can also be a mechanism such as a pin-slot type fitting mechanism, or an electric or hydraulic mechanism.
  • the present invention does not particularly limit the circumferential distribution and number of the associated teeth, and there may be an integer multiple relationship with each other.
  • at least one of the force transmitting teeth 52, 72, the lead teeth 92, 102, the introduction teeth 152, 162, and the ratchet teeth 132 and the pawl 110 should be optimally equal to the same natural number.
  • the other of the other parties is disposed only at a circumferential bisector which is equally divided by the natural number, and does not require a sufficient arrangement.
  • the natural number is 18, but the number of pawls 110 may be equal to one.
  • the number of the ratchet teeth 132 and the force transmitting teeth 52 or 72 are equal (for example, 18)
  • the number of the fitting paths 80 in one week will be equal to the number of the force transmitting teeth, and the force transmitting fitting mechanism M1 can be ensured.
  • the angle of the nose of the associated lead flank 164 should be no less than the helix angle of the leading flank 94 or 104.
  • the derivation mechanism M2 and the introduction mechanism M3 of the present embodiment can respectively rotate the relative rotation between the moving ring 70 and the fixing ring 50 and the relative rotation between the moving ring 70 and the second rotating member 208, respectively. Actuate the motive force to play the action separately or simultaneously. As shown in Fig. 1, the straight spline of the spline tooth pair can be changed to a helical spline, and its rotation direction is opposite to the rotation direction of the auxiliary introduction tooth flanks 164, so as to facilitate the axial movement of the moving ring 70 and the fitting force. Stable state (double insurance).
  • the introduction mechanism M3 can be ensured, regardless of its location. At the office. That is, even if the individual rotation of the moving ring 70 relative to the fixed ring 50 and the second rotating member 208 is self-locking due to the friction pair and cannot be moved axially, it can be introduced/crushed by the relative rotation between the latter two.
  • the helical guiding mechanism in the SSS synchronizing clutch li which uniquely activates the motive force by the relative rotation between the moving ring 70 and the second rotating member 208, still has some of the advantages described above and is still superior to the prior art.
  • the ratchet engaging surface 134 is parallel to the axis 200, and the cross-section of the force transmitting teeth 52, 72 may be rectangular, the force-transmitting fitting mechanism M1 may have no circumferential clearance, and in order to avoid the rotation, the force-fitting is engaged.
  • the ratchet fixed-point actuation mechanism M4 must first fail in an axially displaced manner.
  • the flat end layout form is significantly better than ⁇ , ⁇ when taking other values of the cone or The layout of the cylindrical surface.
  • the force transmitting structure on the fixing ring 50 can also be changed into a force transmission key groove or a belt groove by the gear teeth 202, or directly
  • the embodiment is changed to a shaft-shaft transmission structure as shown in FIG.
  • an introduction ring 150 may be separately provided in accordance with Fig. 8 to split the above-mentioned rigidly integrated guide ratchet mechanism. That is, a separate ratchet type fixed point actuation mechanism ⁇ 4 is formed between the introduction ring 150 and the fixed ring 50, and a separate screw introduction mechanism ⁇ 3 is formed between the introduction ring 150 and the moving ring 70 (see the description of the third embodiment for details). If so, the fixed point actuation mechanism ⁇ 4 can also be in the form of an end face type ratchet mechanism.
  • a spring ball positioning mechanism such as a spring ball positioning mechanism is disposed radially at the crotch portion of the second rotating member 208 to more effectively prevent the accidental collision of the moving ring 70 with the fixing ring 50.
  • the slip angle of the embodiment can be reduced by increasing the circumferential density of the force transmitting teeth and the ratchet teeth, thereby directly reducing the fixed point actuation mechanism ⁇ 4 and the force transmitting fitting mechanism. Ml's meshing impact.
  • the additional damping mechanism M9 can also buffer the rigid impact when the force transmitting fitting mechanism M1 is engaged.
  • a shoulder is formed at the A portion of the second rotating member 208, and an oil-discharge annular seal is formed by the shoulder and the spline base ring 76 of the moving ring 70, or a plurality of circumferential directions as shown in FIG. The cylindrical piston seal of the cloth.
  • Embodiment 2 Wheel-shaft-driven one-way overrunning clutch with package type two C2
  • the overrunning clutch C2 is a deformation of the overrunning clutch C1, which radially outwardly flips the latter moving ring 70 and the second rotating member 208, and adds a pawl holding mechanism M8, a damping mechanism M9, and a fixed
  • the ring 50 is rigidly integrated with the first rotating member 206.
  • the damper mechanism M9 includes a damper through hole parallel to the axis 200 on the second rotating member 208, a cylindrical damping member 226 movably fitted therein, a spring and threaded plug assembly 228, and a damping member
  • the middle section of the through hole corresponds to a breathing passage for the damping oil or gas in and out of the spring portion.
  • this structure is more suitable for a scheme in which the spline tooth pair has an introduction function.
  • the damping mechanism M9 of the present embodiment has only a unidirectional damping function, which is advantageous for the overrun separation of the moving ring 70.
  • the columnar damping member 226 The spline teeth 212 that extend outwardly under the action of the spring and threaded plug assembly 228 up to the second rotating member 208 can of course also be placed over a snap ring retaining ring that is selectively insertable therebetween.
  • the damping member 226 dampens the axial movement of the moving ring 70 by the spline teeth 78 until the fitting reset process ends.
  • the pawl 110 therein has the ordinary swing type with the self-rotating shaft shown in FIG.
  • the engaging surface of the head of the claw body 120 is the introduction tooth flanks 154
  • the back surface of the base body 116 is a semi-cylindrical rotating surface 118
  • the front surface is a limiting surface 144a, 144b, and the two sides are respectively at two rotation limit positions and the limiting member.
  • the inner cylindrical surface of 130 is fitted.
  • the seat groove 112 has a corresponding semi-cylindrical surface 148, and the bottom of the groove in which the claw body 120 is accommodated is formed with a radial type spring hole 115 in which the return spring 114 is disposed, see Fig. 10(a).
  • the pawl retaining mechanism M8 is specifically designed to block the engagement of the pawl 110 in the overrun condition, keeping it in the separation station to eliminate noise and spring fatigue.
  • the mechanism includes an open elastic blocking ring 140 that elastically expands in a bearing bore of the support shell 230, and a swing arm 122 on the pawl 110, see Figs.
  • the swing arm 122 is integrally connected to the claw body 120 through the base 116, and the front surface of the head is a blocking surface 124.
  • the blocking surface 124 is in engagement with the blocking working surface 142 of the blocking ring 140 to effectively block the engagement of the pawl 110 with the ratchet 132 in the radial direction.
  • the blocking ring 140 is an open elastic expansion ring, and a relief notch 143 is formed on the inner cylindrical surface to provide a space for the swing arm 122 to swing in the non-blocking state, and the blocking working surface 142 on the notch side has a blocking working surface. Outer edge 145.
  • the blocking surface 124 and the blocking working surface 142 have a skew angle with respect to the radial line or the circumferential tangent, which is optimally capable of ensuring the self-locking of the friction pair between the two contact faces during the blocking process.
  • it is not self-locking, because the blocking relationship does not disappear instantly, especially in the case of large rotational speed difference, the cost is that the blocking effect is worse.
  • the claw body 120 is swingable about the rotation axis to achieve engagement or disengagement with the ratchet teeth 132, and at the same time, the swing arm 122 is oscillated accordingly.
  • the blocking ring 140 rotates together with the supporting shell 230 in the direction of the arrow in FIG. 6, and the upper blocking surface 142 has the lower pressing of the latter. trend.
  • the ratchet teeth 132 radially compress the pawl 110 to the maximum limit
  • the inclined blocking working surface 142 will rotate synchronously to the extreme position where the swing arm 122 is subjected to maximum compression, and establish a stable relationship with the blocking surface 124. Friction self-locking blocking relationship.
  • the blocking ring 140 since the blocking ring 140 is unlikely to further compress the swing arm 122 in the radial direction, the blocking ring 140 can only be stationary with respect to the swing arm 122, and is slidably rubbed against the support shell 230, see Figs.
  • the blocking ring 140 must rotate synchronously with the moving ring 70 in conjunction with the supporting shell 230, and the relative swinging arm 122 reverses a small angle. After that, the blocking state is completely released, and the swing arm 122 and the pawl 110 resume swinging from By this, the pawl 110 can partially engage the ratchet teeth 132 before this moment.
  • the extension of the life of the spring in this embodiment is at the expense of slight frictional resistance, mechanical wear without substantial influence, and loss of fitting angle less than ⁇ . It is especially suitable for use in clutches that can be exceeded at any speed.
  • the solution of the blocking ring 140 and the centrifugal force can also be an option. This is as long as the swing arm 122 and the claw body 120 in Fig. 5 are respectively disposed on different sides of the rotation axis thereof, as shown in Fig. 7. After the overrunning, the centrifugal force on the weighted swing arm 122 will force the pawl 120 against the resilient return force of the spring 114 to rotate into a position that does not contact the ratchet 132.
  • the centrifugal force of the swing arm 122 can also be obtained by forming a radial blind hole at a portion of the seat groove 112 that accommodates the swing arm 122, and then placing a weight member such as a steel ball, and The orifice is naturally sealed with a swing arm 122.
  • magnetic force can be used instead of elastic restoring force.
  • the magnetic body is placed under the tooth flanks of the ratchet teeth 132 or in the radial blind holes to directly attract the pawl 110 or the swing arm 122 configured as a lever.
  • the compressive stress received by the spline tooth pair in the embodiment is smaller due to the larger radius, and is more than the prior art, and the wear strength of the spline tooth is significantly reduced, and the life of the spline tooth is improved. At the same time, the failure rate of the spline tooth sub-axial sliding is significantly reduced.
  • the deriving mechanism ⁇ 2 and the introduction mechanism ⁇ 3 in this embodiment may also be respectively rotated between the moving ring 70 and the fixing ring 50 and between the moving ring 70 and the second rotating member 208, respectively or simultaneously.
  • the relative rotation is the driving force for its actuation.
  • the moving ring 70 can directly output torque, and the inner annular surface can be used as a radial positioning sliding bearing surface, and the spline teeth 78 can be used as the gear teeth 202, and the gear teeth 202 can be straight or helical cylindrical gear teeth. . Further, the guide mechanism ⁇ 2 and the introduction mechanism ⁇ 3 may be mixed with the helical cylindrical gear mechanism.
  • the damper chamber of the damper mechanism ⁇ 9 is designed to be either oil-discharged or oil-absorbing.
  • the damper mechanism ⁇ 9 can also be changed from the pin-type damper cavity structure to the annular damper cavity structure in the form of an annular member.
  • Embodiment 3 Unpackaged shaft-shaft transmission type two-way overrunning clutch C3
  • the two-way overrunning clutch C3 is essentially an organic superposition of the two-way overrunning clutch C1 (removing the bearing 224 and the support housing 230) in which the two working directions reciprocate, and is added to cause the fixed point actuation mechanism M4a or
  • the force transmitting teeth 52, 72 respectively have two force transmitting flank faces 54, 74 parallel to the axis 200, and the circumferential degrees of freedom in the fitted state, which should be large or not
  • the degree of interference with the operation of the derivation mechanism M2 and the introduction mechanism M3 is different from that of the one-way overrunning clutch C1 in this embodiment, see Figs. 9, 10, and 17.
  • the derivation mechanism M2 includes an accessory derivation ring 100, and an outlet ring 90 that is independent of the stationary ring 50.
  • the lead-out ring 90 is slidably press-fitted onto the bore end face 56 of the retaining ring 50, one end of which is circumferentially evenly distributed with the indexing teeth 92 equal to the number of teeth of the force transmitting teeth 52, the teeth having two leading tooth flanks 94.
  • the auxiliary lead-out ring 100 is common to the spline base ring 76 of the moving ring 70, and has an auxiliary lead-out tooth 102 equal to the number of teeth of the force-transmitting teeth 72, which has two auxiliary lead-flank 104 on the end face.
  • the axial fitting depth between the lead-out tooth 92 and the auxiliary lead-out tooth 102 should be optimally larger than the axial fitting depth between the force-transmitting tooth 52 and the force-transmitting tooth 72.
  • the introduction mechanism M3 of the present embodiment includes the introduction teeth 152 and the auxiliary introduction teeth 162 which are fitted to each other, and the former is formed radially on the inner cylindrical surface of the introduction ring 150, and has two spiral introduction tooth flanks 154, the latter corresponding
  • the ground surface is formed on the outer cylindrical surface of the sub-introduction ring 160, which has two helical auxiliary lead-toothed sides 164, see Figs. 8, 9, 13, and 17.
  • the sub-introduction ring 160 is integrally formed with the sub-extraction ring 100.
  • An auxiliary lead-in slot inlet 168 is formed in the bottom of each of the subsidiary lead-in slots 166. Among them, the slots 166 are not penetrated in the radial direction (may also penetrate).
  • a support member 136 axially defining the position of the introduction ring 150 is also arranged, and the three end pins on the member are slidably passed through the axial through hole 82 formed in the moving ring 70. , topped on the support seat ring 138.
  • the support seat ring 138 is fixed to the end surface of the spline teeth 212 of the second rotating member 208 by the snap ring 220b.
  • the derivation mechanism M2 and the introduction mechanism M3 are the results of the integration of the one-way mechanisms corresponding to two different circumferential directions. Moreover, since the outward force received by the take-up mechanism M2 and the introduction mechanism M3 is small, mechanical damage is less likely to occur even if only one lead-out tooth or lead-in tooth is used for meshing. Therefore, it is entirely possible to arrange only one to three auxiliary lead-out teeth 102, an auxiliary lead-in tooth 162 or a pawl 110 as explained in the first embodiment.
  • two fixed-point actuation mechanisms M4a, M4b corresponding to the first and second rotational directions, respectively, include mutually engageable pawls 110a, 110b and bi-directional ratchet teeth 132.
  • the pawls 110a, 110b are respectively inserted into the seating grooves 112a, 112b of the cylindrical surface of the fixing ring 50 in a circumferentially opposite direction, and the engaging faces of the heads of the claws 120 are parallel to the axis 200, It has a guiding effect.
  • the mounting groove 112 of the pawl has a semi-cylindrical rotating surface 148, and a pawl return spring 114 is disposed in the spring hole 115 at the bottom of the groove accommodating the claw body 120.
  • the circumferential groove 58 extends through the semi-cylindrical groove portion of the seat groove 112a, 112b, and the opening elastic annular restricting member 130 fitted therein restricts the pawl 110a, 110b to the seat groove through the limiting faces 144a, 144b 112a, 112b.
  • the outer cylindrical surface of the introduction ring 150 has a bidirectional ratchet shape, and the circumferentially evenly distributed ratchet teeth 132 have the same number of teeth as the force transmitting teeth 52.
  • the contact surfaces 134a and 134b on the both sides of the tooth and the corresponding pawls 110a and 110b can be respectively contacted.
  • the rotation of the introduction ring 150 relative to the stationary ring 50 in the second or first rotational direction is stopped.
  • the introduction mechanism M3 can cause the force-transmitting fitting mechanism M1 in the axially separated state. Reset to the correct fitting station.
  • the ratchet engaging faces 134a, 134b may also be formed on the one-way ratchet teeth 132a, 132b which are mutually offset from each other in the axial direction.
  • the introduction mechanism M3 of the present embodiment is directly actuated by the relative rotation between the moving ring 70 and the introduction ring 150, the introduction ring 150 at the time of actuation is fixed by the fixed point actuation mechanism M4.
  • the ring 50 is rotated, that is, the actuating motive force of the introduction mechanism M3 is actually still the relative rotation between the moving ring 70 and the fixed ring 50.
  • this embodiment is substantially the same as the pawl holding mechanism M8 of the second embodiment, and the former is equal to the simple superposition of the latter.
  • the blocking ring 140 is still an open elastic ring, and the ring is elastically contracted on the outer cylindrical surface 158 of the introduction ring 150, and the circumferential faces of the two sets of blocking working surfaces 142a and 142b on the outer peripheral surface are opposite to each other to alternate The swing arms 122a, 122b on the pawls 110a, 110b of the two corresponding directions are blocked.
  • Actuation selection mechanism M5 which is one of the key mechanisms of the present embodiment, includes an actuation selector ring 170 and a state pawl 126, and an arcuate avoidance notch 128 is formed in the state jaw 126, see Figs. 8, 11, and 12.
  • Actuation The selector ring 170 is rigidly integrally formed on the outer cylindrical surface of the lead-out ring 90, and has two relief notches 178a, 178b formed on the outer peripheral surface thereof, and two cam faces 176a, 176b having circumferentially opposite ends.
  • the circumferential angle between the avoidance notches 178a, 178b and the circumferential angle between the pawls 110a, 110b differ by a commutation angle value ⁇ .
  • the direct result of actuating the selection ring 170 relative to the rotational angle of the stationary ring 50 is the state of the interchangeable state fingers 126a, 126b.
  • the cam surface 176a completely lifts up the state claw 126a in the radial direction, so that the pawl 110a rigidly integrated with the latter is not in the separation station for a long time and cannot be swung, corresponding to the fixed point of the first rotation direction.
  • the moving mechanism M4a therefore fails, which is equivalent to non-existence.
  • the cam surface 176b removes the radial blocking of the state claw 126b, so that the pawl 110b rigid with the latter resumes the swinging freedom, and the fixed point actuating mechanism M4b corresponding to the second rotational direction is thus restored. Therefore, the two-way overrunning clutch C3 has an introduction capability only in one working rotational direction, and works as a pure one-way overrunning clutch.
  • the reversing mechanism M6 which is the second key mechanism of the present embodiment, is a swinging guide mechanism including a retaining ring 50, a lead-out ring 90, and a swing lever 180, see Figs. 8, 10, 12, and 15.
  • the swinging rod 180 is fitted in a scalloped groove 62 at the bottom of the annular groove 60 at the non-fitting end of the retaining ring 50, and the center pin 182 at one end thereof is rotatably inserted into the rotary hole 64, and the cylindrical reversing pin 186 at the other end passes through
  • the annular through hole 66 in the retaining ring is slidably fitted into the guide groove 196 which is inclined on the endless end surface of the lead ring 90.
  • the self-rotating rocker 180 can be slid by the guide between the reversing pin 186 and the reversing guide groove 196 to drive the outer circumference of the lead-out ring 90 relative to the fixed ring 50.
  • Angle ⁇ Obviously, the actuation selection mechanism ⁇ 5 is also driven to synchronize a corresponding actuation selection.
  • the first relative position corresponding to the first rotational direction is as shown in Fig. 17, and the lead-out ring 90 is rotated upward by a circumferential angle ⁇ with respect to the fixed ring 50, and is located at a second relative position corresponding to the second rotational direction.
  • the embodiment is also optimally arranged with a disc cam type reversing drive mechanism ⁇ 7, which includes fixing Ring 50, swing rod 180, and drive ring 190, see Figures 8, 10, 15, 16.
  • the driving ring 190 is rotatably fitted in the annular groove 60 of the non-fitting end of the fixing ring 50, and is axially positioned by the snap ring 220a embedded in the outer ring snap ring groove 204 of the annular groove 60.
  • the rod-shaped base 184 is confined in the sector groove 62.
  • a cam channel 192 in the annular plane of the drive ring 190 slidably restrains the cylindrical drive pin 188 on the swing link 180 at its inner edge forming end face teeth 191 for applying a rotational force thereto.
  • the cam channel 192 includes a slanted cam driving section 198 in the middle, and three arcuate stopping sections 194a, 194b at both ends which are centered on the axis of curvature but have different radii.
  • the drive pin 188 it is impossible for the drive pin 188 to push the drive ring 190 to rotate by the stop segments 194a, 194b, in which the friction pair can be self-locking, and the radial position of the drive pin 188 when sequentially stopped in the two segments is the swing lever 180.
  • the drive pin 188 will complete a slide between the stop segments 194a, 194b. transposition.
  • the swing lever 180 is unlikely to rotate, naturally, the commutation drive mechanism M7, the reversing mechanism M6 associated therewith. And the working position of the actuating mechanism M5 is locked, and the one-way working state of the two-way overrunning clutch C3 is stabilized.
  • the two-way overrunning clutch C3 can be controlled to be a one-way overrunning clutch that operates in the first or second rotational direction.
  • the drive ring 190 in Fig. 16(a) rotates clockwise with respect to the fixed ring 50
  • the drive pin 188 is pushed up to the stop section 194a and locked
  • the reversing drive mechanism M7 drives the reversing mechanism M6
  • the moving selection mechanism M5 positions the working direction of the two-way overrunning clutch C3 in the first rotational direction and fails the fixed-point actuation mechanism M4b (Fig. 12).
  • the lead-out ring 90 is located at the first relative position as shown in FIG.
  • the lead-out ring 90 is located at the second relative position with respect to the fixing ring 50, and only the b-series member or the feature portion corresponding to the direction is effective in the related member, and the opposite a-series member or feature portion is invalid and is equivalent to the non-existence. See Figure 17.
  • the present invention has only one-way over-separation at the same time, the double-side moving ring 70 without the jaw-type freewheel differential simultaneously reverses and separates, and the cogging width is relatively reduced, and the force-transmitting teeth are increased by a corresponding one. Therefore, the thickness of the flank 94 is derived. Therefore, the torque transmission capability of the two-way overrunning clutch C3 is about 1.3 times that of the latter.
  • the analog torque of the first embodiment is 155 mm, and the calculated torque is about 20,384 ⁇ . Meter.
  • the carrying capacity according to the present invention is doubled higher than that of the prior art and its products of the one-way transmission, and, like the one described in the first embodiment, it still has almost complete power such as high speed and high power.
  • this embodiment can have other optional advantageous conditions.
  • the pure clutch and two-way coupling conditions are implemented as shown in Figure 16(b).
  • the actuation selection ring 170 is mixed with the drive ring 190 into a ring that is directly manipulated (of course, may be a non-combined, separate operation), and the widened relief recesses 178a, 178b are formed in the outer circumference of the drive ring 190.
  • the state claw 126 and the seat groove 112 on the pawl 110 both extend axially to the outer circular surface of the drive ring 190 (the pawl 110 is axially embedded by the end surface of the fixed ring 50 when installed).
  • the station shown in Fig. 16(b) corresponds to the coupling condition, at which time the drive pin 188 of the swing lever 180 is located at the midpoint of the circumferential direction of the cam drive section 198, and the fixed point actuation mechanisms M4a, M4b are simultaneously active.
  • the overrunning clutch C3 cannot be fixed as long as the drive ring 190 is held at this position.
  • the lead-out ring 90 has only the ability to be imported without the ability to export, thereby shifting into the two-way jaw coupling condition.
  • one of the fixed-point actuation mechanisms M4a, M4b can be deactivated, thereby shifting into a one-way overrunning clutch condition corresponding to the first or second direction of rotation.
  • the slewing drive ring 190 can work under the coupling condition, and after entering the overrun condition, continue to rotate the drive ring 190, and can be disabled by the cam faces 176c, 176d.
  • the transition between the above three operating conditions is extremely simple, fast and reliable, and the idea of activating the fixed-point actuator M4 to achieve a pure clutch can also be used in a one-way overrunning clutch, which will not be described in detail herein.
  • the coupling operating condition can also be increased by appropriately increasing the circumferential degree of freedom of the deriving mechanism M2, and correspondingly adding a midpoint stop at the radial midpoint of the cam driving section 198 as shown in Fig. 16(a).
  • the cam channel 192 is realized by a two-step shape to a three-step shape.
  • the above conditions are particularly applicable to and advantageous for drive shafting systems such as dual power drive systems in large ships. That is, after completing the relay replacement of the power machine, the overrunning clutch C3 in the load state is set as the coupling working condition, which can eliminate the possibility of harmful separation in the case of reverse or wind and waves; In the working condition, the overrunning clutch C3 in the non-load state is set to the pure clutch condition, and the replaced power machine can be subjected to unrestricted maintenance and debugging.
  • a gear such as a shoulder or a slip ring mechanism, which is interlocked with the reversing drive mechanism M7 or separately controlled, may be disposed between the outer circular surface of the moving ring 70 and the frame, and the moving ring 70 is axially defined. To prevent accidental separation, fitting or collision.
  • the reversing drive mechanism M7 is separately controlled or linked to the brake mechanism of the motor vehicle, and the embodiment can also be used as a state-controllable two-way glider.
  • the two-way glider when overtaking the vehicle brakes or intentionally controlling it, the two-way glider is turned into a reverse force transmission condition or a coupling condition, causing reverse or full fixed point actuation.
  • the mechanism is effective, the glider is temporarily fitted to reset to end the gliding and reverse loading, and the vehicle engine begins to provide braking force.
  • the glider will automatically revert to the previously set working direction. If the override function of the reverse direction is canceled, for example, the stop section 194b in Fig.
  • the function of the fixing ring 50 in the introduction mechanism M3 or the fixed point actuation mechanism M4 can be replaced by the lead-out ring 90, and the auxiliary lead-out ring 100 can also be attached to the fixing ring 50, and the actuation mechanism M5 is activated to disable the fixed-point actuation mechanism M4.
  • the method is not limited to the above, and it is feasible to axially shift the pawl and the ratchet, or to sequentially connect the two separate ratchet mechanisms with the introduction mechanism M3.
  • the mechanism for the radially raised state jaws 126 is not limited to a form of a disc cam, such as an end face cam disposed on the drive ring 190, which compresses an axially resiliently reset selector lever with a wedge shaped head in the form of a cylindrical cam mechanism.
  • two actuation selection rings 170 having a radius of infinity are respectively arranged in the plane of the shaft), which can also lift the state claws 126 radially.
  • the state jaws 126 and the seat slots 112 do not necessarily extend axially to the outer circular surface of the drive ring 190. Actuating the selector ring 170 is not limited to interlocking a control form.
  • the actuation selector ring 170 can be a separate ring formed with an end face lug that fits over the outlet ring 90, the lug rotatably passing through a corresponding annular through hole in the retaining ring 50, and the drive ring 190 or The groove corresponding to the outer edge of the other control ring is fitted to achieve circumferential fixation of both, and thus the individual control that the selection mechanism M5 can be actuated.
  • the manipulation of the reversing drive mechanism M7, or the rotation of the drive ring 190 can be carried out in a stalled state, by means of the inventor's "relative motion direction sensing device (see 200810080503.9 patent) (Related note) "Adapted adaptively during operation (in this case, the pure clutch condition can be obtained as long as the set rotation direction is always opposite to the actual rotation direction), or the circumferential frictional resistance can be artificially applied. It is also possible to form the rotary guide mechanism by the drive ring 190 and a control ring fixed in the circumferential direction of the fixed ring 50 but axially sliding, in such a manner as to axially move the control ring.
  • the driving rotation direction, the reversing rotation direction and the target working rotation direction can also be realized. Positive or negative opposition, and control of the speed and sensitivity of the commutation. It is not difficult to understand that all the above-mentioned working conditions are actually only effective or invalid by the fixed-point actuation mechanisms M4a, M4b obtained by joint or separate control, and the arrangement and combination of the lead-out rings 90 respectively positioned at the first and second relative positions or free positions. Part of the result.
  • the two embodiments of the fixed ring 50 and the lead-out ring 90 are axially double-backed in a back-to-back rigid manner, respectively, and a new type of fixed-point fitting is obtained.
  • the jaw-mounted self-locking differential In all two-way solutions, there should be no optimal guiding between the moving ring 70 and the second rotating member 208.
  • the present invention is hardly limited to the reversing mechanism M6 and the reversing drive mechanism M7, and it may be any one of mechanisms such as mechanical, hydraulic or electromagnetic, and there are a large number of known techniques. Embodiments may be selected, modified or combined.
  • the inventors' application numbers 200710152152.3 and 200810080503.9 have many related proposals in the Chinese invention patents. Therefore, the entire contents of the two patent applications are hereby incorporated herein by reference.
  • by selectively failing the fixed-point actuation mechanisms M4a, M4b it is easy to obtain a single or two-way guided jaw clutch with precise engagement, or a guided jaw-mounted electronically controlled clutch, which is excellent in performance, structure and use conditions. Corresponding jaw clutch or jaw electromagnetic clutch in the prior art.
  • the reversing mechanism M6 is not limited to a circumferentially positioned full-toothed lead-out ring.
  • an axially positioned half-toothed lead-out ring scheme such as that shown in Figure 18 can also be employed.
  • the lead-out teeth 92a, 92b on the lead-out rings 90a, 90b are only half of those in Fig. 12, and each has a leading flank 94a, 94b whose circumference faces opposite to each other.
  • the mutually disposed lead-out rings 90a, 90b and the fixing ring 50 are axially slidably axially slidable with each other (through, for example, an axial pin-hole fitting mechanism, a sliding key coupling mechanism), and the second ring and the driving ring 190 are formed with each other such as
  • the channel type cylindrical cam reversing mechanism realizes the purpose of alternately extending the two lead-out rings in the axial direction. That is, when the lead-out ring 90a is in the first relative position shown in FIG. 18(a) (equivalent to the position in FIG. 17), the lead-out ring 90b must be axially retracted and hidden, and the overrunning clutch C3 operates at the first rotation. In the direction.
  • the sub-introduction ring 160 can be formed as a separate open elastic ring, with the shoulder of the end portion and the corresponding circumferential groove on the moving ring 70. Axial fixation is achieved, and circumferential fixation is achieved by means of a second rotational member 208 that also extends through the inner bores.
  • Embodiment 4 Wheel-shaft-driven two-way overrunning clutch with package type two C4
  • the present embodiment employs the package form shown in the second embodiment and the main structure of the third embodiment, and the pawl holding mechanism M8 and the independent introduction ring 150 are omitted.
  • a minor modification is that the sub-emission tooth 102 in the derivation mechanism M2 is radially integrated with the force-transmitting tooth 72.
  • the introduction mechanism M3 and the fixed-point actuation mechanism M4 are respectively mixed into two spatially-oriented ratchet mechanisms corresponding to different circumferential directions, and the introduction flank 154 is again the claw-engagement surface of the pawl 110, forming an integral two-way ratchet 132 and the auxiliary introduction teeth 162 are circumferentially evenly distributed on the inner hole surface of the moving ring 70, and are equal to the number of the force transmitting teeth 72.
  • the introduction mechanism M3 of the present embodiment is still actuated by the relative rotation between the moving ring 70 and the fixed ring 50.
  • annular restricting member 130 is formed with a circumferential through hole for avoiding the pawl 110 claw body 120, and the pawl 110 is radially defined to axially define the lead-out ring 90.
  • the actuation selector ring 170 is integrally formed on the inner diameter side of the outlet ring 90. And, the position of the reversing mechanism 196 and the reversing pin 186 in the reversing mechanism M6 is adjusted to facilitate the arrangement of the reversing guide groove 196 on the lead-out ring 90.
  • the scallops on the retaining ring 50 are divided into two stepped portions 62a, 62b, and the annular through-holes 66 at the bottom of the groove are no longer centered on the axis 200 and whose center of curvature/slewing is the center of the swivel hole 64.
  • Figures 23(a) to (c) show three operating conditions of transmitting torque, over-disengagement and fitting reset in the first rotational direction.
  • the working process of the guiding pawl 110 is clearly shown, and obviously, the embodiment can also add the axial supporting member 136 to the pawl 110, as well as the blocking ring 140, and can also be added to the introduction ring 150 according to the idea of FIG.
  • the introduction mechanism M3 is separated from the fixed point actuation mechanism M4. As shown in FIG. 24, the introduction ring 150 is a bidirectional ratchet, and the outer direction is a bidirectional introduction tooth 152.
  • the corresponding working mechanism of the introduction mechanism M3 is most clearly revealed in FIG. 23(d), and the mechanism is substantially axially reversed. Or against the export mechanism M2.
  • the invention can be applied directly to all mechanical transmission fields, especially for overrun and reverse applications in addition to direct indexing, with transmission capability common to almost all speeds, all torques and all power.
  • Torque Converters Automatic Transmissions, Pulsating Continuously Variable Transmissions, Backstops in Hoisting Machinery and Other Machinery, Harvesting / Harvesting Machines, (High Voltage Switching) Vacuum Circuit Breakers, Turbine Generator Sets , power machine starting device (can easily meet the actual requirements and use requirements of the starting device and the flywheel for permanent engagement, to completely abandon the electromagnetic switch), large surface ships, twin-engine helicopters, wheeled motor vehicles, non-slip limited finite differential ratio Differential, glider, etc.

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  • Mechanical Operated Clutches (AREA)

Description

导向式牙嵌超越离合器 技术领域
本发明涉及机械传动领域中的一种离合装置， 以及诸如包含此种离合装置 的牙嵌差速器之类的其它传动装置， 特别涉及但不仅仅涉及一种具备传递转矩 和转动超越功能的牙嵌式超越离合器。 背景技术
摩擦式超越离合器作为现有技术中的主要 /流类型， 其存在承载能力低下、 可靠性差、 传动效率低、 加工装配困难、 成本高、 易磨损、 应用范围小的缺点 (《超越离合器的发展现状及趋势》， 张济政等， 第三届中日机械技术史国际学 术会议， 昆明， 2002年， 398 ~ 403 )。 而嵌入式， 虽具有克服上述不足的可观潜 力， 却由于超越转动时有碰撞、 碰撞噪声和业界固有认识等原因， 除了 SSS ( Synchro-Self-Shifting )同步离合器外， 该类超越离合器几乎未得到应有的重视 或应用， 尤其是对于其中的牙嵌式， 其承载能力巨大， 径向尺寸相对较小， 以 及接合 /嵌合后没有相对滑转等优点更没有得到实质利用。
SSS 同步离合器的主要工作原理是， 依靠单向棘爪的致动作用， 以及位于 移动环与传动轴之间的螺旋花键齿副的相对转动的导向作用， 致使传力齿环双 方于周向相对静止中实现轴向的接合 /嵌合和分离（此即所谓的同步自移动）。 其 中， 螺旋花键齿副兼具传递转矩和转动导向的双重功能。 相对摩擦式超越离合 器， SSS 同步离合器兼具高转速和高转矩的优点 （最大转矩与摩擦式超越离合 器的一百万牛米相当）。 但其技术发明至今已有不少于 40年的历史， 或囿于技 术上的惯性思维， 或受现有技术的原理和结构的束縛， 其核心结构或工作模式 未得到实质的改进、 提高或革新。 其转矩传递仍只有依靠圓柱面上的直齿这一 种形式（其周向间隙量致使承载能力与同轴度、 接合可靠性相矛盾）， 传力齿的 接合 /嵌合或分离仍唯一地依靠同步模式， 转动导向机构的致动原动力仍唯一地 依赖于移动环与传动轴之间的相对转动。 因此， 其固有缺点一直没有得到克服。 例如， 制造、 装配的精度和工艺要求过高， 同轴度要求过严， 成本过高， 移动 环轴向移动量较大， 核心结构的轴向、 径向尺寸及独占空间都较大， 不易小型 化， 致使其应用范围过窄 （主要应用于舰船和大型发电机组等少数传动轴系）， 通用性较差。 而且， 其花键齿的螺旋升角特性在导致花键齿面摩擦强度过高的 同时， 更因材料的机械 /力学潜能的有效利用率显著低于 100%, 而降低了其承载 能力。 另外， 因结构和工作原理所致， 其承载能力还不够大， 更无法实现双向 承载以及工作状况的人为可控， 实用中接合 /嵌合或分离可靠性也不够高。 发明内容
本发明致力于解决上述问题。
本发明的目的在于， 提供一种嵌合、 分离均由转动导向机构完成， 分离和 嵌合轨迹相对固定， 传力齿布局和导向机构的致动原动力多样， 齿顶间无滑动 磨损且承载能力更大的导向式牙嵌超越离合器， 其可选择地具有单向、 双向超 越离合器工况， 以及受控制的滑行器、 离合器或联轴器工况。
为达成上述发明目的， 本发明之导向式牙嵌超越离合器， 包括用于传递转 矩的轴向嵌合式传力嵌合机构， 其具有绕同一轴线转动且分别形成有传力齿的 固定环和移动环， 移动环与离合器外的第二转动构件组成既可轴向移动又可传 递转矩的外部传动机构； 至少一个转动导向式导出机构， 其在第二转动构件与 固定环间开始超越转动时， 致使移动环轴向远离固定环， 以解除传力嵌合机构 的轴向嵌合； 至少一个转动导向式导入机构， 其在第二转动构件与固定环间开 始反超越转动时， 致使移动环轴向移向固定环， 以恢复传力嵌合机构的轴向嵌 合； 至少一个定点致动机构， 其用于上述反超越转动过程中， 在移动环相对固 定环处于特定圓周位置时致动导入机构， 以完成导入运动； 其特征在于： 传力 嵌合机构嵌合时， 其全部传力齿的实际啮合表面的啮合中线至少大体上位于同 一啮合圓锥面上。
优选地， 上述啮合圓锥面与回转轴线之间的夹角大于等于 0。 而小于等于 180° ， 导出机构和导入机构中的至少一个机构以移动环与固定环之间的相对转 动为其致动原动力， 传力嵌合机构嵌合时的入口裕度大于零。
本发明的更多的优良改进方案， 以及受控制的双向超越离合器方案， 由其 它从属权利要求给出。
需要特别说明的是， 本发明文件所用相关概念或名词的含义如下： 失效： 指机构或零件由于人为或客观原因不能正常工作， 丧失其基本功能 之含义。 例如， 以破坏构件基本关系的方式将棘轮机构中棘爪与棘齿轴向相互 错位， 或者， 以破坏构件基本运动的方式将其棘爪强制约束在分离位置上， 该 棘轮机构都将因二者丧失啮合的可能而失效。
转动导向机构： 由圓周相对转动为致动原动力以产生 /得到轴向相对移动的 机构。 既包括螺旋升角严格一致的同轴螺旋机构、 异轴外啮合斜齿轮机构等， 也包括螺旋升角不严格一致的径向销槽机构、 端面棘轮机构、 端面嵌合机构及 圓柱凸轮机构等。 本发明中导入机构和导出机构均为转动导向机构。
轴向嵌合： 嵌合机构中嵌合双方的嵌合或分离均伴随有双方轴向相对移动 的一种嵌合模式； 其嵌合齿可以形成在双方的端面、 锥面或圓柱面上。
超越转动和反超越转动： 都是转矩传递路径下游一方的转动构件相对转矩 传递路径上游一方转动构件的转动， 只是前者的相对转动方向与超越离合器所 要传递的圓周力方向一致， 而后者的相对转动方向却与之正好相反。
根据本发明的导向式牙嵌超越离合器， 其转动导向式导出机构和导入机构 的致动原动力， 穷尽了所有可能， 可以是移动环与第二转动构件间的相对转动， 也可以是移动环与固定环间的相对转动。 另外， 传力嵌合机构的传力齿不再唯 一地呈圓柱形分布， 丰富了传力齿的布局形式， 致动选择机构、 换向及换向驱 动机构令控制其工况成为可能并简单快捷。 不仅实现了本发明的目的， 而且相 比现有技术， 还提升了承载能力， 实现了大转矩高转速高功率的传动， 更可受 控地具有双向超越、 纯离合器、 联轴器及滑行器工况。 同时， 其接合和分离更 可靠， 同比尺寸更小， 适用性更广， 制造和装配工艺成熟， 安装、 使用和维护 简单方便。 借助下述实施例的说明和附图， 本发明的目的和优点将显得更为清 楚和明了。 附图说明
图 1是根据本发明的封装形式一的单向超越离合器的简化的轴向剖面图。 图 2是图 1中移动环的示意图， （a)是仰视图的轴向半剖图， （b)是主视图。 图 3是图 1 中的传力嵌合机构、 导出机构以及导入机构三者的齿廓， 在同 一外圓柱面上的径向投影的局部系列展开图， 以及对应的导入机构的假想齿廓 在棘爪所处轴截面轴向投影的局部展开图， 其中， （a)对应于传力工况， （b)对应 于超越工况， （c)对应于嵌合复位过程， 图中箭头表示移动环相对固定环的转动 方向。
图 4是根据本发明的封装形式二的单向超越离合器的简化的轴向剖面图。 图 5是图 4中棘爪的示意图， （a)是仰视图， （b)是主视图， （c)是俯视图。 图 6是图 4中棘爪阻挡环的局部结构的放大示意图。
图 7是图 4中一种可选棘爪的示意图， （a)是仰视图， （b)是主视图， （c)是俯 视图。
图 8是根据本发明的轴一轴传动形式的双向超越离合器的轴向剖面图。 图 9是图 8中移动环的示意图， （a)是右视图的轴向半剖图， （b)是主视图。 图 10是图 8中固定环的示意图， （a)是主视图的轴向半剖图， （b)是 (a)中 H 方向局部结构的放大示意图。
图 11是图 8中棘爪的示意图， （a)是仰视图， （b)是主视图， （c)是俯视图。 图 12是图 8中导出环的示意图， （a)是简化的主视图的轴向剖面图， （b)是左 视图的局部结构的放大示意图。
图 13是图 8中导入环的示意图， （a)是仰视图的轴向半剖图， （b)是主视图。 图 14是图 8中棘爪阻挡环的局部结构的放大示意图。
图 15是图 8中摆杆的示意图， （a)是主视图， （b)是俯视图。 图 16(a)是图 8中驱动环的局部结构的放大示意图， （b)是其可选结构。
图 17是图 8中的传力嵌合机构 (1)、导出机构 (2)以及导入机构 (2)三者的齿廓， 在同一外圓柱面上的径向投影的局部系列展开图， 以及对应的定点致动机构的 假想轴向投影轮廓的局部展开图（3), 其中， （a)对应于传力工况， （； b)对应于超越 工况， （C)对应于嵌合复位过程， 图中箭头表示移动环相对固定环的转动方向。
图 18是双向超越离合器中采用半齿导出环换向方案时， 以传力嵌合机构以 及导出机构二者齿廓间的关系表示的换向原理图， （a)对应于方向一的超越工况， (b)对应于方向二的超越工况， 图中箭头表示超越转动方向。
图 19是图 20中摆杆的示意图， （a)是主视图， （b)是俯视图。
图 20是根据本发明的封装形式二的双向超越离合器的轴向剖面图。
图 21是图 20中移动环的示意图， （a)是右视图的简化的轴向剖面图， （b)是 简化的主视图。
图 22是图 20中固定环的示意图， （a)是主视图的轴向半剖图， （b)是 (a)中 H 方向局部结构的放大示意图， （c)是 (b)中 T T截面的局部放大图。
图 23是图 20 中的传力嵌合机构 (1)、 导出机构 (2)以及导入机构 (2)三者的齿 廓， 在同一外圓柱面上的径向投影的局部系列展开图， 以及对应的定点致动机 构在棘爪所处轴截面的假想轴向投影轮廓的局部展开图 (3), 其中， （a)对应于传 力工况， （b)对应于超越工况， （c)对应于嵌合复位过程， （d)对应于加入可选导入 环后传力工况的示意图； 图中箭头表示移动环相对固定环的转动方向。
图 24是图 20中可选构件导入环的示意图， （a)是主视图， （b)是左视图的轴 向半剖图。 具体实施方式
必要说明： 本说明书的正文及所有附图中， 相同或相似的构件及特征部位 均釆用相同的标记符号， 并只在它们第一次出现时给予必要说明。 同样， 也不 重复说明相同或类似机构的工作机理。 另外， 为区别布置在对称或对应位置上 的两个相同的构件或特征部位， 本说明书在其编号后面附加了字母， 而在泛指 说明或无需区分时， 将不作区分也不附加字母。
实施例一： 具有封装形式一的轮一轴传动式单向超越离合器 C1
参见图 1 ,超越离合器 C1的固定环 50与支撑壳 230通过螺钉 218紧固成一 体， 二者借助两个轴承 224轴向固定在第二转动构件 208上， 以将传力嵌合机 构 Ml、 导出机构 M2、 导入机构 M3和定点致动机构 M4封装起来， 并可单独 或一体地绕轴线 200转动。
传力嵌合机构 Ml包括轴向相对嵌合的固定环 50与移动环 70, 参看图 1 ~ 图 3。移动环 70上周向均布的传力齿 72最佳地具有一个平行于轴线 200的传力 齿侧面 74, 其另一齿侧面是相对轴线 200倾斜的附属导出齿侧面 104。 显然， 传力齿侧面 74的齿根也可内缩或外凸 （但无益于制作、 转矩传递和啮合稳定）。 固定环 50上嵌合端面的结构和齿数完全对等于移动环 70， 其传力齿 52相应地 具有传力齿侧面 54和导出齿侧面 94。 也就是说， 传力齿 52与导出齿 92共体， 传力齿 72与附属导出齿 102共体， 周向上各占据齿体的一半各对应于不同的圓 周方向， 横截面总体上呈锯齿状。 因此， 嵌合双方在组成单向传力嵌合机构 Ml 的同时也组成了单向导出机构 M2, 并分别对应不同的圓周方向。 这一点在图 3 中显示得最为清楚。 其中， 固定环 50、 移动环 70的齿槽槽口宽度显著大于对方 的齿顶宽度， 二者之差即入口裕度显著大于零。
传力嵌合机构 Ml的实际啮合表面的啮合中线 210的全体位于同一啮合圓锥 面上， 其与回转轴线 200的夹角分别记为 α和 β , 0。 α 180° ， 0。 （3 180。 ， 且 α + β = 180。 ， 参见图 1。 本实施例中， α = β = 90。 ， 即啮合圆锥 面具体为啮合平面， 这显然更有利于制造， 更有利于结构布置和提升性能。 这 里， 啮合圓锥面的概念与直齿圓锥齿轮中的节圆锥相类似。
参看图 1、 图 2, 导入机构 Μ3包括可以相互啮合的柱销式棘爪 110和螺旋 型附属导入齿 162,前者与复位弹簧 114一起以周向和轴向均固定的方式嵌装在 固定环 50的径向型座槽 112中，后者周向均布在附属导入环 160的外圓柱面上， 该环与移动环 70的花键基体环 76形成为一体。 附属导入齿 162的齿侧面 164 的升角可确保其对应的摩擦副不自锁， 并具有与附属导出齿侧面 104相同的螺 旋旋向 （例如均为左旋）， 但两齿侧面的圓周朝向正好互反。
定点致动机构 Μ4包括棘爪 110 , 复位弹簧 114, 以及与附属导入齿 162形 成为一体的棘齿 132 , 其齿数等于传力齿 72的齿数， 其啮合面 134就是附属导 入齿侧面 164。 也就是说， 定点致动机构 Μ4与导入机构 Μ3混合成一个空间导 向棘轮机构， 参看图 1〜图 3。 并且关键地， 当传力嵌合机构 Ml完全啮合时， 定点致动机构 M4仍处于啮合状态， 以制止前者的非超越分离。
于是， 在移动环 70通过其内花键齿 78与第二转动构件 208的花键齿 212 相连组成花键式外部传动机构后，超越离合器 C1便可在后者轴孔中的第二转轴 与通过轮齿 202相连到固定环 50的齿轮（均未示出 )之间传递转矩。 例如， 经 轮齿 202传给固定环 50的转矩， 再经传力嵌合机构 Ml传至移动环 70, 并最终 由花键副传递给第二转动构件 208， 或者正好相反， 参见图 3(a)。 传力工况中， 定点致动机构 M4的棘爪 110和棘齿 132始终处于啮合状态，具有可靠的轴向锁 定作用， 所以， 移动环 70不可能相对固定环 50轴向分离， 除非超越转动时。 因此， 传力工况是稳定和可靠的。
超越转动开始后 , 随着移动环 70的转速开始快于固定环 50, 导出机构 M2 中的附属导出齿侧面 104相对导出齿侧面 94按图 3(b)所示的箭头方向滑转， 致 使移动环 70相对固定环 50轴向分离导出，传力嵌合机构 Ml及导出机构 M2先 后失效， 超越离合器 C1进入超越工况， 复位弹簧 114压迫下的棘爪 110沿棘齿 132的齿背无阻挡地滑转摩擦或无接触空转， 参见图 3(b)。
当移动环 70的转速开时慢于固定环 50的瞬间，也就是其按图 3(c)所示的箭 头方向开始反超越转动的零时刻，超越离合器 C1便由超越工况即刻转入嵌合复 位的导入过程。定点致动机构 M4的棘爪 110必与遇到的第一个棘齿 132啮合以 致动导入机构 M3。 之后， 在棘爪 110的带动下， 反超越转动的移动环 70将顺 着螺旋型附属导入齿侧面 164规定的轨迹滑转， 其传力齿面 74的最高缘则沿对 应的嵌合路径 80滑转， 直至传力齿 72与 52轴向上周向上完满啮合。 于是， 嵌 合复位过程结束， 反超越转动停止， 超越离合器 C1回复到如图 3(a)所示的传力 工况。 并且， 棘爪 110与棘齿 132仍处于啮合状态。
由上述说明可见， 本实施例中， 传力齿 52、 72及花键齿 78上的周向压应 力和弯矩可以 100%用于转矩传递， 即， 在表面压应力和抗弯强度两方面， 材料 的机械潜能的有效利用率可以达到 100%， 这显著高于现有技术中螺旋花键副的 约 70 ~ 90%的水平。 尤其是， 移动环 70 的轴向移动量与端面型传力齿 52、 72 的齿长无关， 不像其在 SSS 同步离合器中那样直接制约了传力齿的长度也就是 转矩传递能力， 令该长度与快速接合并减小接合冲击， 即减小溜滑角直接矛盾。 因此， 相比现有技术， 本实施例具有更大承载能力和更高抗冲击能力， 更小径 向和轴向尺寸， 以及相应更高工作转速和更广应用范围的优点。 可应用于大型 传动场合， 更可应用于小型传动场合， 通用性更好。
例如， 作为现有技术和产品代表的英国 SSS Gears Limited公司的同步离合 器 55T的具体参数为： 工作转矩 5,000牛米， 破坏转矩 15,000牛米， 传力齿最 大接合外径 155毫米， 独占轴向长度 207.5毫米（移动环位移量应不低于 13毫 米）； 而 《汽车车桥设计》一书（刘惟信， 清华大学出版社 2004年 4月， p273 ~ 277)介绍的 3 H JI -164 型牙嵌式自由轮差速器在传力齿最大接合外径同为 155 毫米时， 其单侧超越离合器的计算转矩就已高达 15,680牛米， 而其 18个传力齿 还是横截面呈倒梯形的双向齿，轴向共用长度不到 60毫米，移动环位移量仅 5.5 毫米。 如果换成本实施例的锯齿形单向传力齿， 其承载能力还要放大约 4 ~ 5倍 左右。 也就是说， 简单类比即可知， 本实施例在同等传力外径时的承载能力已 达 62,720 ~ 78,400牛米， 约为 55T的 12 ~ 15倍。
另外， 假设本实施例与现有技术的诸如产品 55T的棘齿齿数相同， 那么， 双方在定点致动机构 M4致动前反超越转过的（统计或概率平均）圓周角便相同 , 于是， 对比双方的啮合冲击强度将仅仅取决于导入过程所需转动的角度。 毋庸 置疑， 在具有相同的导向螺旋升角的情况下， 本实施例的转动角度成倍地小于 现有技术的。 一是如上所述， 移动环 70的移动量已成倍降低， 二是本实施例的 导入机构 M3所处半径显著较大，而半径较大就意味着同样的圓周距离所对应的 圓周角较小 （本实施例的导入机构 M3更可布置到移动环 70的外圓柱面上， 参 见下述说明）。 因此，本实施例为完成嵌合复位所需反超越转过的总的圆周角度， 也就是溜滑角， 传力嵌合机构 Ml啮合瞬间双方的转速之差以及啮合冲击强度， 都显著或至少明显小于现有技术的。
显然， 本实施例没有独立的非对称回转构件， 没有与离心力成正比的摩擦 阻力， 因此， 实现相对更高转速更高转矩的同时也相对大幅提高了传递的功率。 而获得这些性能优点的同时， 因可选用锯齿形传力齿以及直齿花键等具有相对 更成熟和更低成本的工艺结构， 其制造、 装配、 安装反而更简单， 成本更低。 并且， 参见图 3(c), 因传力嵌合机构 Ml的轴向嵌合采用的是始于齿顶的具有大 入口裕度的楔形模式， 而非始于齿侧的小的或零入口裕度的等宽模式； 因作为 控制部分的导出机构 M2和导入机构 M3、 作为传力部分的传力嵌合机构 Ml和 花键齿副 （即外部传动机构） 四者各自独立， 且均不受其它机构磨损的影响； 因导入、 导出过程中， 定点致动机构 M4始终处于有效状态， 更可与传力嵌合机 构 Ml同时处于啮合状态而无承受转矩的可能，其组成双方无需像现有技术那样 为确保上述啮合状态的互反而反复地轴向错位； 所以， 本实施例相对现有技术 具有更高的工作可靠性， 更长的工作寿命， 更低的使用和维护要求。 并且， 传 力嵌合机构 Ml轴向嵌合的可靠性更不再敏感于同轴度，不再敏感于花键副或者 轴 的磨损。
应该注意到， 由于本实施例中移动环 70的质量相对较小， 所以， 定点致动 机构 M4啮合的瞬间， 棘爪 110受到的沖击强度相对现有技术较小。 另外， 本实 施例中的定点致动机构 M4实质上是一个导向棘轮机构，啮合时，棘爪 110上存 在轴向分力。 因此， 只要如图 8所示以一个轴向支撑构件 136来单向限定棘爪 110, 上述冲击的危害就会减小， 尤其是在支撑构件 136本身具有轴向弹性， 或 者其间间隔有弹性元件的情况下。 于是， 相比现有技术， 本实施例的定点致动 机构 M4不仅所受冲击更小， 而且其抗冲击能力更强， 刚性损坏的可能性更显著 降低。 与关键之处显著提高了整体的工作可靠性和寿命。
分析图 1 容易明了， 导向棘轮机构的棘爪和棘齿不仅可以径向换位， 而且 还可以布置在或重复布置在移动环 70的 C部位与支撑壳 230的 D部位，以分别 或同时适应高低工作转速， 且其棘爪或棘齿仍可通过螺钉 218实现与固定环 50 的周向固定。 对于具备明确超越规律的应用部位， 例如低速接合高速分离的起 动机的传动轴系， 或者高速接合低速分离的双发直升机的传动轴系， 借助离心 力将很容易实现无空载转矩的超越转动。 由此可见， 导向棘轮机构中的复位弹 簧 114并非必需， 其功能完全可以借助棘爪 110的周向分布密度（如大于 4个） 而由其本身的重力或离心力提供。 同样容易明了的是， 本发明并未限定棘爪 110 的结构形式， 以圓柱形出现 在图 1 中只为方便说明而已， 它完全可以具有其它任何一种结构形式， 比如图
11 所示的形成有自转定位柱面或孔面的普通摆动式棘爪。 同理， 定点致动机构 M4也可以是诸如销槽式嵌合机构， 或电动、 液动等机构。
应说明的是， 本发明并未对相关齿的圓周分布和数量给予特別限定， 相互 间完全可以存在整数倍关系。但必须强调的是，传力齿 52、 72, 导出齿 92、 102, 导入齿 152、 162, 及棘齿 132、 棘爪 110四组中的至少一方， 其数量应最佳地 分别等于同一自然数， 并布置在按该自然数所等分的圓周等分点上， 而各自的 另一方只要布置在按该自然数所等分的圓周等分点上即可， 并不要求足数布置。 比如， 本实施例中该自然数是 18, 但棘爪 110的数量就可以等于 1。
所以， 当棘齿 132与传力齿 52或 72的个数相等时 （如 18个）， 一周内嵌 合路径 80的个数将等于传力齿的个数， 可以确保传力嵌合机构 Ml具有最快最 优的嵌合复位性能， 参看图 3(c)。 另夕卜， 为可靠工作， 附属导入齿侧面 164的螺 旋升角应不小于导出齿侧面 94或 104的螺旋升角。
必须指出的是， 本实施例的导出机构 M2和导入机构 M3 , 可分别或同时以 移动环 70与固定环 50之间的相对转动 , 以及移动环 70与第二转动构件 208之 间的相对转动为其致动原动力， 以分别或同时发挥致动作用。 如图 1 所示， 花 键齿副的直齿花键可以改为螺旋花键， 其旋向相反于附属导入齿侧面 164 的旋 向， 以利于移动环 70 的轴向移动及嵌合传力状态的稳定 （双保险）。 并且， 只 要附属导入齿侧面 164 与同一圓周部位的花键齿啮合面之间的夹角大于该两个 摩擦副联合自锁的摩擦角， 就可确保导入机构 M3的有效， 而不论其位于何处。 也就是说， 即便移动环 70相对固定环 50和第二转动构件 208的单独转动因摩 擦副分别自锁而不能轴向移动，但仍可被后两者之间的相对转动导入 /挤入， 即，
Figure imgf000010_0001
SSS同步离合 li中的 螺旋导向机构， 唯一地以移动环 70与第二转动构件 208之间的相对转动作其致 动原动力， 其仍具有如上所述的部分优点， 仍优于现有技术。 届时， 棘齿啮合 面 134平行于轴线 200, 传力齿 52、 72的横截面可以呈矩形， 传力嵌合机构 Ml 可无周向间隙，并且，为避免承受转拒，在传力嵌合机构 Ml完全嵌合复位之前， 棘轮式定点致动机构 M4须以轴向错位的方式先行失效。
显然， 本实施例的传力齿的布局可以有圓柱面、 圓锥面以及平端面多种形 式。 但不难理解， 无论是对于制造装配， 还是对于性能、 可靠性或寿命等， α = (3 = 90。 的平端面布局形式， 都要显著优于 α、 β取其它值时的锥面或圓柱 面的布局形式。 特别地， 当 α = 180° , β = 0° , 即传力嵌合机构 Ml的传力齿 均以直齿形式形成在圓柱面上， 啮合圓锥面具体为啮合圓柱面， 而所述导出机 构 M2和导入机构 M3又唯一地与花键齿副混合成为一个螺旋导向机构时，得到 的将是现有技术的 SSS 同步离合器。 即， 现有技术仅仅是本实施例 α或 β扩展 取值的特例， 而且是失去上述所有有益效果的特例。
另外很显然地， 本实施例中的第二转动构件 208 并不是必需的， 同时， 固 定环 50上的传力结构也可由轮齿 202改换成传力键槽或者皮带沟槽， 或者直接 将本实施例改换为如图 8所示的轴一轴传动结构形式。 并且更可仿照图 8单独 设置一个导入环 150 , 以拆分上述刚性合一的导向棘轮机构。 即， 在导入环 150 与固定环 50之间形成单独的棘轮式定点致动机构 Μ4， 在导入环 150与移动环 70之间形成单独的螺旋导入机构 Μ3 (详见实施例三的说明）， 如是， 定点致动 机构 Μ4还可以呈端面型棘轮机构的形式。
再有， 由于没有轴向强制压合力， 所以， 本实施例几乎没有分离阻力， 并 在超越工况中显然地不存在轴向间的摩擦接触及与之对应的空载转矩。 而在第 二转动构件 208的 Β部位径向地布置诸如弹簧滚珠定位机构， 以保持住移动环 70的轴向分离工位， 更可有效防止其与固定环 50的意外碰撞。
进一步地， 为减少棘爪 110 的磨损、 碰撞和噪音， 降低棘爪复位弹簧 114 的疲劳速度， 延长寿命， 还可以通过加入如图 4所示的阻挡环 140的方法来阻 止超越工况中棘爪 110 的嵌合复位， 相关说明参见实施例二。 另外， 引入图 8 所示的实施例三中的致动选择机构 Μ5 , 便可人为地失效定点致动机构 Μ4 , 从 而获得纯离合器工况。 具体说明参见实施例三。
还应该说明的是， 与现有技术一样， 可以通过增加传力齿和棘齿的圓周密 度来减小本实施例的溜滑角，从而直接减小定点致动机构 Μ4以及传力嵌合机构 Ml的啮合冲击。 而且， 增设阻尼机构 M9还可緩沖传力嵌合机构 Ml啮合时的 刚性冲击。 比如， 在第二转动构件 208的 A部位形成一个轴肩， 并以该轴肩与 移动环 70的花键基体环 76组成一个排油式环形封腔， 或者如图 4所示的多个 周向均布的圓柱活塞式封腔。 具体说明参见实施例二。
实施例二： 具有封装形式二的轮一轴传动式单向超越离合器 C2
对照图 1、 4, 超越离合器 C2是对超越离合器 C1的变形， 它径向上外翻了 后者的移动环 70及第二转动构件 208, 追加了棘爪保持机构 M8、 阻尼机构 M9 以及与固定环 50刚性一体的第一转动构件 206。
参看图 4 , 其阻尼机构 M9包括第二转动构件 208上的平行于轴线 200的阻 尼通孔， 可移动地嵌装于其中的柱状阻尼构件 226， 弹簧和螺纹堵头组件 228， 以及形成在阻尼通孔中段对应弹簧部位的供阻尼油或气进出的呼吸通孔。 显然 地， 本结构更适合于花键齿副具有导入功能的方案。
与现有技术的双方向阻尼不同，本实施例的阻尼机构 M9只有单方向阻尼功 能， 这显然有利于移动环 70的超越分离。 进入超越工况后， 柱状阻尼构件 226 在弹簧和螺纹堵头组件 228 作用下外伸直至顶到第二转动构件 208 的花键齿 212, 当然也可以顶在可选择地嵌装于其间的卡环式挡环上。 嵌合复位导入过程 的后半段， 阻尼构件 226通过花键齿 78对移动环 70的轴向移动实施阻尼， 直 至嵌合复位过程结束。
与超越离合器 C1稍有不同的是， 虽然定点致动机构 M4与导入机构 M3仍 混合成一个空间导向棘轮机构， 但其中的棘爪 110 已呈图 5所示的具有自回转 轴的普通摆动式结构， 而且还安装有将其可转动地约束在其位于第一转动构件 206上的座槽 112中的环状限制构件 130。 爪体 120头部的啮合面是导入齿侧面 154, 其基体 116的背面是半圓柱形回转面 118, 正面则是限位面 144a、 144b, 二面分别在两个自转极限位置上与限制构件 130 的内圆柱面贴合。 对应地， 座 槽 112具有相应的半圓柱形回转面 148,槽中容纳爪体 120的凹槽底部形成有布 置复位弹簧 114的径向型弹簧孔 115 , 参见图 10(a)。
棘爪保持机构 M8专门用于阻挡超越工况中棘爪 110的啮合，令其保持在分 离工位上， 以消除噪音和弹簧疲劳。 该机构包括弹性地膨胀在支撑壳 230 的轴 承孔中的开口弹性阻挡环 140, 以及棘爪 110上的摆动臂 122, 参见图 5、 6。 其 中， 摆动臂 122通过基体 116与爪体 120连成一体， 其头部正面为阻挡面 124。 阻挡面 124与阻挡环 140的阻挡工作面 142相贴合， 便可径向上有效阻挡棘爪 110与棘齿 132的啮合。 这里， 阻挡环 140是一个开口的弹性膨胀环， 其内圓柱 面上形成有避让缺口 143， 以提供非阻挡状态中摆动臂 122摆动的空间， 该缺口 一侧的阻挡工作面 142具有阻挡工作面外边缘 145。 贴合时， 阻挡面 124和阻挡 工作面 142相对径向线或者圓周切线均有一个偏斜角， 其最优地应能保证阻挡 过程中的两接触面之间摩擦副的自锁。 当然， 不自锁也行， 因为阻挡关系不会 瞬间消失， 尤其是大转速差的情况下， 代价是阻挡效果变差。
工作过程中， 爪体 120可绕自转轴摆动以实现与棘齿 132的啮合或分离， 同时， 摆动臂 122 随之作相应角度的摆动。 超越转动开始时， 在膨胀产生的摩 擦力的驱使下， 阻挡环 140会随着支撑壳 230按图 6中的箭头方向一同相对摆 动臂 122转动， 其上阻挡面工作 142有下压后者的趋势。 而当棘齿 132将棘爪 110径向压缩到最大极限的瞬间，倾斜的阻挡工作面 142将同步乘势转动到对摆 动臂 122实施最大压缩的极限位置上， 并与阻挡面 124建立起稳定的摩擦自锁 式阻挡关系。 此时， 由于阻挡环 140不可能在径向上进一步压缩摆动臂 122， 因 此， 阻挡环 140只能相对摆动臂 122静止不动， 而相对支撑壳 230滑转摩擦， 参见图 4、 6。
一旦进入嵌合复位的导入过程， 在膨胀摩擦力的驱使下， 阻挡环 140 必随 同支撑壳 230—起与移动环 70同步反超越转动， 在相对摆动臂 122反转过一个 很小的角度 Θ后， 阻挡状态便被完全解除， 摆动臂 122和棘爪 110恢复摆动自 由， 而在此时刻之前， 棘爪 110便可以部分地与棘齿 132啮合了。
由上述说明可见， 本实施例中弹簧的寿命的延长是以微小的摩擦阻力， 无 实质影响的机械磨损， 以及小于 Θ的嵌合角度损失为代价的。 其特别适用于任 意转速上都可以超越的离合器中。 当然， 如实施例——样， 舍弃阻挡环 140 而 借助离心力的方案也可以是一个选择。这只要将图 5中的摆动臂 122与爪体 120 分别布置在其自转轴的不同侧即可， 如图 7 所示。 超越转动后， 经过配重设计 的摆动臂 122上的离心力将迫使爪体 120克服弹簧 114的弹性复位力， 转动到 与棘齿 132不相接触碰撞的位置上。 另外， 摆动臂 122的离心力还可以通过这 样的方式获得， 即， 在座槽 112的容纳摆动臂 122的部位形成一径向盲孔， 再 置入一个诸如钢球之类的配重构件， 并以摆动臂 122 自然地封住该孔口。 另夕卜， 还可利用磁性力替代弹性复位力。 例如， 将磁性体置入棘齿 132 齿面之下或上 述径向盲孔中， 以直接吸引棘爪 110或与其成杠杆配置的摆动臂 122。
相比实施例一， 本实施例中花键齿副受到的压应力因半径较大而较小， 更 成倍低于现有技术的， 在显著降低花键齿的磨损强度， 提升其寿命的同时， 更 显著降低花键齿副轴向滑动时的故障率。
与实施例——样， 本实施例中导出机构 Μ2和导入机构 Μ3 , 也可分别或同 时以移动环 70与固定环 50之间的相对转动，以及移动环 70与第二转动构件 208 之间的相对转动为其致动原动力。
必须说明的是， 限制构件 130、 第二转动构件 208、 支撑壳 230等不是必需 的。 移动环 70可以直接输出转矩， 其内环面可以当作径向定位滑动轴承面， 其 花键齿 78可当作轮齿 202, 该轮齿 202呈直齿或斜齿圓柱轮齿均可。 而且， 导 出机构 Μ2和导入机构 Μ3也可以与该斜齿圓柱轮齿机构混合成一体。
另外， 阻尼机构 Μ9的阻尼腔设计成排油式或吸油式均可。 本实施例中， 在 径向尺寸增大至少一个花键齿 78的齿高后， 阻尼机构 Μ9也可以由柱销型阻尼 腔体结构改换为环状构件形式的环形阻尼腔体结构。
实施例三： 无封装的轴一轴传动式双向超越离合器 C3
参见图 1、 8 ,双向超越离合器 C3实质上是两个工作方向互反的单向超越离 合器 C1的有机叠加（去掉轴承 224和支撑壳 230 ),并增加了用于致使定点致动 机构 M4a或 M4b失效的致动选择机构 M5 , 用于规定离合器工作转动方向的换 向机构 M6， 控制该机构的换向驱动机构 M7， 以及非必需的棘爪保持机构 M8 和少量辅助构件。
除了为传递两个圓周方向的转矩而导致传力齿 52、 72分别具有两个平行于 轴线 200的传力齿侧面 54、 74, 以及嵌合状态下的周向自由度， 应大到不妨碍 导出机构 M2和导入机构 M3工作的程度外， 本实施例与单向超越离合器 C1二 者中的传力嵌合机构 Ml没有不同， 参见图 9、 10、 17。 参见图 9、 12, 导出机构 M2 包括附属导出环 100, 以及独立于固定环 50 的导出环 90。 导出环 90可滑转地嵌压在固定环 50内孔端面 56上， 其一端周向 均布有与传力齿 52齿数相等的导出齿 92， 该齿具有两个导出齿侧面 94。 附属 导出环 100与移动环 70的花键基体环 76共体， 其端面上周向均布有与传力齿 72齿数相等的附属导出齿 102, 该齿具有两个附属导出齿侧面 104。 其中， 导出 齿 92与附属导出齿 102间的轴向嵌合深度， 应最佳地大于传力齿 52与传力齿 72间的轴向嵌合深度。
本实施例的导入机构 M3 , 包括相互嵌合的导入齿 152和附属导入齿 162， 前者径向地形成在导入环 150内圓柱面上， 其具有两个螺旋形导入齿侧面 154， 后者对应地形成在附属导入环 160 外圓柱面上， 其具有两个螺旋形附属导入齿 侧面 164, 参见图 8、 9、 13、 17。 这里， 附属导入环 160与附属导出环 100形 成为一体。 每个附属导入齿齿槽 166的槽底都形成有附属导入齿槽入口 168。 其 中， 齿槽 166径向上并未贯通（也可贯通）。 为求得最佳工作效果， 还专门布置 有轴向限定导入环 150位置的支撑构件 136，该构件上的三个端面柱销可滑动地 穿过形成在移动环 70上的轴向贯通孔 82, 顶在支撑座环 138上。 支撑座环 138 被卡环 220b固定在第二转动构件 208上花键齿 212的端面上。
实际上，导出机构 M2和导入机构 M3就是对应于两个不同圓周方向的单向 机构分别综合的结果。 而且， 由于导出机构 M2和导入机构 M3受到的周向力都 很小， 因此， 即便只有一个导出齿或导入齿用于啮合， 也不易造成机械损害。 所以， 完全可以如实施例一中的说明， 仅仅布置 1 ~ 3个附属导出齿 102、 附属 导入齿 162或棘爪 110。 特别是在具有独立的导入环 150时， 由于其与定点致动 机构 M4的联动， 以及导入机构 M3中嵌合关系的恒久存在， 因此， 其导入齿更 不受数量和圆周等分点的约束， 参见图 17、 23(d)。
参见图 10、 11、 13、 17, 两个分别对应于第一、 第二转动方向的定点致动 机构 M4a、 M4b， 分别包括可相互啮合的棘爪 110a、 110b以及双向的棘齿 132。 其中，棘爪 110a、 110b以圓周朝向互反的方式分别嵌装在位于固定环 50内孔圓 柱面上的座槽 112a、 112b中， 其爪体 120头部的啮合面平行于轴线 200 , 不再 具有导向作用。与实施例二相同，棘爪的安装座槽 112具有半圓柱形回转面 148, 容纳爪体 120的凹槽底部的弹簧孔 115 内布置有棘爪复位弹簧 114。 周向槽 58 贯穿座槽 112a、 112b的半圓柱形凹槽部分， 嵌装在其中的开口弹性环状的限制 构件 130通过限位面 144a、 144b将棘爪 110a、 110b可摆动地限制在座槽 112a、 112b中。 导入环 150的外圓柱面呈双向棘轮状， 其周向均布的棘齿 132与传力 齿 52的齿数相等， 该齿双侧的啮合面 134a、 134b与对应棘爪 110a、 110b的接 触 , 可分别停止住导入环 150相对固定环 50沿第二或第一转动方向上的转动。 在该两个停止位置上， 导入机构 M3可致使轴向分离状态中的传力嵌合机构 Ml 复位到正确的嵌合工位。 显然， 如稍微增加点轴向长度，棘齿啮合面 134a、 134b 也可分别形成在轴向上相互错开的圓周朝向互反的单向棘齿 132a、 132b上。
由上述说明可见， 尽管本实施例的导入机构 M3以移动环 70与导入环 150 之间的相对转动为其致动直接动力， 但致动时的导入环 150 是借助定点致动机 构 M4与固定环 50—体转动的， 即， 导入机构 M3的致动原动力实际上仍然是 移动环 70与固定环 50之间的相对转动。
如图 11、 13、 14所示， 本实施例与实施例二中的棘爪保持机构 M8实质相 同， 且结构上前者等于两个后者的简单叠加。 其中， 阻挡环 140 仍为开口弹性 环， 该环弹性收缩地套装在导入环 150的外圓柱面 158上， 其外缘面上的两组 阻挡工作面 142a 与 142b 的圓周朝向互反， 以交替阻挡两个对应方向的棘爪 110a, 110b上的摆动臂 122a、 122b。
作为本实施例关键机构之一的致动选择机构 M5,其包括致动选择环 170和 状态爪 126, 状态爪 126上形成有弧状避让缺口 128, 参见图 8、 11、 12。 致动 选择环 170刚性一体地形成在导出环 90的外圓柱面上， 其外缘面上形成有两个 避让缺口 178a、 178b, 且分别具有两个圓周朝向互反的凸轮面 176a、 176b。 避 让缺口 178a、 178b间的圓周夹角与棘爪 110a、 110b间的圓周夹角相差一个换向 角度值 ε。 于是， 致动选择环 170相对固定环 50转动 ε角的直接结果， 就是互 换始终互反的状态爪 126a、 126b的状态。 以图 12的逆时针转动为例， 凸轮面 176a径向上完全顶起状态爪 126a，令与后者刚性一体的棘爪 110a长久地处于分 离工位不能摆动， 对应于第一转动方向的定点致动机构 M4a因此失效， 等同于 不存在。 同时， 凸轮面 176b撤除对状态爪 126b的径向阻挡， 令与后者刚性一 体的棘爪 110b恢复摆动自由，对应于第二转动方向的定点致动机构 M4b因此恢 复有效。 所以， 双向超越离合器 C3只在一个工作转动方向上具有导入能力， 工 作如纯粹单向超越离合器。
作为本实施例关键机构之二的换向机构 M6， 是包括固定环 50、 导出环 90 和摆杆 180的摆动导杆机构， 参见图 8、 10、 12、 15。 摆杆 180嵌装在位于固定 环 50非嵌合端的环形槽 60底部的扇形槽 62中， 其一端的中心销 182可转动地 嵌入回转孔 64 内， 其另一端的圓柱换向销 186穿过固定环上的环形通孔 66可 滑动地嵌入导出环 90无齿端面上倾斜状的换向导槽 196中。 在环形通孔 66允 许的径向区间内， 自转的摆杆 180， 可通过其换向销 186与换向导槽 196间的导 向滑动， 驱使导出环 90相对固定环 50转动换向所需的圆周角度 ε。 显然， 致 动选择机构 Μ5也被驱动着同步完成一次相应的致动选择。本实施例中，对应于 第一转动方向的第一相对位置如图 17所示， 导出环 90相对固定环 50向上转动 圓周角度 ε , 便位于对应第二转动方向的第二相对位置。
另外， 本实施例还最佳地布置有盘形凸轮式换向驱动机构 Μ7, 其包括固定 环 50、 摆杆 180以及驱动环 190, 参见图 8、 10、 15、 16。 驱动环 190可转动地 嵌装在固定环 50非嵌合端的环形槽 60内， 被嵌装于环形槽 60外端卡环槽 204 中的卡环 220a轴向定位的同时，也将摆杆 180的杆状基体 184限制在扇形槽 62 中。 驱动环 190环形平面上的凸轮槽道 192可滑动地径向约束住摆杆 180上的 圓柱驱动销 188 , 其内缘处形成有用以对其施加旋转力的端面凸齿 191。 其中， 凸轮槽道 192 包括位于中间的倾斜状凸轮驱动段 198, 以及位于两端的以轴线 200为曲率中心但半径不同的弧状停止段 194a、 194b三部分。 这样， 驱动销 188 不可能通过停止段 194a, 194b反过来推动驱动环 190转动， 其间摩擦副可以实 现自锁， 而驱动销 188在该两段内依次停留时的径向位置， 就是摆杆 180将导 出环 90依次停留在第一和第二相对位置时的径向位置。 于是， 当驱动环 190相 对固定环 50转动的圓周角大于 λ时 , 也就是至少转过凸轮驱动段 198所对应的 圓周角时， 驱动销 188就会在停止段 194a、 194b之间完成一次滑动换位。 其后， 只要驱动环 190 不再转动， 例如， 借助诸如弹簧滚珠之类机构的圓周限定或定 位， 摆杆 180就不可能自转， 自然地， 与其相关的换向驱动机构 M7、 换向机构 M6以及致动选择机构 M5的工作定位便得以锁定， 双向超越离合器 C3的单向 工作状态便得以稳定。
如上所述，双向超越离合器 C3可被控制为按第一或第二转动方向工作的单 向超越离合器。 例如， 当图 16(a)中的驱动环 190相对固定环 50顺时针转动， 将 驱动销 188径向推升至停止段 194a并予以锁定， 换向驱动机构 M7便驱动换向 机构 M6和致动选择机构 M5 , 将双向超越离合器 C3的工作方向定位在第一转 动方向上， 并失效定点致动机构 M4b (图 12 )。 此时， 导出环 90相对固定环 50 位于如图 17所示的第一相对位置， 相关构件中只有与该方向对应的 a系列构件 或特征部位有效， 而相反的 b 系列构件或特征部位失效并等同于不存在一样。 即， 等效于导出环 90与固定环 50组合体的端面上布置有一圈具有传力齿侧面 54a和导出齿侧面 94a的锯齿形齿， 移动环 70端面上布置有一圏具有传力齿侧 面 74a和附属导出齿侧面 104a的锯齿形齿， 完全等效于单向超越离合器 Cl。 因 此， 无需重复说明其工作过程， 可参看图 17。 但应该指出的是， 因为导入齿 152 始终嵌合在附属导入齿 162的齿槽中， 所以， 除导出和导入过程外， 导入环 150 与移动环 70始终同步转动。
对应地， 当图 16(a)中的驱动环 190相对固定环 50逆时针转动， 将驱动销
188径向压低至停止段 194b并予以锁定，换向驱动机构 M7便驱动换向机构 M6 和致动选择机构 M5 ,将双向超越离合器 C3的工作方向定位在第二转动方向上， 并失效定点致动机构 M4a。 此时， 导出环 90相对固定环 50位于第二相对位置， 相关构件中只有与该方向对应的 b 系列构件或特征部位有效， 而相反的 a系列 构件或特征部位失效并等同于不存在一样， 参见图 17。 即， 等效于导出环 90与 固定环 50组合体的端面上布置有一圈具有传力齿侧面 54b和导出齿侧面 94b的 锯齿形齿， 移动环 70布置有一圈具有传力齿侧面 74b和附属导出齿侧面 104b 的锯齿形齿， 完全等效于反向工作的单向超越离合器 Cl。
由于本发明同一时刻仅存在单向超越分离情况， 没有牙嵌式自由轮差速器 那样的双侧移动环 70同时互反超越分离的情况， 其齿槽宽度相对缩小， 传力齿 增加一个相应于导出齿侧面 94的厚度， 因此， 双向超越离合器 C3转矩传递能 力是后者的约 1.3倍， 类比实施例一的传力齿外径为 155毫米的情况， 其计算转 矩约为 20,384牛米。 可见， 即便是双向传动， 根据本发明的承载能力也成倍高 于单向传动的现有技术及其产品， 而且， 与实施例一中所述相同， 仍具有诸如 高转速大功率等几乎完全相同的优异性能和特点， 以及结构、 工艺、 装配、 操 作和维护简单的优点， 尽管增加了方向和状态控制机构。
实际上， 本实施例除具有双向超越离合器基本功能外， 还可再具有其它可 选择的有益工况。 例如， 按图 16(b)控制方式所实现的纯离合器和双向联轴器工 况。 其中， 致动选择环 170与驱动环 190混合成一个被直接操纵 (当然也可以 是不佳的非联合的单独操纵） 的环， 其加宽的避让缺口 178a、 178b形成在驱动 环 190外圓面上， 同时， 棘爪 110上的状态爪 126及座槽 112均轴向延伸至驱 动环 190外圓面 （安装时棘爪 110由固定环 50的端面轴向嵌入）。 图 16(b)给出 的工位对应于联轴器工况，此时摆杆 180的驱动销 188正好位于凸轮驱动段 198 的周向中点， 且定点致动机构 M4a、 M4b同时有效。 由于该段的周向宽度足以 保证驱动销 188在其中可分别达到其在停止段 194a, 194b上对应的径向高度， 因此，只要保持住驱动环 190此时的位置，超越离合器 C3便不能固定导出环 90, 只有导入能力而没有导出能力， 从而转入双向牙嵌式联轴器工况。 而适当转动 驱动环 190, 就可失效定点致动机构 M4a、 M4b 中的一个， 从而转入对应于第 一或第二转动方向的单向超越离合器工况。 所以， 在任一工作方向上的荷载状 态下， 回转驱动环 190便可工作在联轴器工况下， 而在进入超越工况后， 继续 转动驱动环 190,借助凸轮面 176c、 176d便可失效唯一有效的定点致动机构 M4a、 M4b, 以取消所有导入功能， 致使超越离合器 C3进入纯离合器工况。
很显然， 上述三工况间的转换极为简单、 快捷和可靠， 并且， 失效定点致 动机构 M4 以实现纯离合器的思想也可以用于单向超越离合器中， 在此不再详 述。 另外， 联轴器工况也可通过适当加大导出机构 M2的周向自由度， 并对应地 在如图 16(a)所示的凸轮驱动段 198的径向中点处加入一个中点停止段， 将凸轮 槽道 192由两级台阶状变成三级台阶状的方式实现。
现实中， 上述工况特別适用于和有利于诸如大型舰船中的双动力驱动系统 的传动轴系。 即， 在完成动力机的接力置换后， 将荷载状态的超越离合器 C3置 为联轴器工况， 可消除其在倒车或风浪等情况中产生有害分离的可能； 将超越 工况中非荷载状态的超越离合器 C3置为纯离合器工况，可对置换出来的动力机 进行不受限制的维修和调试。 当然， 为做到万无一失， 还可在移动环 70外圓面 与机架间布置一个与换向驱动机构 M7 联动的或单独控制的诸如挡肩或滑环式 机构， 轴向限定住移动环 70以防意外分离、 嵌合或碰撞。
另夕卜，单独控制换向驱动机构 M7或者令其与机动车辆的制动机构联动，本 实施例还可用作状态可控的双向滑行器。 参见图 16(b), 当超越滑行中的车辆制 动或者人为特意控制时， 双向滑行器便转入反向传力工况或联轴器工况， 致使 反向的或全部的定点致动机构有效， 滑行器即刻临时性地嵌合复位以结束滑行 并反向荷载， 车辆发动机开始提供制动力。 车辆再次驱动行使时， 滑行器将立 即自动回复到先前设定的工作方向上。 如果取消倒车方向的超越功能， 例如取 消图 16(b)中的停止段 194b, 得到的就是单向滑行器。 关于滑行器的结构、 换向 以及操作等的更进一步的说明，可参阅本发明人的申请号为 200710152152.3 (压 合式牙嵌超越离合器） 的中国在审发明专利， 该专利申请的全部内容及构思以 参引方式包含在本专利申请中。
不难理解， 导入机构 M3或定点致动机构 M4中固定环 50的作用可由导出 环 90替代， 附属导出环 100也可以附属到固定环 50上，致动选择机构 M5失效 定点致动机构 M4的方法也不止上述一种，轴向错开棘爪和棘轮， 或以两单独棘 轮机构与导入机构 M3 依次相联等就是可行方案。 同样， 径向抬升状态爪 126 的机构也不止盘形凸轮一种形式， 如， 在驱动环 190 上布置端面凸轮， 以圓柱 凸轮机构的形式压缩一带楔形头部的可轴向弹性复位的选择杆 （实质就是在轴 平面内分别布置两个半径为无穷大的致动选择环 170 ), 该头部同样可以径向抬 升状态爪 126。 状态爪 126及座槽 112也不必需轴向延伸至驱动环 190外圓面， 致动选择环 170也不止联动一种控制形式。 例如， 致动选择环 170可以是形成 有端面凸销的套装在导出环 90外的独立环， 该凸销可转动地穿过固定环 50上 相应的环状的通孔， 与驱动环 190 或其它控制环外缘对应的凹槽嵌合以实现两 者的周向固定， 于是， 致动选择机构 M5可以得到的单独的控制。
应该指出的是， 对换向驱动机构 M7的操纵， 或者说对驱动环 190的旋转， 可以于停转状态下实施， 可以借助本发明人的 "相对运动方向传感装置 （参见 200810080503.9专利中的相关说明）" 于工作中自适应地进行（此时， 只要令设 定转动方向总是相反于实际转动方向， 也可得到纯离合器工况）， 也可以人为施 加周向摩擦阻力的方式进行， 还可由驱动环 190和一个与固定环 50周向固定但 轴向滑动的控制环组成转动导向机构， 以轴向移动该控制环的方式进行。 另外， 通过控制凸轮驱动段 198、换向导槽 196二者相对径向线的夹角的方向以及 λ与 ε比值的大小， 还可实现驱动转动方向、 换向转动方向与目标工作转动方向间 的正对应或反对应关系， 及控制换向的速度和灵敏度。 不难理解， 上述所有工况实际上不过是通过联合或单独控制得到的定点致 动机构 M4a、 M4b分别有效或失效， 导出环 90分别定位在第一、 第二相对位置 或自由位置的排列组合结果的一部分。 而且， 如果去掉其中的摆杆 180 和换向 驱动机构 M7， 以固定环 50、 导出环 90分别背靠背刚性一体的方式轴向双联两 个本实施例， 就会得到一种新型的定点嵌合的牙嵌式自锁差速器。 当然， 所有 双向方案中， 移动环 70与第二转动构件 208间应最佳地不具有导向作用。
必须说明的是， 对于换向机构 M6和换向驱动机构 M7 , 本发明几乎未作具 体限制， 它可以是诸如机械、 液压或电磁等机构中的任何一种， 而且， 公知技 术中已有大量实施例可供选择、 变形或组合。 例如， 本发明人提出的申请号为 200710152152.3及 200810080503.9的两项在审中国发明专利中就有许多相关方 案。 因此， 该两项专利申请的全文被引用在此， 不再详细说明。 而且， 通过选 择性地失效定点致动机构 M4a、 M4b , 可轻易得到接合精准的单、 双向导向式 牙嵌离合器， 或者导向式牙嵌电控离合器， 其性能、 结构和使用条件等均显著 优于现有技术中对应的牙嵌离合器或牙嵌式电磁离合器。
本发明中，换向机构 M6并不局限于周向定位的全齿导出环一种方案。例如， 还可采用诸如图 18所示的轴向定位的半齿导出环方案。 其中， 导出环 90a、 90b 上的导出齿 92a、 92b仅为图 12中的一半， 分别具有一个圓周朝向互反的导出 齿侧面 94a、 94b。 相互套装的导出环 90a、 90b与固定环 50三者相互间周向固 定轴向滑动（通过诸如轴向销孔式嵌合机构、滑键联接机构），二环与驱动环 190 间形成有诸如槽道式圓柱凸轮换向机构， 以实现二导出环轴向上交替伸出的目 的。 即， 当导出环 90a处于图 18(a)所示的第一相对位置时（等同于图 17中的位 置）， 导出环 90b必是轴向缩回隐藏的， 超越离合器 C3工作在第一转动方向上。 而转动驱动环 190 , 导出环 90a轴向缩回隐藏的同时， 导出环 90b必伸出至图 18(b)所示的第二相对位置， 超越离合器 C3工作在第二转动方向上。 明显地， 图 18所显示的换向关系较图 17所示易于理解， 且对应方案更便于电磁控制， 以及 更易于得到纯离合器工况（两导出环同时伸出， 定点致动机构 M4全失效）和联 轴器工况 （两导出环同时缩回， 定点致动机构 M4全有效）。
如前所述， 阻挡环 140和导入环 150都不是必需的。 另外， 为缩小轴向尺 寸， 以及方便附属导入齿 162的加工制作， 可以将附属导入环 160制作成单独 的开口弹性环， 借助其端部的轴肩与移动环 70上相应周向槽的配合实现轴向固 定， 再借助同时贯穿于二者内孔的第二转动构件 208实现周向固定。
实施例四： 具有封装形式二的轮一轴传动式双向超越离合器 C4
参见图 19 ~ 22 , 本实施例采用了实施例二所示的封装形式和实施例三的主 要结构， 并省去了棘爪保持机构 M8和独立导入环 150。 其微小改动之处在于， 导出机构 M2中的附属导出齿 102与传力齿 72连成径向一体。 并且同于实施例 一，导入机构 M3与定点致动机构 M4又分别混合成为两个对应于不同圓周方向 的空间导向棘轮机构， 导入齿侧面 154再次成为棘爪 110的爪头啮合面， 形成 为一体的双向棘齿 132与附属导入齿 162周向均布在移动环 70的内孔面上， 且 与传力齿 72的数量相等。 如实施例三中的说明， 本实施例的导入机构 M3仍以 移动环 70与固定环 50之间的相对转动为其致动原动力。 另外， 环状的限制构 件 130上形成有用以避让棘爪 110爪体 120的周向通孔， 在径向限定棘爪 110 的同时还轴向限定导出环 90。 致动选择环 170一体形成在导出环 90内径侧。 以 及，对调了换向机构 M6中换向导槽 196和换向销 186的位置， 以利于在导出环 90上不便于布置换向导槽 196的情况。 相应地， 固定环 50上的扇形槽分成台阶 状的 62a、 62b两部分， 槽底的环形通孔 66以轴线 200而不再以回转孔 64的轴 心为其曲率 /回转中心。
图 23(a) ~ (c)示出了其在第一转动方向上传递转矩、 超越分离和嵌合复位三 种工作状况。 其中清楚地示出了导向棘爪 110 的工作过程， 并且显然地， 本实 施例也可以为棘爪 110添加轴向支撑构件 136， 以及阻挡环 140， 还可以按图 8 的思想加入导入环 150以将导入机构 M3与定点致动机构 M4分开。 如图 24所 示， 导入环 150内为双向棘轮， 外为双向导入齿 152, 与其对应的图 23(d)对导 入机构 M3的工作机理揭示得最为清晰，该机构实质上就是轴向反装或反对应的 导出机构 M2。 对比图 23(a) ~ (c)和图 23(d)不难发现， 加入导入环 150将有利于 提升导入机构 M3的可靠性， 毕竟， 相对于单向， 用作棘齿 132的双导附属导入 齿 162所能提供给棘爪 110的啮合机会已明显降低。
工业适用性
本发明可直接应用于儿乎所有机械传动领域， 尤其是除直接分度外的超越 和逆止应用场合， 具备通用于几乎所有转速、 所有转矩和所有功率的传动能力。 例如： 液力变矩器， 自动变速箱， 脉动式无级变速器， 起重机械及其它机械中 的逆止装置， 各类作物收割 /获机， （高压开关）真空断路器， 汽轮发电机组， 动 力机起动装置 （可轻易满足起动装置和动力机飞轮恒久啮合的现实需求和使用 要求， 以彻底抛弃电磁开关）， 大型水面舰船， 双发动机直升机， 轮式机动车辆 的防滑转的有限差速比差速器、 滑行器等等。
以上仅仅是本发明针对其有限实施例给予的描述和图示， 具有一定程度的 特殊性， 但应该理解的是， 所提及的实施例都是用来进行说明的， 其各种变化、 等同、 互换以及更动结构或各构件的布置， 都将被认为未脱离开本发明构思的 青神和范围。

Claims

权利要求
1. 一种导向式牙嵌超越离合器， 包括：
用于传递转矩的传力嵌合机构， 其具有绕同一回转轴线转动且分别形成有 传力齿的固定环和移动环；
至少一个导出机构， 其至迟在所述移动环与所述固定环二者间开始超越转 动时， 致使所述移动环轴向远离所述固定环， 以将所述移动环的传力齿从所述 固定环的传力齿齿槽中导出， 从而解除所述传力嵌合机构的轴向嵌合；
至少一个导入机构， 其至迟在所述移动环与所述固定环二者间开始反超越 转动时， 致使所述移动环轴向移向所述固定环， 以将所述移动环的传力齿导入 到所述固定环的传力齿齿槽中， 从而恢复所述传力嵌合机构的轴向嵌合；
至少一个定点致动机构， 其用于所述反超越转动过程中， 在所述移动环相 对所述固定环转动至特定的圓周位置时致动所述导入机构， 以使该导入机构完 成所述导入运动；
其特征在于：
所述传力嵌合机构嵌合时， 其全部传力齿的实际啮合表面的啮合中线至少 大体上位于同一个啮合圓锥面上。
2. 按权利要求 1所述的超越离合器， 其特征在于：
( a ) 所述啮合圓锥面与所述回转轴线之间的夹角大于 0° , 小于 180° ；
( b )还包括一个与所述移动环以耦合和啮合中的一种方式可轴向滑动地相 连接的第二转动构件；
( c ) 所述导出机构和所述导入机构数量上均只有一个， 该二机构合并成一 个传力螺旋机构， 并以所述移动环与所述第二转动构件之间的相对转动为其致 动原动力。
3. 按权利要求 2所述的超越离合器， 其特征在于：
( a ) 所述定点致动机构是棘轮机构， 其棘齿和棘爪分别与所述固定环及所 述移动环二者中的一个至少间接地周向固定； 所述传力嵌合机构轴向嵌合时的 相对移动， 致使所述定点致动机构以所述棘爪和所述棘齿轴向错位的方式失效；
( b ) 所述传力螺旋机构是螺旋花键齿机构。
4. 按权利要求 1所述的超越离合器， 其特征在于：
( a ) 所述啮合圓锥面与所述回转轴线之间的夹角大于等于 0 小于等于 180。 ； ( b )所述导出机构和所述导入机构中的至少一个机构， 以所述移动环与所 述固定环之间的相对转动为其致动原动力；
( C ) 所述传力嵌合机构嵌合时的入口裕度大于零。
5. 按权利要求 4所述的超越离合器， 其特征在于：
( a ) 所述定点致动机构是棘轮机构， 其棘齿和棘爪分别与所述固定环及所 述移动环二者中的一个至少间接地周向固定；
( b )至少一个所述导出机构， 其包括所述固定环与所述移动环， 并与所述 传力嵌合机构混合成一个嵌合机构； 所述固定环沿一个圓周方向的转动致使所 述移动环一体转动， 而沿相反圆周方向的转动则致使双方相对转动， 并导出所 述移动环以解除轴向嵌合关系； 即， 所述二环上的传力齿的两个齿侧面分别具 有传递转矩和转动导向的功能；
( C ) 至少一个所述导入机构， 其包括所述固定环与所述移动环， 并与所述 定点致动机构混合成一个导向棘轮机构， 即， 所述棘爪和所述棘齿的啮合面与 所述回转轴线不平行， 且具有转动导向功能。
6. 按权利要求 4所述的超越离合器， 其特征在于：
( a ) 所述导入机构及所述定点致动机构数量上均为两个， 分别对应于两个 不同的圓周转动方向， 所述传力嵌合机构可传递两个不同圓周方向的转矩； ( b )所述导出机构是包括附属导出环和至少一个导出环的绕所述回转轴线 转动的轴向嵌合机构， 其嵌合双方都形成有导出齿； 在两个不同圓周转动方向 上， 所述导出机构均具有所述功能； 所述附属导出环与其属主环形成为一体， 该属主环是所述传力嵌合机构嵌合双方中的任意一方构件， 导出过程中， 所述 导出环和所述传力嵌合机构中与所述属主环相对的一方构件至少具有轴向上的 单向限定关系；
( C ) 所述导入机构和所述导出机构， 均以所述移动环与所述固定环之间的 相对转动为其致动原动力；
( d )还包括换向机构， 其用于变换所述导出环相对所述固定环的位置， 以 规定所述超越离合器传递转矩和超越转动的方向；
( e )还包括至少一个致动选择机构， 其用于致使特定的所述定点致动机构 失效。
7. 按权利要求 6所述的超越离合器， 其特征在于：
( a ) 所述换向机构包括所述固定环和所述导出环， 该机构可以将所述导出 环分别变换到相对所述固定环的二个不同的特定位置上， 在该两个位置上， 所 述超越离合器分别具体为对应于两个不同圓周方向的单向超越离合器；
( b )所述致动选择机构是凸轮机构， 其包括形成有凸轮轮廓面的致动选择 环， 以及与所述棘爪形成为一体的状态爪， 该凸轮机构可驱动所述棘爪绕其自 身的转动轴线转动一定角度， 以致使所述棘爪失去与对应棘齿相啮合的能力； ( C )还布置有换向驱动机构， 其用于驱动所述换向机构运动， 以实现所述 换向机构的所述功能和锁定所述换向机构的运动状态。
8. 按权利要求 7所述的超越离合器， 其特征在于：
( a ) 所述定点致动机构是棘轮机构， 其棘齿和棘爪分别与所述固定环及所 述移动环二者中的一个至少间接地周向固定；
( b )对应于同一圓周转动方向上的所述导入机构与所述定点致动机构合并 成一个导向棘轮机构， 即， 所述棘爪和所述棘齿的啮合面与所述回转轴线不平 行， 且具有转动导向作用。
9. 按权利要求 4、 6或 7任一项所述的超越离合器， 其特征在于：
( a )还包括至少一个导入环， 该环位于所述固定环与所述移动环之间， 绕 所述回转轴线转动；
( b )所述定点致动机构是棘轮机构， 其棘齿和棘爪分别与所述固定环及所 述导入环二者中的一个至少间接地周向固定； 入环。
10. 按权利要求 9所述的超越离合器， 其特征在于： 所述两个导入机构合并 成一个轴向嵌合机构， 其所述导入环和所述附属导入环上均形成有导入齿， 该 导入齿具有两个分别对应于不同圓周方向的导向面。
11. 按权利要求 7、 8任一项所述的超越离合器， 其特征在于： 所述换向机 构驱动所述致动选择机构运动， 以实现所述致动选择机构的所述功能。
12. 按权利要求 4 ~ 7任一项所述的超越离合器， 其特征在于： 所述传力嵌 合机构， 所述导出机构， 所述定点致动机构三者的各组成构件中， 至少各有一 方构件的所述传力齿， 所述导出齿， 以及所述棘齿或所述棘爪， 其数量分别等 于同一自然数， 并布置在按该自然数所等分的圓周等分点上。
13. 按权利要求 8所述的超越离合器， 其特征在于: ( a )还包括一个与所述移动环可轴向滑动地耦合的第二转动构件；
( b )所述棘爪受到与所述第二转动构件具有轴向限定关系的一转动构件的 轴向支撑。
14. 按权利要求 9所述的超越离合器， 其特征在于：
( a )还包括一个与所述移动环可轴向滑动地耦合的第二转动构件；
( b )所述导入环受到所述第二转动构件或与该构件具有轴向限定关系的一 转动构件的轴向支撑。
15. 按权利要求 7、 8任一项所述的超越离合器， 其特征在于： 所述换向驱 动机构是凸轮机构。
16. 按权利要求 15所述的超越离合器， 其特征在于： 所述致动选择环与所 述凸轮机构的凸轮形成为一体。
17. 按权利要求 7所述的超越离合器， 其特征在于： 所述换向机构是摆动导 杆机构。
18. 按权利要求 6 ~ 9任一项所述的超越离合器， 其特征在于： 所述导出机 构的轴向嵌合深度， 大于等于所述传力嵌合机构的轴向嵌合深度。
19. 按权利要求 4、 7 ~ 9任一项所述的超越离合器， 其特征在于： 嵌合复位 过程中， 所述移动环的轴向移动受到一个阻尼机构的制约， 以降低其与所述固 定环之间的冲击力。
20. 按权利要求 7、 8任一项所述的超越离合器， 其特征在于： 还包括相对 运动方向传感装置， 其用于所述超越离合器的工作转动方向发生改变时， 自适 应地操纵所述换向驱动机构以实现其所述驱动功能。
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