Abstract:
In a continuously-variable-ratio-device of the type having races between which drive is transferred by rollers which are movable in accordance with changes of variator ratio, it is necessary to apply a traction load urging the rollers and races into engagement. Each rollers is also subject to a transverse reaction force by a hydraulic reaction roller actuator receiving a controlled reaction pressure. It is desirable to create a relationship between traction and reaction force, and also to provide for adjustment of this relationship. In the present invention this is achieved hydraulically. A traction pressure related to the reaction pressure is applied to an actuator which creates the traction load. The hydraulics also include a working chamber, which is selectively connectable to and disconnectable from either (or both) of (i) the reaction pressure and (ii) the traction pressure, the traction force being dependent upon pressure in the working chamber.

Description:
CROSS-REFERENCE TO OTHER APPLICATIONS 
     This is a National Phase of International Application No. PCT/GB2004/002500, filed on Jun. 15, 2004, which claims priority from Great Britain Patent Application No. 0317499.2, filed on Jul. 25, 2003. 
     FIELD 
     The present invention relates to control of traction load in a continuously variable transmission device (“variator”) of rolling traction type. 
     BACKGROUND 
     In a rolling traction variator, drive is transmitted through at least one roller (and more typically a set of rollers) running upon at least one a pair of rotary races. To provide traction between rollers and races they must be biased into engagement with each other. The biasing force is referred to herein as the “traction load”. In known variators the rollers and races do not make contact with each other since they are separated by a thin film of “traction fluid”. It is shear of this fluid which, given suitably high pressure, provides the requisite roller/race traction. 
     Control of traction load is important to variator performance. One reason for this is that energy losses taking place at the roller/race interface vary with traction load, which thus has a bearing on variator efficiency. These losses are due to (1) spin at the interface—i.e. rotation of one surface relative to the other, due to the fact that the two surfaces are following circular paths about different axes and (2) shear at the interface—i.e. speed difference between the two surfaces, producing the shear in the fluid. It is found that excessively high traction loads increase spin, losses while low traction loads lead to high shear losses, optimal efficiency lying between the two extremes. 
     Another reason why traction load control is important is that excessive slip of rollers relative to the races, in response to inadequate traction loading, can result in failure of the variator and damage to it. 
     It is known to vary traction load in sympathy with “reaction force”. To explain first of all what reaction force is, consider that due to the torque being transmitted the races exert a tangential force upon each of the rollers. This force must be reacted back to the transmission casing. In known rolling traction variators the rollers are typically movably mounted and the force exerted by the races is opposed by, and reacted to the casing through, an actuator acting upon the roller&#39;s mountings. The reaction force applied by the actuator is adjustable for the purpose of controlling the variator and is equal but opposite to the tangential force exerted by the races. 
     The variator&#39;s traction coefficient μ can be defined as follows 
             μ   =     RF   TL           
where TL is the traction load and RF is the reaction force. This is strictly a simplification, since the true coefficient of traction at any chosen roller/race interface depends upon the magnitude of the forces perpendicular and parallel to the interface, and the traction load is not generally perpendicular to the interface. However this simple definition will suffice for the present discussion.
 
     Variators are known in which traction load is varied along with reaction force to provide a constant traction coefficient. Reference is directed in this regard to Torotrak&#39;s European patent EP 894210 wherein reaction force is provided by double acting hydraulic roller actuators and the two pressures at these actuators are also led to a hydraulic traction load actuator. The hydraulic coupling of roller and traction load actuators is advantageous because it allows the traction load to be very quickly varied along with reaction force. This is important in responding to “torque spikes”—rapid fluctuations in transmission torque occurring for example upon emergency braking of the vehicle. A torque spike produces a rapid change in reaction force which could lead to slip between rollers and races, were if not for the fact that, in the known arrangement, increased pressures which are created in the roller actuators are passed on to the traction load actuator to correspondingly increase traction load with little time lag. 
     To increase still further the speed of response of the traction load to the reaction force, Torotrak&#39;s International Patent Application PCT/GB02/01551, published under No. WO 02/079675, teaches how pressure from the roller actuators can be used to control a pilot operated valve which in its turn controls application of fluid from a high pressure source to the traction load actuator. The same document recognises the desirability of adjusting the traction coefficient and provides some ways in which this can be achieved. 
     One reason for adjusting the traction coefficient (as opposed to maintaining, so far as possible, a constant ratio of traction load to reaction force) is that the properties of the traction fluid, and consequently the appropriate traction coefficient, vary with temperature. It is also desirable to adjust μ with variator rolling speed and with variator ratio. 
     WO 02/079675 suggests that traction coefficient adjustment can be carried out by applying an adjustable force to the spool of the pilot operated valve using a solenoid. The effect is to add an offset to the traction load so that 
             TL   =       RF   μ     -   OF           
where μ is the traction coefficient which would be obtained without the solenoid force and OF is the offset produced by the solenoid force. It will be apparent that the ratio of traction load to reaction force varies as the magnitude of the reaction force varies and this is undesirable, particularly because an offset which produces an appropriate traction coefficient at high reaction force/traction load produces, at much lower levels of reaction force, a large error in traction load. Inaccuracy in this large offset could, furthermore, potentially result in the traction load being too small when the reaction force is low.
 
     Rather than adding an offset to the traction load, it would be desirable to provide for adjustment of the traction coefficient itself, so that:— 
             TL   =     RF     (     μ   +   K     )             
where μ is once more the traction coefficient which would be provided in the absence of the adjustment and K is an adjustment determined by the control electronics associated with the variator. Given this relationship, changes in reaction force RF do not produce discrepancies in traction load. The desirability of this type of traction load control was recognised in WO 02/079675 but devising a practical hydraulic arrangement for achieving the relationship is problematic. That document did show one possible circuit which used a series pair of flow restrictors, one of which was variable, in a manner analogous to a potential divider in an electrical circuit, to modify the traction load pressure. This arrangement introduces certain problems of its own, particularly as it relies on a continuous flow of pressurized fluid out of the hydraulics, adding to the burden placed upon the associated pump or pumps.
 
     The desirability of adjusting traction coefficient has also been recognised in U.S. Pat. No. 6,162,144, assigned to General Motors Corporation. However the circuit drawn in that patent simply uses a pulse width modulated valve to feed a percentage of the end load pressure to a second chamber of the traction load actuator, working in opposition to the main traction load pressure, to thereby adjustably modify the reaction load. It is believed that this would not provide a practical system capable of reacting with sufficient speed to rapid reaction force changes, the bandwidth of the pulse width modulated valve being too low. 
     SUMMARY 
     An aim of the present invention is to provide an improved means of controlling traction load, making provision for adjustment of the relationship between reaction force and traction load. 
     In accordance with the present invention, there is a hydraulic control arrangement for a variator of the type having a pair of races, at least one roller which is arranged to engage both races to transfer drive from one race to the other and is movable in accordance with changes in variator drive ratio, a hydraulic traction loading actuator arranged to apply a traction load urging the roller and races into engagement to provide traction therebetween and so enable the transfer of drive, and at least one hydraulic roller actuator arranged to apply a reaction force to the roller, the control arrangement comprising hydraulics for applying fluid at an adjustable reaction pressure to the roller actuator to control the reaction force and for applying fluid to the traction loading actuator at a traction pressure which is related to the reaction pressure, thereby maintaining a relationship between reaction force and traction load, characterised in that the control arrangement further comprises at least one working chamber which is selectively connectable to and disconnectable from the reaction pressure or the traction pressure, and in that traction force is dependent upon pressure in the working chamber, so that by connecting/disconnecting the working chamber to/from the relevant pressure, the relationship between reaction force and traction force is changed. 
    
    
     
       BRIEF DESCRIPTION OF THE DRAWINGS 
       Specific embodiments of the present invention will now be described, by way of example only, with reference to the accompanying drawings in which:— 
         FIG. 1  is a simplified illustration of a toroidal race, rolling traction variator of a type which is in itself known and which is operable in accordance with the present invention; 
         FIG. 2  is a diagramatic representation of a hydraulic arrangement for controlling the variator in accordance with the present invention, incorporating a traction control valve; 
         FIG. 3  is a cross section through a practical embodiment of the traction control valve; and 
         FIG. 4  is a cross section through a traction load actuator constructed in accordance with the present invention. 
     
    
    
     DETAILED DESCRIPTION 
     Variators operable in accordance with the present invention are known in the art and an example will be only briefly described herein with reference to  FIG. 1 . 
     Two outer races, formed as shaped discs  12 ,  14  are mounted upon a drive shaft  16  for rotation therewith and have respective part-toroidal surfaces  18 ,  20  facing toward corresponding part toroidal surfaces  22 ,  24  formed upon an inner disc  26 , two toroidal cavities being thus defined between the discs. The inner disc is journalled such as to be rotatable independently of the shaft  16 . Drive from an engine or other prime mover, input via the shaft  16  and outer discs  12 ,  14  is transferred to the inner disc  26  via a set of rollers disposed in the toroidal cavities. A single representative roller  28  is illustrated but typically three such rollers are provided in each cavity. A traction load applied across the outer discs  12 ,  14  by a hydraulic traction load actuator  15  provides pressure between rollers and discs to enable such transfer of drive. Drive is taken from the inner disc to further parts of the transmission, typically an epicyclic mixer, as is well known in the art. Each roller is journalled in a respective carriage  30  which is itself coupled to a hydraulic actuator  32  whereby an adjustable reaction force can be applied to the roller/carriage combination. As well as being capable of translational motion the roller/carriage combination is able to rotate about an axis  33  determined by the hydraulic actuator  32  to change the “tilt angle” of the roller and to move the contacts between rollers and discs, thereby allowing variation in the variator transmission ratio, as is well known to those skilled in the art. 
     The illustrated variator is of the type known in the art as “torque controlled”. The hydraulic actuator  32  exerts a controlled reaction force on the roller/carriage and this is balanced by an equal but opposite force upon the roller resulting from the torques transmitted between the disc surfaces  18 ,  20 ,  22 ,  24  and the roller  28 . The reaction force upon the carriage  30  is adjusted by means of a hydraulic circuit through which fluid is supplied to opposite sides of piston  31  of the hydraulic actuator  32  at different, adjustable pressure through hydraulic feed lines  36  and  38 . The principal control input to the variator is thus provided in the form of two hydraulic pressures, applied to the roller control pistons through the hydraulic feed lines  36  and  38 . These pressures are manipulated through the hydraulic control circuit. 
       FIG. 2  shows a hydraulic circuit which serves to control both reaction force and traction load. A set of rollers  28 , each with an associated roller actuator  32 , is indicated schematically although other parts of the variator itself are omitted. Opposite sides of each actuator  32  are connected via the feeds lines  36 ,  38  to respective supply lines S 1 , S 2 . Pressure in the supply lines is adjustable by means of pressure reducing control valves V 1 , V 2  fed with high pressure fluid by a pump  39 . As indicated in the drawing, both control valves have a respective solenoid  50 ,  52  whose force upon the valve spool is opposed by a pilot signal from the associated supply line. Hence by setting the solenoid force, an electronic controller E.C. sets desired control pressures in the supply lines and in the roller actuators, thereby controlling desired reaction force. Note however that reaction force and the control pressures are also subject to external influences, as will now be explained. 
     Roller motion is accompanied by flow in the hydraulics. If there were no restrictions upon flow of fluid then such flow would not result in any pressure change since the valves V 1 , V 2  would act to maintain constant pressure. However any hydraulic arrangement provides some degree of restriction upon flow and in fact the present circuit incorporates restrictions in the form of dampers  54 ,  56 ,  58 ,  60  each with a restricted cross section for fluid flow, the dampers serving to create back pressure when flow takes place. Their purpose is to damp oscillation of the variator rollers  28 . First dampers  54 ,  56  disposed in the supply lines S 1 , S 2  damp a mode of oscillation in which the rollers move in unison with each other. Second dampers  58 ,  60  in the feed lines to individual actuators  32  damp a different mode of oscillation in which the rollers move out of phase with each other. The dampers may be formed as orifices or may take other forms. The important point for present purposes is that when the rollers move and flow takes place in the circuit, the dampers create back pressure in the hydraulics tending to resist the roller motion. 
     Consider, therefore, what happens in the event of a transmission torque spike created for example by hard vehicle braking. The brakes apply a large torque causing the vehicle wheels and consequently the variator output disc to decelerate. Variator ratio consequently falls and the variator rollers are required to rapidly move and precess to positions corresponding to lower ratio. Fluid in the hydraulics is displaced by the rapidly moving pistons of the actuators  32  and resistance to the resulting flow in the hydraulics, created by the dampers and by other parts of the hydraulics including the valves V 1 , V 2 , produces an increase in fluid pressure on one side of each actuator and a reduction on the other side, tending to resist the roller movement. Reaction force is dramatically increased in a manner which is not initiated nor directly predictable by the electronic control. 
     The traction load must be modified in sympathy with the reaction force if excessive roller slip is to be avoided and this is achieved by means of traction control valve  62  which receives two pilot control pressures. 
     A higher-pressure-wins valve arrangement  64 , which may be formed using back-to-back non-return valves, connects its own output  66  to whichever of the two sides of the hydraulic circuit is at higher pressure, this output being led to the traction control valve  62  to serve as a first pilot pressure which is referred to below as the “reaction pressure”. Note that the higher-pressure-wins valve arrangement is connected to points in the hydraulics close to the ports of one of the actuators  32 , and between the chosen actuator and its dampers  58 , 60 , so that the pressures it receives are as close as possible to the prevailing pressures in the actuator itself. The higher of the two actuator pressures is used as an indication of the reaction force. Pressure on the opposite side of the actuator is typically low and is neglected in this embodiment. 
     A second pilot pressure is led to the traction control valve  62  from working chamber  70  of the traction load actuator, which is indicated in highly schematic form at  72  in  FIG. 2 . The first and second pilot pressures act in opposition to each other upon the spool of the valve  62 . 
     The traction valve  62  is a 3 port, 3 position valve, One port is connected to a high pressure fluid source, formed in this embodiment by the pump  39 . A second port leads to the working chamber  70  of the traction load actuator. A third port leads via a non-return valve  74  to a pressure sink, which in the present embodiment takes the form of transmission drain  76 . In dependence upon the pilot pressures, the traction valve either:—
         i. connects the traction load actuator to the drain  76  via a restrictor  78 , venting pressure therefrom, while closing the high pressure port;   ii. closes all three ports to sustain pressure in the traction load actuator; or   iii. connects the traction load actuator to the high pressure source—the pump  39 —to boost pressure therein and increase traction load.       

     Because this control is carried out in dependence upon reaction pressure, the valve serves to vary traction load in sympathy with, and more specifically in proportion to, reaction pressure and hence reaction force. In this way the hydraulics can serve to maintain traction coefficient substantially at a chosen level. However the traction coefficient set by the valve  62  is also adjustable in accordance with the present invention, as will now be explained. 
     Note that the traction control valve has two working chambers within which reaction pressure can act to influence valve state, labelled C 1  and C 2 . Similarly there are two working chambers within which traction load pressure can act to influence valve state, labelled C 3  and C 4 . A constantly open connection leads reaction pressure to the first of the reaction pressure chambers C 1 . However a reaction pressure valve, formed as a two port, two position valve  80  under control by the E.C., serves to connect the second reaction pressure chamber C 2  either to reaction pressure or to drain. Likewise a constantly open connection leads tractions pressure to first traction pressure chamber C 3  and a traction pressure valve  82  connects second traction pressure chamber C 4  either to traction pressure or drain, under control by the E.C. 
     Opening and closing the reaction pressure and traction pressure valves changes the areas upon which the pilot pressures act. In this way the ratio of traction pressure to reaction pressure provided by the traction control valve is adjusted—i.e. traction coefficient is adjusted. The adjustment to traction coefficient is discrete rather than continuous: in the illustrated embodiment four different values of traction coefficient μ can be provided, since there are four possible states for the combination of the two valves. 
     Let us refer to the working areas within the chambers C 1  to C 4  as A 1  to A 4 . Typically the pilot pressures act to move the valve spool and it is on the spool that the areas A 1  to A 4  are formed, although it is possible to devise alternative valves in which the areas are formed instead upon other valve parts, such as a moveable sleeve. Also let us define a “design” traction coefficient to be the value μ d  of traction coefficient obtained if traction and reaction pressures are equal. μ d  is dependent e.g. upon the area of a piston in the end load actuator. The actual traction coefficient μ is found from 
             μ   =         A   T       A   R       ×     μ   d             
where A R  is the total working area (A 1 +A 2 ) of the valve upon which reaction pressure is exerted and A T  is the total working area (A 3 +A 4 ) upon which traction pressure is exerted. Suppose further that it is desired to provide for a range of traction coefficients from 1.5μ d  for efficient operation to
 
               μ   d     2         
for high speed operation.
 
This can be achieved by forming the valve such that
 
                   A   1     +     A   2         A   3       =   2             and                   A   3     +     A   4         A   1       =   1.5         
if we take
 
A R =A T  
 
then a suitable solution is that
 
               A   1     =       2   ⁢     A   T       3                   A   2     =       A   T     3                   A   3     =       A   T     2                   A   4     =       A   T     2           
The four possible combinations of states of the reaction pressure and traction pressure valves  80 ,  82  can be represented in a truth table:—
 
                                                           Reaction Pressure Valve   Traction Pressure Valve   μd                                Open   Closed   0.5   μd       Closed   Closed   0.75   μd       Open   Open   1   μd       Closed   Open   1.5   μd                    
Of course the above calculations serve as examples only. Different valve areas may be chosen in accord with different design criteria.
 
     The practical embodiment of the traction control valve  62  illustrated in  FIG. 3  has a valve spool  100  slidably received in a stepped longitudinal bore of a valve body  102 . The three valve ports are labelled as follows:—
         port P LOAD  is connected to the traction load actuator   port P HIGH  is connected to the high pressure source and   port P DRAIN  is connected to the drain
 
All three valve ports are formed as stepped, transverse bores in the valve body  102  communicating with the longitudinal bore  102 . The four chambers in which pilot pressures act are once more labelled C 1  to C 4  and communicate with respective connection passages P 1  to P 4 , two of which are formed as transverse bores and the other two of which are formed by end regions of the bore  102 . The valve spool has narrowed end regions  104 ,  106  which slide, and seal, in respective bushes  108 ,  110  so that pilot pressures in chambers C 1  and C 3  act only upon respective end areas A 1 , A 3  of the spool. The bushes are retained by respective externally threaded end plates  112 ,  114  both having through bores  116  to permit fluid to pass. Bolts  118 ,  120  passing through the respective end plates limit the spool&#39;s travel by abutting with its ends A 1 , A 3 . Outer piston portions of  122 ,  124  of the spool seal and slide in an enlarged centre portion of the longitudinal bore, their respective outer faces providing the areas A 2  and A 4 . Switching between the three valve states is carried out by an inner piston portion  126  which, depending upon spool position, connects P LOAD  to P HIGH  or to P DRAIN , or alternatively closes P LOAD .
       

     It will be appreciated by the reader that by connecting/disconnecting the controlling pressures to the working chambers C 2 , C 4  a discontinuous adjustment to traction coefficient is effected. In the embodiments illustrated in  FIGS. 2 and 3  the relevant working chambers are found in the traction control valve  62 . Another way to provide for the desired traction coefficient adjustment is to connect/disconnect chambers within the traction load actuator. 
     A suitable traction load actuator  100  is illustrated in  FIG. 4 . This drawing shows only half of the actuator, lying to one side of its rotational axis  99 . In this drawing one of the variator discs is indicated at  102  and is received in the manner of a piston within an annular, shaped sleeve  104  which is carried upon variator shaft  106  and rotates along with it Splines  107  between the disc  102  and the shaft  105  ensure that they rotate together but that the disc is able to move along the shaft under the influence of the end load. The actuator  100  comprises first, second and third annular pistons P 1 , P 2 , P 3  within the sleeve  104  which together define first, second and third working chambers CH 1 , CH 2  and CH 3 . Pressurizing any of the three working chambers creates a traction load urging disc  102  toward the right hand side of the drawing. However each working chamber provides a different working area, so that the magnitude of the traction load differs depending upon which of the working chambers is pressurized. In this embodiment the traction control valve  62  ( FIG. 2 ) may be dispensed with. The reaction pressure taken through the output line  66  of the higher-pressure-wins valve  64  is instead led, through an arrangement of switching valves (not illustrated), to a chosen working chamber CH 1 , CH 2  or CH 3 , or to a combination of chambers. Because reaction pressure is being used to create the traction load, a relationship is once more maintained between reaction force and traction load: one is proportional to the other (at least while pressure changes, due e.g. to flow through higher-pressure-wins valve  64  and line  66 , can be neglected so that the pressures in the actuators  32  and  100  can be taken to be identical). Furthermore, because there is the facility to change the combination of working chambers in which the reaction pressure acts, this relationship can be adjusted, varying the traction coefficient. 
     Looking in more detail at the construction of the traction load actuator  100 , the first chamber CH 1  is defined between the rear face of the variator disc  102  and front surfaces of first piston P 1 , sleeve  104  and an annular spigot portion  106  of piston P 2 . First piston P 1  rides upon the spigot portion  106  and seals against that part and against a cylinder bore of sleeve  104  by virtue of sealing rings. Second piston P 2  carries first stubs  108  (only one of which is seen in the drawing) which project forward through first piston P 1 , forming a seal with it, so that by virtue of abutment of the stubs  108  against the rear face of the variator disc  102 , the second piston P 2  is able to exert a force upon the disc and so contribute to the traction load. The second piston P 2  rides upon a hub portion  109  of the sleeve  104  and is also received within an annular flange portion  111  of the third piston P 3 , forming seals with both through respective sealing rings. Third piston P 3  carries second stubs  110  which project forward through second piston P 2 , forming a seal with it, and can abut against the rear face of first piston P 1  to exert a force upon it. Third piston P 3  rides upon an enlarged diameter portion of the hub portion  109  and seals against that and also against the cylinder bore of the sleeve  104  by virtue of respective sealing rings. A frusto-conical spring  112  is pre-stressed between an end wall  114  of the sleeve  104  and rear face  116  of the third piston P 3 . 
     When none of the three working chambers CH 1 , CH 2 , CH 3  is pressurized, a pre-loading traction force is applied to the disc  102  by the pre-stressed spring  112  acting through the piston P 3 , the second set of stubs  110  and the first piston P 1  which bears upon the rear of the disc. This pre-load provides traction upon start-up of the transmission. Subsequently applying sufficient pressure to any of the working chambers causes the piston P 3  to move leftward until a flange  118  projecting from its rear face abuts the sleeve&#39;s end wall  114 , and in this condition the effect of the spring  112  is removed and traction load is determined by pressures in the working chambers. 
     If working chamber CH 1  alone is pressurized then piston P 1  moves leftward to the limit of its travel, defined by its abutment with the second set of stubs  110 . Fluid pressure acts upon the entire area of the rear face of the disc  102 . Pressurizing the second working chamber CH 2  alone causes fluid pressure to be applied to the smaller area formed by the rear face of the second piston P 2 , this force being transmitted to the disc  102  through the first set of stubs  108 . Pressurizing only the third working chamber CH 3  causes fluid pressure to be applied to the still smaller working area formed by the rear face of the first piston P 1 , which consequently advances to bear upon the disc  102  and so apply a force to it. Hence by applying the same reaction pressure to any one of the three working chambers a different traction load, and a different traction coefficient, is obtained. 
     Still more different values of traction coefficient can be obtained by applying pressure to two or more of the working chambers. Pressurizing the first and second chambers CH 1  and CH 2  provides a traction force determined by the sum of the areas of the rear faces of the disc  102  and the second piston P 2 . A different traction force again can be provided by pressurizing all three chambers. 
     Supply of fluid to the working chambers CH 1 , CH 2 , CH 3  can be provided for through bores in the shaft  106  (not shown in this simplified drawing). 
     The aforegoing embodiments serve as examples only of ways to implement the present invention. Numerous variants are possible. One such combines some of the virtues of the two embodiments shown in  FIGS. 2 and 4  by using a pilot operated valve to supply pressure to the working chambers CH 1 , CH 2 , CH 3  of the  FIG. 4  actuator. The valve in question has a spool subject to opposed reaction and traction load pressures. It does not require the switchable chambers C 2 , C 4  of valve  62  ( FIG. 2 ), since instead traction coefficient adjustment is achieved by supplying its output, through any of three switching valves, to the selected working chamber(s) in the  FIG. 4  traction load actuator.