Abstract:
A hydraulic control circuit is disclosed for a continuously variable transmission comprising a continuously variable ratio transmission unit (“variator”)  10 , the circuit comprising a supply line  106  and means (which may take the form of a pump  110 ) for providing a flow of pressurised fluid in the supply line, means (which may comprise a pressure control valve  116 ) for generating a back pressure in the supply line, and at least one connection for feeding fluid from the supply line to a hydraulic actuator  100  acting on a movable torque transmission  23 - 37  element (which may comprise a roller  28 ) of the variator  10 . The valve means  152  connected to the supply line allows pressure in the supply line to be selectively modified in response to rate of flow in the supply line.

Description:
BACKGROUND OF THE INVENTION 
   1. Field of the Invention 
   The present invention relates generally to continuously variable transmissions and more specifically to hydraulic control thereof. 
   2. Background Art 
   Continuously variable transmissions comprise, inter. alia an input shaft rotatable by a prime mover, an output shaft connected to vehicle wheels and a ratio varying component (hereinafter referred to as a “variator”) disposed between the input and output shafts. The variator is typically controlled by means of hydraulic pressures which can adjust the effective ratio of the variator in accordance with a driver&#39;s demands, road conditions and the like. Under steady or smoothly-changing conditions the required flows of hydraulic fluid are relatively small. However, problems can sometimes occur during rapid ratio changes of the variator, for example during braking, rapid engine acceleration or abuse conditions. During such rapid ratio changes, the flows of fluid can be excessive, resulting in adverse system controllability. 
   This is particularly, but not exclusively, applicable in the case of torque control variators. 
   Major components of a known torque-control variator  10  of the “full toroidal”, toroidal-race rolling traction type are illustrated in  FIG. 1 . Here, two input 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 a central output disc  26 . The output 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 input discs  12 ,  14 , is transferred to the output 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 both cavities. An end load applied across the input discs  12 ,  14  by a hydraulic end loading device  15  provides contact forces between rollers and discs to enable the transfer of drive. Drive is taken from the output disc to further parts of the transmission, typically an epicyclic mixer, as is well known in the art and described eg. in European patent application 85308344.2 (published as EP 0185463). Each roller is journalled in a respective carriage  30  which is itself coupled to a hydraulic actuator  32  whereby a controlled translational 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 determined by the hydraulic actuator  32  to change the “tilt angle” of the roller and so move the contacts between rollers and discs, thereby varying the variator transmission ratio, as is well known to those skilled in the art. 
   As mentioned above, the illustrated variator is of the type known in the art as “torque control”. The hydraulic actuator  32  exerts a controlled force on the roller/carriage and for equilibrium this must be balanced by the reaction force upon the roller resulting from the torques transmitted between the disc surfaces  18 ,  20 ,  22 ,  24  and the roller  28 . As is well known in the art, the centre of the roller is constrained to follow the centre circle of the torus defined by the relevant pair of discs. The axis determined by the actuator  32  is angled to the plane of this centre circle. This angle is referred to as the “castor angle”. The well known result of this arrangement is that in use each roller automatically moves and precesses to the location and tilt angle required to transmit a torque determined by the biasing force from the actuator  32 . 
   The biasing force is controlled by means of a hydraulic circuit through which fluid is supplied to the actuators at variable pressure. 
   It will be appreciated that while the equilibrium position of the rollers is determined by the balance of the reaction force and the applied biasing force, there is the potential for unwanted oscillatory motion of the roller/carriage combination about this position, with resulting impairment of transmission function. More than one mode of oscillation is possible but in the simplest such mode all rollers oscillate in unison and this oscillatory motion is accompanied by a corresponding flow of fluid in the hydraulic circuit. 
   Damping of such oscillation can be provided by means of the hydraulic circuit and specifically by restricting or throttling fluid flow to and from the actuators  32 . During a change in variator transmission ratio, the rollers  28  must move and precess to new positions, fluid thus being expelled from one side of the pistons of the actuators  32  and taken in on the other side. Under these conditions, if fluid flow in the hydraulic circuit is suitably restricted, pressure is increased in the hydraulic circuit on the side of fluid expulsion and diminished on the other side of the circuit, modifying the net force exerted on the rollers by the actuators such as to tend to resist roller motion and thus to create a torque between the variator input and output discs. 
   The effect is two-fold:
         i. damping is provided, which helps to deliver smooth non-oscillatory variator response, particularly when installed in a mechanical power train; but   ii. the torque created resists required ratio change, which can impair transmission performance during rapid transient events such as rapid braking and rapid acceleration.       

   Particularly stringent requirements are imposed on the transmission by such “transients”—rapid changes in the operating conditions of the vehicle requiring correspondingly rapid changes of transmission ratio. An emergency stop or “brake to rest” is one example. In order to maintain engine speed and to avoid stalling the engine during a brake to rest, rapid ratio change is required of the variator. This is particularly significant in a transmission of the “geared neutral” type in which the variator remains coupled to the vehicle&#39;s wheels even while the wheels are stationary—that is, in vehicles lacking a clutch or other means to isolate wheels and engine. The high rate of ratio change required during rapid brake to rest corresponds to a rapid motion of the variator rollers and their associated pistons. Large flows are created in the hydraulic control circuit. If adequate hydraulic flow to accommodate such motion is not available—particularly because such flow is restricted—the rollers can fail to move with sufficient speed, leading eg. to an engine stall. Within the hydraulic circuit the effect can be a large increase in pressure on one side of the circuit and a large fall in pressure on the other side of the circuit. The result must be a large net biasing force on the roller/carriage combinations and this is reflected in a large variator torque which is the cause of the engine stall. 
   SUMMARY OF THE INVENTION 
   It has been found in practice that the level of damping required to achieve smooth transmission operation can unacceptably inhibit variator response when rapid transients occur. Achieving the necessary balance of stability against response has proved problematic. Overcoming or at least alleviating this problem is an object of the present invention. 
   This need is not unique to toroidal-race rolling-traction type variators but is also applicable to many other types of hydraulically controlled variators, for example variators of the band-and-sheave type where the separation of the sheaves of each of two pulley units around the band is entrained is controlled by hydraulic pressure. 
   In accordance with the present invention, there is a hydraulic circuit for a continuously variable transmission comprising a continuously variable ratio transmission unit (“variator”) which is controllable by application of fluid pressure to at least one hydraulic actuator acting on a movable torque transmission element of the variator, the circuit comprising a fluid supply line connected to the hydraulic actuator for feeding fluid to and from the hydraulic actuator, means for providing a flow of fluid through the fluid supply line, variable control valve means in the fluid supply line downstream of the connection to the hydraulic actuator for generating an adjustable back pressure therein, and further valve means connected to the fluid supply line to selectively modify pressure in the supply line in response to rate of fluid flow in the supply line. 
   In such a circuit the path controlled by the valve means can serve to vent and/or supplement fluid flow through the flow line and thus to allow increased flow in response to transients. 
   By reacting to flow rate in the line rather than pressure therein, the further valve means allows transient contditions to be accomodated without imparing control of the variator by the variable control valve means. 
   Where the invention is applied to a toroidal-race rolling traction type variator, the movable torque transmission element takes the form of a variator roller. The roller  28  illustrated in  FIG. 1  is an example. 
   Preferably the further valve means comprises a valve which controls a further connection to the supply line to modify pressure in the supply line. 
   Preferably, the valve means is controlled such that its degree of opening is substantially constant while the rate of flow in the flow line is within a chosen range. Still more preferably the valve means is closed while the rate of flow is within the chosen range. 
   It has been found, somewhat unexpectedly, that in this way stable controllable operation of the variator can be achieved while also enabling rapid response to transients. 
   Preferably the valve has a variable opening controlled as a function of the supply line flow rate. A variable opening valve is of further assistance in maintaining variator stability. 
   It is preferred that the valve means comprises a pilot operated valve having a spool which is controlled by pilot pressures from upstream and downstream of the flow restrictor. 
   The valve means preferably control a path to a low pressure area in order to vent excess fluid from the flow line. In this way the valve means can prevent excessive pressure build up in the flow line. 
   Additionally or alternatively, the valve means may control a path connected to a pressurised accumulator in order to selectively supplement fluid flow into the flow line. 
   An important advantage of such an embodiment is that the capacity required of the means for providing flow of pressurised fluid (typically formed as a pump) can be reduced as compared with existing circuits in which this capacity is chosen to meet the maximum requirements of the circuit. Provision of the accumulator allows the flow in the line to be supplemented when required, from the accumulator, so that the pump need not supply the maximum required rate of flow. 
   In certain preferred embodiments of the present invention the further valve means comprises a flow limiting valve in the supply line. Flow limiting valves are known in the art of hydraulics. This valve may serve to limit the maximum flow to the variable control valve means or to ensure a minimum flow thereto. Two such valves may be provided in the supply line to serve both functions. 
   Practical circuits generally control double acting actuators and in a further preferred embodiment of the present invention the circuit comprises a pair of fluid supply lines, the hydraulic actuator being double acting and receiving opposed fluid pressures from the two supply lines, each supply line having a said variable control valve means and a said further valve means. 

   
     BRIEF DESCRIPTION OF THE DRAWINGS 
     A specific embodiment 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 known toroidal-race rolling traction type variator which is suitable for control by the hydraulic circuit to be described below; 
       FIG. 2  is a schematic diagram of a hydraulic circuit embodying the present invention; 
       FIG. 3  is a schematic illustration of a control valve utilised in the  FIG. 2  circuit and of connections thereto; 
       FIG. 4  is a schematic diagram of an accumulator circuit forming part of the  FIG. 2  circuit; 
       FIG. 5  is a graph of total forward flow in a flow line of the  FIG. 2  circuit against flow through a flow restrictor of the circuit; 
       FIG. 6  is a graph of pressure against flow through the restrictor; 
       FIG. 7  is a schematic diagram of a further hydraulic circuit embodying the present invention; 
       FIG. 8  is a schematic diagram of still a further hydraulic circuit embodying the present invention; and 
       FIG. 9  is a schematic diagram of yet a further hydraulic circuit embodying the present invention. 
   

   DETAILED DESCRIPTION OF THE PREFERRED EMBODIMENTS 
   The hydraulic circuit illustrated in  FIG. 2  is suitable for use with a torque control variator of the type described above with reference to  FIG. 1 .  FIG. 2  shows, by way of illustration, a set of three hydraulic actuators  100 ,  100 ′ and  100 ″ (typically in a variator of the above described twin cavity type, six such actuators would be provided—three per cavity—but remaining actuators are omitted for the sake of clarity). Each actuator comprises a piston  102  whose two faces are exposed to control pressure in first and second working chambers  104 ,  204  so that the biasing force applied by each actuator is determined by the difference in these control pressures. Each actuator  100  is coupled to a corresponding roller/carriage assembly of the type illustrated in  FIG. 1 . 
   The hydraulic circuit provides a first flow line  106  for supplying hydraulic fluid to the first working chambers  104  and a second flow line  206  for supplying fluid to the second working chambers  204 . 
   The first flow line  106  comprises a supply line  112  and a drain line  114 . A pump  110  draws hydraulic fluid from a sump  111  (and it should be noted that while the diagram shows the symbol for the sump in several places, these are all the same item: the circuit has one sump only) and supplies a pressurised flow of fluid through the supply line  112  to the first working chambers  104  of the actuators  100 . The supply line is branched to connect to all of the first working chambers  104 . The drain line however is only directly connected to one of these chambers—chamber  104 ′ of actuator  100 ′, referred to as the first master actuator. Pressure in the supply line  112 —and hence in the actuator working chambers  104 —is adjusted by means of a first pressure control valve  116  incorporated in the drain line  114 . This valve is controlled by an electronic control unit (ECU)  117 . It is again to be understood that while the symbol for the ECU is seen in two places on the diagram for the sake of representational convenience, these symbols both refer to a single such unit. From the downstream side of the pressure control valve  116 , the drain line leads back to the sump  111  from which the hydraulic fluid is recirculated. 
   The second flow line  206  is similarly formed to the first, comprising a second supply line  212  which supplies pressurised hydraulic fluid from a second pump  210  to the second working chambers  204  and a second drain line  214  leading via a second pressure control valve  216  to the sump  111 . The second drain line  214  is connected to working chamber  204 ″ of a second master actuator  100 ″. 
   The master actuators  100 ′ and  100 ″ provide limits to the actuator travel, as is known in the art. When the pistons  102  move sufficiently far to the left, piston  102 ′ of the first master actuator  100 ′ covers the mouth of the drain line  114 , preventing further exhaustion of fluid therethrough and thus preventing further motion of the pistons to the left. The second master actuator  100 ″ limits travel of the pistons to the right in an analogous manner. 
   The ECU  117  monitors pressure in both of the flow lines  106 ,  206  by means of respective pressure transducers  118 ,  218  and controls the opening of the pressure control valves  116 ,  216 . The purpose of the valves is to create an adjustable back pressure in the flow line. On the downstream (drain) side of the valves pressure in the circuit is low. Typically this region is close to or at atmospheric pressure. Upstream of the valves is normally a higher pressure, controlled by means of the valves  116 ,  216 . In this way, by setting the pressure control valves  116 ,  216 , the ECU influences the pressures applied to, and the forces applied by, the actuators  100 . Note however that these pressures and forces are determined not only by the valves  116 ,  216  but also by the torques to which the variator is subject and the consequent motions of the rollers/pistons. This will be explained further below. 
     FIG. 2  also shows a valve arrangement  121  of the “highest pressure wins” type having a respective input connected to both of the supply lines  112 ,  212 . The arrangement supplies via its output  123  hydraulic fluid from whichever supply line is at higher pressure to a hydraulic actuator (item  15  in  FIG. 1 ) for applying required end load to the variator discs. Also shown in  FIG. 2  are first and second pressure limiting valves  124 ,  224  connected respectively to the first and second supply lines  112 ,  212 . 
   Reference has been made above to the need to damp oscillatory motion of the pistons  102  and hence of the rollers to which they are connected. While a degree of damping exists in the circuit by virtue in particular of the pressure control valves  116 ,  216 , which tend to restrict surges of fluid flow, damping is also provided in the illustrated circuit by first and second restrictors  150 ,  250  incorporated respectively in the first and second flow lines. More specifically, each restrictor is connected in the respective drain line  114 ,  214 , downstream of connections of the flow line to the actuator working chambers but upstream of the respective pressure control valve  116 ,  216 . The restrictors are formed as metering orifices, having a reduced cross section relative to other regions of the flow line and a small extent in the direction of fluid flow (ie. a sharp edged orifice). The viscosity of the hydraulic fluid is, unavoidably, dependent to some degree on temperature. The metering effect of an orifice having a small extent along the flow direction is less dependent on viscosity, and hence less variable with temperature, than that of a longer restricted passage. In use, flow through each restrictor creates a pressure differential across it, the downstream side being at lower pressure. This pressure differential is approximately proportional to the square of the rate of fluid flow. The effect is to oppose surges of fluid flow, including those due to variator roller oscillation, and thus to damp the roller&#39;s movement. 
   As pointed out above, a circuit which provides adequate damping can prove problematic during transient rapid variator ratio change. The circuit illustrated in  FIG. 2  incorporates compensation valves  152 ,  252  which obviate such problems. Consider first of all, however, the effect of a rapid ratio change on a circuit which lacked these valves. Suppose that the actuator pistons are required to move in the direction to the left in  FIG. 2 . Fluid must be expelled from the first actuator chambers  104  into the first flow line  106  and this extra fluid can only flow away through the first restrictor  150  and the first pressure control valve  116 . The increased flow rate thus required through these parts creates increased back pressure across them, raising pressure in the first flow line  106  and opposing the required movement of the pistons, and so of the variator rollers. On the other side of the circuit, there is a requirement for flow of fluid into the second actuator working chambers  204  and, unless back flow takes place in the drain line  214  of the second flow line  206 , this increased flow rate must be supplied by the pump. In known circuits, the required flow rate can exceed the pump capacity, leading to a major drop in pressure on this side of the circuit which again opposes the required piston/roller movement and may give rise to cavitation. Even without the restrictors  150 ,  250 , such problems of excessive flow requirements can arise due to the limited opening of which the pressure control valves  116 ,  216  are capable and the limited capacity of the pumps  110 ,  210 . 
   The compensation valves  152 ,  252  allow such problems to be avoided by venting fluid from the high pressure flow line and also in the illustrated embodiment by injecting fluid to the low pressure flow line as required to keep pressures within acceptable limits. While other valve constructions are possible within the scope of the present invention, in the illustrated embodiment each compensation valve  152 ,  252  is formed as a double pilot operated directional control valve. This valve has a proportional response. Its operation will now be described with reference to  FIGS. 2 and 3 . 
   The compensation valve  152 ,  252  illustrated in  FIG. 3  responds to the pressure differential across its associated restrictor  150 ,  250 . It comprises a valve cylinder  300 , a valve spool  302  movable along the cylinder and three ports:
         i. a common port  304  which is selectively connectable through respective communication chambers  306 ,  308  defined in the cylinder  300  by the spool  302 , on either side of a head  309 , to either of   ii. an outlet port  310  connected to a path to the sump  111  and   iii. an inlet port  312  connected to an accumulator  154 ,  254  (see  FIG. 2 ). The accumulator is constantly maintained at high pressure (eg. in the range 45-55 bar) and will be described in more detail below.       

   The common port  304  is connected to the associated flow line  106  or  206  upstream of its restrictor  150 ,  250 . In the illustrated embodiment this connection is made to a point in the drain line  114 ,  214 , downstream of connections to the hydraulic actuators  100 . 
   The position of the spool  302  in the cylinder  300  is controlled by pilot pressure signals taken from upstream and downstream of the associated restrictor  150 ,  250  and applied to respective opposed faces  314 ,  316  of the spool  302 . The force exerted on the spool by these pilot pressure signals is not nomnally equal, since flow through the restrictor  150 ,  250  produces a back pressure across it, but the resultant net force on the spool is opposed by a spring  318  acting on the spool (an additional or alternative way to balance the force on the spool would be to change the relative areas of the faces  314 ,  316 , as will be apparent to one skilled in the art). The valve characteristics are in the present embodiment chosen such that during normal, non-transient variator operation the spool head  309  closes the common port  304 . That is, during non-transient operation the rate of flow through the associated flow line  106 ,  206  and the resultant pressure differential across the restrictor  150 ,  250  are such as to maintain the compensation valve  152 ,  252  in a closed state. Consequently during such operating conditions the compensation valves  152 ,  252  do not, in the present embodiment, significantly affect the function of the hydraulic circuit. The valve head is longer, in the direction of valve spool movement, than the common port  304  which it controls, as indicated by a double-headed arrow in  FIG. 3 . Consequently the valve has a “deadband”. That is, some valve spool movement can take place (corresponding to non-transient variator operation) without opening of the compensation valve  152 ,  252 . 
   Consider however what happens during transient, rapid, variator ratio change, the actuator pistons  102  again moving to the left as viewed in  FIG. 2 . Fluid expelled from the first actuator working chambers  104  again causes pressure rise in the first flow line  106  and an increased rate of flow from it. The pressure differential across the first restrictor  150  consequently increases (because, as explained above, the pressure differential relates to the square of the flow rate through the restrictor) and the pilot pressure signals to the first compensation valve  152  are correspondingly altered. Looking again at  FIG. 3 , pressure on spool face  314  (connected to the upstream side of the first restrictor  150 ) increases more than pressure on opposed face  316  (connected to the downstream side of the first restrictor  150 ). Thus the spool is displaced, to the right as viewed in  FIG. 3 , and if this displacement is sufficient to move the spool beyond its deadband then the common port  304  of the compensation valve  152  is connected to its outlet port  310 . A path is thereby opened for flow of hydraulic fluid out of the first flow line  106 , to the sump  111 . Note that the degree of opening of the valve is variable and is related to the pressure differential across the restrictor  150 . That is the valve has a progressive response. A large back pressure, corresponding to a requirement for a large flow rate, produces a correspondingly large valve opening suitable to accommodate the large flow. This progressive response is of assistance in maintaining transmission stability and the currently preferred valve characteristics will be explained below. 
   Note also that due to the configuration of the illustrated circuit, the first compensation valve  152  (and similarly the second valve  252 ) can be thought of as operating in a negative feedback loop. Fluid vented from the flow line through the first compensation valve  152  reduces rate of flow through the first restrictor  150  and hence tends to reduce the valve opening. The result is a stable, progressive control of the valve. The valve  152  in effect reacts to and regulates flow through its line  106 . Note however that the valve  152  does not react to changes in line pressure as such except so far as these result in changes of flow rate. 
   Looking now at the other side of the hydraulic circuit, the leftward displacement of the pistons  102  produces a requirement for rapid fluid flow into the second actuator working chambers  204 . Consequently flow through the second restrictor  250 , and hence pressure differential across it, are reduced and again the pilot pressure signals to the compensation valve  252  are correspondingly altered. The pressure differential across the spool  302  of the second compensation valve  252  is reduced and the spool consequently moves, under the biasing force of the spring  318 , to open a path from the second flow line  206  to the accumulator  254 . The accumulator  254  is maintained at a pressure higher than that of the second flow line  206 , so that the effect of opening of the second compensation valve  252  is to cause fluid flow into the flow line as required to prevent excessive pressure drop therein. Again this is done in a progressive manner and is subject to control through a negative feedback loop effect. 
   The ECU  117  can also be programmed to react to differential pressure produced by transients by appropriate control of the pressure control valves  116 ,  216 . For example, the pressure control valve  216  in the low pressure side may be closed while the pressure control valve  116  in the high pressure side is fully opened, to assist in compensating for—and reducing—the pressure changes during transients. 
   Of course when the requirement is for motion of the pistons  102  in the opposite direction, to the right in  FIG. 2 , the roles of the two compensation valves  152 ,  252  are swapped, but in other respects the function of the circuit is as described above. Opening of the two compensation valves  152 ,  252  need not be simultaneous and some transients may for example cause opening of one but not the other. 
   Various possible accumulator constructions will be known to those skilled in the art but preferably the accumulator has an accumulator vessel, having a resiliently variable volume, connected in an unloader circuit which tops up the vessel without continuously requiring high pressure pump flow.  FIG. 4  illustrates a suitable circuit, the vessel being seen at  400  and connected via a non-return valve  402  to a pump  404 . A valve  406  diverts the pump output to the sump  111  when the required pressure is achieved in the accumulator vessel and the circuit&#39;s output is at  408 . 
   The characteristics of the compensation valves  152 ,  252  must be chosen to suitably balance the requirements for hydraulic damping (necessarily somewhat reduced by valve opening) against the requirement for rapid flow during transients. It is also important to retain transmission stability during valve opening and closing. The currently preferred characteristics can be understood from  FIGS. 5 and 6 , although it must be understood that continued development may in time show other characteristics to be favoured. 
   In  FIG. 5  the vertical axis corresponds to forward flow through the pressure control valve  116  or  216  and the horizontal axis to forward flow through the corresponding restrictor  150 ,  250 . In the control deadband region  500  the relevant compensation valve  152 ,  252  is closed. The flow through the restrictor and the pressure control valve are thus the same and the graph has the form of a straight line through the axis. When the flow into the regulator exceeds the deadband region, opening of the compensation valve vents fluid and, in the illustrated case, the consequent flow through the pressure control valve is controlled by the progressive valve response such as to be largely constant (or in fact subject to a very gradual rise with restrictor flow). Similarly when flow into the regulator falls below the deadband, the compensation valve maintains a largely constant minimum level of flow. 
   The way this flow characteristic relates to pressure, given constant opening of the pressure control valves  116 ,  216 , is seen in  FIG. 6 , which shows flow line pressure (vertical axis) against flow into the restrictor. In the deadband region the graph is roughly a square function, back pressure being related to the square of the flow rate. Outside this region, the compensation valve serves essentially to set a maximum and a minimum pressure in the flow line, related of course to the opening of the respective control valve  116 ,  216 . It is important to note that while the illustrated flow characteristic represents a situation in which opening of the pressure control valve  116  or  216  is kept constant, in practice these valves are typically to be adjusted in response to a transient so that although flow rates in the two flow lines may be affected by the transient, differential pressure in the two lines need not be greatly altered. 
   Numerous possible variations of the above described exemplary circuit will be apparent to those skilled in the art. For example, while the illustrated circuit allows both excessively high and excessively low flow rates to be compensated, the circuit may instead provide only for venting of fluid from the circuit&#39;s high pressure side, the pressure drop on the low pressure side (which of course cannot amount to a hydraulic lock) being accepted or compensated in some other way. Various suitable constructions for the compensation valves  152 ,  252  will be known to the skilled person and whereas the illustrated valve is hydraulically controlled, an electronically controlled valve could conceivably be used. 
   In further embodiments of the present invention, flow control valves may be used in the flow lines to enable required flow rates without undesired extremes of pressure. 
     FIG. 7  illustrates such an embodiment in schematic form and shows a plurality of double-acting actuator pistons  510 , each of which is arranged to control the position of a respective roller (not illustrated) of a toroidal-race rolling-traction type variator, as before. Only two of the pistons  510  are illustrated in  FIG. 7 , for the sake of clarity. 
   As in the earlier circuits, hydraulic fluid is fed to each of the two faces of each piston  510  from a respective one of two flow lines  512 ,  514 . Hydraulic fluid is supplied under pressure from a sump  516  into each of the left and right hand flow lines  512 ,  514  by means of an associated pump  518 ,  520  (typically operating at an output between 0 and 50 bar) and is supplied to the faces of each piston  510 . In  FIG. 7 , the uppermost piston  510 ′ is the master piston and is fed with hydraulic fluid directly from the pumps  518 ,  520 . The remaining pistons  510  are known as “slave” pistons which are connected to the left and right-hand flow lines  512 ,  514  by supply branches  522 ,  524  respectively, the pressures in the branches  522 ,  524  following the pressures applied to the respective piston faces of the master piston  510 ′. In the arrangement as illustrated, a single piston  510 ′ forms the master piston. However, and as will be appreciated by those skilled in the art, one side of a first piston for the other flow lines  512 ,  514 . 
   The pumps  518 ,  520  provide hydraulic fluid to the pistons at the pump outlet pressure when control valves  526 ,  528  located downstream of the pistons  510  in the left and right-hand flow lines  512 ,  514  are sufficiently restricted. Thus by controlling the degree to which each of the valves  526 ,  528  is closed, pressures are applied to the opposite faces of the pistons to control the variator. The left and right-hand flow lines  512 ,  514  combine downstream of the control valves  26 ,  28  and normally lead to drain  530 . 
   The  FIG. 7  arrangement as described thus far is generally conventional. However, it will be observed that a respective one of two identical flow control valves  532 ,  534  is located in each of the left and right-hand flow lines  512 ,  514  between the pistons  510  and the respective control valves  526 ,  528 . The flow control valves  532 ,  534  are of conventional construction and operate to limit or cap the flow of hydraulic fluid which can pass through the valves to the respective control valve  532 ,  534 . If fluid flow into the flow control valves  532 ,  534  exceeds the predetermined limit or cap value (which, typically, may be 5 l/min) any excess flow is discharged to drain  536 ,  538 , the capped flow being supplied to the respective control valve  526 ,  528 . Flow control to the control valves  532 ,  534  is unaffected by rates of flow below the predetermined limit or cap value. 
   When transient ratio changes occur due to manoeuvres such as braking, rapid input (engine) acceleration or abuse, the pistons  510  move rapidly in response, ejecting large flows of hydraulic fluid into one of the flow lines  512 ,  514 . However, as a result of the flow control valves  532 ,  534 , the flow rate received by the control valves  526 ,  528  never exceeds the predetermined value set by the flow control valves  532 ,  534 . Thus, extremes of flow line pressure and loss of system controllability and abuse tolerance associated with the prior art can be avoided. 
   The construction of the embodiment illustrated in  FIG. 8  is similar to that of  FIG. 7  and the same reference numerals are used to denote the same features as in  FIG. 7 . However, the significant difference is the replacement of the flow control valves  532 ,  534  of  FIG. 7  with two identical flow boost valves  542 ,  544 , a respective one valve being located in each of the left- and right-hand flow lines  512 ,  514 , and a hydraulic fluid accumulator  546  being connected to the flow boost valves  542 ,  544 . 
   The flow boost valves  542 ,  544  are of conventional construction and each valve is arranged to connect the accumulator  546  to the flow lines  512 ,  514  in which the valve  542 ,  544  is located when the flow rate in the associated flow line  512 ,  514  falls below a preset value, typically 5 l/min. This ensures that the fluid flow in the flow lines does not fall below a minimum value. For flow rates in excess of the preset value in the flow line  512 ,  514 , the accumulator is isolated from the flow line by the valve  542 ,  544 . 
   Thus, additional pressure from the accumulator  546  is supplied to either or both of the flow lines  512 ,  514  whenever the flow rate in the flow line  512 ,  514  falls below a predetermined value. During rapid ratio changes, one side of the pistons  510  will be evacuated. The rate of evacuation may be so great that cavitation or relative flow would normally occur, thereby contributing to control problems. However, by using the flow boost valves,  542 ,  544 , hydraulic pressure from the accumulator  546  is supplied to one or both of the flow lines  512 ,  514  whenever the evacuation rate exceeds a predetermined level, thereby ensuring that the flow lines  512 ,  514  are maintained at or above a miminum pressure. 
   In the embodiment of  FIG. 8 , the flow boost valves  542 ,  544  are connected downstream of the pistons  510  and upstream of the main flow control valves  526 ,  528 . However, the flow boost valves may be located at other positions within the valve circuit, upstream of the main control valves  526 ,  528 . For example, and as variator of the toroidal-race rolling-traction type, the invention is equally applicable to other types of variators (e.g. the band-and-sheave type variators) as indicated previously.