Patent Abstract:
An axial bearing between a first part and a second part that presses with an axial load against the first part and can rotate around a rotation axis relative to the first part comprising a circular or arc-shaped ridge on the first part centered around the rotation axis, a pressure source for providing pressurized hydraulic fluid on a first side of the circular or arc-shaped ridge, an adjustable gap between the circular or arc shaped ridge and a bearing surface on the second part, wherein the pressurized hydraulic fluid flows through the adjustable gap to a second side of the circular or arc-shaped ridge. In accordance with the invention the circular or arc-shaped ridge or the bearing surface include a ridge chamber for locally creating a larger adjustable gap between the circular or arc-shaped ridge and the bearing surface.

Full Description:
CROSS-REFERENCE TO RELATED APPLICATIONS 
     The present application is a continuation of pending International patent application PCT/EP2010/054702 filed on Apr. 9, 2010 which designates the United States and claims priority from European Patent applications EP 09161738.1 filed on Jun. 2, 2009 and EP 09158296.5 filed on Apr. 20, 2009. The content of all prior applications is incorporated herein by reference. 
    
    
     FIELD OF THE INVENTION 
     The invention concerns an axial bearing between a first part and a second part that presses with an axial load against the first part and can rotate around a rotation axis relative to the first part comprising a circular or arc-shaped ridge on the first part centered around the rotation axis, a pressure source for providing pressurized hydraulic fluid on a first side of the circular or arc-shaped ridge, an adjustable gap between the circular or arc shaped ridge and a bearing surface on the second part, wherein the pressurized hydraulic fluid flows through the adjustable gap to a second side of the circular or arc-shaped ridge. 
     BACKGROUND OF THE INVENTION 
     Such axial bearings are known for instance from hydraulic devices such as pumps, motors or transformers where they are used as bearing between the rotor and the port plate. Other known uses are hydrostatic bearings in a gearbox with helical gears or hydrostatic bearings in other machinery. 
     In the axial bearing, the pressure difference between both sides of the circular or arc-shaped ridge causes the oil pressure to fall in the adjustable gap when going from the first side to the second side. How the pressure falls and whether a pressure profile from one side of the ridge to the other side of the ridge is linear, progressive or digressive determines the force that this pressure generates to counteract the axial load and with a given axial load determines the gap-height. 
     For the pressure profile, the shape of the gap is very important. Especially important is whether the walls of the gap, seen in flow direction, are parallel, diverging, or converging. As the gap-height is small, from 2 to 15 microns, minor changes in temperature distribution in the walls of the gap create changes in the diverging or converging of the walls so that the pressure profile in the gap often is unpredictable. The walls of the gap influence the flow in such narrow gaps considerably and the flow theories using laminar or turbulent flow models do not describe the situation properly. As the walls of the adjustable gap move relative to one another, there is a viscous friction. The viscous friction increases with the speed of the relative movement as the gap gets narrower and/or the speed increases and decreases with increasing gap-height. The viscous friction generates heat in the oil that might influence the gap-height due to change in dimensions of the ridge or the bearing surface. 
     In the known axial bearings, it is very difficult to optimize the axial bearing. A too small axial load leads to a large height of the adjustable gap due to the oil pressure in the adjustable gap between the circular or arc-shaped ridge and the bearing surface. This can lead to a too large oil flow through the gap. This large oil flow will arise if, for average oil viscosity, the average height of the gap is more than 10-20 micron and the pressure of the pressurized hydraulic fluid is more than 10 MPa. 
     If the axial load is too large, there is too much friction during rotation of the rotor combined with heating of the oil flow due to viscous losses in the adjustable gap. In addition, in an adjustable gap that is very narrow, local deformations or local disturbances in the flow through the gap may occur which might lead to further local heat generation. Local heat generation leads to deformations of the circular of arc-shaped ridge or the bearing surface and to further narrowing of the gap. These deformations might lead to undesired wear as metallic contact between the rotating and stationary parts may occur. 
     SUMMARY OF THE INVENTION 
     In order to reduce the disadvantages the axial bearing, the circular or arc-shaped ridge and the bearing surface comprise a ridge chamber for locally creating a larger adjustable gap between the circular or arc-shaped ridge and the bearing surface. Because of this feature, the pressure profile in the adjustable gap is more stable as the pressure in the ridge chamber is constant. The pressure changes from the one side of the circular of arc-shaped ridge to the other side take place over a considerably reduced distance so that variations in the pressure profile have less influence on the force counteracting the axial load and have less influence on the gap-height. A further result is that over a considerable surface of the adjustable gap the gap-height is higher which strongly reduces the viscous friction in the adjustable gap. In situations where the walls of the adjustable gap have a high relative speed, locally increasing the gap-height strongly reduces the friction and heat generation. This leads to less energy loss and less deformation due to local high temperatures in the walls of the adjustable gap on the circular or arc-shaped ridge and the bearing surface. This reduces the risk of metallic contact and so reduces wear. 
     In an embodiment, the ridge chamber has a surface that is at least 50% of the surface of the circular or arc shaped ridge. This ensures that for at least half the surface the friction between rotating parts is considerably reduced, which means that there is a considerable reduction or possibly halving of the viscous friction between the two parts. 
     In an embodiment, the ridge chamber has a depth of more than 10-30 microns. This ensures that in the ridge chamber there is sufficient oil of a constant pressure. In this way oil pressure in the ridge chamber counteracts the axial load with a constant force that is little influenced by the gap-height. 
     In an embodiment, a first slot connects the ridge chamber with the first side of the circular or arc-shaped ridge. This ensures that always a certain amount of oil flows into the ridge chamber and that the pressure of the oil in the ridge chamber can have a value that is more or less between the pressures on both sides of the circular or arc-shaped ridge. The oil pressure in the ridge chamber is now less dependent on the gap-height of the sides of the ridge chamber and is therefore less dependent on the deformations or shape of the walls of the adjustable gap. This reduces the risk that the axial load might reduce the gap-height too much and cause too much viscous friction or metallic contact and wear. 
     In an embodiment, a second canal connects the ridge chamber with the second side of the circular or arc-shaped ridge. This ensures that always a certain amount of oil flows out of the ridge chamber and that the pressure of the oil in the ridge chamber can have a value that is more or less between the pressures on both sides of the circular or arc-shaped ridge. The oil pressure in the ridge chamber is now less dependent on the gap-height of the sides of the ridge chamber and so is less dependent on the deformations or shape of the walls of the adjustable gap. This reduces the risk of too much flow of oil of high-pressure through the adjustable gap and with that of unnecessary energy loss. 
     In an embodiment, a slot from the ridge chamber to the first side or the second side forms the first canal or the second canal respectively and the width of the slot is less than half its height. This ensures that the gap-height has only little influence on the opening of the canal so that changing the cap height does not change the inflow in the ridge chamber or the outflow from the ridge chamber and ensures that the changing gap-height has only little influence on the pressure in the ridge chamber. 
     In an embodiment, the first or the second canal has valve means to adjust the flow resistance of the canal. This ensures that the height of the adjustable gap is adapted to the actual situation. In situations that the rotation speed is high, it is advantageous to reduce the friction in the adjustable gap. In that situation, reducing the flow resistance in the first canal and/or increasing the flow resistance in the second canal leads to a higher pressure in the ridge chamber and to a higher gap, which gives less friction. 
     In situations with high pressure of the pressurized fluid and relative low rotation speed, the major source of energy loss is leakage of oil through the adjustable gap. Increasing the flow resistance in the first canal and/or reducing the flow resistance in the second canal leads to a lower pressure in the ridge chamber and to a narrower gap. The narrow gap has less leakage and so reduces the energy loss. 
     In an embodiment, the rotation speed of the first part or the second part controls the valve means, preferably the valve means are set by a centrifugal force generated by the rotation in the part. This ensures in a simple way that the axial bearing adapts to a large range in the rotation speed. 
     In an embodiment, the pressure of the hydraulic fluid on the first side controls the valve means, preferably the valve means are set by the pressure of the pressure source on the first side. This ensures in a simple way that the axial bearing adapts to a large range of the pressure of the hydraulic fluid in the pressure source. 
     In an embodiment, the axial load depends on the pressure of the hydraulic fluid on the first side. This ensures that the gap-height is independent of the pressure of the hydraulic fluid in the pressure source. 
     In an embodiment, the pressure source provides hydraulic fluid between two concentric circular or arc-shaped ridges and two radial ridges connecting the circular or arc-shaped ridges. This ensures a small area with the high pressure of the hydraulic fluid in the pressure source. This small area limits the length of the ridges surrounding it so that oil leakage through the gaps between the ridges and the bearing surface is smaller. 
     The invention also concerns a hydraulic transformer with 4-quadrant operation for use in vehicle drive system. In the known hydraulic transformers, the rotors are in the centre and the barrel plate rests against the port plate, the covers support the inclined port plates. The rotor and the shaft guide the radial forces on the pistons via a bearing to the covers. The radial forces generated by the port plates on the covers counteract in the covers these radial forces. However, the forces on the shaft are considerable and lead to bending and elastic deformations that are a disadvantage as this might lead to oscillations and leakage. In addition, the setting of the hydraulic transformer by rotating both port plates synchronously is complicated. 
     In order to overcome these disadvantages the hydraulic transformer comprises a housing with covers at opposite sides, in the housing a shaft with a rotation axis, two rotors each in axial direction supported by a first axial bearing, pistons mounted in the rotors, two inclined barrel plates that rotate with the rotors, barrel sleeves supported by the barrel plates, a chamber formed by a barrel sleeve and a piston, wherein the volume of the chamber changes during rotation of the rotor and a swath plate with a second axial bearing between the swath plate and the barrel plate characterized in that the covers each support or are part of a port plate that is part of the first axial bearing, further that the rotors are between the first axial bearings and that both swash plates are located between the rotors and support or are part of a swash block, the hydraulic transformer further comprising an actuator for rotating the swash block. In this way, the swash block leads radial forces from the barrel plate via a short way to the rotor and pistons so that the deformations are minimal. Further, the setting of the hydraulic transformer is easy by rotating the swash block. Rotating the swash block sets the top dead centre angle on both sides of the swash block simultaneously. 
     In an embodiment, the actuator comprises a rotary cylinder mounted in the swash block. This ensures a direct hydraulic rotation of the swash block that ensures quick setting of the hydraulic transformer. 
     In an embodiment, the housing comprises a sensor for detecting the rotary position of the swash block. In this way, an accurate setting of the hydraulic transformer is possible. 
     The invention also concerns a vehicle with a hydraulic drive system. In the known systems the motor/pump unit and the hydraulic transformer are coupled directly. This leads to the situation when the setting of the hydraulic transformer has as result that the motor/pump unit exerts a braking torque on the wheel that after the wheel has stopped rotating the braking torque starts acting as a driving torque in reverse direction if the setting of the hydraulic transformer is not changed immediately. For instance during parking of the vehicle, this could lead to undesirable situations. 
     In order to overcome this disadvantage the hydraulic drive system comprises a common high-pressure rail with an high-pressure accumulator, a common low-pressure rail with a low-pressure accumulator, an internal combustion engine driving a constant displacement pump connected to the common high-pressure rail and the common low-pressure rail, for each front wheel or for each rear wheel a motor/pump unit and a hydraulic transformer with 4-quadrant operation with connections to the common high-pressure rail and via a first motor line and a second motor line to the motor/pump unit characterized in that the hydraulic transformer comprises a forward propulsion valve or a reverse propulsion valve connecting the common low-pressure rail respectively to the first motor line or to the second motor line, which propulsion valves have a spring to hold the valve in a first position wherein they act as check valve blocking the flow to the common low-pressure rail and an actuator that can switch the propulsion valve to a second position connecting the common low-pressure rail to one of the motor lines. In this way, a wheel can only rotate in one direction unless the control system changes the setting of a valve. This prevents undesired or unexpected rotations of the wheels. 
    
    
     
       BRIEF DESCRIPTION OF THE DRAWINGS 
       The invention will be explained in more detail below with reference to several exemplary embodiments by means of a drawing, in which: 
         FIG. 1  schematically shows the components of a hydraulic drive of a car, 
         FIG. 2  shows a diagram of a drive and brake system of a hydraulic driven wheel of a car driving forward, 
         FIG. 3  shows a diagram of the drive and brake system of a hydraulic driven wheel of a car driving reverse, 
         FIG. 4  shows a diagram of the drive and brake system of a hydraulic driven wheel of a car riding forward and braking, 
         FIG. 5  shows a diagram of the drive and brake system of a hydraulic driven wheel of a car riding reverse and braking, 
         FIG. 6  shows a perspective view of a hydraulic transformer assembly for use in the hydraulic drive of a car, 
         FIG. 7  shows a perspective view of the hydraulic transformer of  FIG. 6  with a cut out and opened housing showing the internal parts, 
         FIG. 8  shows an exploded view of the main parts of the hydraulic transformer of  FIGS. 6 and 7  excluding the housing, 
         FIG. 9  shows a perspective view with a cut out of the housing of the hydraulic transformer assembly of  FIG. 6-8  without the rotating parts and end covers, 
         FIG. 10  shows a section through the hydraulic transformer of  FIGS. 6-9  with an actuator for setting the transformer control angle, 
         FIG. 11  shows a longitudinal section through the hydraulic transformer of  FIGS. 6-10 , 
         FIG. 12  shows graphically the pressure quotient of the pressure of the operation pressure and the high pressure in dependence of a transformer control angle, 
         FIG. 13  shows a perspective view of sealing area on a rotating part, and 
         FIG. 14  shows a schematic section through a sealing area between a rotating and a stationary part. 
     
    
    
     DETAILED DESCRIPTION OF THE INVENTION 
       FIG. 1  shows a passenger car  12  with the various components of a hydraulic drive system for the car wherein all four wheels of the car  12  are driven. The drive system comprises an internal combustion engine  2  that drives a constant displacement pump  4  that pumps hydraulic fluid from a common low-pressure rail  6  to a common high-pressure rail  5 . The common low-pressure rail  6  is connected to a low-pressure accumulator  8  and the common high-pressure rail  5  is connected to a high-pressure accumulator  9 . A drive control system  1  controls the internal combustion engine  2  and this drive control system  1  maintains by controlling the rotation speed and/or of the starting or stopping of the internal combustion engine  2  that the hydraulic pressure in the common high-pressure rail  5  is between a high and a low value. 
     The front wheels of the passenger car  12  each have a front wheel motor/pump  3  that is connected to a front axle hydraulic transformer  7 . Document WO97/31185 describes the operation principle of a hydraulic transformer; hereafter the design of the hydraulic transformer is further elucidated. The front axle hydraulic transformer  7  is also connected to the common high-pressure rail  5  and the common low-pressure rail  6  and is controlled by the drive control system  1 . The rear wheels of the passenger car  12  each have a rear wheel motor/pump  11  that is connected to a rear axle hydraulic transformer  10 . The rear axle hydraulic transformer  10  is connected to the common high-pressure rail  5  and the common low-pressure rail  6  and is controlled by the drive control system  1 . In other embodiments of passenger cars  12 , only the front wheels are driven or only the rear wheels are driven. The hydraulic drive systems for these cars are similar and form a simplified version of the described embodiment. Hydraulic drive systems for commercial vehicles are similar with front wheel drive, rear wheel drive, or four-wheel drive as well. The wheel motor/pump  3 ,  11  is designed such that it acts as a motor for driving the wheel  22  and that it acts as a pump for braking the wheel  22 . 
       FIGS. 2-5  schematically show how a wheel motor/pump  3 ,  11  drives and brakes an attached wheel  22  with a wheel rotation direction  23 ; the shown design is for all wheels similar. Braking the rotation of the motor/pump  3 ,  11  and the wheel  22  by pumping hydraulic fluid back into the common high-pressure rail  5  recuperates the kinetic energy of the speeding vehicle. The wheels  22  have an additional brake system of conventional design used for emergency braking if required and for braking during standstill or parking. An interrupted line in the  FIGS. 2-5  indicates the hydraulic transformer  7 ,  10  as such. The motor/pump  3 ,  11  is directly coupled to the wheel  22 . A first motor/pump connection  26  and a second motor/pump connection  25  connect the motor/pump  3 ,  11  to the hydraulic transformer  7 ,  10 . The first motor/pump connection  26  connects to a first user connection port  13  of the hydraulic transformer  7 ,  10 . The second motor/pump connection  25  connects to a second user connection port  18  of the hydraulic transformer  7 ,  10 . The common high-pressure rail  5  connects via a high-pressure connection HP to the high-pressure port  15  of the hydraulic transformer  7 ,  10 . 
     The first motor/pump connection  26  further connects via a reverse propulsion valve  24  and a low-pressure connection LP to the common low-pressure rail  6  and the second motor/pump connection  25  connects via a forward propulsion valve  20  and the low-pressure connection LP to the common low-pressure rail  6 . The forward propulsion valve  20  and the reverse propulsion valve  24  each have two positions. A spring  19  pushes the valves  20 ,  24  in a first position and an actuator  21  controlled by the drive control system  1  can bring the valves  20 ,  24  in the second position. In the first position, a check valve in each valve  20 ,  24  prevents flow from the first, respective the second motor/pump connection  25 ,  26  to the low-pressure connection LP and in the second position the first, and respective the second motor/pump connection  25 ,  26  have an open connection to the low-pressure connection LP. 
     In the first position of the valves  20 ,  24  oil flow is only possible from the low-pressure connection to the hydraulic transformer  7 ,  10  so that the wheel motor/pump  3 ,  11  can only act as pump and the wheel  22  has to supply energy and brakes independent of the setting of the hydraulic transformer  7 ,  10 . This means that with the valves  20 ,  24  in the first position inadvertently driving the wheels  22  is not possible. 
       FIGS. 2-5  show the hydraulic transformer  7 ,  10  schematically with the three ports  13 ,  15  and  18  that are part of a port plate  30  (see  FIGS. 8 ,  9 ,  11 ) and shown as arcs around a circle indicating a rotation group  17 . A top dead centre TDC indicates the setting of a top dead centre of piston movement  14  in the rotation group  17  at varying transformer control angles δ. In the circle, an arrow  16  indicates the direction of rotation of the rotation group  17 . The areas p and m indicate where in the rotation group  17  a volume of a chamber  65  (see  FIG. 11 ) above the pistons  42  (see  FIGS. 8 ,  9 ,  11 ) decreases or increases during rotation of the rotation group  17  areas and acts as in a pump or motor respectively. 
       FIG. 2  shows the transformer control angle δ set so that hydraulic pressure in the high-pressure port  15  drives the rotation of the rotation group  17  of the hydraulic transformer  7 ,  10 . The pistons in the area p in the rotation group  17  pump the hydraulic fluid via the first user connection port  13  and the first motor/pump connection  26  to the wheel motor/pump  3 ,  11 . The setting of the transformer control angle δ determines the pressure of the hydraulic fluid in the first motor/pump connection  26  and so determines the driving torque. The reverse propulsion valve  24  is closed so that the hydraulic fluid flows only to the wheel motor/pump  3 ,  11  and causes the wheel  22  to rotate in the rotation direction  23  and the passenger car  12  starts moving at an increasing speed. The forward propulsion valve  20  is in the second position so that hydraulic fluid flowing at low-pressure from the wheel motor/pump  3 ,  11  through the second motor/pump connection  25  can flow to the low-pressure connection LP and to the second user connection port  18  of the hydraulic transformer  18 . 
       FIG. 4  shows the transformer control angle δ set at an opposite angle as compared to the situation shown in  FIG. 2  and the forward propulsion valve  20  is closed as well. In this setting, the wheel motor/pump  3 ,  11  exerts a braking torque on the rotating wheel  22  so that its speed reduces. The wheel motor/pump  3 ,  11  now acts as pump and it pumps hydraulic fluid through the second motor/pump connection  25  to the second user connection port  18 . In the hydraulic transformer  7 ,  10 , the hydraulic fluid expands in the chambers above the pistons of the rotation group  17  in the area m. These pistons drive the rotation group  17  in the direction indicated with the arrow  16 . The chambers above the pistons connect first to the second pump user connection port  18  and after that to the high-pressure port  15 . When the chambers are connected to the high-pressure port  15 , the pistons in the rotation group  17  compress hydraulic fluid to the high-pressure connection HP. The wheel motor/pump  3 ,  11  supplies the energy required for this compression by pumping hydraulic fluid at a raised pressure in the second motor/pump connection  25  and this results in a braking torque on the wheel  22 . The setting of the transformer control angle δ determines the pressure of the hydraulic fluid in the second motor/pump connection  25  and so determines the braking torque. The first user connection port  13  and the low-pressure connection LP via the check valve in the forward propulsion valve  20  provide the hydraulic fluid that the wheel motor/pump  3 ,  11  pumps in the second motor/pump connection  25 . 
       FIGS. 3 and 5  show the settings of the hydraulic transformer  7 ,  10 , the forward propulsion valve  20  and the reverse propulsion valve  24  respectively in the situation that wheel motor/pump  3 ,  11  exerts a reverse driving torque on the wheel  22  and the situation that the wheel motor/pump  3 ,  11  brakes the reverse rotating wheel  22 . The various settings and flows of hydraulic fluid are similar to those described for  FIGS. 2 and 4 . 
       FIG. 4  shows braking of the wheel  22  when the vehicle is driving forward. The setting of the transformer control angle δ is similar to the situation as shown in  FIG. 3  wherein the wheel motor  3 ,  11  exerts a reverse driving torque on the wheel  22 . The difference is the setting of the reverse propulsion valve  24 . During braking as shown in  FIG. 4 , at the moment of standstill of the wheel  22  the rotor in the hydraulic transformer  7 ,  10  stops rotating. The rotation group  17  cannot start to rotate in the opposite direction (as is possible in the situation shown in  FIG. 3 ) due to the settings of the propulsion valves  20 ,  24  and the wheel remains stationary. In this way the propulsion valves  20 ,  24  act to release a driving torque in the desired direction of rotation of a wheel independent of the setting of the hydraulic transformer  7 ,  10 . In a situation that the drive control system  1  is switched off the springs  19  will set the propulsion valves  20 ,  24  in a position that the wheel motor/pump  3 ,  11  can only generate a braking torque so that undesired acceleration of the wheels  22  is prevented under all circumstances. 
       FIGS. 6 and 7  show external views of a hydraulic transformer assembly  27 , which comprises the hydraulic transformer  7 ,  10  with the propulsion valves  20 ,  24 .  FIGS. 8 and 9  show the various components inside the housing  52  of the hydraulic transformer assembly  27  in perspective view.  FIGS. 10 and 11  show respectively a cross section and a longitudinal section of the hydraulic transformer assembly  27 . 
     The hydraulic transformer assembly  27  includes the components as shown in  FIGS. 2-5  such as the hydraulic transformer  7 ,  10 , the propulsion valves  20 ,  24  and an actuators  21  for each propulsion valve  20 ,  24 . The first motor/pump connection  26  and the second motor/pump connection  25  each connect the transformer assembly  27  to two front wheel motor/pumps  3  or to two rear wheel motor/pumps  11 . A housing  52  has at both ends covers  28 , a rim aligns the covers  28  inside the housing  52 . Bearings  31  are mounted in the covers  28 , the bearings  31  support a shaft  34 . At both ends of the shaft  34  there is a rotor  32 . The shaft has outer splines  37  that cooperate with the inner splines  39  of the rotor  32  so that both rotors  32  rotate with the shaft  34 . Both rotors  32  have pistons  42  whereby the inner and outer splines  37 ,  39  are set in such a way that the rotative positions of the pistons  42  of one rotor  32  are between the rotative positions of the pistons  42  of the other rotor  32 . 
     A pin  76  synchronizes the rotation of a barrel assembly  33  comprising a barrel plate  56  and cups  40  with the rotation of the rotor  32 . The shaft  34  supports a swivel bearing sphere  64  that supports a spherical swivel bearing  44  of the barrel plate  56  so that the barrel plate  56  can swivel relative to the rotor  32 . A spring  62  pushes at one side against a support ring  61  that is fixed on the inside of the rotor  32 . The spring  62  pushes at its other side against pressure pins  63  that push against the swivel bearing sphere  64  and so push the barrel plate  56  and the rotor  32  in opposite directions. The barrel plate  56  supports cups  40  which are mounted side by side and between cup positioners  55 . A cup holding plate  54  holds the cups  40  and the cup positioners  55  on the barrel plate  56 . 
     Pistons  42  are mounted on rotor  32  and each forms with the cup  40  a chamber  65  that has a changing volume. The piston  42  has a piston canal  38  that extends through the rotor  32  and forms a canal with a port  43  in a port plate  30 . The port plate  30  has a pin  66  that maintains the port plate  30  in a fixed rotative position in the cover  28  and with that relative to the housing  52 . From the port  43  the canal continues as a canal in the cover  28  and a canal  29  in the housing  52  to the first user connection port  13 , second user connection  18  or the high pressure connection HP (as shown in  FIGS. 2-5 ). 
     Bearings  35  are mounted on the shaft  34  and support a swash block  36  that can rotate a limited angle in the housing  52 . The swash block  36  has at both sides inclined swash plate surfaces  41  that support the barrel plates  56 . The barrel plates  56  swivel around the swivel bearing sphere  64  and rest against the inclined swash plate surfaces  41  so that the pistons  42  move in and out the cup  40  during rotation of the shaft  34 . Due to the swiveling movement the volume of the chamber  65  changes between a minimum and a maximum value. By rotating the swash block  36  in the housing  52  the rotative position of the rotor  32  where the volume of the chamber  65  is minimal, which is the top dead centre TDC indicated with 53 can be set to a desired value. 
       FIG. 10  shows the top dead centre  53  of the swash plate surfaces  41 . In the shown embodiment the swash plate surfaces  41  at both sides of the swash block  36  intersect in a line perpendicular to the rotation axis of shaft  34  so that the top dead centre  53  for the volume of the chambers  65  at both sides of the swash block  36  is at the same rotative position and as the pistons  42  on the one side of the swash block  36  are between the pistons  42  on the other side of the swash block  36 , the minimum value at both sides of the swash block  36  follow each other. 
     In the outer circumference of the swash block  36  there is a groove with moving vanes  45  diametrically opposite each other and sealing against the inner surface of the housing  52 . In the housing  52  there are diametrically opposed stationary vanes  47 . The stationary vanes  47  and the moving vanes  45  form in the housing four pressure chambers  46  that have a TDC control connection ports  48  connected to a swash block control valve (not shown). The pressure chambers  46  rotate the swash block  36  in the housing  52 . The swash block  36  has a detector groove  49  that cooperates with a sensor (not shown) for detecting the rotative position of the swash block  36 . 
     The moving vanes  45  are mounted on the swash block  36  in such a way that the top dead centre of the swash plate  53  can rotate over 97 degrees in one direction and 69 degrees in the opposite direction. This asymmetry makes it possible to set the hydraulic transformer assembly  27  in such a way that the first user connection port  13  has a higher pressure than the high-pressure port  15 . In this way it is possible when the common high-pressure rail  5  has a lower pressure than the maximum pressure on which the hydraulic transformer assembly  27  can operate, which occurs during normal driving in order to be able to recuperate kinetic energy during braking, to bring full the maximum hydraulic pressure on the first motor/pump connection  26  and make maximum acceleration of the vehicle possible. 
       FIG. 12  shows the quotient of the first user connection port  13  and the high-pressure port HP in dependence of the angle δ of the top dead centre  53  of the swash plate surfaces  41 . A line  51  shows the pressure quotient in dependence of the transformer control angle δ. An operating range  50  of the hydraulic transformer is chosen such that although the transformer can be used for driving and braking in both directions of rotation (four-quadrant use) the settings of the transformer are asymmetrical so that the driving torque can be higher than the braking torque. 
     The oil pressure in the chambers  65  pushes the barrel plate  56  against the swivel block  36  and the rotor  32  against the port plate  30 . This is the main axial force, except in situations where the oil pressure is very low. In that situation the force of the spring  62  presses the rotor  32  and the barrel plate  56  against respective the port plate  30  and the swivel block  36  in order to prevent oil leakage and facilitate starting. The forces on the rotor  32  in the axial direction of the rotation axis of the shaft  34  created by the oil pressure in the chambers  65  are necessary for creating a seal in the second axial bearing  59  and are in part balanced by forces of oil pressure in the piston canal  38  and the port  43  in the second axial bearing  59  between the rotor  32  and the port plate  30 . 
     The forces on the barrel plate  56  caused by the oil pressure in the chambers  65  and are necessary for creating a seal in the first axial bearing  57 . These forces are in part balanced by forces of oil pressure in the first axial bearing  57 . For this a barrel plate canal  58  connects the chamber  65  and the first axial bearing  57 . The forces in axial direction on both sides of the swivel block  36  are more or less identical in opposite direction so that this brings no load on the bearings  35 . 
     The forces in radial direction on the swivel block  36  are guided through the respective bearing  35  and the outer splines  37  via the inner splines  39  to the pistons  42  where they are counteracted by the radial hydraulic forces on the pistons  42  that are caused by the asymmetric surface to which the hydraulic pressure subjects those piston  42 . Due to the slight inclination of the swath plate surface  41  these forces are limited and cause no undesirable loads or deformations. 
     The hydraulic transformer has two first axial bearings  57  and two second axial bearings  59 . In these bearings  57 ,  59  a rotating part, the rotor  32  or the barrel plate  58 , with a number of canals with fluid of high pressure, respectively the piston canal  38  and the barrel plate canal  57 , seals against a stationary part, respectively the port plate  30  and the swivel block  36 . In prior art the sealing comprises a rim that is pressed against a flat surface with a narrow gap in the range from 2to 14micron between them. A narrow gap of limited height reduces the leakage over the sealing. The disadvantage of a too narrow gap is that it brings the risk that local deformation in one of the parts, for instance due to local heat generation, leads to local metallic contacts and so to lack of lubrication and to undesired wear. 
       FIGS. 13 and 14  show the first and second axial bearing  57 ,  59  of the hydraulic transformer assembly  27 .  FIG. 13  shows a perspective view of the rotor  32  showing the second axial bearing  59 . An outer ridge  67 , inner ridge  68  and radial ridges  69  surround a recess that forms the end of the piston canal  38  in the rotor  32 . In  FIG. 13  the ridges  67 ,  68  and  69  are indicated in black and this black surface is the surface that seals against a sealing surface of the port plate  30 . Each recess around a piston canal  38  connects intermittently to one of the three ports  43  in the port plate  30  and during the passage from one port  43  to the next port  43  the radial ridge  69  blocks the oil flow between the ports  43  by sealing on the bridge between the ports  43 . The outer ridge  67  and the inner ridge  68  are provided with ridge chambers  70  that have a surface of approximately 50% of the surface of the ridges  67 ,  68  respectively. When the piston canal  38  is connected to a pressure source the ridges  67 ,  68  and  69  form an adjustable gap with the sealing surface of the port plate  30 . Where there is a ridge chamber  70 , which has a depth of at least 10-30 micron, the gap is higher and the viscous friction between the parts when rotating is reduced. The radial ridges  69  interrupt the ridge chambers  70 . The depth of the ridge chambers  70  is at least 10-30 micron so that the viscous friction is reduced. 
     The oil pressure in the chambers  70  will be average between the hydraulic pressure on the both sides of the inner or outer ridge  67 ,  68  if the gaps on both sides of the chamber  70  are identical. In practice this is often not the case. If for instance the gap on the side of the piston canal  38  is a smaller than the gap on the other side of the chamber  70  the pressure in the chamber can be very low and the rotor  32  might be pressed towards the port plate  30  and the viscous friction increases. If the situation is the other way round the pressure in the chamber  70  might be high and the gaps get higher so that the leakage increases. The difference in the height of the gaps of a few microns might lead to these situations and also slight deformation in the ridges  67 ,  68  and  69  might lead to instability in the height of the gaps. In order to stabilize this, a slot  73  connects the chamber  70  with the high pressure side of the ridge  67 ,  68 . The width of the slot  73  must be small and it is relatively deep in order minimize the influence of a changing gap-height. In practice the slot  73  is 30 micron wide and 30 micron deep, preferably its width is half of its depth. 
       FIG. 14  shows in a schematic section the second axial bearing  59 . The schematic section show for each ridge  67 ,  68  a low-pressure side rim  71  and a high-pressure side rim  72 . A canal  76  connects the chamber  70  between the low-pressure side rim  71  and the high-pressure side rim  72  with the piston canal  38  via a connecting line  74  with a restriction  75 . This restriction can be adjusted by the control system, or in the parts are mechanical means that set the restriction in dependence of the circumstances of use. The restriction can for instance depend on the pressure in the piston canal  38  or it can depend on the rotation speed of the rotor  32 . 
     In addition to the above described embodiment of the axial bearing, wherein the hydraulic pressure is supplied between ridges that form a short arc near each piston canal  38 , other embodiments of axial bearings can have two concentric rings between which an oil flow with hydraulic pressure is supplied. Such embodiments can be used in machinery that has no pistons but where axial loads are generated and where the axial bearing guides these loads to a housing. In this machinery the pressure of the axial load causes a hydraulic pressure in the axial bearing, there will be control means to set the adjustable gap so that oil loss and friction resistance are optimized.

Technology Classification (CPC): 5