Abstract:
A radial piston pump includes cylinders oriented radially to an axis of rotation of an eccentric shaft and pistons arranged radially movably in the cylinders against the force of a spring member such that the pistons are pressed radially outwards by the rotational movement of an eccentric and radially inwards by the spring members. The pistons have at least one inlet opening connected to an inlet chamber of a pumping medium in the radially inner position of the pistons, and the pumping medium is pressed into a pressure area during the radially outward movement of the pistons. An eccentric shaft is mounted in sliding bearings arranged on both sides of the eccentric and the shaft is traction driven. A pressure connection between an annular duct and one of the sliding bearings advantageously provides a bearing gap between the sliding bearing and the eccentric shaft to constantly supply a close film of oil which has a damping effect upon the radial movements of the eccentric shaft. Thus, there is less noise due to the mechanical contact of the eccentric shaft with the sliding bearing.

Description:
BACKGROUND OF THE INVENTION 
     The invention relates to a radial piston pump, with cylinders oriented radially to an axis of rotation of an eccentric shaft, and with pistons arranged radially movably in the cylinders against the force of a spring. The pistons are pressed radially outwards by the rotational movement of an eccentric and are pressed radially inwards by the spring. The pistons have an inlet opening connected to an inlet chamber of a pumping medium when the pistons are in the radially inner positions. The pumping medium is pressed into a pressure area during the radially outward movement of the pistons. The eccentric shaft is mounted in sliding bearings arranged on both sides of the eccentric and is drivable by a traction means. 
     Radial piston pumps of this type are known. The alternating radial inward and outward movements of the pistons in the cylinders pump a medium, for example oil, is conveyed in a known manner. Radial piston pumps of this type are used for levelling systems in motor vehicles for example. In that case, the radial piston pump is driven by a belt drive which is driven by an internal-combustion engine of the motor vehicle. The belt engages on a drive wheel of the radial piston pump in order to rotate the eccentric shaft of the radial piston pump. The arrangement of the radial piston pump applies a belt force having a radial direction vector upon the eccentric shaft by the belt drive. The direction vector and the amount of the belt force are substantially constant. 
     In addition, the eccentric shaft is loaded by hydraulic forces which are introduced by the pistons of the radial piston pump and which likewise have a radial direction vector. A resulting hydraulic force of the radial piston pump, formed from partial hydraulic forces, is produced in accordance with the number of pistons of the radial piston pump. In this case the level and the direction vector of the resulting hydraulic force vary during use of the radial piston pump for its intended purpose in accordance with a rotational speed of the eccentric shaft. The constant belt force is overlaid by the variable hydraulic force, causing the eccentric shaft to be acted upon with a varying radial force. The resulting hydraulic force (also referred to as the “bearing force” below) has to be removed by the sliding bearings in which the eccentric shaft is mounted. 
     With large volumes in the radial piston pump and high hydraulic pressures, the resulting hydraulic forces can have a greater total than the belt force and, depending upon their operative direction, the hydraulic forces can cause a change in direction of the resulting force acting upon the eccentric shaft. In this way, the eccentric shaft can be pressed onto the sliding bearing against the belt force by the hydraulic forces. In this case, the actual resulting hydraulic force determines the direction vector of the resulting bearing force of the eccentric shaft and thus specifies a position of the eccentric shaft in the sliding bearing. 
     A drawback of this is that the change in position of the eccentric shaft in the sliding bearings can generate noise, a so-called knocking, as well as increased wear. In particular, if the radial piston pump is suction-throttled and is operated heavily regulated, phases can occur in which none of the pistons of the radial piston pump conveys the pumping medium, so that the eccentric shaft is oriented exclusively by the belt force as a result of the absence of hydraulic forces. At the beginning and the end of this phase, the resulting bearing force changes abruptly with respect to its direction vector, so that a reciprocating movement of the eccentric shaft occurs in the sliding bearings. 
     In addition, the hydraulic force acting upon the eccentric shaft does not change continuously, but changes abruptly, with respect to both the amount and the direction vector. Depending upon whether a piston of the radial piston pump begins or ceases to convey, the hydraulic force and thus the resulting bearing force produced by the superimposition with the belt force suddenly change. 
     It is known to lubricate the sliding bearings of the eccentric shaft in radial piston pumps with the pumping medium, for example oil. This oil is generally heavily foamed, particularly in the case of suction-regulated radial piston pumps, so that mixed friction of the eccentric shaft in the sliding bearings occurs as a result of air inclusions in the pumping medium. The mixed friction is not sufficient to damp the above-mentioned knocking of the eccentric shaft in the sliding bearings. 
     SUMMARY OF THE INVENTION 
     The object of the invention is to provide a radial piston pump of the above type which is simple in design and which prevents an eccentric shaft in a sliding bearing from knocking as a result of varying hydraulic forces which act upon the eccentric shaft. 
     This object is attained through a pressure connection present between the pressure area of the radial piston pump and at least one of the sliding bearings. It is advantageously possible for a bearing gap between the sliding bearing and the eccentric shaft to be constantly supplied with a closed film of oil which has a damping effect upon the radial movements of the eccentric shaft. This prevents the production of noise due to mechanical contact of the eccentric shaft with the sliding bearing. The radial piston pump as a whole operates more quietly. In particular, it is possible to counteract knocking by the superimposition of the hydraulic force acting upon the eccentric shaft and the belt force. 
     In a preferred embodiment of the invention, the pressure connection is formed by a duct which is formed in a housing of the radial piston pump and which opens with at least one outlet opening into the sliding bearing. This makes it possible to build up a volume flow of the pumping medium from the pressure area of the radial piston pump to the sliding bearing, and that volume flow performs the lubrication and damping of the sliding bearing. 
     In particular, the pumping medium is preferably conveyed into a radially central region of the sliding bearing. This makes a satisfactory distribution over the entire bearing surface of the sliding bearing possible, so that particularly good damping and lubrication can be achieved. 
     In a further preferred embodiment of the invention, the pressure connection opens in a range of ±90°, preferably ±50°, and in particular ±30°, with respect to a direction vector of the force of a traction means, in particular a belt traction force, acting upon the eccentric shaft. This advantageously makes it possible for the pressure build-up to first occur in particular in the region of the sliding bearing in which the eccentric shaft can be pressed against the bearing shell by the belt traction force, so that particularly good damping of the sliding bearing is provided in the direction of the belt traction force. 
     In addition, in a preferred embodiment of the invention, the pressure connection opens into a plurality of openings arranged preferably symmetrically over the periphery of the sliding bearing. This advantageously makes it possible for a uniform film of oil to be built up in the bearing gap between the eccentric shaft and the sliding bearing, enabling a high degree of damping of the sliding bearing in all radial directions, particularly in the case of radial piston pumps with high hydraulic forces which can be superimposed on the belt traction forces in an opposite manner. 
     Other objects and features of the invention are explained below in embodiments with reference to the accompanying drawings. 
    
    
     BRIEF DESCRIPTION OF THE DRAWINGS 
     FIG. 1 is an elevational sectional view of a radial piston pump; 
     FIG. 2 is an enlarged sectional view of the radial piston pump according to FIG. 1, and 
     FIGS. 3 to  6  are diagrammatic cross-sections through a sliding bearing of a radial piston pump in different embodiments. 
    
    
     DESCRIPTION OF PREFERRED EMBODIMENTS 
     FIG. 1 is a sectional view of a radial piston pump  10 . The radial piston pump  10  comprises a housing  12  in which a stepped bore  14  is formed. In order to form the stepped bore  14  the housing  12  may comprise a plurality of parts not explained individually below. The parts are connected to one another in a pressure-tight manner by suitable means. The stepped bore  14  receives an eccentric shaft  16  which carries an eccentric  18  located toward the axial center of the pump. 
     Sliding bearings  20  and  22  respectively, which mount the eccentric shaft  16 , are arranged on axially opposite sides of the eccentric  18 . Each sliding bearing comprises a respective bearing shell  24  which is inserted, for example pressed, into the stepped bore  14  of the housing  12 . In the regions of the sliding bearings  20  and  22 , the eccentric shaft  16  has portions  26  and  28  respectively of greater diameter, with external diameters adapted to the internal diameters of the respective bearing shells  24 . The diameters are adapted to one another in such a way that a slight bearing gap  30  remains between the shaft portions  26 ,  28  and the bearing shells  24 , respectively. Each bearing gap  30  is used to receive, in a manner to be explained below, a lubricant for the sliding bearings  20  and  22 , respectively. In addition, the eccentric shaft  16  is guided in seals  32  and  34  respectively (FIG. 2) which provide a pressure-tight mounting for the eccentric shaft  16 . 
     Cylinders  36 , which are oriented radially to an axis of rotation  38  of the eccentric shaft  16 , are inserted into the housing  12  in the axial region of the eccentric  18 . The number of the cylinders  36  can vary with different radial piston pumps  10 . In this way, it is possible for only one cylinder  36  or for a plurality of cylinders  36  to be provided, optionally arranged uniformly over the periphery of the eccentric  18 . A piston  40 , which is pressed against the eccentric  18  by the force of a spring  42 , is guided inside each cylinder  36 . The spring  42  is supported at one radially outward end on a plug  44  closing the cylinder  36  and at the other radially inward end on a base  46  of the piston  40 . The piston  40  has the shape of a cup, with one opening oriented in the direction of the plug  44 . At least one inlet opening  48  is provided in a peripheral wall of the piston  40 . In the example illustrated, four inlet openings  48  are arranged symmetrically around the periphery of the piston  40 . 
     A bore  50  leads from the cylinder  36  to an annular duct  52  in the housing  12 . A valve  54  is arranged between the bore  50  and the annular duct  52 . In the valve  54 , a closure member closes a connection between the bore  50  and the annular duct  52  against the force of a spring. The annular duct  52  is connected to a pressure connection  56  of the radial piston pump  10 . 
     In the region of the eccentric  18  the stepped bore  14  forms an inlet chamber  58  which is connected by at least one duct  60  to a suction connection  57  of the radial piston pump  10 . 
     The annular duct  52  is connected to a stepped bore  62  which extends substantially parallel to the axis of rotation  38 . A branch duct  66  leads from a portion  64  of the stepped bore  62  of smaller diameter to the sliding bearing  20 . A throttle  68  or diaphragm is arranged in the portion  64 . A step  70  of the stepped bore  62  receives a screen  72 . A diameter of the throttle  68  preferably amounts to from 0.1 to 0.5 mm, in particular from 0.15 to 0.3 mm. A mesh width of the screen  72  is somewhat finer than the diameter of the throttle  68  and preferably amounts to from 0.1 to 0.4 mm. 
     The shell  24  of the sliding bearing  20  has a through opening  74  which at one end is connected to the branch duct  66  and at the other end opens into a coaxial annular groove  76  in the bearing shell  24 , which is open in the direction of the portion  26  of the eccentric shaft  16 . 
     An extension  78  of the eccentric shaft  16  carries a flange  80  to which a drive wheel  82  is fastened by at least one fastening means  84 . The drive wheel  82  is pot-shaped and surrounds the housing  12  of the radial piston pump  10 . The free end of the drive wheel  82  is provided with a receiving means  86  for a drive belt (not shown). 
     The bases  46  of pistons  40  are supported on a bearing race  110  (FIG. 2) which is constructed in the form of a steel ring for example. The bearing race  110  is supported on the eccentric  18 . A plain bearing bush  112 , which is pressed into the bearing race  110 , is arranged between the eccentric  18  and the bearing race  110 . The eccentric shaft  16  has a through opening  114  which at one end opens on the periphery of the eccentric  18  and at the other end is connected to a pressure area inside the radial piston pump  10 . The pressure area is connected to the suction connection  57 . In this way, a pressure which corresponds to the pressure at the suction connection  57 , for example a tank pressure, is present in the through opening  114 , which is formed, for example, as a bore extending at an angle to the axis of rotation  38 . The through opening  114  preferably opens, as viewed in the axial extension of the eccentric  18 , in the middle region thereof. 
     The radial piston pump  10  shown in FIG. 1 operates as follows: 
     The general operation of a radial piston pump  10  is known, so that within the scope of the present description there is no need to go into this in greater detail. The drive wheel  82  and thus the eccentric shaft  16  are set in rotation by the traction means. The eccentric  18  mounted in a rotationally fixed manner on the eccentric shaft  16  rotates jointly in accordance with the rotation of the eccentric shaft  16 , so that in accordance with the eccentricity, the pistons  40  in abutting contact with the eccentric  18  have a radial lifting movement imparted to them. In this case, the pistons  40  are held at all times in abutting contact with the eccentric  18  by the spring  42 , so that an alternating radial movement directed inwards and outwards takes place. Upon inward movement the inlet openings  48  overlap with the inlet chamber  58 , so that the inner space of the piston  40  is filled with a medium to be conveyed, for example oil. This pumping medium is forced, through a decreasing volume of a space surrounded by the cylinder  36  in the piston  40 , into the bore  50  by the subsequent movement of the piston radially outwards. In this way the valve  54  is opened, so that the pumping medium passes into the annular duct  52  and from there through the stepped bore  62  to the pressure connection  56  of the radial piston pump  10 . When a plurality of pistons  40  are provided, they pump all the medium into the annular duct  52  in accordance with the principle described. The annular duct  52  is thus situated in a pressure area of the radial piston pump  10 . 
     A pressure connection is built up with the sliding bearing  20  by way of the stepped bore  62 , its portion  64  and the branch duct  66 . In this case the throttle  68  arranged in the portion  64  limits a volume flow of the pumping medium which flows from the pressure area of the pump to the sliding bearing  20 . Since the sliding bearing  20  is not sealed off in the direction of the inlet chamber  58  (see FIG.  2 ), circulation occurs between the pressure area and the suction area of the radial piston pump  10  by the sliding bearing  20 . In this case an exact volume flow can be set in accordance with the setting of the throttle  68 . The penetration into the sliding bearing  20  of impurities possibly taken up is prevented by the screen  72  positioned upstream of the throttle  68 . These impurities are deposited on the screen  72 . This prevents clogging of the throttle  68 . 
     The bearing gap  30  is provided with an oil film (with oil as the pumping medium) by the adjusted volume flow by way of the sliding bearing  20 . The oil film is distributed over the bearing gap  30  by the annular groove  76  which is preferably arranged coaxially with the axis of rotation  38  and is situated centrally with respect to an axial extension of the portion  26 . In this case, the oil under pressure is forced into the annular groove  76  through the through opening  74 , so that the oil is distributed over the annular groove  76 . The oil under pressure present in the annular gap  30  causes the sliding bearing  20  to be lubricated in a reliable manner. Since the sliding bearing is lubricated satisfactorily with oil foamed to an insignificant extent, knocking movements of the eccentric shaft  16 , which occur as a result of the superimposition of a belt traction force (to be explained hereinafter) and a hydraulic force acting upon the eccentric shaft  16 , are damped. 
     In the embodiment illustrated, only the sliding bearing  20  is acted upon with an oil flow under pressure. In further embodiments, the sliding bearing  20  can likewise be acted upon, additionally or optionally exclusively, with pressure oil. For this purpose, suitably adapted connecting paths have to be provided from the pressure area of the radial piston pump  10  to the sliding bearing  20 . 
     The through opening  114  in the eccentric shaft  16  improves lubrication between the eccentric  18  and the plain bearing bush  112 . Because of a relatively high relative speed between the bearing race  110  and thus the plain bearing bush  112  and the eccentric  18 , it is necessary to lubricate this area in order to prolong the service life and to damp noise. Since the medium to be conveyed (oil) is heavily foamed in the inlet chamber  58 , this medium alone would not be sufficient to perform adequate lubrication. The oil in the eccentric space  58  is heavily foamed, since the oil flow drawn in is already throttled upstream of the inlet chamber  58 . In this way, an under-pressure is present at the same time in the inlet chamber  58 . Oil, which is insignificantly foamed and which is at the starting pressure (tank pressure), now passes through the through the opening  114  between the eccentric  18  and the plain bearing bush  112 . As a result of a pressure drop between the inlet chamber  58  and the through opening  114 , a constant oil flow is made available for lubricating the plain bearing bush  112 . 
     FIG. 2 is a detailed view of an enlargement in part of the radial piston pump  10 , the arrangement of the pressure connection between the pressure area of the radial piston pump  10  and the sliding bearing  20  being shown in particular. The same parts are provided with the same reference numerals as in FIG.  1  and are not explained further. 
     In particular, the pressure connection between the pressure area (annular duct  52 ) and the suction area (inlet chamber  58 ) of the radial piston pump  10  is labeled as reference numeral by an arrow  88  in FIG.  2 . The pressure connection  88  is made to the inlet chamber  58  through the stepped bore  62 , the portion  64  thereof, the branch duct  66 , the through opening  74 , the annular groove  76  and the bearing gap  30 . 
     Radial sections through the portion  26  of the eccentric shaft  16  and thus the sliding bearing  20  are shown in each case in FIGS. 3 to  6 . 
     The through opening  74  opening into the annular groove  76  of the bearing shell  24  is shown in FIG.  3 . The through opening  74  is connected to the branch duct  66  which in turn opens into the portion  64  of the stepped bore  62 . The pressure oil is distributed over the entire periphery of the portion  26  of the eccentric shaft  16  by the annular groove  76 . The bearing gap  30 , the size of which is dependent upon a bearing clearance, is distributed over the annular groove  76 . In this way, a thin film of an oil under pressure is built up as it were between the portion  26  and the bearing shell  24 . Sufficient oil is thus present, which, in addition, is only moderately foamed, so that a hydrodynamic lubricating film can be built up in the sliding bearing. 
     In addition, an arrow  90 , which corresponds to a direction vector of a belt traction force F, is indicated in FIG.  3 . The belt traction force F acts upon the eccentric shaft  16  and has a direction vector which is dependent upon the action of a belt drive upon the drive wheel  82 . The direction vector of the belt traction force F is dependent upon the installation point of the radial piston pump  10 , for example in a motor vehicle with respect to an internal-combustion engine, which drives the belt. The direction vector and an amount of the belt traction force F are ideally constant. In FIG. 3, the through opening  74  opens into the annular groove  76  substantially opposite the operative direction of the belt traction force F. In further embodiments, the through opening  74  can open at any point in the annular groove  76  and thus with respect to the operative direction of the belt traction force F. 
     With a known fitted position of the radial piston pump  10 , the through opening  74  can open into the bearing gap  30  in a defined position with respect to the operative direction of the belt traction force F by insertion of the pressure connection in a desired manner between the pressure area of the radial piston pump  10  and the sliding bearing  20 . 
     A preferred area  91 , inside which the through opening  74  opens with respect to the operative direction of the belt traction force F, is indicated in FIG.  4 . The area  91  encloses an angle α in and opposite a direction of rotation of the eccentric shaft  16  by the direction vector  90 . In FIG. 4, the direction of rotation is assumed to be in the clockwise direction (arrow  92 ). The angle α amounts for example to 90°, preferably to 50° and in the embodiment illustrated in particular to 30°. In accordance with the illustration shown, inside the angle α the through opening  74  is arranged offset by an angle β of about 10° in the direction of rotation  92  with respect to the operative direction  90  of the belt traction force F. This makes it possible for the pressure oil to flow into the bearing gap  30 , into an area which—as viewed from the axis of rotation  38 —is situated in the radial direction which is substantially in the operative direction of the belt traction force F. The pressure oil is distributed from this area  91  through the bearing gap  30  over the entire periphery of the sliding bearing  20 . Since the cross-section for the volume flow of the pressure oil increases from the cross-section of the through opening  74  to the inlet chamber  58  (FIG. 2) in accordance with the design of the bearing gap  30 , a slight build-up of pressure will occur at an increasing distance from the opening of the through opening  74 . If the said opening is now situated in the said area  91  with respect to the belt traction force F, the greatest build-up of pressure will occur there, so that the belt traction force F can be compensated. In particular, if the belt traction force F is superimposed by an hydraulic force acting in the same operative direction as the belt traction force F, satisfactory damping of the clearance of the eccentric shaft  16  is achieved in the sliding bearing  20 . The operative direction of the hydraulic force is not indicated in FIGS. 3 and 4, since it rotates, in terms of both the amount and the direction vector, in accordance with the rotational speed of the eccentric shaft  16 , the volume flow of the radial piston pump  10  and the number of the pistons  40  following simultaneously and/or in succession. The hydraulic force is superimposed upon the belt traction force F so as to form a resulting bearing force by which the portion  26  of the eccentric shaft  16  is pressed against the bearing shell  24 . This resulting bearing force likewise has a rotating direction vector with a different amount which is dependent upon the momentary direction vector of the hydraulic force from the constant direction vector of the belt traction force F. If viewed graphically, it produces an elliptical curve of the resulting bearing force about the axis of rotation  38 . As a result of the pressure oil introduced into the bearing gap  30 , a damping of the radial movement of the portion  26  of the eccentric shaft  16  in the sliding bearing  20  is achieved independently of the amount and the direction vector of the resulting bearing force. 
     In the embodiment illustrated in FIG. 4, the arrangement of the annular groove  76  is omitted. The through opening  74  thus opens directly as a lubrication bore relief into the bearing gap  30 . In accordance with a further embodiment, an annular groove corresponding to the through opening  74  can be arranged in the portion  26  of the eccentric shaft  16 . 
     The arrangement of the through opening  74  with respect to a maximum pressure point P max  of the eccentric shaft  16  is shown in FIG.  5 . In this case the pressure point P max  corresponds to the point at which the greatest resulting bearing force F L  can occur, which is derived from the superimposition of the belt traction force F and the hydraulic force. The pressure point P max  can be determined from the fitted position of the radial piston pump  10  and the theoretically calculable maximum hydraulic forces. In this case the through opening  74  opens into an area  96  which is situated either in or opposite the direction of rotation  92  by an angle γ about a point  98  (radial), the point  98  being situated in front of the pressure point P max  by an angle δ opposite the direction of rotation  92 . As a result, the pressure oil in the bearing gap  30  flows into the bearing gap  30  in the angular range ±γ with respect to the angle δ and is taken up by the rotational movement of the eccentric shaft  16  into the area of the maximum pressure point P max . In this way, a constant high pressure, which results in a reliable damping of the movement of the eccentric shaft  16  in the sliding bearing  20 , can build up in the bearing gap  30  in the area of the maximum pressure point P max . The angle δ preferably amounts to 30° and the angle γ preferably amounts to 15°. 
     FIG. 6 shows a further variant of embodiment, in which an annular groove  100  is formed in the housing  12 . The branch duct  66  opens into the annular groove  100 . The annular groove extends coaxially around the bearing shell  24 . In the region of the annular groove  100  the bearing shell  24  is provided with at least one through opening  102 , six through openings  102  in the example illustrated, by way of which the pressure oil arrives in the bearing gap  30 . In this case the through openings  102  are arranged symmetrically over the periphery of the bearing shell  24 . In accordance with further embodiments the arrangement of the through openings  102  can be made in such a way that they are arranged at smaller intervals in the area of the maximum pressure point P max  and/or the area of the operative direction of the belt traction force F. 
     A combination of the different variants of embodiment illustrated in FIGS. 3 to  6  is possible. In this way, in particular in accordance with a further embodiment, it can be provided that the bearing shell  24  comprises two partial bearing shells which are arranged at a slight axial distance from each other in order to form the annular groove  76 . 
     Although the present invention has been described in relation to particular embodiments thereof, many other variations and modifications and other uses will become apparent to those skilled in the art. It is preferred, therefore, that the present invention be limited not by the specific disclosure herein, but only by the appended claims.