Abstract:
A sliding vane pump comprising a rotor housing having a pumping chamber, a rotor in the pumping chamber having a plurality of radially disposed slots, and a plurality of sliding vanes disposed in the slots configured to extend to follow an inner wall of said pumping chamber.

Description:
TECHNICAL FIELD  
         [0001]    This invention relates to sliding vane pumps.  
         BACKGROUND OF THE INVENTION  
         [0002]    Sliding vane pumps are typically used to provide hydraulic pressure and flow to various types of hydraulic systems, such as hydraulic power assist steering systems in automobiles. One example of a common sliding vane pump includes a rotor eccentrically mounted in a cylindrical chamber. As the rotor rotates, vanes within the rotor slide in and out to follow the contour of the housing, pushing fluid from an inlet port to an outlet port in the process.  
           [0003]    Another style is referred to as a hydraulically balanced sliding vane pump, which uses a rotor configuration such as that shown in FIG. 2. In this configuration, rotor  12  is centrally located in an oblong, or elliptical chamber  15  defined by pump ring  13 . Chamber  15  includes two inlets  16  and two outlets  18 , with rotor  12  rotating counter-clockwise as shown. Vanes  22  are radially disposed in radial slots  24 . Under the influence of fluid pressure from down stream of the outlets  18 , vanes  22  are urged out of slots  24  to follow the contour of chamber  15 . Vanes  22  therefore urge fluid along in spaces, or pumping cavities, between the vanes from the inlets  16  to outlets  18  as rotor  12  rotates.  
           [0004]    The function of the pumping cavity is to transfer a discrete volume of fluid at low pressure to high pressure. This happens repeatedly during a rotation of the pump shaft due to the presence of multiple pumping cavities. The end result is a steady flow of fluid discharged from the discharge port. Ideally for this to occur, the volume of fluid in the pumping cavity will be compressed and just reach the particular discharge pressure as it is allowed to enter the discharge port, providing a smooth transition from low to high pressure. However, this is seldom the case in practice. During operation, the pressure of the fluid in the pumping cavity is not the same as the pressure of the fluid in the discharge port just prior to the leading vane passing the discharge port. If the pressure in the cavity is lower than the pressure in the discharge port, fluid will quickly flow into the pumping cavity as the leading vane passes the discharge port. Conversely, if the pressure in the pumping cavity is higher than that in the discharge port, then fluid will quickly flow out of the pumping cavity as the leading vane passes the discharge port. This flow pulse is superimposed upon the steady flow of oil discharged from the discharge port. This small but quick flow pulse results in a corresponding pressure pulse (positive or negative) in the discharge port when the leading vane passes the discharge port.  
           [0005]    Since the pressure pulse occurs every time the leading vane passes the discharge port, the pulse occurs at vane passage frequency. Since there are multiple vanes passing the discharge port during one revolution of the pump shaft, and the pump shaft is rotated at a constant speed, the vane passage frequency will be an integer multiple of the pump shaft rotation frequency.  
           [0006]    The pressure pulse acts upon components within the pump, and components located downstream of the pump, causing these components to vibrate at the corresponding frequency of the pulse. Vibration of these components can radiate sound that is undesirable.  
           [0007]    The annoyance of pump noise is due not only because it is loud, but also because it is tonal in nature, due to the repeating of discrete pumping cycles, which occur with equal time intervals between them, every time a vane passes the outlet ports of the pump.  
           [0008]    Another drawback to sliding vane hydraulic pumps is that their speed range is limited by cavitation of the fluid within the pumping chamber. Cavitation is the formation and collapse of low-pressure bubbles in liquids. These bubbles are caused by air or vapor absorbed or otherwise entrained in the hydraulic fluid. Cavitation greatly increases pressure ripple which causes excessive noise and vibration, as well as loss of performance. The use of a jet supercharger to increase the inlet pressure to the pump has been used to increase the speed at which cavitation becomes audible, but these efforts have not sufficiently increased cavitation speed for many applications. Another approach has been to remove the air from the hydraulic fluid, but this has proven to be difficult in practice. Yet another drawback is that sliding vane pumps have a fixed capacity, i.e., they pump a fixed amount of fluid in each revolution of the rotor. This is a serious drawback of this type of pump in certain applications. For example, in the automotive industry, the hydraulic pump is often driven by an internal combustion engine that operates at a speed independent of the needed hydraulic power.  
         SUMMARY OF THE INVENTION  
         [0009]    The above-listed drawbacks and disadvantages of the prior art sliding vane pumps are overcome or alleviated by a high speed dual discharge sliding vane hydraulic pump with two external discharge ports for varying the capacity of the pump and/or irregularly spaced vanes to reduce the tonal characteristics of noise caused by pressure ripple effects and/or increasing the inlet slot length to increase the cavitation speed. 
       
    
    
     BRIEF DESCRIPTION OF THE DRAWINGS  
       [0010]    The present invention will now be described, by way of example, with reference to the accompanying drawings, in which:  
         [0011]    [0011]FIG. 1 is a cut-away perspective view of a hydraulic pump;  
         [0012]    [0012]FIG. 2 is a schematic representation of a typical sliding vane hydraulic pump of the prior art;  
         [0013]    [0013]FIG. 3 is another schematic representation of a pump;  
         [0014]    [0014]FIG. 4 is a graph comparing the pressure ripple amplitudes for the sliding vane hydraulic pump of FIG. 2 with that of FIG. 3;  
         [0015]    [0015]FIG. 5 shows a plan view of a conventional pump ring;  
         [0016]    [0016]FIG. 6 shows a plan view of another pump ring;  
         [0017]    [0017]FIG. 7 shows a graph comparing the pressure ripple performance of the pump rings of FIGS. 5 and 6;  
         [0018]    [0018]FIG. 8 is a cross section view of a conventional discharge housing;  
         [0019]    [0019]FIG. 9 is a cross section view of another discharge housing; and  
         [0020]    [0020]FIG. 10 is a schematic representation of an exemplary hydraulic system making use of a pump having the discharge housing of FIG. 9. 
     
    
     DESCRIPTION OF THE PREFERRED EMBODIMENT  
       [0021]    [0021]FIG. 1 shows a cut-away view of pump  100 . Pump  100  includes a rotor housing  102  supporting a pressure plate  108 , pump ring  106  and thrust plate  110 , which define the pump chamber  114  in which rotor  104  resides. Rotor rotor housing  102  and discharge housing  112  are preferably formed as different portions of a unitary structure, but are treated separately herein so that individual portions may be referred to more easily. Fluid enters rotor housing  102  through an external inlet (not shown) and is directed to annular space  116 . From annular space  116 , fluid enters internal inlets  118 , which are located on either side of rotor  104 . Internal inlets  118  are formed by notches in pump ring  106 , thrust plate  110 , and pressure plate  108 , described in further detail below. Fluid radially enters pump chamber  114  through these notches and is motivated by vanes  120  to axial internal discharge ports  122  (only one shown). There are actually four internal discharge ports, two located in pressure plate  108 , and two more in thrust plate  110 . A pressure plate cover (not shown) encloses the space immediately above pressure plate  108  and directs hydraulic fluid through axial ports  124  in pump ring  106  to discharge housing  112 . Pressure plate  108  has elongated curved slots  126  to direct high pressure fluid to spaces  128  behind each vane  120 , causing each vane  120  to slide out until the tip reaches the inside surface of pump ring  106 .  
         [0022]    Referring to FIG. 2, a schematic of a typical balanced hydraulic sliding vane pump  10  is shown, including rotor  12 , pump ring  13 , inlet ports  16 , discharge ports  18 , and  10  equally spaced sliding vanes  22 . During operation, the pressure of fluid in a pumping cavity  17  is sometimes not the same as the pressure of fluid in the discharge port just prior to the leading vane  14  passing the discharge port. If the pressure in cavity  17  is lower than the pressure in discharge port  18 , fluid will quickly flow into the pumping cavity as the leading vane  14  opens pumping cavity  17  to the discharge port. If the pressure in cavity  17  is greater than the pressure in discharge port  18 , fluid will quickly flow out of the pumping cavity as the leading vane  14  opens pumping cavity  17  to the discharge port. This process is repeated as each vane opens the next pumping cavity to the discharge port.  
         [0023]    This small but quick flow pulse results in a corresponding pressure pulse (positive or negative) in the discharge port when each vane opens a pressure cavity to the discharge port. Since the pressure pulse occurs every time the leading vane passes the discharge port, the pulse occurs at vane passage frequency, the frequency being an integer multiple of the pump shaft rotation frequency. In the example shown in FIG. 2 having ten vanes, the frequency of the pressure pulse will be ten times the shaft rotation frequency.  
         [0024]    The pressure pulse acts upon components within the pump, and components located downstream of the pump, causing these components to vibrate at the corresponding frequency of the pulse, as well as harmonics thereof.  
         [0025]    [0025]FIG. 3 shows schematic representation of pumping chamber  114  having rotor  104  disposed therein having  12  unequally spaced vane slots  126  carrying vanes  120 . Slots  126  are located such that the angles between the first six consecutive slots, θ n , are not duplicated with the first six slots. However, the angles between the second six consecutive slots are identical to the angles between the first six slots as shown in the diagram. This repeating of the angles between slots  126  in the second set of six slots  126  provides for mechanical and hydraulic balance of rotor  104 . In other words, for each vane slot  126 , there is another vane slot  126  located 180 degrees, or on the opposite side of rotor  104  providing for mechanical balance. Where the pressure differential is not so great that perfect balancing must be maintained, the vanes may be at varying angles without the repetition described above. In such a configuration, it may be desirable to off-set vanes so that vanes on opposite sides of the rotor do not clear the outlet port simultaneously, thus further reducing pressure ripple effects.  
         [0026]    The uneven spacing of the slots  126  minimizes the periodicity of the pressure ripple that causes noise. By placing the vanes at unequal angles, the pump activity within one revolution of the pump is repeated at multiple frequencies, thereby spreading the sound energy to an increasing number of fundamental frequencies and their corresponding harmonics. Since this spread-spectrum, or broadband noise is much easier to mask by other ambient sounds than tonal noise, the pump noise is perceived to be lower. While the  12  vane configuration shown in FIG. 3 has proven advantageous in reducing tonal noise, it should be noted that a rotor having just one or two vanes set off-set from an equally spaced configuration would noticeably reduce the tonal noise generated by the pump.  
       EXAMPLE  
       [0027]    [0027]FIG. 4 compares the frequency spectrum of pressure ripple from a pump that has  12  unevenly spaced vanes (dashed line) to the conventional pump having  10  evenly spaced vanes (solid line). Note the existence of an increased number of harmonic tones that are interspersed in the spectrum for the pump with unevenly spaced vanes (dashed line). Note also the increase in the amount of energy in the spectrum. Even though the overall energy (spectral content) of the pressure ripple has increased, the annoyance is reduced because the source of the sound (pressure ripple) is more broadband and much less tonal in nature. The presence of the extra harmonics is indicative of the spreading of energy among many frequencies.  
         [0028]    Turning to FIG. 5, a conventional pump ring  13  is shown in plan view. Pump ring  13  includes two notches  26  which form part of the inlet  16  to chamber  15  (FIG. 2). For reasons unknown, these notches have traditionally matched the notch length in pressure plate  108  and thrust plate  106  shown in FIG. 1. FIG. 6 shows a pump ring  106 , having notches  130  of approximately 68 degrees. The shape of notches  130  can be seen clearly in FIG. 1. FIG. 1 also shows that notches  121  in pressure plate  108  and thrust plate  106  have not been extended, and remain at about 59 degrees.  
         [0029]    The inventors found that by lengthening notches  130  to approximately 68 degrees, the cavitation speed of the pump, i.e., the speed at which cavitation is initiated, is greatly increased, thus greatly increasing the operating speed range of the pump. In fact, pump  100  has reliably operated without cavitation at speeds as high as 7,000 rpm with a pump ring having a 68 degree notch. The inventors found that any lengthening of the inlet notches improves performance of the pump up to a maximum length where the inlets and outlets are not spaced apart by more than the width of a pumping cavity. At this inlet notch length, an effective seal cannot be maintained, and performance is adversely affected.  
       EXAMPLE  
       [0030]    [0030]FIG. 7 shows test result data comparing cavitation speed (the approximate speed at which cavitation is initiated) with notch length, in terms of the angle that the notch extends around a pump ring. The graph shows the pressure ripple in pounds per square inch for each speed from 600 to 6000 rpm. FIGS.  7  shows a pump with a 59 degree notch compared with a pump having a 72 degree notch. Note that, for the 59 degree notch, the pressure ripple greatly increases after 4000 rpm, indicating an inception of cavitation somewhere between 4,000 and 4,500 rpm. This is consistent with prior art pumps of this type. However, the pump having a 72 degree notch exhibits no cavitation all the way to 6000 rpm.  
         [0031]    These test results show that an unexpected significant increase in cavitation speed is realized by simply increasing the notch size. Further investigation may show that changing the size and/or shape of notches  121  (FIG. 1) of pressure plate  108  or thrust plate  110  may also be beneficial in increasing the cavitation speed.  
         [0032]    A conventional discharge housing  32  is shown in cross-section in FIG. 8. Here, the dual internal discharge ports  18  are in communication with a single external discharge port  34 . This combines the flows from both discharge parts  18  to provide a single output of pump  10  (FIG. 2) and ensuring that the rotor remains balanced.  
         [0033]    [0033]FIG. 9 shows a cross section of discharge housing  112  in which each internal discharge port  122  is connected to a separate external discharge port  134 ,  136 . The external discharge ports include primary external discharge port  134  and secondary external discharge port  136 . Having separate external discharge ports  134 ,  136  allows pump  100  to operate at one-half or full capacity. When operating at one half capacity, only primary external discharge port  134  is connected to a load while the secondary external discharge port  136  is connected to a low-pressure reservoir. Since only one side of rotor  104  is doing the actual work of pumping, the torque required to operate the pump is reduced by approximately one half.  
         [0034]    Of course, external discharge ports  134  and  136  are interchangeable and are designated “primary” and “secondary” only distinguish them, i.e., either port may be designated “primary” and be connected to the load when operating at half-capacity.  
         [0035]    An exemplary system  150  utilizing pump  100  will now be described with reference to FIG. 10. FIG. 10 schematically shows pump  100  providing pressure and flow to system  162 , which constitutes a load. System  162  may be any type of hydraulic power system, such as a hydraulic actuator, e.g., a lift, or a power transfer system such as an automotive variable transmission. Pump  100  is driven by shaft  103  which in turn is driven by motive power source  152 . Motive power source  152  may be an electric motor, an internal combustion engine, or other source of mechanical power. Fluid exits pump  100  by primary external discharge port  134  and secondary external discharge port  136 . Flow from primary external discharge port  134  passes directly to system  162  via path  137 .  
         [0036]    Flow discharged from system  162  is discharged to low pressure reservoir  168 , which is in communication with pump inlet  169 , from which it is divided and passed to respective internal inlets  118  (FIGS. 1, 3) to be repressurized.  
         [0037]    Flow from secondary external discharge port  136  passes to valve  156  which directs the flow to path  137  and/or jet supercharger  164 . Valve  156  includes an actuator (not shown) that receives signals along line  160  from control unit  158 . Control unit  158  operates to adjust valve  156  depending on the flow requirements of system  162 . When operating at full capacity, valve  156  directs all of the flow from secondary external discharge port  136  to path  137  to combine with the flow from primary external discharge port  134 , which will then be directed to system  162 . When operating at half capacity, all of the flow from secondary external discharge port  136  is directed to jet supercharger  164 , via supercharger inlet  167 . Jet supercharger  164  includes a nozzle  166 . As fluid passes through nozzle  166 , the fluid accelerates and entrains additional fluid from reservoir  168 , increasing the pressure at internal inlets  118  (FIGS. 1, 3), thereby improving performance and increasing its operating range. If the use of jet supercharger  164  is not desirable, it is of course contemplated that valve  156  could instead direct flow to another low pressure location, such as reservoir  168 , or to another system that requires hydraulic power.  
         [0038]    Referring again to FIG. 1, note that supercharger  164  is located just beneath discharge housing  112  and forms part of the structure of pump  100  or is otherwise fixedly attached to it. Low pressure inlet  169  is not visible in FIG. 1, but is located just to the left of supercharger inlet  167 . The flows are combined as schematically represented in FIG. 10, and the combined flow exits supercharger  164  and passes through opening  171  (FIG. 9) in discharge housing  112  to annular space  116  described above with reference to FIG. 1.  
         [0039]    It is contemplated that valve  156  could be also incorporated into the pump housing that comprises rotor housing  102  and discharge housing  112 . This would necessitate connecting only one line from reservoir  168  to pump  100  and only one line from pump  100  to system  162 . This would reduce installation time and improve reliability by reducing the number connections and hoses required. Such a configuration would include a housing for all the elements encompassed by box  200  in FIG. 10.  
         [0040]    Valve  156  is an on/off valve so that it can direct all the fluid from external discharge port  136  to either system  162  or supercharger  164 .  
         [0041]    Supercharger  164  presents a lower back-pressure to secondary external discharge port  136 , thereby reducing the overall torque required to drive pump  100  when operating at less than full capacity.  
         [0042]    While preferred embodiments have been shown and described, various modifications and substitutions may be made thereto without departing from the spirit and scope of the invention. Accordingly, it is to be understood that the present invention has been described by way of illustration only, and such illustrations and embodiments as have been disclosed herein are not to be construed as limiting to the claims.