Patent Publication Number: US-9890782-B2

Title: Fluid pump with radial bearing between inner rotor and rotary shaft and lubrication groove in outer peripheral surface of radial bearing

Description:
CROSS REFERENCE TO RELATED APPLICATION 
     This application is based on and incorporates herein by reference Japanese Patent Application No. 2015-81916 filed on Apr. 13, 2015. 
     TECHNICAL FIELD 
     The present disclosure relates to a fluid pump that draws and discharges fluid by changing a volume of respective pump chambers formed between external teeth of an inner rotor and internal teeth of an outer rotor. 
     BACKGROUND 
     A previously proposed fluid pump has an inner rotor, an outer rotor, a pump housing and a rotatable shaft. The inner rotor includes external teeth, and the outer rotor includes internal teeth for meshing with the external teeth. The pump housing receives the inner rotor and the outer rotor. The rotatable shaft drives the inner rotor to rotate the same. When the inner rotor is rotated by rotating the rotatable shaft, a rotational force of the inner rotor is transmitted from the external teeth to the internal teeth. Thereby, the outer rotor is also rotated. When the inner rotor and the outer rotor are rotated, the volume of the respective pump chambers, which are formed between the external teeth and the internal teeth, changes. In response to increasing of the volume of the pump chamber, the fluid is drawn into the pump chamber. Thereafter, in response to decreasing of the volume of the pump chamber, the fluid is compressed in the pump chamber and is discharged from the pump chamber (see, for example, JP2013-60901A). 
     In a case where a repulsive force, which is applied from the fluid to the inner rotor, is large, like in a case where viscosity of the fluid is high, a force (tilting force), which is applied from the fluid to the inner rotor in a direction for tilting the inner rotor relative to the rotatable shaft, is increased. Thereby, a slide resistance between a radial bearing, which rotatably and slidably supports the rotatable shaft, and the rotatable shaft is increased to cause an increase in the energy loss or generation of damage at a sliding portion between the radial bearing and the rotatable shaft. 
     With respect to the above point, the inventors of the present application have studied a structure for coupling the inner rotor to the rotatable shaft through a joint member rather than directly coupling the inner rotor to the rotatable shaft. With this structure, the above-described tilting force can be absorbed through resilient deformation of the joint member, and thereby the slide resistance between the radial bearing and the rotatable shaft can be reduced. 
     In the above coupling structure, since the inner rotor is not directly coupled to the rotatable shaft, it is necessary to provide a member that rotatably and slidably supports the inner rotor. The inventors of the present application have studied a structure that slidably supports the rotatable shaft through a cylindrical inner peripheral surface of a radial bearing and also slidably supports the inner rotor through a cylindrical outer peripheral surface of the radial bearing. 
     However, the inventors of the present application have noticed that the above-described bearing structure poses the following new disadvantage. That is, the rotatable shaft is placed to extend over both of a high pressure passage, which conducts the fluid discharged from each corresponding one of pump chambers, and an inside of the pump housing. Thereby, the fluid in the high pressure passage penetrates into an area between the cylindrical inner peripheral surface of the radial bearing and the rotatable shaft to implement lubricating function. In contrast, it is difficult to provide a structure, which enables penetration of high pressure fluid between the cylindrical outer peripheral surface of the radial bearing and the inner rotor, so that the lubricating function of the fluid cannot be expected. Therefore, the slide resistance of the inner rotor cannot be sufficiently reduced in comparison to the slide resistance of the rotatable shaft. 
     That is, in the case where the above structure is adapted, although the tilting force can be absorbed through the joint member, there is required a structure that slidably supports the inner rotor. In this case, there is the new disadvantage of that the slide resistance of the inner rotor cannot be sufficiently reduced. 
     SUMMARY 
     The present disclosure is made in view of the above point. According to the present disclosure, there is provided a fluid pump that includes an inner rotor, an outer rotor, a pump housing, a rotatable shaft, a joint member and a radial bearing. The inner rotor is shaped into a cylindrical tubular form and has a plurality of external teeth. The outer rotor has a plurality of internal teeth for meshing with the plurality of external teeth. The pump housing receives the outer rotor and the inner rotor and forms a plurality of pump chambers between the plurality of internal teeth and the plurality of external teeth. Each of the plurality of pump chambers draws and compresses fluid by changing a volume of the pump chamber. The rotatable shaft is placed to extend over both of: a high pressure passage, which conducts the fluid discharged from each corresponding one of the plurality of pump chambers; and an inside of the pump housing. The joint member couples between the inner rotor and the rotatable shaft to transmit a rotational torque of the rotatable shaft to the inner rotor. The radial bearing is shaped into a cylindrical tubular form. The radial bearing rotatably and slidably supports the rotatable shaft through a cylindrical inner peripheral surface of the radial bearing and rotatably and slidably supports an inner peripheral surface of the inner rotor through a cylindrical outer peripheral surface of the radial bearing. At least one lubrication groove is formed in the cylindrical outer peripheral surface of the radial bearing and accumulates the fluid, which is present in the inside of the pump housing. 
    
    
     
       BRIEF DESCRIPTION OF THE DRAWINGS 
       The drawings described herein are for illustration purposes only and are not intended to limit the scope of the present disclosure in any way. 
         FIG. 1  is a partial cross-sectional view indicating a fuel pump according to an embodiment of the present disclosure; 
         FIG. 2  is a cross-sectional view taken along line II-II in  FIG. 1 ; 
         FIG. 3  is a cross-sectional view taken along line in  FIG. 1 ; 
         FIG. 4  is a cross-sectional view taken along line IV-IV in  FIG. 1 ; 
         FIG. 5  is a partial enlarged view of  FIG. 1 ; 
         FIG. 6  is a cross-sectional view of a radial bearing shown in  FIG. 5 ; 
         FIG. 7  is a cross-sectional view taken along line VII-VII in  FIG. 6 ; 
         FIG. 8  is a cross-sectional view showing a modification of the radial bearing shown in  FIG. 7 ; 
         FIG. 9  is a cross-sectional view showing another modification of the radial bearing shown in  FIG. 7 ; 
         FIG. 10  is a cross-sectional view showing another modification of the radial bearing shown in  FIG. 7 ; 
         FIG. 11  is a cross sectional view showing a modification of the radial bearing shown in  FIG. 6 ; 
         FIG. 12  is a cross sectional view showing another modification of the radial bearing shown in  FIG. 6 ; 
         FIG. 13  is a partial perspective view showing a lubrication groove formed in the radial bearing shown in  FIG. 6 ; 
         FIG. 14  is a partial perspective view showing a modification of the radial bearing shown in  FIGS. 6 and 13 ; and 
         FIG. 15  is a partial perspective view showing another modification of the radial bearing shown in  FIGS. 6 and 13 . 
     
    
    
     DETAILED DESCRIPTION 
     An embodiment of a fluid pump according to the present disclosure will be described with reference to the accompanying drawings. The fluid pump of the present embodiment is installed in a vehicle. A subject fluid to be pumped with the fluid pump is liquid fuel used for combustion in an internal combustion engine. Specifically, in the present embodiment, light oil (diesel fuel), which is used for combustion in a compression self-ignition internal combustion engine, is used as the subject fluid to be pumped. The fluid pump is received in an inside of a fuel tank. 
     As shown in  FIG. 1 , the fluid pump  101  of the present embodiment is a rotary internal gear pump of a positive displacement type. The fluid pump  101  includes a pump body  102 , a pump main body  103 , an electric motor  104  and a side cover  105 . The pump main body  103  and the electric motor  104  are received in an inside of the pump body  102 , which is shaped into a cylindrical tubular form, such that the pump main body  103  and the electric motor  104  are arranged one after another in an axial direction. The side cover  105  is installed to an opening of one of two axially opposite end parts of the pump body  102 , which is located on the electric motor  104  side. The side cover  105  includes an electric connector  105   a , which supplies an electric power to the electric motor  104 , and a discharge port  105   b , through which fuel is discharged from the fluid pump  101 . In the fluid pump  101 , a rotatable shaft  104   a  of the electric motor  104  is rotated when the electric power is supplied from an external circuit through the electric connector  105   a . Thus, an outer rotor  130  and an inner rotor  120  of the pump main body  103  are rotated by a drive force of the rotatable shaft  104   a  of the electric motor  104 , and thereby fuel is drawn into and compressed in the fluid pump  101  and is then discharged from the fluid pump  101  through the discharge port  105   b . The fluid pump  101  pumps the light oil, which has the higher viscosity in comparison to gasoline, as the fuel. 
     In the present embodiment, the electric motor  104  is an inner rotor brushless motor and includes magnets  104   b , which form four magnetic poles, and coils  104   c , which are installed in six slots. For example, at a start preparation time (e.g., a time of turning on of an ignition switch of the vehicle), a positioning control operation of the electric motor  104  is executed to rotate the rotatable shaft  104   a  toward a drive rotation side or a counter-drive rotation side (the counter-drive rotation side being opposite from the drive rotation side). Thereafter, the electric motor  104  executes a drive control operation, which rotates the rotatable shaft  104   a  from the position, at which the rotatable shaft  104   a  is positioned in the positioning control operation, toward the drive rotation side. 
     Here, the drive rotation side is a positive direction side of a rotational direction Ri of the inner rotor  120  in a circumferential direction of the inner rotor  120 . The counter-drive rotation side is a negative direction side of the rotational direction Ri of the inner rotor  120 , which is opposite from the positive direction side. 
     Hereinafter, the pump main body  103  will be described in detail. The pump main body  103  includes a pump housing  110 , the inner rotor  120 , the outer rotor  130  and a joint member  160 . The pump housing  110  includes a pump cover  112  and a pump casing  116 , which are placed one after another in the axial direction. 
     The pump cover  112  is made of metal and is shaped into a circular disk form. The pump cover  112  axially projects outward from the end part of the pump body  102 , which is located on the side of the electric motor  104  that is opposite from the side cover  105 . 
     In order to draw the fuel from an outside of the fluid pump  101 , the pump cover  112  shown in  FIGS. 1 and 2  has a suction passage  112   a , which is formed as a cylindrical hole, and a suction groove  113 , which is shaped into an arcuate form. In the pump cover  112 , the suction passage  112   a  is communicated with the suction groove  113  at a predetermined opening location Ss, which is eccentric from a central axis (hereinafter referred to as an inner central axis) Ci of the inner rotor  120 . The suction groove  113  is axially grooved, i.e., formed in an inside wall surface of the pump cover  112  and opens on the pump casing  116  side of the pump cover  112 . A communicating portion of the suction groove  113 , which is communicated with the suction passage  112   a , extends through the pump cover  112  in the axial direction. A non-communicating portion of the suction groove  113 , which is not directly communicated with the suction passage  112   a , is shaped into a cup form having a bottom. As shown in  FIG. 2 , the suction groove  113  has a circumferential extent, which is less than one half (less than 180 degrees) of an entire circumference of the inner rotor  120  in the rotational direction Ri (also see  FIG. 4 ). 
     The suction groove  113  extends from a start end part  113   c  to a terminal end part  113   d  in the rotational direction Ri, Ro such that a radial extent (hereinafter referred to as a width) of the suction groove  113 , which is measured in a radial direction of the rotational axis, progressively increases in the rotational direction Ri, Ro from the start end part  113   c  to the terminal end part  113   d . The suction passage  112   a  opens in a groove bottom portion  113   e  of the suction groove  113  at the opening area Ss, so that the suction groove  113  is communicated with the suction passage  112   a . As shown particularly in  FIG. 2 , in an entire range of the opening area Ss, in which the suction passage  112   a  opens, the width of the suction groove  113  is smaller than a width (diameter) of the suction passage  112   a.    
     Furthermore, the pump cover  112  forms an installation space  158  at an area that is opposed to the inner rotor  120  along the inner central axis Ci. The installation space  158  is shaped into a recessed hole. A main body  162  of the joint member  160  is rotatably installed in the installation space  158 . 
     The pump casing  116  shown in  FIGS. 1, 3, 4 and 5  is made of metal and is shaped into a cylindrical tubular form having a bottom. An opening portion  116   a  of the pump casing  116  is covered with the pump cover  112  such that an entire circumferential extent of the opening portion  116   a  is tightly closed by the pump cover  112 . As shown particularly in  FIGS. 1 and 4 , an inner peripheral portion  116   b  of the pump casing  116  is formed as a cylindrical hole that is eccentric relative to the inner central axis Ci of the inner rotor  120 . The pump casing  116  forms a discharge passage  117 , which is formed as an arcuate hole, to discharge the fuel from the discharge port  105   b  through a high pressure passage  106  defined between the pump body  102  and the electric motor  104 . The discharge passage  117  axially extends through a recessed bottom portion  116   c  of the pump casing  116 . Particularly, as shown in  FIG. 3 , the discharge passage  117  has a circumferential extent, which is less than one half (i.e., less than 180 degrees) of the entire circumference of the inner rotor  120  in the rotational direction Ri. A radial extent (hereinafter referred to as a width) of the discharge passage  117 , which is measured in the radial direction, progressively decreases in the rotational direction Ri, Ro from a start end part  117   c  to a terminal end part  117   d.    
     Furthermore, the pump casing  116  includes a reinforcing rib  116   d  in the discharge passage  117 . The reinforcing rib  116   d  is formed integrally with the pump casing  116  such that the reinforcing rib  116   d  extends across the discharge passage  117  in a crossing direction, which crosses the rotational direction Ri of the inner rotor  120 , and thereby the reinforcing rib  116   d  reinforces the pump casing  116 . 
     An opposing suction groove  118  shown in  FIG. 3  is formed in the recessed bottom portion  116   c  of the pump casing  116  at a corresponding area that is opposed to the suction groove  113  in the axial direction while pump chambers  140  (described later in detail) are interposed between the opposing suction groove  118  and the suction groove  113  in the axial direction. The opposing suction groove  118  is an arcuate groove that corresponds to a shape, which is produced by projecting the suction groove  113  onto the pump casing  116  in the axial direction. In this way, in the pump casing  116 , the discharge passage  117  is formed to be symmetric to the opposing suction groove  118  with respect to the symmetry axis located between the discharge passage  117  and the opposing suction groove  118 . As shown particularly in  FIG. 2 , an opposing discharge groove  114  is formed in the pump cover  112  at a corresponding area that is opposed to the discharge passage  117  in the axial direction while the pump chambers  140  are interposed between the opposing discharge groove  114  and the discharge passage  117  in the axial direction. The opposing discharge groove  114  is formed as an arcuate groove that is shaped to correspond with a shape, which is produced by projecting the discharge passage  117  onto the pump cover  112  in the axial direction. In this way, in the pump cover  112 , the suction groove  113  is formed to be symmetric to the opposing discharge groove  114  with respect to the symmetry axis located between the suction groove  113  and the opposing discharge groove  114 . An outline (contour) of the suction groove  113 , an outline (contour) of the opposing discharge groove  114 , an outline (contour) of the discharge passage  117 , and an outline (contour) of the opposing suction groove  118  are shaped to extend in parallel with a rotational path of the external teeth  124   a  and a rotational path of the internal teeth  132   a.    
     As shown in  FIG. 1 , a radial bearing  150  is securely fitted to the recessed bottom portion  116   c  of the pump casing  116  along the inner central axis Ci to radially support the rotatable shaft  104   a  of the electric motor  104  in a manner that enables rotation of the rotatable shaft  104   a . Furthermore, a thrust bearing  152  is securely fitted to the pump cover  112  along the inner central axis Ci to axially support the rotatable shaft  104   a  in a manner that enables the rotation of the rotatable shaft  104   a.    
     As shown in  FIGS. 1 and 4 , a receiving space  156 , which receives the inner rotor  120  and the outer rotor  130 , is formed by the recessed bottom portion  116   c  and the inner peripheral portion  116   b  of the pump casing  116  and the pump cover  112 . 
     The inner rotor  120 , which is indicated in  FIGS. 1 and 4 , is centered at the inner central axis Ci and is thereby coaxial with the rotatable shaft  104   a  (i.e., coaxial with a rotational axis of the rotatable shaft  104   a ), so that the inner rotor  120  is eccentrically placed in the receiving space  156 . An inner peripheral portion  122  of the inner rotor  120  is radially supported by the radial bearing  150 , and two slide surfaces  125  of the inner rotor  120 , which are respectively formed at two opposed axial ends of the inner rotor  120 , are supported by the recessed bottom portion  116   c  of the pump casing  116  and the pump cover  112 , respectively, in a manner that enables rotation of the inner rotor  120 . 
     The inner rotor  120  has a plurality of insertion holes  127  that extend in the axial direction at a corresponding area of the inner rotor  120 , which is opposed to the installation space  158 . In the present embodiment, the number of the insertion holes  127  is five, and these insertion holes  127  are arranged one after another at equal intervals in the circumferential direction along the rotational direction Ri. The insertion holes  127  extend through the inner rotor  120  from the installation space  158  side to the recessed bottom portion  116   c  side in the axial direction. Legs (projections)  164  of the joint member  160  are inserted into the insertion holes  127 , respectively, so that the drive force of the rotatable shaft  104   a  is transmitted to the inner rotor  120  through the joint member  160 . Thereby, the inner rotor  120  is rotated in the circumferential direction about the inner central axis Ci in response to the rotation of the rotatable shaft  104   a  of the electric motor  104  while the slide surfaces  125  of the inner rotor  120  are slid along the recessed bottom portion  116   c  and the pump cover  112 , respectively. 
     The inner rotor  120  includes a plurality of external teeth  124   a , which are formed in an outer peripheral portion  124  of the inner rotor  120  and are arranged one after another at equal intervals in the circumferential direction along the rotational direction Ri. Each of the external teeth  124   a  can axially oppose the suction groove  113 , the discharge passage  117 , the opposing discharge groove  114  and the opposing suction groove  118  in response to the rotation of the inner rotor  120 . Thereby, it is possible to limit sticking of the inner rotor  120  to the recessed bottom portion  116   c  and the pump cover  112 . 
     As shown in  FIGS. 1 and 4 , the outer rotor  130  is eccentric to the inner central axis Ci of the inner rotor  120 , so that the outer rotor  130  is coaxially received in the receiving space  156 . In this way, the inner rotor  120  is eccentric to, i.e., is decentered from the outer rotor  130  in an eccentric direction De, which is the radial direction. An outer peripheral portion  134  of the outer rotor  130  is radially supported by the inner peripheral portion  116   b  of the pump casing  116  in a manner that enables rotation of the outer rotor  130 . Furthermore, the outer peripheral portion  134  of the outer rotor  130  is axially supported by the recessed bottom portion  116   c  of the pump casing  116  and the pump cover  112  in a manner that enables the rotation of the outer rotor  130 . The outer rotor  130  is rotatable in the rotational direction (certain rotational direction) Ro about an outer central axis Co, which is eccentric to the inner central axis Ci. 
     The outer rotor  130  has a plurality of internal teeth  132   a  for meshing with the external teeth  124   a  of the inner rotor  120 . The internal teeth  132   a  are formed in an inner peripheral portion  132  of the outer rotor  130  and are arranged one after another at equal intervals in the rotational direction Ro. Each of the internal teeth  132   a  can axially oppose the suction groove  113 , the discharge passage  117 , the opposing discharge groove  114  and the opposing suction groove  118  in response to the rotation of the outer rotor  130 . Thereby, it is possible to limit sticking of the outer rotor  130  to the recessed bottom portion  116   c  and the pump cover  112 . 
     A fuel pressure (discharge pressure) in an inside of the discharge passage  117  is axially exerted against the inner rotor  120  and the outer rotor  130  toward the suction passage  112   a . A fuel pressure in the opposing discharge groove  114  is also the discharge pressure and is axially exerted against the inner rotor  120  and the outer rotor  130  toward the electric motor  104  side. Since the opposing discharge groove  114  is axially opposed to the discharge passage  117 , the fuel pressure of the opposing discharge groove  114  and the fuel pressure of the discharge passage  117  are balanced with each other. Therefore, it is possible to limit tilting of the inner rotor  120  and the outer rotor  130 , which would be otherwise caused by the discharge pressure. 
     Similarly, since the opposing suction groove  118  is axially opposed to the suction groove  113 , the fuel pressure (the suction pressure) of the opposing suction groove  118  and the fuel pressure (the suction pressure) of the suction groove  113  are balanced with each other. Therefore, it is possible to limit tilting of the inner rotor  120  and the outer rotor  130 , which would be otherwise caused by the suction pressure. The external teeth  124   a  and the internal teeth  132   a  are shaped to have a trochoid tooth profile. The number of the internal teeth  132   a  is set to be larger than the number of the external teeth  124   a  by one. The inner rotor  120  is meshed with the outer rotor  130  due to the eccentricity in the eccentric direction De. In this way, the pump chambers  140  are radially formed between the internal teeth  132   a  and the external teeth  124   a  in the receiving space  156 . A volume of each pump chamber  140  is increased and decreased through the rotation of the outer rotor  130  and the rotation of the inner rotor  120 . 
     The volume of each of opposing ones of the pump chambers  140 , which are axially opposed to and communicated with the suction groove  113  and the opposing suction groove  118 , is increased in response to the rotation of the inner rotor  120  and the rotation of the outer rotor  130 . Thereby, the fuel is drawn from the suction passage  112   a  into the corresponding pump chambers  140  through the suction groove  113 . At this time, since the width (radial extent) of the suction groove  113  progressively increases from the start end part  113   c  to the terminal end part  113   d  in the rotational direction Ri, Ro (also see  FIG. 2 ), the amount of fuel drawn into the pump chamber  140  through the suction groove  113  corresponds to the amount of increase in the volume of the pump chamber  140 . The corresponding ones of the pump chambers  140 , each of which draws the fuel by increasing its volume in the above-described manner, are referred to as negative pressure portions (or negatively pressurized pump chambers)  140 L. 
     The volume of each of opposing ones of the pump chambers  140 , which are axially opposed to and communicated with the discharge passage  117  and the opposing discharge groove  114 , is decreased in response to the rotation of the inner rotor  120  and the rotation of the outer rotor  130 . Therefore, simultaneously with the suctioning function discussed above, the fuel is discharged from the corresponding pump chamber  140  into the high pressure passage  106  through the discharge passage  117 . At this time, since the width (radial extent) of the discharge passage  117  progressively decreases from the start end part  117   c  to the terminal end part  117   d  in the rotational direction Ri, Ro (also see  FIG. 3 ), the amount of fuel discharged from the pump chamber  140  through the discharge passage  117  corresponds to the amount of decrease in the volume of the pump chamber  140 . The corresponding ones of the pump chambers  140 , each of which compresses the fuel by decreasing its volume in the above-described manner, are referred to as high pressure portions (or highly pressurized pump chambers or positively pressurized pump chambers)  140 H. 
     The joint member  160  is made of synthetic resin, such as poly phenylene sulfide (PPS). The joint member  160  relays the rotatable shaft  104   a  to the inner rotor  120  to rotate the inner rotor  120  in the circumferential direction. The joint member  160  includes the main body  162  and the legs  164 . 
     The main body  162  is installed in the installation space  158 , which is formed in the pump cover  112 . A fitting hole  162   a  is formed in a center of the main body  162 , and thereby the main body  162  is shaped into a circular ring form. When the rotatable shaft  104   a  is fitted into the fitting hole  162   a , the main body  162  is securely fitted to the rotatable shaft  104   a  to rotate integrally with the rotatable shaft  104   a.    
     The number of the legs  164  corresponds to the number of the insertion holes  127  of the inner rotor  120 . Specifically, in order to reduce or minimize the influence of the torque ripple of the electric motor  104 , the number of the legs  164  is different from the number of the magnetic poles and the number of the slots of the electric motor  104  and is thereby set to five (5), which is a prime number, in the present embodiment. The legs  164  axially extend from a plurality of locations (five locations in the present embodiment), respectively, on a radially outer side of the fitting hole  162   a , which is a fitting location of the main body  162 . The legs  164  are arranged one after another at equal intervals in the circumferential direction. Each leg  164  is resiliently deformable because of the resilient material and the axially elongated shape of the leg  164 . When the rotatable shaft  104   a  is rotated, each leg  164  is flexed through the resilient deformation thereof in conformity with the corresponding insertion hole  127 . Thereby, the leg  164  contacts an inner wall of the insertion hole  127  while absorbing circumferential dimensional errors of the insertion hole  127  and the leg  164  generated at the manufacturing. In this way, the joint member  160  transmits the drive force of the rotatable shaft  104   a  to the inner rotor  120  through the legs  164 . 
     Next, with reference to  FIGS. 5 to 7 , a structure of the radial bearing  150  will be described in detail. 
     As shown in  FIG. 5 , the radial bearing  150  is shaped into a cylindrical tubular form. The radial bearing  150  is made of metal and is coated with resin. The rotatable shaft  104   a  is inserted into the inside of the radial bearing  150  such that a cylindrical inner peripheral surface  150   i  of the radial bearing  150  rotatably and slidably supports the rotatable shaft  104   a.    
     An axial portion of the radial bearing  150 , which is located on the pump cover  112  side in the axial direction, will be referred to as a slide portion  1502 . Furthermore, another axial portion of the radial bearing  150 , which is located on the pump casing  116  side in the axial direction, will be referred to as a seal portion  1501 . An inner diameter of an axial portion of the cylindrical inner peripheral surface  150   i , which is located in the slide portion  1502 , is equal to an inner diameter of an axial portion of the cylindrical inner peripheral surface  150   i , which is located in the seal portion  1501 . In contrast, an outer diameter of an axial portion of a cylindrical outer peripheral surface  150   o , which is located in the seal portion  1501 , is larger than an outer diameter of an axial portion of the cylindrical outer peripheral surface  150   o , which is located in the slide portion  1502 . 
     The slide portion  1502  is inserted into the inside of the inner rotor  120 , which is shaped into the cylindrical tubular form, such that the cylindrical outer peripheral surface  150   o  of the slide portion  1502  rotatably and slidably supports the inner rotor  120 . The seal portion  1501  is securely press fitted into a through-hole  116   e  of the pump casing  116 . The radial bearing  150  is non-rotatably fixed to the pump casing  116  through this pressing fitting. The outer peripheral surface of the seal portion  1501  tightly contacts the inner peripheral surface of the through-hole  116   e  to seal between the inner peripheral surface of the through-hole  116   e  and the cylindrical outer peripheral surface  150   o.    
     An axial location of an end surface of the slide portion  1502  coincides with an axial location of an end surface of the pump casing  116 , which contacts the pump cover  112 . Furthermore, an axial location of an end surface of the seal portion  1501  coincides with an axial location of a wall surface of the pump casing  116 , which forms the high pressure passage  106 . In other words, an axial length of the pump casing  116  coincides with an axial length of the radial bearing  150 . 
     As shown in  FIGS. 4, 6 and 7 , a lubrication groove G 1 , which accumulates the fuel, is formed in the cylindrical outer peripheral surface  150   o  of the radial bearing  150 . The lubrication groove G 1  is located in the portion of the cylindrical outer peripheral surface  150   o , which forms the slide portion  1502  and is displaced from the seal portion  1501 . The lubrication groove G 1  is shaped such that the lubrication groove G 1  extends from the end surface of the slide portion  1502  toward the seal portion  1501  in the axial direction (see  FIG. 6 ). The lubrication groove G 1  is formed by cutting a portion of the slide portion  1502  in a cutting process such that the portion of the cylindrical outer peripheral surface  150   o  is cut and is thereby radially inwardly recessed (see  FIG. 7 ). 
     The high pressure fuel of the high pressure passage  106  penetrates into an area (slide surface) between the cylindrical inner peripheral surface  150   i  of the radial bearing  150  and the outer peripheral surface of the rotatable shaft  104   a  and thereafter leaks from this area (slide surface) into the installation space  158  after dropping of the pressure of the high pressure fuel in this area (slide surface). Therefore, the installation space  158  accumulates the fuel (intermediate pressure fuel) that has the pressure, which is lower than the pressure of the high pressure fuel of the high pressure passage  106  and is higher than the pressure of the fuel (suction fuel) of the suction passage  112   a.    
     As shown in  FIGS. 4 and 5 , a first groove  1201  is formed in a surface of the inner rotor  120 , which is axially opposed to the pump casing  116 . The first groove  1201  is shaped into a ring form (annular form) and circumferentially extends about the radial bearing  150 . Furthermore, a second groove  1202  is formed in an opposite surface of the inner rotor  120 , which is axially opposite from the pump casing  116 . The second groove  1202  is shaped into a ring form (annular form) and circumferentially extends about the radial bearing  150 . An outer diameter of the second groove  1202  is the same as an outer diameter of the first groove  1201 . 
     The high pressure fuel of the discharge passage  117  penetrates into an area (slide surface) between the inner rotor  120  and the pump casing  116  and thereafter leaks form this area (slide surface) into the first groove  1201  after dropping of the pressure of the high pressure fuel in this area (slide surface). Therefore, the first groove  1201  accumulates the fuel (intermediate pressure fuel) that has the pressure, which is lower than the pressure of the high pressure fuel of the high pressure passage  106  and is higher than the pressure of the fuel (suction fuel) of the suction passage  112   a . The second groove  1202  is filled with the intermediate pressure fuel of the installation space  158 . Since both of the first groove  1201  and the second groove  1202  are shaped into the ring form and have the same outer diameter, the pressure (the intermediate pressure) of the fuel accumulated in the first groove  1201  and the pressure (the intermediate pressure) of the fuel accumulated in the second groove  1202  are balanced with each other. Therefore, it is possible to limit tilting of the inner rotor  120 , which would be otherwise caused by the intermediate pressure fuel. 
     As discussed above, the fuel accumulated in the first groove  1201  and the fuel accumulate in the second groove  1202  have the identical pressure (the intermediate pressure). Therefore, penetration of the fuel into the area (slide surface) between the cylindrical outer peripheral surface  150   o  of the radial bearing  150  and the inner peripheral surface of the inner rotor  120  is less probable in comparison to the penetration of the high pressure fuel into the cylindrical inner peripheral surface  150   i . However, since the lubrication groove G 1 , which accumulates the fuel, is formed in the cylindrical outer peripheral surface  150   o , the intermediate pressure fuel can relatively easily penetrate into the lubrication groove G 1 . 
     Next, a location of the lubrication groove G 1  will be described in detail with reference to  FIGS. 2 to 4 . 
     With reference to  FIGS. 2 and 3 , a region of the pump housing  110 , in which the corresponding ones of the pump chambers  140  suction the fuel (i.e., a region, in which the corresponding ones of the pump chambers  140  function as the negative pressure portions  140 L), is defined as a suction region  11 . Furthermore, another region of the pump housing  110 , in which the corresponding ones of the pump chambers  140  compress the fuel (i.e., a region, in which the corresponding ones of the pump chambers  140  function as the high pressure portions  140 H), is defined as a compression region  21 . Each of two boundary lines  11   a ,  11   b  between the suction region  11  and the compression region  21  is a straight line that connects between a corresponding halfway point, which is circumferentially located between the opposing discharge groove  114  and the suction groove  113 , and the inner central axis Ci. Specifically, the boundary line  11   a  is the straight line that radially connects between the left side halfway point, which is circumferentially located between the opposing discharge groove  114  and the suction groove  113  at the left side thereof in  FIG. 2 , and the inner central axis Ci. The boundary line  11   b  is the straight line that radially connects between the right side halfway point, which is circumferentially located between the opposing discharge groove  114  and the suction groove  113  at the right side thereof in  FIG. 2 , and the inner central axis Ci. 
     The lubrication groove G 1  is located in a rotational angular range, throughout which the suction region  11  is present, in the rotational direction (see  FIG. 7 ). That is, the lubrication groove G 1  is located in the angular extent of the suction region  11  in the rotational direction. For example, it is desirable that the lubrication groove G 1  is entirely placed in this rotational angular range. More specifically, the lubrication groove G 1  is located on a maximum negative pressure line Csa, which connects between a suction center line Cs of the suction passage  112   a  and the inner central axis Ci. For example, a circumferential center part of the lubrication groove G 1 , which is centered in the circumferential direction (the rotational direction), is located on the maximum negative pressure line Csa (see  FIGS. 2 and 4 ). 
     Now, advantages of the present embodiment will be described. 
     In the case where the temperature of the fuel is low, the viscosity of the fuel is increased. Particularly, in the case where the fuel is the light oil, the viscosity of the fuel becomes very high. Therefore, in such a case, a reaction force, which is applied from the fuel to the inner rotor  120 , is increased. This reaction force is not uniformly applied to the entire inner rotor  120 . Thus, the reaction force is applied to the inner rotor  120  as a force (tilting force) that is exerted to tilt the inner rotor  120  relative to the rotatable shaft  104   a  (the rotational axis of the rotatable shaft  104   a ). As a result, if the joint member  160  is eliminated from the fluid pump  101  unlike the present embodiment to directly engage the rotatable shaft  104   a  to the inner rotor  120 , the tilting force is directly applied to the rotatable shaft  104   a . Thus, the slide resistance between the radial bearing  150  and the rotatable shaft  104   a  is increased to cause an increase in the energy loss or generation of damage at the sliding portion between the radial bearing  150  and the rotatable shaft  104   a.    
     With respect to the above-described disadvantage, according to the present embodiment, the inner rotor  120  is coupled to the rotatable shaft  104   a  through the joint member  160 , so that the above-described tilting force is absorbed through the resilient deformation of the joint member  160 , and thereby the slide resistance between the radial bearing  150  and the rotatable shaft  104   a  is reduced. 
     Furthermore, according to the present embodiment, the rotatable shaft  104   a  is placed to extend over both of the inside of the pump housing  110  and the high pressure passage  106 . Therefore, the high pressure fuel of the high pressure passage  106  can penetrate into the area between the cylindrical inner peripheral surface  150   i  of the radial bearing  150  and the rotatable shaft  104   a  to perform its lubricating function, so that the slide resistance of the rotatable shaft  104   a  can be sufficiently reduced. 
     Furthermore, the lubrication groove G 1  is formed in the cylindrical outer peripheral surface  150   o  of the radial bearing  150 , and the lubrication groove G 1  accumulates the intermediate pressure fuel that is present in the pump housing  110 . Therefore, the intermediate pressure fuel, which is accumulated in the lubrication groove G 1 , can leak from the lubrication groove G 1  in the circumferential direction along the cylindrical outer peripheral surface  150   o  and can enter the area (slide surface) between the cylindrical outer peripheral surface  150   o  and the inner rotor  120  to perform the lubricating function therebetween. Thus, the slide resistance of the inner rotor  120  can be sufficiently reduced. 
     In this type of fluid pump  101 , it is identified which ones of the pump chambers  140  function as the high pressure portions  140 H and which ones of the pump chambers  140  function as the negative pressure portions  140 L. Therefore, the corresponding ones of the pump chambers  140 , which are located in the corresponding predetermined area in the rotational direction, function as the high pressure portions  140 H, and the other corresponding ones of the pump chambers  140 , which are located in the other corresponding predetermined area in the rotational direction, function as the negative pressure portions  140 L. That is, the predetermined area in the rotational direction becomes the compression region  21 , and the other predetermined area in the rotational direction becomes the suction region  11 . For example, in the case of  FIG. 5 , the right half side area (the pump chambers  140  located at the right side), which is located on the right side of the rotatable shaft  104   a , always functions as the negative pressure portions  140 L (the suction region  11 ), and the left half side area (the pump chambers  140  located at the left side), which is located on the left side of the rotatable shaft  104   a , always functions as the high pressure portions  140 H (the compression region  21 ). For example, in the case of  FIG. 4 , the lower half side area (the pump chambers  140  located at the lower side), which is located on the lower side of the rotatable shaft  104   a , always functions as the negative pressure portions  140 L (the suction region  11 ), and the upper half side area (the pump chambers  140  located at the upper side), which is located on the upper side of the rotatable shaft  104   a , always functions as the high pressure portions  140 H (the compression region  21 ). 
     The fuel pressure is applied to the inner rotor  120  from the high pressure portions  140 H (the compression region  21 ) toward the negative pressure portions  140 L (the suction region  11 ) in the radial direction of the rotational axis. Therefore, the fuel pressure is always continuously applied in the same direction, i.e., the direction from the compression region  21  side toward the suction region  11  side. Thus, as shown in  FIG. 7 , an urging force F is always applied from the inner rotor  120  to the radial bearing  150  in the direction that is from the compression region  21  toward the suction region  11 . 
     In the present embodiment, which is made in view of the above point, the lubrication groove G 1  is present in the rotational angular range, throughout which the suction region  11  is present, in the rotational direction. Thereby, it is possible to avoid concentration of the urging force F to edges G 1   e  of the lubrication groove G 1 . Thus, it is possible to limit an increase in the slide resistance in the cylindrical outer peripheral surface  150   o , which would be caused by the formation of the lubrication groove G 1 . Furthermore, since the urging force F is not exerted in the rotational angular range of the cylindrical outer peripheral surface  150   o , in which the suction region  11  is present, a small gap is formed between the inner rotor  120  and the cylindrical outer peripheral surface  150   o . Thus, the fuel in the lubrication groove G 1  can more easily leak from the lubrication groove G 1  in the circumferential direction of the cylindrical outer peripheral surface  150   o , and thereby the reliability of implementing the lubricating function can be improved. 
     Furthermore, in the present embodiment, since the lubrication groove G 1  is located on the maximum negative pressure line Csa, the lubrication groove G 1  is located in the location where the size of the above-described gap is maximized. Thus, the above-described advantage, which is implemented by the absence of the urging force F, can be maximized. 
     Furthermore, in the present embodiment, the lubrication groove G 1  is located in the portion of the cylindrical outer peripheral surface  150   o , which forms the slide portion  1502  and is displaced from the seal portion  1501 . In this way, a seal length of the seal portion  1051  measured in the axial direction can be increased in comparison to the case where the lubrication groove is formed in a portion of the seal portion  1501 . Thus, it is possible to limit leakage of the high pressure fuel of the high pressure passage  106  to the first groove  1201  through the cylindrical outer peripheral surface  150   o  of the radial bearing  150 . 
     OTHER EMBODIMENTS 
     The present disclosure has been described with respect to the one embodiment. However, the present disclosure is not limited to the above embodiment, and the above embodiment may be modified in various ways within a principal of the present disclosure. 
     In the embodiment shown in  FIG. 2 , each of the boundary lines  11   a ,  11   b  between the suction region  11  and the compression region  21  is set to be the straight line that connects between the corresponding halfway point, which is between the opposing discharge groove  114  and the suction groove  113 , and the inner central axis Ci. Alternatively, as shown in  FIG. 4 , each of boundary lines  10   a ,  10   b  between a suction region  10  and a compression region  20 , which respectively correspond to the suction region  11  and the compression region  21  of the above embodiment (see  FIG. 2 ), may be a straight line that extends parallel to the eccentric direction De and passes through the inner central axis Ci. 
     In the embodiment shown in  FIG. 2  and the above modification (the suction region  10  and the compression region  20 ) shown in  FIG. 4 , the lubrication groove G 1  is located on the maximum negative pressure line Csa. Alternatively, the lubrication groove G 1  may be located at a location that is circumferentially displaced from the maximum negative pressure line Gsa as long as the lubrication groove G 1  is located in the suction region  10 ,  11 . 
     However, it is desirable that the lubrication groove G 1  is located in a rotational angular range  12 , throughout which the suction groove  113  is present, in the rotational direction to further improve the above-described advantage, which is implemented by the absence of the urging force F. That is, it is desirable that the lubrication groove G 1  is located in the angular extent of the suction groove  113  in the rotational direction. 
     For example, it is desirable that the lubrication groove G 1  is entirely received in this rotational angular range  12  (the angular extent of the suction groove  113 ). The rotational angular range  12 , throughout which the suction groove  113  is present, is a range that is circumferentially defined between a line  12   a , which connects between one circumferential end of the suction groove  113  and the inner central axis Ci, and a line  12   b , which connects between the other circumferential end of the suction groove  113  (see  FIG. 2 ). 
     Furthermore, it is desirable that the lubrication groove G 1  is located in a rotational angular range  13 , throughout which the suction passage  112   a  is present, in the rotational direction to further improve the above-described advantage, which is implemented by the absence of the urging force F. That is, it is desirable that the lubrication groove G 1  is located in the angular extent of the suction passage  112   a  in the rotational direction. For example, it is desirable that the lubrication groove G 1  is entirely received in this rotational angular range  13  (the angular extent of the suction passage  112   a ). The rotational angular range  13 , throughout which the suction passage  112   a  is present, is a range that is circumferentially defined between a tangent line  13   a , which is tangent to the suction passage  112   a  on one circumferential side of the suction passage  112   a  and extends through the inner central axis Ci, and a tangent line  13   b , which is tangent to the suction passage  112   a  on the other circumferential side of the suction passage  112   a  and extends through the inner central axis Ci (see  FIG. 2 ). 
     In the embodiment shown in  FIG. 7 , the lubrication groove G 1  has a planar cross-sectional shape. Alternative to the lubrication groove G 1  of  FIG. 7 , as shown in  FIG. 8 , a lubrication groove G 2 , which has a triangular cross section, may be formed in the cylindrical outer peripheral surface  150   o  of the radial bearing  150 . Further alternatively, as shown in  FIG. 9 , a lubrication groove G 3 , which has an arcuate cross section, may be formed in the cylindrical outer peripheral surface  150   o  of the radial bearing  150 . Further alternatively, as shown in  FIG. 10 , a lubrication groove G 4 , which has a rectangular cross section, may be formed in the cylindrical outer peripheral surface  150   o  of the radial bearing  150 . 
     In the embodiment shown in  FIG. 6 , an end part G 1   a  of the lubrication groove G 1 , which is axially opposite from the pump cover  112 , is shaped into a right-angled edge. Alternatively, as shown in  FIG. 11 , the cylindrical outer peripheral surface  150   o  of the radial bearing  150  may have a lubrication groove G 5 , which has an end part G 5   a  that is located on the axial side opposite from the pump cover  112  and is shaped into an arcuately curved form. Further alternatively, as shown in  FIG. 12 , the cylindrical outer peripheral surface  150   o  of the radial bearing  150  may have a plurality of lubrication grooves G 6 , which are arranged one after another in the axial direction. 
     In the embodiment shown in  FIG. 6 , the lubrication groove G 1  is formed to extend in parallel with the axial direction, as shown in  FIG. 13 . Alternatively, as shown in  FIGS. 14 and 15 , the lubrication groove G 1  may be formed to extend in a crossing direction that crosses the axial direction. 
     In the embodiment shown in  FIG. 5 , the radial bearing  150  is made of the metal and is coated with the resin. Alternatively, the radial bearing  150  may be made of the metal without the resin coating. Further alternatively, the radial bearing  150  may be made of resin. 
     In the embodiment shown in  FIG. 4 , the external teeth  124   a  and the internal teeth  132   a  are shaped to have the trochoid tooth profile. Alternatively, the external teeth  124   a  and the internal teeth  132   a  may be shaped to have any other suitable type of tooth profile, such as a cycloid tooth profile or a profile of a combination of various curved lines. 
     The subject fluid to be pumped with the fluid pump  101  is not limited to the light oil (diesel fuel) and may be any other liquid fuel, such as gasoline or alcohol. Furthermore, the subject fluid to be pumped with the fluid pump  101  is not limited to the fuel and may be liquid, such as hydraulic oil used in a hydraulic actuator or any of various lubricant oils. The fluid pump  101  is not limited to the fluid pump installed in the vehicle. 
     In the embodiment shown in  FIG. 1 , the present disclosure is implemented in the fluid pump  101  that has the pump main body  103  and the electric motor  104 , which are integrated together. However, the electric motor  104  may not be provided in the fluid pump  101  of the present disclosure, and the electric motor  104  may be formed separately from the rest of the fluid pump  101 . In the embodiment shown in  FIG. 1 , the inner rotor  120  is driven by the electric motor  104 . Alternatively, the inner rotor  120  may be driven to rotate by a portion of a drive force for driving the vehicle, such as a drive force of a crankshaft of an internal combustion engine of the vehicle. 
     In the embodiment shown in  FIG. 1 , the discharge passage  117  is located on the opposite side of the pump housing  110 , which is opposite from the suction passage  112   a  in the axial direction. Alternatively, the discharge passage  117  and the suction passage  112   a  may be placed on the same axial side of the pump housing  110 .