Patent Publication Number: US-8118579-B2

Title: Gear pump

Description:
CROSS REFERENCE TO RELATED APPLICATIONS 
     This application is a Division of application Ser. No. 11/357,523 filed Feb. 21, 2006 now U.S. Pat. No. 7,479,000 which is a division of application Ser. No. 10/452,827 filed Jun. 2, 2003 now Pat. No. 7,014,436 issued Mar. 21, 2006. 
     PRIORITY INFORMATION 
     This application claims priority under 35 U.S.C. §119(e) of Provisional Application 60/385,689, filed Jun. 3, 2002 and Provisional Application 60/464,395 filed Apr. 18, 2003, the entirety of these applications are herein incorporated by reference. 
    
    
     BACKGROUND OF THE INVENTION 
     1. Field of the Invention 
     The present invention relates to pumps, and, in particular, to gear pumps. 
     2. Description of the Related Art 
       FIG. 1  is a schematic illustration of an exemplary prior art gear pump  100 . Such a pump  100  typically includes a casing  111  and a pair of rotors  113 ,  115 , with intermeshing gear teeth  117 . The casing  111  defines an inlet port  107  and an outlet port  108 , which extend in a generally radial direction with respect to the rotors  113 ,  115 . Fluid is carried from the inlet port  108  in spaces (or chambers)  102  that are formed between the gear teeth of the rotors. The fluid in these chambers  102  is displaced as the teeth engage with the teeth of the opposing rotor and the fluid is displaced out the discharge port  108 . 
     Such conventional gear pumps are simple and relatively inexpensive, but suffer from a number of performance limitations. A source of problems with conventional gear pumps is in the area where the teeth  117  mesh and create a seal  104  between the inlet and discharge ports  107 ,  108 . Conventional gear pumps use conventional gear tooth profiles such as would be used in a geared power transmission device. This type of gear configuration is well suited for power transmission, but has significant limitations when used to pump incompressible fluid. 
     A need therefore exists for an improved gear pump which addresses at least some of the problems described above. 
     SUMMARY OF THE INVENTION 
     In one embodiment having certain features and advantages according to the present invention, a gear pump is configured to address the tendency of conventional gear pumps to show significant reductions in performance as the teeth experience wear. In such an embodiment, the gear pump may utilize a modified gear tooth profile and a corresponding inlet and discharge port design to provide a number of performance characteristics including reduced turbulence, reduced vibration, and reduced noise, while providing a pump with the ability to experience significant wear between the gear teeth with minimal effect on volumetric efficiency and pressure capability. 
     Another aspect of the present inventions comprises a pump having a driving rotor and a driven rotor that are positioned in a housing such that, as the driving rotor and the driven rotor rotate, the teeth of the driving rotor and the teeth of the driven rotor mesh to form a positive displacement chamber. The teeth of the driving rotor and the driven rotor are configured such a seal between the inlet side and the discharge side of the pump is formed between only the leading surfaces of the driving rotor and the trailing surfaces of the driven rotor. 
     Another aspect of the present inventions comprises a pump having a driving rotor and a driven rotor that are positioned in a housing such that, as the driving rotor and the driven rotor rotate, the teeth of the driving rotor and the teeth of the driven rotor mesh with sufficient backlash to form a seal between the inlet side and the discharge side of the pump, which is formed only between the leading surfaces the driving rotor and the trailing surfaces of the driven rotor. 
    
    
     
       BRIEF DESCRIPTION OF THE DRAWINGS 
         FIG. 1  is a schematic illustration of a top plan view of a prior art pump. 
         FIG. 2  is a schematic illustration of a top plan view of an exemplary embodiment of a pump having certain features and advantages according to the present invention. 
         FIG. 2   b  is a schematic illustration of a top plan view of another exemplary embodiment of a pump having certain features and advantages according to the present invention. 
         FIG. 3  is a closer view of a portion of the pump of  FIG. 2  with a zero degree dwell angle. 
         FIG. 4  is a closer view of a portion of the pump of  FIG. 2  with greater than zero degree dwell angle. 
         FIG. 5  is a side perspective view of a casing of the pump of  FIG. 2 . 
         FIG. 6  is a modified embodiment of the casing of  FIG. 5  having certain features and advantages according to the present invention. 
         FIG. 6   a  is a cross-sectional view of the casing of  FIG. 6 . 
         FIG. 7  is a modified embodiment of the casing of  FIG. 6  having certain features and advantages according to the present invention. 
         FIG. 7   a  is a cross-sectional view of the casing of  FIG. 7 . 
         FIG. 8  is a schematic illustration of a top plan view of another exemplary embodiment of a pump having certain features and advantages according to the present invention. 
         FIG. 9  is a schematic cross-sectional illustration of the pump shown in  FIG. 8  running in the opposite direction. 
         FIG. 10  is a closer view of a portion of the pump of  FIG. 8  with a zero degree dwell angle. 
         FIG. 11  is a closer view of a portion of the pump of  FIG. 8  with a zero degree dwell angle and running in the direction shown in  FIG. 9 . 
         FIG. 12  is a closer view of a portion of the pump of  FIG. 9  with a greater than zero degree dwell angle. 
         FIG. 13  is a closer view of a portion of the pump of  FIG. 9  with material removed from the smallest diameter of the gear teeth. 
         FIG. 14   a  is a closer view of a portion of a modified embodiment of the pump of  FIG. 8 . 
         FIG. 14   b  is a side perspective view of a rotor of the pump of  FIG. 14   a.    
         FIG. 15  is a closer view of a portion of a modified embodiment of the pump of  FIG. 2 . 
         FIGS. 16   a - c  illustrate various embodiments of rotors having certain features and advantages according to the present invention. 
         FIG. 17  is a schematic top plan view of another exemplary embodiment of a pump having certain features and advantages according to the present invention. 
         FIG. 18  is a schematic top plan view of an exemplary embodiment of a pump with four rotors having certain features and advantages according to the present invention. 
         FIG. 19  is a top plan view of the casing of the pump of  FIG. 18 . 
         FIG. 20  is a top plan view of the pump of  FIG. 18 . 
         FIG. 21  is a modified embodiment of the casing of the pump of  FIG. 18 . 
         FIG. 22  is a schematic top plan view of exemplary embodiment of an internal gear pump having certain features and advantages according to the present invention. 
         FIG. 23  is a side perspective view of an exemplary embodiment of a rotor of the internal gear pump of  FIG. 22 . 
         FIG. 24  is a schematic top plan view of the pump of  FIG. 22  showing additional features of the design. 
         FIG. 25  is a side perspective view of an exemplary embodiment of a casing of the internal gear pump of  FIG. 22 . 
         FIG. 26  is a schematic top plan view of another exemplary embodiment of an internal gear pump having certain features and advantages according to the present invention. 
         FIG. 27  is a schematic top plan view of another exemplary embodiment of an internal gear pump having certain features and advantages according to the present invention. 
         FIG. 28  is a schematic top plan view of modified embodiment of an internal gear pump of  FIG. 27 . 
         FIG. 29  is a schematic top plan view of exemplary embodiment of a top plate that may be used with the embodiments of  FIGS. 27 and 28 . 
         FIG. 30  is a side perspective view of exemplary embodiment of an outer rotor that may be used with the embodiments of  FIGS. 27 and 28 . 
         FIG. 31  is a side perspective view of the rotor of  FIG. 30  attached to a drive shaft. 
         FIG. 32  is a schematic top plan view of another exemplary embodiment of planetary gear pump having certain features and advantages according to the present invention. 
         FIG. 33  is a side perspective view of the gear pump of  FIG. 32 . 
         FIG. 34  is a partial cross-sectional view of the gear pump of  FIG. 32 . 
         FIG. 35  is an exploded side view of another exemplary embodiment of planetary gear pump having certain features and advantages according to the present invention. 
         FIG. 36  is another exploded side view of the pump of  FIG. 35 . 
         FIG. 37  is a top plan view of the pump of  FIG. 35 . 
         FIG. 38  is an exploded side view of another exemplary embodiment of internal gear pump having certain features and advantages according to the present invention. 
         FIG. 39  is another exploded side view of the pump of  FIG. 38 . 
         FIG. 40  is a top plan view of the pump of  FIG. 38 . 
         FIG. 41  is a side perspective view of another exemplary embodiment of an internal gear pump having certain features and advantages according to the present invention. 
         FIG. 42  is another side view of the pump of  FIG. 41 . 
         FIG. 43  is a top plan view of the pump of  FIG. 41  with a top cover removed. 
         FIG. 44  is a partial cross-sectional view of the pump of  FIG. 41 . 
     
    
    
     DETAILED DESCRIPTION OF THE PREFERRED EMBODIMENT 
       FIGS. 2-5  illustrate an exemplary embodiment of an internal gear pump  200  having certain features and advantages according to the present invention. The term “pump” is used broadly, and includes its ordinary meaning, and further includes a device which displaced fluid or which turns as the result of the displacement of fluid, either compressible or incompressible. As such, the term “pump” is intended to include such applications as hydraulic motors or other devices which require expanding chambers of compressing chambers or both. In addition, throughout this description reference is made to certain directions (e.g., forward, backward, up, down, etc.) and relative positions (e.g., top, bottom, lower, upper, side, etc.). However, it should be appreciated that such directions and relative positions are intended merely to help the reader and are not intended to limit the invention. 
     The exemplary pump  200  comprises a casing  199  and a pair of opposing rotors  202 ,  203 , with intermeshing gear teeth  223   a ,  223   b . As seen in  FIGS. 2 and 5 , the casing  199  defines an inlet port  210 , an outlet port  211  and a pair of annular recesses  221   a ,  221   b  with circular bearing surfaces  227   a ,  227   b  or other similar structures for supporting the rotors  202 ,  203  for rotation about a shaft  225   a ,  225   b.    
     With particular reference to  FIG. 2 , the design of the teeth  223   a ,  223   b  has certain similarities to the prior art embodiment described above. However, in the exemplary embodiment, a side  201  of the gear teeth is relieved or removed as indicated by the dashed lines. By removing material from the gear teeth, a trailing face  204  of the driving rotor  202  and/or a leading face  205  of the driven rotor  203  are recessed with respect to their corresponding leading and trailing faces  208 ,  209 . As will be explained in more detail below, the casing  199  may be provided with an inlet axial-port relief  206  and/or a discharge axial-port relief  207  such that a positive seal  196  and/or  198  is formed between the two rotors  202 ,  203  and the casing  199  with seal surfaces between the rotors  202 ,  203  being formed only between the leading faces  208  of the driving rotor  202  and the t railing faces  209  of the driven rotor  203 . 
     The exemplary embodiment has several advantages. For example, an improved operating principle may be established which provides an improved seal between the rotors  202 ,  203  even if manufacturing tolerances are low. In addition, as will be explained in more detail below, any wear that occurs between the seal surfaces  208 ,  209  will not increase the clearance between these faces because a contact seal will exist between these faces  208 ,  209  due to the discharge pressure, which will cause the driven rotor to resist forward rotation. This allows the rotor faces to “wear in” to each other during initial service which will reduce the need for high manufacturing tolerances which will, in turn, reduce the cost of the pump. The ability of the gear teeth  223   a ,  223   b  to maintain a positive seal even with significant wear is believed to enable the pump  200  to operate far longer without maintenance and/or replacement than a conventional gear pump, especially when pumping abrasive fluids. 
     With continued reference to  FIG. 2 , the leading faces  208  of the driving rotor  202  maintain a positive contact pressure against the trailing faces  209  of the driven rotor  203  due to the pressure of the fluid in the discharge port  211 , which press the faces  208 ,  209  together thereby providing an efficient seal. As a result, this embodiment allows the sealing faces  208  of the driving rotor  202  and/or the sealing faces  209  of the driven rotor  203  to experience significant wear without reducing the seal effectiveness between the sealing faces  208 ,  209  of the rotors  202 ,  203 . 
       FIG. 2   b  illustrates the pump  200  of  FIG. 2  with significant wear on the contact faces  208 ,  209  of the rotors  202 ,  203 . As the sealing faces  208 ,  209  of one or both rotors  202 ,  203  wear down from contact with each other or from the presence of abrasives in the fluid being pumped, the driving rotor  202  will advance slightly relative to the driven rotor  203  and/or the driven rotor  203  will rotate backward slightly relative to the driving rotor  202  so that a contact seal  196  and/or  198  is maintained between the teeth  223   a ,  223   b . This relative rotation of one or both rotors  202 ,  203  will allow the pump  200  to seal effectively until there is no longer sufficient material left on the teeth  223   a ,  223   b  to provide the strength to pump at the discharge pressure or until one or more of the sealing faces  208 ,  209  wears enough to reduce the rotor tip diameter so it no longer provides an adequate seal against the casing  199  at the gear tooth tips  220 . 
     The exemplary pump  200  may utilize different configurations of inlet and outlet ports each having particular advantages. In the exemplary embodiment illustrated in  FIGS. 2-5 , the pump  200  utilizes radial ports  210 ,  211 , which define an inlet and outlet flow axis that extend in a generally radial direction with respect to the rotors  202 ,  203 . As will be explained in more detail below,  FIG. 6  illustrates a modified embodiment that includes axial ports  213 ,  216 , which define a flow path that is generally perpendicular to the radial direction and parallel to the axis of rotation of the rotors  202 ,  203 . 
     In the embodiments illustrated in  FIGS. 2B and 5 , the radial ports,  210 ,  211  allow fluid to flow to and from the chambers  212  formed between the meshing rotor teeth  223   a ,  223   b  during the beginning of the volume reduction of these chambers  212  on the discharge side, and during the end of volume increase of these chambers on the intake side. 
     As each chamber nears the lowest volume position  212  (see e.g.,  FIG. 2 ), however, the chamber becomes sealed to the discharge port by the engagement of the subsequent meshing teeth. Therefore, the illustrated embodiment includes an axial port recess  207  (see  FIG. 5 ) for the fluid to displace into if a high pressure spike between the rotors is to be avoided. Similarly, as each chamber moves away from the lowest volume position, the chamber  212  remains sealed to the intake port  210  by the engagement of the proceeding teeth on each of the rotors  202 ,  203  and requires an axial port recess  206  (see  FIG. 5 ) from which to draw in fluid if a low pressure spike between the rotors is to be avoided. 
       FIGS. 6 and 6   a  illustrate an embodiment of the pump  200   b , which includes axial ports  213   b ,  216   b , which define a flow path that is generally perpendicular to the radial direction. As shown, the casing  199   b  includes the axial ports  213   b ,  214   b  radial port casing recesses  215   b ,  216   b  and axial port recesses  206   b ,  207   b  as described above. 
       FIG. 7  illustrates another embodiment of the pump  200   c . In this embodiment, the pump  200   c  includes a modified casing  199   c  with purely axial ports  213   c ,  214   c  with no axial port recesses (as compared to the embodiment illustrated in  FIG. 6   a ). This embodiment may result in higher fluid flow resistance as compared to the embodiment of  FIG. 6   a.    
     In addition to the embodiments described above, various port combinations and sub-combinations are also possible. For example, the pump may include radial ports only or axial ports only or various combinations of these two port types. In most embodiments, it is only required that there be an axial intake port  215  or port recess  206  to avoid a vacuum spike between the rotors just after the chamber  212  is momentarily or briefly formed for part of the rotation, which could cause the driven rotor  203  to advance rotationally and disengage the sealing surfaces  196 ,  198 . This situation tends to happen if the negative pressure of the vacuum spike exceeded the discharge pressure. As such, the preferred embodiment utilizes an axial intake port  213  or port recess  206  at one end face of the rotors  202 ,  203  or more preferably at both ends of the rotors. A discharge axial port  214  or axial port recess  207  would also increase certain performance characteristics of the pump but may not be necessary for operation in all situations. 
     Radial ports as described above with reference to  FIGS. 2-5  may offer convenience benefits for plumbing depending on the application. As mentioned above, a purely axial port casing design  FIG. 7  could have a radial port effect of reduced flow resistance by providing casing recesses in the areas  215 ,  216  ( FIG. 6 ) of the rotor engagement and disengagement. Purely axial ports  213   c ,  214   c  are shown in  FIG. 7 . Purely axial ports may be advantageous for certain pump configurations. 
     With initial reference to  FIGS. 2   b  and  3 , a consideration in the design of the axial port recesses  206 ,  207  or axial port  210 ,  211  is what will be referred to as the dwell angle. The dwell angle is the angular rotation of the rotors  202 ,  203  on one side or the other of the lowest chamber volume position when the chamber  212  is sealed between the contact surfaces  208 ,  209  of the teeth of the two rotors  202 ,  203  and between the end faces  1601 ,  1602  (see  FIG. 16   a ) of the rotor teeth and the casing  119 . The dashed line in  FIG. 3  shows inlet and discharge axial port recesses  206 ,  207  with a dwell angle of 0 degrees. In  FIG. 4 , the dashed line shows inlet and discharge port recesses  206 ,  207  with a dwell angle of approximately 2 degrees. 
     Generally speaking, a dwell angle of 0 degrees or less will result in a smoother running pump, but will exhibit reduced volumetric efficiency as more leakage will occur. A dwell angle of greater than 0 degrees will result in increased noise and vibration due to pressure and vacuum spikes in the chamber  212 , but in certain embodiments this may be preferable to increase volumetric efficiency and pressure capability. In one preferred embodiment, the pump includes a positive dwell angle of several degrees combined with the addition of rounded edges  501  (see  FIG. 5 ) on the axial port recesses  206 ,  207 , or axial ports  210 ,  211 . Such rounded edges  501  will help prevent wear of the port  210 ,  211  or port recess  206 ,  207  edges over time, especially when pumping abrasive fluids or slurries. As shown in  FIG. 5 , in the preferred embodiment, the rounded edges  501  generally follow the contour of the leading edges  208 ,  209 , which form the chamber  212 ; however, in other embodiments of the contour may be modified from this shape. 
     It should also be noted that certain embodiments may use different dwell angles on the inlet and discharge sides of the pump to achieve different operating characteristics. For example, to prevent cavitation at higher operating speeds or lower inlet charge pressures, the inlet dwell angle may be reduced to 0 degrees or less to reduce or eliminate any vacuum spikes in the chamber  212  while increasing the discharge dwell angle to 2 or 3 degrees to assure that a positive seal is maintained at all times. This example of a different dwell angle on the inlet and discharge sides of the pump will operate with slightly higher levels of noise and vibration but this may be an acceptable compromise in applications where cavitation is a concern. Of course, for many applications, some routine experimentation or optimization may be beneficial to determine the ideal dwell angle to achieve the desired performance and to maintain a consistent fluid “creep” and “backflow” at all times during the rotation of the rotors. 
       FIGS. 8 and 9  illustrate another exemplary embodiment of a pump  800  having certain features and advantages according to the present inventions. In this embodiment, similar reference numbers have been provided for parts that are similar to parts described above. As shown in  FIGS. 8 and 9 , the rotors  802 ,  803  are designed with gear teeth  805  that are similar in shape on the leading and trailing edges (e.g., the gear teeth  805  are generally symmetrical). To achieve the effect of removing material from the trailing face  204  of the driving rotor  202  and/or the leading face  205  of the driven rotor  203  as described above, the rotors  802 ,  803  are provided with sufficient “backlash” to allow relatively unrestricted flow of fluid through the space between the unsealed areas between the trailing surface  802  of the teeth  805  of the driving rotor  802  and the leading surface  802  of the teeth  805  of the driven rotor  802 . As shown in  FIG. 9 , such a pump  800  would have the ability to pump equally or nearly equally as well when operated in a reversed direction. 
     In this embodiment it may be advantageous to use a “universal” port recess shape which seals the lowest volume position of the chambers  212  with the desired dwell angle when the pump is pumping forward ( FIG. 8 ) as well as when the pump is pumping in reverse ( FIG. 9 ). A universal reversible port shape with a dwell angle of approximately 1 degree is shown in  FIG. 10  with the pump operating in the forward direction and in  FIG. 11  with the pump operating in the reverse direction. In both directions it can be see that the area  212  is sealed momentarily at the lowest volume position and for 1 degree on either side of this position because the edge  1001 ,  1002  of the axial ports (not shown) or axial port recesses  206 ,  207  is aligned with the edge of the meshing teeth at 1 degree of rotor rotation on either side of the position which forms the chamber  212  in  FIG. 10  and  FIG. 11 . 
     This axial port or axial port recess edge  1001 ,  1002  alignment is advantageous in order to achieve as large an area as possible for the fluid to enter and exit the chamber between the rotors on either side of the lowest volume  212  position.  FIG. 12  shows the increased backlash embodiment with the rotors  802 ,  803  at approximately 3 degrees past the lowest chamber volume position  212 . In this position the trailing edge  1201  of the driven rotor  803  has just entered the axial inlet port recess  206  allowing fluid  1202  to flow into the chamber  1212  through the opening  1203 . 
     To reduce turbulence and fluid flow resistance, it is advantageous for this opening  1203  to become as large as possible as quickly as possible. Another method of accomplishing this is shown in  FIG. 13  where material has been removed from the rotors  802 ,  803  in the space between the teeth  1302 ,  1303 . The effect of this material removal is to increase the size of the opening  1203  as the trailing edge  1301  of the driven rotor  803  enters the intake axial port recess  206  or the leading edge  1304  of the driving rotor  802  leaves the discharge axial port recess  207 . This material removal could be advantageous for many different rotor configurations and gear tooth profiles. 
       FIGS. 14   a  and  14   b  show a preferred rotor embodiment to increase the opening  1202  size. In this embodiment, very little gear tooth strength is lost because only a recess  1401  is removed from the rotors. These recesses  1401  can be any depth and at one end or both ends of one or both rotors. The recesses  1401  depth is shown in  FIG. 14   b  allows significant reduction of fluid turbulence and velocity resulting in reduced pressure and vacuum spikes in the chamber  1202  without significantly reducing the strength of the gear teeth. In one embodiment which is particularly suited for gear pumps that require tight clearances, the recess  1401  has a depth of 0.005 to 0.050 inches. In another embodiment, the recess  1401  has a depth of approximately 0.1 inches for a 1 inch long rotor. 
       FIG. 14   a  shows the alignment of this rotor recess  1401  with the edge of the axial port  206  and how it more than doubles the size of the opening  1503 . For example, the reference number  1503   a  indicates the opening size that would exist without the recess  1401  while the reference number  1503   b  indicates the opening size with the recess  1401 . As such, the recess  1401  together with the port shape illustrated in  FIG. 14   a  produces approximately twice the cross-sectional area that would exist without the recess  1401 . 
       FIG. 15  shows a modified port recess or port shape  1606 ,  1607  which increases the size of the opening  1603  without having to remove any material from the rotors. Specifically, as indicated by the hatched area in  FIG. 15 , the proximity of the recess edges  1608   a ,  1608   b  to the chamber  1202  increases the size of the opening  1603 . 
       FIG. 16   a  through  16   c  show various embodiments of rotors  700   a - c  with different gear tooth profiles that may provide at least some of the advantages described in above. These embodiments are merely exemplary and many other shapes and configurations of the rotor teeth which utilize such recesses are also conceivable. As explained above, in these embodiments, the gear teeth on one or both of the rotors are configured such that each rotor engagement zone has a sufficient space between the trailing face of the drive rotor teeth and the leading face of the driven rotor teeth so that a seal is not established between these faces. This space may be for the entire length of one or both rotors as shown in  FIG. 2 , and  FIG. 13 , or part of the length of one or both rotors as shown in  FIG. 14 ,  FIG. 16   a ,  FIG. 16   b ,  FIG. 16   c.    
     It should be noted that the above description and drawings are of a simplified nature for clarity of explanation and have been used to represent pump configurations with many variations including greater or lesser number of gear teeth and rotors which could be larger or smaller in size. Also, port shapes and sizes are representative and in an actual pump could be smaller or larger or of a different shape as will be apparent to one of skill in the art. 
     A number of examples of pump configurations which would benefit from the port shapes and configurations and/or the gear tooth shapes and configurations as described above, will now be discussed. It should be noted that these examples do not comprise a complete list of possible pump configurations, but are only intended to demonstrate the wide range of potential applications, which may utilize the port shapes and configurations and/or the gear tooth shapes and configurations described above. As such, the gear tooth profiles mentioned above could be used for any of the following examples of pump configurations; however, for each of discussion, the partially relieved gear teeth  202 ,  203  from  FIG. 2  will be used in the following description and drawings. 
       FIG. 17  shows an example of a three gear configuration pump  1700  with the top cover removed. The pump  1700  includes three rotors  1701 ,  1702 ,  1703  with intermeshing teeth and a casing  1704 , which defines a pair of inlet and outlet ports  1705 ,  1706  and recesses  1707 ,  1708 . As mentioned above, the pump  1700  may be formed with various rotor sizes and gear tooth numbers on each rotor. In addition, the number of rotors may also be varied. 
       FIG. 18  shows an example of a four rotor design pump  1800  with a top cover removed. This embodiment includes a casing  1806  in which three outside rotors  1802 ,  1803 ,  1804  that are driven by a central driving rotor  1801  are positioned. In modified embodiments, one or more of the outside rotors may be used to drive the remaining motors. Flow in and out of the pump could be through radial ports  1807 ,  1808 , with axial port recesses  1811 ,  1815 , as shown or any combination of ports or port recesses as described above. 
       FIG. 19  shows the casing from the example pump  1800  of  FIG. 18  with both casing covers and the rotors  1801 ,  1802 ,  1803 ,  1804  removed. The discharge ports  1808  are located in the top cover  1810  and the dashed lines show the location of the inlet ports  1807  in the bottom cover (not shown). 
     With reference back to  FIG. 18 , fluid is drawn into the pump  1800  through axial openings  1807 . The fluid then travels through intake radial conduits  1814  and the axial port intake recesses  1815  to the area  1813  where the rotor teeth are disengaging and drawing fluid into the expanding space between the teeth of the meshing rotors. The fluid then travels around between the teeth of the rotors and the casing  1806  to where these chambers are reduced in volume as the rotor teeth engage in area  1816 . The fluid is then discharged from between the engaging rotor teeth and out through the discharge axial ports  1811  and the discharge radial port conduits  1812  and finally out the discharge ports  1808 . 
     In this example embodiment, the larger inner rotor  1801  allows the use of multiple outer rotors  1802 ,  1803 ,  1804 . In the embodiment of  FIG. 17 , multiple outer rotors  1703  ( FIG. 17 ) can be used with an inner rotor  1701  of the same size. However, the larger inner rotor  1801  of the embodiment of  FIG. 18  may advantageously provide more sealing length between the inner rotor  1801  and the casing  1806  along the interior face  1805  of the casing  1806 . This area will be referred to as the “tooth tip to casing seal zone”. In the illustrated, three rotor configuration there are always at least three teeth providing a seal between the inner rotor  1801  and the casing  1806  along the face of the casing  1805 . This is advantageous for increased pressure capability and increased volumetric efficiency. More outside rotors  1802 ,  1803 ,  1804  can be used as long as the inner driving rotor  1801  is of sufficient size to provide a seal of at least one tooth at all times in the “tooth tip to casing seal zone.” 
     It should be noted that any of the rotors could be the driving rotor, and that even more than one of the rotors could be a driving rotor at the same time. In the preferred embodiment, the inside rotor  1801  would be the only driving rotor for simplicity and minimized cost. 
     Many other combinations of the casing and port designs are also possible with the four rotor design described above.  FIG. 20  illustrates a modified pump  2100  embodiment wherein the fluid enters and discharges from the pump  2100  from axial ports without the radial conduits  1812 ,  1814  of the embodiment shown in  FIG. 18 .  FIG. 20  shows an example of this port configuration with the top cover removed so as to expose the inlet port recesses  207 , discharge port recesses  206 , and discharge axial ports  2114 . Such a pump  2100  may have the advantage of reduced flow resistance as it does not require the fluid to change directions as many times as the previous embodiment and therefore may require less input power to do the same amount of hydraulic work. 
     In the example in  FIG. 18 , the number of teeth on the inside rotor  1801  is not divisible by the number of outside rotors  1802 ,  1803 ,  1804  so the rotational engagement of each of the outside rotors  1802 ,  1803 ,  1804  with the driving rotor  1801  will be different from each other at all times. This has the advantage of further reducing noise and vibration by staggering any output pulsation that may be inherent in a particular configuration. 
       FIG. 21  shows how a staggered effect can be accomplished if the number of teeth on the driving rotor  2001  can be divided by the number of outside driven rotors  2002 ,  2003 ,  2004 . In this embodiment, the axis of rotation of the outside driven rotors  2002 ,  2003 ,  2004  are positioned at various angles  2005 ,  2006 ,  2007  to each other to stagger the engagement of each outer rotor  2002 ,  2003 ,  2004  with the teeth of the inner driving rotor  2001 . In this manner, a similar effect to the configuration in  FIG. 18  can be accomplished. 
     It should be noted that it may be beneficial to have a non-staggered effect in some configurations. An example embodiment of such a pump is illustrated in  FIG. 32  and  FIG. 33  and will be described in more detail below. A non staggered effect may have the advantage of causing any pressure variations or pressure spikes to act in all directions equally at the same time providing a more balanced force on all pump components. 
       FIG. 22  shows an exemplary embodiment of an internal gear pump  2200 , which includes an internal gear  2201 , an outer gear  2002 , an a inner casing  2203  and an outer casing  2204 . In this embodiment, the internal gear  2201  may be provided with less than half the teeth of the outer gear  2202 .  FIG. 23  shows the outer rotor  2202  of the pump in  FIG. 22  with an example of radial “rotor ports” which, as is known in the art, allow the fluid to flow radially through the rotor  2202 .  FIG. 24  is a cross section of the assembled pump of  FIG. 22  showing the alignment of the outer rotor ports  2301  with radial perimeter port recesses  2401 ,  2402  and the radial perimeter ports  2403 ,  2404 , which are provided in the outer casing  2204 . The radial perimeter port recesses  2401 ,  2402  have a dwell angle of approximately 1 degree. 
       FIG. 25  shows the casing for the pump in  2200  described above with axial port recesses  2501 ,  2502 , axial ports  2503 ,  2504 , radial perimeter port recesses  2401 ,  2402  and the radial perimeter ports  2403 ,  2404 . Both types of ports and port recesses or a combination of these port and port recesses may be used together depending on the requirements of the application. 
       FIG. 26  shows an exemplary embodiment of an internal pump  2600  that is similar to the previous embodiment. However, in this embodiment, the pump  2600  includes an inner rotor  2601  with more than half as many teeth as the outer rotor  2602 . For simplicity, no ports or port recesses are shown in  FIG. 26 . 
       FIG. 27  illustrates another exemplary embodiment of an internal gear pump  2700 . In this embodiment, the inner driven gear  2701  has half as many teeth as the outer drive rotor  2702 . With this 2:1 tooth ratio, a unique seal surface interface shape is possible. The outer rotor seal face  2703  is a flat surface which is offset from a radial line from the rotational center of the outer rotor  2702  by the radius dimension of the arc seal surface  2704  of the inner rotor  2701 . (see  FIG. 43 , dimensions labeled R and r) 
     As mentioned above, there are many different conventional and unconventional gear tooth shapes that could be used with the embodiments described above. Such configurations include the gear tooth shapes in  FIG. 27 , helical gear shapes and bevel gears etc. When using such conventional and unconventional gear shapes, due consideration should be given to the principles of the present invention as described above. For example, the chamber, which is established between the teeth as they mesh, is preferably defined by the leading faces only of the driving rotor and the trailing faces only of the driven rotors. In the case of a multi-rotor design such as the exemplary planetary gear pump  3200 ,  3300  shown in  FIG. 32  and  FIG. 33  (described in more detail below), driven planet gears  3205 ,  3311  also act as driving gears against a ring gear  3206 ,  3306 . In such an embodiment, both the leading and trailing faces are used as sealing faces at the same time but on different meshing gears. 
     It is understood that these drawings are simplified and do not contain detailed information about how the rotors are supported by shafts or bearings or fluid film bearing effects with the casing or engaging rotors. However, in light of the teachings of the present application, such features can be readily determined by one of skill in the art given through routine experimentation or modeling. For example, the gap clearance between the two rotors, and between the rotors and the casing is also not specified but could be anywhere from a contact fit to lesser or greater than 0.005″. It is believed by the inventor that a gap clearance of 0.0005″ to 0.005″ is the range that will be useful for a wide range of applications. A gap clearance of approximately 0.003″ has been tested with SAE  30  weight oil with very good pressure capability and very good volumetric efficiency. 
     Several things must be considered when determining which rotor is to drive and which rotor is to be driven in an internal rotor configuration. Specifically, the displacement of the pump will be increased if the outer rotor is driven. Another consideration is that the drive must be in the opposite direction if the outer rotor is used to drive the pump rather than the inside rotor unless the rotor teeth are designed to be reversible. 
     An aspect of the present inventions is the prevention or reduction of wear in abrasive or high pressure or other applications by the “contact force reduction” of the sealing surfaces if the outer rotor drives the inner rotor. This effect is most easily illustrated in the example configuration in  FIG. 27 . To achieve this “contact force reduction” effect, the outer drive rotor  2702  is driven clockwise in this embodiment which in turn causes the inner driven rotor  2701  to turn clockwise as well by the contact points  2705 . Any hydraulic pressure that results in the areas  2706  and  2707  will act on the inner rotor in the clockwise direction against the trailing face  2708  of the inner rotor  2701  and in the counterclockwise direction against the leading face  2709 . As a result of the greater area of the leading surface  2709  being exposed to the discharge pressure as compared to the trailing surface  2708 , the total rotational force which will result from the hydraulic discharge pressure will be in the counterclockwise direction on the inner rotor  2701  but only by the difference between the two surfaces  2709  and  2708 . This difference is very slight and therefore, the contact pressure which results from the rotational force of the inner rotor  2701  seal surface  2704  against the outer rotor  2702  seal surfaces  2703  is much less than if the inner rotor is used to drive the outer rotor. 
     The contact force that results from driving the outer rotor  2702  will ideally be large enough to establish a satisfactory seal, but small enough to establish a fluid film between the seal surfaces. This contact force is adjustable by increasing or decreasing the diameter of the inner rotor largest diameter surface  2710  as well as the interior casing seal surface  2711 . This changes the difference between the leading surface  2709  and the trailing surface  2708  which are exposed to the discharge pressure. 
       FIG. 28  is a cross sectional view of an example of a unique port configuration which could be used on any of the internal gear pumps described herein. The advantage of this port configuration includes movement of intake fluid through an axial port  2801  and the discharge fluid through a discharge axial port  2802  ( FIG. 29 ). This port arrangement allows the ports  2801   2802  to be aligned at 180 degrees to each other in the inner casing seal member  2803 . This has advantages for access restricted and size restricted applications such as down-hole pumps for water or oil. Another advantage of this configuration is the ability to stack the pump rotors in series stages to increase pressure capability by stacking the stages at 180 degrees to each other. The pump stages could also be stacked in parallel to increase flow volume by stacking the stages in the same position in line with each other. A combination of parallel and series stages could be implemented to achieve both increased pressure and increased flow. 
     The example configuration in  FIG. 28  is a single stage which draws fluid in through the axial intake port  2801  and then through the radial inlet conduit  2808  to the rotor disengagement area  2804 . The expanding chamber  2805  is sealed from the rotor disengagement area  2804  so it is necessary to provide an alternate path for the fluid to flow into this area. In the example embodiment of  FIG. 28 , radial rotor ports  2806  allow fluid to flow from the perimeter port recesses  2807  which are supplied by fluid from the radial intake conduit  2803  through the radial rotor ports  2806 . The fluid goes through the reverse cycle on the discharge side of the pump where it is discharged out the port  2802  ( FIG. 29 ). Axial port recesses could also be used in this configuration to further reduce fluid flow resistance but are not shown in  FIG. 28 . 
     An outer rotor with radial rotor ports with a simplified manufacturing design is shown in  FIG. 30 . This outer rotor would have to be driven by the inner rotor. A simplified manufacturing design of an outer rotor which can be mounted to a drive shaft is shown in  FIG. 31 . This rotor design has manufacturing advantages that will not be capable of as high pressure or speeds as some of the other configurations described in this patent description. 
       FIG. 32  shows an exemplary planetary gear pump having certain features and advantages according to the present invention. In this example embodiment, the inner rotor  3201  drives the planet gears  3205  which, in turn, drive the ring gear  3206 . The fluid is drawn into the pump through the intake ports  3207 ,  3208  in and then discharged from the pump through the discharge ports  3209 ,  3211  in the upper casing (not shown) represented by the dashed lines. As mentioned above, there are many possible variations of this and other pump embodiments that can be achieved using the teachings of this patent application. For example, different sizes of rotors, different numbers of rotors, different gear face shapes, different port and casing configurations may be integrated into the configurations described herein. It should be appreciated that the example embodiment in  FIG. 32  does not show any axial port recesses for simplicity of the drawing, but the round axial ports approximate the ideal shape of the axial ports and should therefore be acceptable for some applications. The inner driving gear  3201  and outer ring gear  3206  are single direction configurations as in  FIG. 2  while the planet gears are of a reversible design with increased backlash as in  FIG. 8 . Only the planet gears  3205  need to be of a reversible shape in this embodiment because the opposite side of the gear teeth are in contact with the inner rotor  3201  as they are with the outer rotor  3206 . 
       FIG. 33  shows a variation of this example embodiment which uses a stationary ring gear  3306  and a rotating inner casing/planet gear carrier  3310 . Advantages of this configuration may include a reduced outer diameter as the ring gear  3306  could serve as the outer casing. Also, by allowing the inner casing/planet gear carrier  3310  to rotate freely, the radial load on the planet gears  3311  may reduce the side load on the bearings and shafts of the planet gears and allow the use of abrasive resistance sleeve bearings which would not need to be sealed from the fluids and which would have reduced wear due to the reduced load. The inner gear  3301  is used to drive the pump in  FIG. 33 . 
     In  FIG. 34  the inlet ports which are located in the spinning inner casing/planet carrier  3310  could use inertia charge conduits  3401  on the inlet ports  3402  to increase the inlet charge pressure to avoid cavitation at higher speeds or with higher viscosity fluids. 
     With respect to the embodiment described above, planetary gear tooth profiles can be a challenge to designers because the ideal planet tooth shape will be different for the ring gear than it will be for the sun gear. The relationship of the planet gear to the ring gear is of an internal gear set. The relationship of the planet gear to the sun gear is of an external gear set. 
     In one embodiment, for a single direction planetary gear pump such as for a down hole pump, a planet gear tooth shape on the leading edge which is ideally shaped to engage with the ring gear can be used with a gear tooth shape on the trailing edge of the planet gears which is ideally shaped to engage with the sun gear. When combined with the sufficient backlash designs described above, a pump design can be simplified and the manufacturing cost reduced. Unconventional gear tooth shapes can also be used in this asymmetric planet gear tooth profile configuration, but with the configuration, conventional gear tooth profiles and manufacturing processes can be utilized to create pump rotors. This configuration will operate in reverse but may not provide as an ideal seal as when operated in the forward direction. 
       FIG. 35  and  FIG. 36  show exploded views and  FIG. 37  shows a front cross section view of a three inner rotor  3501  pump using the unconventional gear tooth shape as shown in  FIG. 16   c . In this configuration, the outer rotor  3502  is the drive rotor. The shafts  3503  of the inner rotors  3501  are held between the cover  3504  and the cover plate  3506 . The fluid enters and exits the pump through the axial inlet ports  3507  which provide fluid to the radial casing inlet port recesses  3509 . The radial casing inlet port recesses  3509  supply fluid to the outer rotor radial rotor ports  3510  and to the axial port recesses  3601  in the casing cover  5304  ( FIG. 36 ). The fluid is discharged through the axial discharge port recesses  3602 , the outer rotor radial rotor ports  3510 , and the radial casing discharge port recesses  3511 , and finally out through the axial discharge ports  3508 . 
     In the pump configuration of  FIGS. 35-37 , the inner driven rotors  3501  rotate in response to a driving force applied to a helical trailing surface  3513  of one of a plurality of teeth  3515  of each of the inner driven rotors  3501  by matching opposing helical leading surfaces  3517  of a plurality of teeth of the outer driving rotor  3502 . In this case, as is best shown in  FIG. 37 , the teeth  3519  of the outer driving rotor  3502  are interfaced with the teeth  3515  of the inner driven rotors  3501  with sufficient backlash to form first and second seals  3523  and  3524  between the axial inlet port  3507  (of  FIG. 35 ) at the inlet side of the pump and the axial discharge port  3508  (also of  FIG. 35 ) at the outlet side of the pump. A first seal  3523  is formed only between helical leading surfaces  3517  of first teeth  3519  of the outside driving rotor  3502  and matching opposing helical trailing surfaces  3513  of respective teeth  3515  of the inside driven rotors  3501 . The second seal  3524  is formed only between helical leading surfaces  3518  of different teeth  3520  of the outside driving rotor  3502  and matching opposing helical trailing surfaces  3514  of different teeth  3516  of the inside driven rotors  3501 . Thus, the first and second seals  3523  and  3524  shown in  FIG. 37  prevent communication between the inlet port recess  3509  and the discharge port recess  3511 . The backlash is located at a positive displacement chamber  3521  between the trailing surfaces  3525  of the teeth  3519  of the driving rotor  3502  and the leading surfaces  3527  of the teeth  3515  of the driven rotors  3501  so as to prevent sealing therebetween. 
       FIG. 38  through  FIG. 40  show an exemplary embodiment of an internal gear pump  3800  having certain features and advantages according to the present invention. This pump  3800  has a gear tooth configuration similar to that of  FIG. 27 . This example embodiment uses the inner gear  3801  as the drive gear and the outer gear  3802  s the driven gear. It should be noted that significant material can be worn off the seal face  4001  of the inner rotor  3801  ( FIG. 40 ) and the seal face  4002  of the outer rotor  3802  ( FIG. 40 ). Fluid is drawn into this embodiment through the intake axial port  4002  (shown in dashed lines in  FIG. 40 ) in the casing cover  3901  (not shown in  FIG. 40 ) and the axial inlet port recess  4004 . Fluid is discharged from the pump through the axial inlet port  4005  and finally out through the axial discharge port  4006 . The inner rotor  3801  is supported and driven by the inner rotor shaft  3803 . The outer rotor  3802  in this example embodiment is supported by a fluid film bearing effect between the outer rotor outer surface  3804  and the casing inner surface  3805 . 
       FIG. 41  through  FIG. 44  show a preferred embodiment of a pump  4100  having certain features and advantages according to the present invention. This embodiment has advantageously reduced manufacturing and design costs, while still producing excellent pressure capability and high volume output. In addition, both rotors  4301 ,  4302  can experience significant wear and still maintain a seal between the two rotor seal surfaces  4303 ,  4304 . The inner rotor  4301  is driven by the inner rotor drive shaft  4101  which is rotationally supported by a bearing in the casing cover  4201  and the casing  4102 . Torque is transferred from the shaft  4101  to the inner rotor  4301  by the drive shaft keyways  4105  and the drive dowels  4103 . 
     Fluid is drawn into the pump through the radial port  4402  into the radial casing port recess  4403 . The fluid is then drawn into the rotor disengagement area  4404  through the outer rotor radial rotor ports  4405 . The fluid then travels in the chamber  4406  between the inner rotor teeth  4408  and the inner casing seal member  4407  and inner surface  4413 . Fluid also travels in the chamber  4410  between the outer rotor teeth  4409  and the outer casing inner surface  4411  and the inner casing seal member outer surface  4412 . When the fluid reaches the rotor engagement area  4414 , it is displaced through the outer rotor radial ports  4405  and then through the casing radial discharge recess  4415  and finally out through the casing radial discharge port  4416 . 
     As the inner rotor seal surface  4303  and/or the outer rotor seal surface  4304  wears, it will advance rotationally relative to the outer rotor  4302 . 
     Although this invention has been disclosed in the context of certain exemplary and preferred embodiments, it will be understood by those skilled in the art that the present invention extends beyond the specifically disclosed embodiments to other alternative embodiments and/or uses of the invention and obvious modifications and equivalents thereof. In addition, while a number of variations of the invention have been shown and described in detail, other modifications, which are within the scope of this invention, will be readily apparent to those of skill in the art based upon this disclosure. It is also contemplated that various combination or subcombinations of the specific features and aspects of the embodiments may be made and still fall within the scope of the invention. Accordingly, it should be understood that various features and aspects of the disclosed embodiments can be combined with or substituted for one another in order to form varying modes of the disclosed invention. Thus, it is intended that the scope of the present invention herein disclosed should not be limited by the particular disclosed embodiments described above, but should be determined only by a fair reading of the claims that follow.