Patent Publication Number: US-6210139-B1

Title: High efficiency gear pump for pumping highly viscous fluids

Description:
CROSS REFERENCE STATEMENT 
     This application claims the benefit of Provisional Application No. 60/102,730, filed Oct. 1, 1998. 
    
    
     FIELD OF THE INVENTION 
     This invention relates to apparatus for conveying highly viscous fluids and, more particularly, to gear pumps. 
     BACKGROUND OF THE INVENTION 
     Gear pumps are used for conveyance of highly viscous fluid, such as polymer melts. For example, gear pumps are typically used for conveying a viscous polymer melt from a vessel, such as a devolatilizer, to another unit operation, such as a pelletizer. In most cases, the highly viscous polymer melt enters the pump inlet under the influence of gravity with essentially no positive pressure. Known gear pumps are susceptible to a number of difficulties in their operation. In particular, for any given pump geometry, known gear pumps are extremely limited with respect to the range of viscosity of fluids that they can handle. Generally, as fluid viscosity increases, the throughput rate of the gear pump decreases, often resulting in a production bottleneck. Also, in general, as gear pump speed (RPM) increases, pump throughput initially increases, but eventually reaches a plateau level, wherein further increases in pump speed do not result in any significant increase in throughput and can lead to a production bottleneck. Heretofore it has generally not been possible to effectively overcome a production bottleneck of this type once the plateau level of the pump speed verses pump throughput has been reached without replacing the existing pump with a larger pump. However, the devolatilizer is typically specially configured to be coupled to a gear pump of a particular size, and it is not generally possible to switch to a larger capacity gear pump of conventional design without also replacing or significantly modifying the devolatilizer. Accordingly, it would be highly desirable to provide a gear pump which operates more efficiently to eliminate such production bottlenecks without requiring replacement or significant modification of the devolatilizer. 
     Various attempts have been made to design gear pumps which are capable of operating efficiently over a wider range of fluid viscosity and over a wider range of pump speeds. These efforts have focused primarily on pump geometry, particularly at the inlet side of the pump. However, the known pump designs have not been entirely satisfactory and further improvements are desirable. 
     SUMMARY OF THE INVENTION 
     The invention provides a gear pump having an improved geometry which attenuates the limitations relating to the viscosity of the fluid being pumped and the pump speed. More specifically, the gear chamber has been designed to provide compression zones which enable more fluid to be compressed over a longer path length into the teeth of the pump gears, and, therefore, provide higher production rates and higher fill efficiency. The improved geometry allows the gear pumps of this invention to operate more efficiently over a relatively broader range of pump speed and with a relatively broader range of fluid viscosity. 
     The gear pumps of this invention include a compression zone defined between each of a pair of pump gears and internal walls of a gear chamber, in which the compression zones have a non-uniform thickness, that is, the spacing between the teeth of the pump gears and the internal walls of the gear chamber in the vicinity of the compression zones varies along the length of the gears. 
    
    
     BRIEF DESCRIPTION OF THE DRAWINGS 
     FIG. 1 is an elevational, cross-sectional, schematic representation of a prior art gear pump, the cross section being perpendicular to the rotational axes of the pump gears; 
     FIG. 2 is a cross-sectional, schematic representation of the pump shown in FIG. 1, the view being along line II—II of FIG. 1; 
     FIG. 3 is a cross-sectional, schematic representation of a gear pump according to the invention, the cross section being perpendicular to the axes of the pump gears; and 
     FIG. 4 is a cross-sectional, schematic representation of the gear pump shown in FIG. 3, with the view being along lines IV—IV of FIG. 3; 
     FIG. 5 is a top plan view of the gear pump shown in FIG. 3 with the pump gears and inlet side of the pump removed; 
     FIG. 6 is an elevational, cross-section of the gear pump shown in FIGS. 3-5 with the pump gears removed, as seen along view lines VI—VI of FIG. 5; 
     FIG. 7 is an elevational, cross-section of the pump shown in FIGS. 3-6 with the pump gears in place, as seen along view lines VII—VII of FIG. 4; 
     FIG. 8 is a top plan view of the pump shown in FIGS. 3-7 with herringbone pump gears in place and with the inlet side of the pump removed; 
     FIG. 9 is a top plan view of an alternative embodiment of the invention configured for use with helical gears, with the inlet side of the pump and the gears removed; 
     FIG. 10 is an elevational, cross-sectional view of the pump shown in FIG. 9 with the gears and inlet side of the pump in place as seen along view lines X—X of FIG. 9; 
     FIG. 11 is a top plan view of the pump shown in FIGS. 9 and 10 with the gears in place and with the inlet side of the pump removed; 
     FIG. 12 is a top plan view of a second alternative embodiment of the invention which utilizes spur gears, with the inlet side of the pump removed and with the spur gears in place; and 
     FIG. 13 is a top plan view of the pump shown in FIG. 12 with the inlet side of the pump and the spur gears removed. 
    
    
     DESCRIPTION OF THE PRIOR ART 
     A typical gear pump in accordance with the prior art is schematically illustrated in FIGS. 1 and 2. The prior art gear pump  10  includes a housing  12  defining internal walls  14 . Gear pump  10  includes an inlet passage  16 , an outlet passage  18 , and a gear chamber  20  disposed between the inlet passage and the outlet passage. Pump gears  22 ,  23  are rotatably supported within gear chamber  20 . The directions of rotation of pump gears  22 ,  23  are indicated by arrows  24 ,  25 . Pump gears  22  and  23  have intermeshing teeth, such as herringbone style teeth. Compression zones  26 ,  27  are defined between pump gears  22 ,  23  and internal wall  14  of gear chamber  20 . Compression zones  26  and  27  have a maximum thickness adjacent inlet passage  16 . The thickness of compression zones  26 ,  27  decrease in the direction of outlet passage  18 , and reach a minimum thickness at about a location on a plane defined by the parallel axes of pump gears  22 ,  23 . The thickness of a compression zone refers to the distance from the outer surfaces of the teeth of the pump gears to the nearest surface of the internal walls of the gear chamber. 
     As can be seen by reference to FIG. 2, the thickness of compression zones  26 ,  27  does not vary along a direction parallel with the rotational axes of pump gears  22 ,  23 . 
     DESCRIPTION OF THE PREFERRED EMBODIMENTS 
     A gear pump having a design in accordance with the principles of this invention is shown in FIGS. 3 through 4. Gear pump  110  includes a housing  112 , having internal walls  114  defining an inlet passage  116 , an outlet passage  118 , and a gear chamber  120  disposed between inlet passage  116  and outlet passage  118 . Pump gears  122 ,  123  are rotatably supported within gear chamber  120 . Pump gears  122 ,  123  include intermeshing teeth, which, in the case of the embodiment shown in FIGS. 3-8, are herringbone style teeth. The direction of rotation of pump gears  122 ,  123  are indicated by arrows  124 ,  125 . Gear chamber  120  is generally divided into two compression zones  126 ,  127  and two seal zones  128 ,  129 . Compression zones  126 ,  127  are defined as those portions of the internal volume of gear chamber  120  which are disposed between the teeth of gears  122 ,  123  and the internal walls of gear chamber  120 , and which are located above seal zones  128 ,  129 . Seal zones  128 ,  129  refers to that portion of the internal volume of gear chamber  120  in which the clearance between the teeth of the gears  122 ,  123  is so small as to effectively prevent any significant fluid movement through the space between the teeth of gears  122 ,  123  and the internal walls of gear chamber  120 , thereby providing an effective seal against the flow of fluid past the outer surfaces of the teeth of gears  122 ,  123 . Each of the compression zones  126 ,  127  has a non-uniform thickness. The thickness of each of the compression zones  126 ,  127 , which is the distance from the outer surfaces of the teeth of gears  122 ,  123  to the surface of the internal walls of the gear chamber, is greatest at a location adjacent inlet passage  116 . The thickness of each of the compression zones  126 ,  127  continuously decreases from inlet passage  116  toward outlet passage  118 . Preferably, the thickness of the compression zones  126 ,  127  smoothly decrease from inlet passage  116  toward outlet passage  118 . The expression “smoothly decrease” as used herein means that internal walls  114  defining compression zones  126 ,  127  do not have any abrupt or sharp edges defined by intersecting planes, but instead are continuously curved. 
     As can be seen by reference to FIG. 4, compression zones  126 ,  127  have a non-uniform thickness along the longitudinal direction of gears  122 ,  123 , which is greatest at a location centered between axially opposite ends of pump gears  122 ,  123  and which is smallest at locations adjacent each of the ends of pump gears  122 ,  123 . Preferably, the thickness of the compression zones continuously decreases from the location centered between the opposite ends of pump gears  122 ,  123  toward each of the ends of pump gears  122 ,  123 . Further, it is desirable that the thickness of the compression zones  126 ,  127  continuously and smoothly decrease from the location centered between the opposite ends of gear pumps  122 ,  123  toward each of the ends of gear pumps  122 ,  123 . 
     Compression zones  126 ,  127  and seal zones  128 ,  129  are preferably further defined by the following criteria: the area of the compression zone is maximized subject to the constraint that the areas of the seal zones  126 ,  127  be sufficient to maintain a reliable seal between the teeth of gears  122 ,  123  and the internal walls of gear chamber  120 . Maximizing the surface area of the compression zone maximizes filling of the volume bounded by adjacent teeth and the internal walls of the gear chamber  120  at the areas of seal zones  126 ,  127 , which, in turn, results in greatly improved pump efficiency. This means that higher flow rates can be achieved for a given size gear pump. Higher pump efficiency for a given size pump will result in substantial capital savings, as it will not be necessary to replace or substantially modify as associated equipment, such as a devolatilizer, in order to accommodate a larger size pump. The option of replacing a conventional gear pump with an improved gear pump which is, in accordance with the principles of this invention, capable of achieving greater fill efficiency and higher throughput rates for a given size pump, will also result in reduced labor costs relating to modification or replacement of equipment associated with a particular size pump, and a reduced period during which a production unit is taken out of service. 
     Illustrated gear pump  110  can be described as having a double compression zone wherein the fluid being pumped is compressed in both the direction of rotation of pump gears  122 ,  123  and in the direction parallel to the rotational axes of pump gears  122 ,  123 . The geometry of the double compression zones  126 ,  127  provide a mechanism whereby the fluid is induced by rotation of pump gears  122 ,  123  through a progressively narrowing gap which generates increasing pressure in the direction of rotation of gears  122 ,  123  ending in a final smooth pinch-off at the start of seal zones  128 ,  129 . A key difference between the invention and the prior art is that the continuous and smooth variation of the boundary of the compression zone in both the axial and radial direction provides more time to fill the space between teeth and, thus, enables more fluid to be compressed over a longer path length into the teeth of pump gears  122 ,  123 , thus providing for higher product rates and higher fill efficiency. 
     As previously mentioned, an important constraint on the area of compression zones  126 ,  127  is that a reliable seal must be maintained between the teeth of gears  122 ,  123  and internal walls of gear chamber  120 . This generally means that seal zones  128 ,  129  must be sized, shaped and contoured so that the entire length of at least one tooth of each of gears  122 ,  123  is sufficiently closely spaced to its associated seal zone to maintain an effective seal between the compression zone and the pump discharge. However, as illustrated in FIG. 7, it is generally preferred to size, shape and contour seal zones  128 ,  129  so that at least two adjacent teeth on each of gears  122 ,  123  are sufficiently closely spaced to their respective seal zones to maintain an effective seal (that is, one in which very little, if any, fluid can flow between the teeth and the walls of the gear chamber in the area of the seal zones) along the entire length of two adjacent teeth. This will prevent minor damage, such as from excessive wear or abrasion, to any single tooth from significantly affecting overall pump performance, thus ensuring longer, reliable service life without significantly reducing pump efficiency and throughput. 
     Because seal zones  128 ,  129  are shaped to follow the length of at least one tooth and preferably two adjacent teeth of gears  122 ,  123 , the shape of seal zones  128 ,  129  is determined by the tooth pattern of gears  122 ,  123 . In the case of herringbone gears, the teeth wind around the gears  122 ,  123  in a helical path in a first direction (for example, in a clockwise direction) from a first end of the gears to the lengthwise mid-section of the gear and then take a sharp turn and wind around the gear in a helical path in a direction opposite to the first direction (for example, in a counter-clockwise direction) from the lengthwise mid-section of the gear to a second end of the gear opposite the first end, as shown in FIG.  8 . Thus, in the case of pump  110 , which has a double tunnel discharge with two discharge ports  130 ,  131  (FIGS. 5 and 6) and which has herringbone gears  122 ,  123 , maximization of the area of the compression zone while maintaining an effective seal between at least two teeth and the portion of the internal walls of gear chamber  120  defining seal zones  128 ,  129  results in a V-shaped seal zone as indicated in FIG. 5 by seal zone boundaries  132 ,  133 . It should be noted that the seal zone boundaries  132 ,  133  are shown for purposes of illustration only, as there is a smooth transition from the compression zone to the seal zone which would not be readily visible, if at all. 
     A double tunnel discharge (as shown in FIGS. 5 and 6) is preferred because it provides a larger area for the compression zone on the suction side of pump  110  without violating the requirement that at least one tooth, and more preferably two teeth, of each of gears  122 ,  123  will seal against the portion of the gear chamber walls defining the seal zone. The double tunnel discharge also allows a larger angle of rotation of gears  122 ,  123  before the teeth break the seal. 
     In FIGS. 9 through 11, an alternative embodiment of the invention utilizing helical gears is shown. As with gear pump  110 , gear pump  210  includes a housing  212  defining internal walls  214 , inlet passage  216 , outlet passage  218  and gear chamber  220  disposed between the inlet passage and the outlet pump. Gears  222 ,  223  are rotatably supported within gear chamber  220 . Gears  222 ,  223  have intermeshing teeth which are helically wound around the entire length of gears  222 ,  223 . As with pump  110 , compression zones  226 ,  227  and seal zones  228 ,  229  are defined by the principle of providing a double compression zone wherein the fluid is compressed in both the direction of rotation of gears  222 ,  223  and in the direction parallel to the rotational axes of pump gears  222 ,  223 , and compression zones  226 ,  227  provide a mechanism whereby the fluid is induced by rotation of gears  222 ,  223  through a progressively narrowing gap in the direction of rotation to generate increasing pressure until the fluid reaches smooth pinch-off at the start of seal zones  228 ,  229 . Applying the same principles to pump  210  as pump  110 , the thickness of each of the compression zones  226 ,  227  continuously decreases from inlet passage  216  toward outlet passage  118 , and each of the compression zones has a non-uniform thickness along the longitudinal (axial) direction of gears  222 ,  223 . However, as can be seen by reference to FIG. 9, the thickness of the compression zone is greatest at a point near one end of each of gears  222 ,  223 , and continuously decreases toward the opposite end. This modification is provided to adapt the principle of this invention to a pump  210  having helical gears  222 ,  223  rather herringbone gears. Likewise, seal zone  228 ,  229  and compression zones  226 ,  227  are defined by seal zone boundaries  232 ,  233 , which follow the contour of the helical teeth of gears  222 ,  223 . Accordingly, seal zones  228 ,  229  are approximately triangular in shape. 
     The principles of this invention can also be applied to gear pump  310  (FIGS.  12  and  13 ), which utilizes spur gears  322 ,  323  having teeth which extend along straight lines parallel with the axial directions of gears  322 ,  323  as shown in FIG.  12 . Pump  310  is similar to pump  110  with respect to the shape of housing  312 , with the primary difference being that seal zones  332 ,  333  and compression zones  326 ,  327  are defined by seal zone boundary lines  332 ,  333 , which are straight lines which are parallel with the rotational axis of gears  322 ,  323  to maximize the area of compression zones  326 ,  327  while maintaining a seal between at least one tooth, and more preferably two teeth of each gear  322 ,  323  and the internal walls of housing  312  in the area of seal zone  328 ,  329 . 
     The invention has been tested in the laboratory and evaluated in the manufacture of polystyrene for a given material and a given pressure differential (between the pump inlet and outlet) fill. Efficiency (ratio of the volume of product pumped to base volume of pump defined by tooth volume) as a function of pump speed (RPM) was shown to remain relatively high (greater than 85 percent) over a broader range of pump speed as compared with conventional gear pumps. 
     It will be apparent to those skilled in the art that various modifications to the preferred embodiment of the invention as described herein can be made without departing from the spirit or scope of the invention as defined by the appended claims.