HIGH EFFICIENCY GEAR ROTARY PUMP FOR PUMPING VISUAL ALTAM FLUIDS
Field of the Invention The present invention relates to apparatuses for transporting highly viscous fluids and more particularly, to rotary gear pumps. Antecedents of the Invention. Rotary gear pumps are used for the transport of highly viscous fluids, such as polymer melts. For example, rotary gear pumps are typically used to transport a viscous polymer melt from a container (such as a devolatilizer), to another operating unit, such as a centrifuge. In most cases, the highly viscous polymer melt enters the pump inlet under the influence of gravity essentially without positive pressure. Known rotary gear pumps are susceptible to a number of difficulties in their operation. In particular, for any given pump geometry, the known rotary gear pumps are extremely limited with respect to the viscosity range of fluids they can handle. Generally, as the viscosity of the fluid increases, the range of performance of the rotary gear pump decreases, often resulting in a bottleneck in production. Likewise, in general, as it increases the speed of the gear pump (RPM), initially increases the performance of the pump, but eventually reaches a high plateau level, where the additional increase in the speed of the pump does not result in some significant increase in performance, and can lead to a production bottleneck. From then on, it has generally not been possible to effectively overcome such a bottleneck in production, since the high-level level of the pump speed against pump performance has been reached without replacing the pump. existing by a larger pump. However, the devolatilizer is usually configured in a special manner to be coupled to a rotary gear pump of a particular design, and it is generally not possible to change to a rotary gear pump of larger capacity of conventional design, without also replacing or modifying the meaningfully the devolatilizer. Therefore, it would be highly desirable to provide a rotary gear pump which operates more efficiently to eliminate such bottlenecks in production, without requiring significant replacement or modification of the devolatilizer. Several attempts have been made to design rotary gear pumps, which have the ability to operate efficiently over a wider range of fluid viscosity and over a wider range of pump speeds. These efforts have focused mainly on the geometry of the pump, particularly on the pump inlet part, for example see U.S. Patent No. 3,476,481. However, the known pump designs have not been completely satisfactory and further improvements are needed. The present invention provides a rotary gear pump having an improved geometry which attenuates the limitations related to the viscosity of the fluid being pumped and the speed of the pump. More specifically, the gear chamber has been designed to provide compression zones, which allow more fluid to be compressed during a longer path in the gear teeth of the pump, and therefore, provide higher production ranges and a superior filling efficiency. The improved geometry allows the gear pumps of the present invention to operate more efficiently in a relatively higher speed range of the pump and with a relatively wider viscosity range of the fluid. The gear pumps of the present invention, include a defined compression zone between each pair of gears of the pump and internal walls of the gear chamber, in which the compression zones have a non-uniform thickness, that is, that the spacing between the teeth of the gears of The pump and the internal walls of the gear chamber in the vicinity of the compression zone, varies along the length of the gears. Brief Description of the Drawings. Figure 1 is a schematic representation of the cross-section in elevation of a gear pump of the prior art, the cross section being perpendicular to the axes of rotation of the gears of the pump; Figure 2 is a schematic representation of the cross section of the pump shown in Figure 1, the view being along the line I-I of Figure 1;
Figure 3 is a schematic representation of a crsection of a gear pump according to the present invention, the crsection being perpendicular to the axes of the pump gears; and Figure 4 is a schematic crsectional representation of the gear pump shown in Figure 3, the view being along the lines 11-11 of Figure 3; Figure 5 is a top plan view of the gear pump shown in Figure 3, with the pump gears and the inlet part of the pump removed; Figure 6 is a crsectional elevation of the gear pump shown in Figures 3 to 5, with the pump gears removed, as seen along the VI-VI lines view of FIG. Figure 5; Figure 7 is a crsection in elevation of the pump shown in Figures 3 to 6, with the gears of the pump in place, as seen throughout the view of the lines VI I- VI I of Figure 4; Figure 8 is a top plan view of the pump shown in figure 3 through 7, with the double helical gears of the pump in place and with the inlet part of the pump removed; Figure 9 is a top plan view of an alternative embodiment of the present invention, configured to be used with helical gears, with the inlet portion of the pump and the gears removed;
Figure 10 is a crsectional elevation view of the pump shown in Figure 9, with the gears and the inlet portion of the pump in place as seen throughout the view of the XX lines of figure 9; Figure 1 1 is a top plan view of the pump shown in Figures 9 and 1 0 with the gears in place and with the inlet part of the pump removed; Figure 1 2 is a top plan view of a second alternative embodiment of the present invention, which utilizes cylindrical gears, with the inlet portion of the pump removed and with the cylindrical gears in place; and Figure 1 3, is a top plan view of the pump shown in Figure 1 2, with the inlet part of the pump and the cylindrical gears removed. Detailed Description of the Invention. In FIGS. 1 and 2, a typical gear pump according to the prior art is illustrated schematically. The prior art gear pump 1 0 includes a housing 1 2 defining internal walls 14. The gear pump 1 0 includes an inlet passage 1 6, an exit passage 18 and a gear chamber 20 positioned between the passage of entrance and the exit passage. The gears of the pump 22, 23 are supported in a rotatable manner within the gear chamber 20. The directions of rotation of the gears of the pump 22, 23 are indicated by arrows 24, 25. The gears of the Bomb 22 and 23 have inter-gear, such as double helical style teeth. The compression zones 26, 27 are delineated between the gears of the pump 22, 23 and the inner wall 14 of the gear chamber 20. The compression zones 26 and 27 have an adjacent input passage of maximum thickness 1 6. thickness of the compression zones 26, 27 decreases in the direction of the exit passage 1 8 and reach a minimum thickness in about one location on a plane delineated by the parallel axes of the pump gears 22, 23. The thickness of the A compression zone refers to the distance from the outer surfaces of the teeth of the pump gears to the closest surface of the inner walls of the gear chamber. As can be seen with reference to Figure 2, the thickness of the compression zones 26, 27 does not vary along a direction parallel with the axes of rotation of the pump gears 22, 23. In the figures of FIG. 3 to 4 a gear pump is shown having a design according to the principles of the present invention. The gear pump 1 1 0 includes a housing 1 12, having internal walls 1 14 defining an entry passage 1 16, an exit passage 1 1 8, a gear chamber 120 placed between the entry passage 1 16 and the exit passage 1 1 8. The gears of the pump 122, 123 are rotatably supported within the gear chamber 120. The gears of the pump 122, 1 23 include internal gear teeth, which, in the case of the embodiment shown in FIGS. 8, are helical double style teeth. The direction of rotation of the gears of the pump 1 22, 123, are indicated by the arrows 1 24, 125. The gear chamber 1 20 is generally divided into two compression zones 1 26, 1 27 and two seal zones 128 , 129. The compression zones 126, 127 are defined as the internal volume portions of the gear chamber 120, which are positioned between the teeth of the gears 122, 123 and the inner walls of the gear chamber 1 20, and which are located above the seal areas 1 28, 129. The seal zones 1 28, 129 refer to the internal volume part of the gear chamber 1 20 in which the space between the teeth of the gears 1, 22, 123, is small enough to effectively prevent any significant fluid movement through the space between the teeth of the gears 1, 22, 123 and the inner walls of the gear chambers 120, thus providing, an effective seal against the flow of fluid passing through the upper surfaces of the teeth of the gears 122, 123. Each of the compression zones 126, 127 has a non-uniform thickness. The thickness of each of the compression zones 1 26, 127, which is the distance from the upper surfaces of the teeth of the gears 1 22, 1 23 to the surface of the inner walls of the gear chamber, is greater at a location of the adjacent entrance passage 1 16. The thickness of each of the compression zones 126, 127 decreases continuously from the inlet passage 1 1 6 towards the exit passage 1 1 8. Preferably, the thickness of the compression zones 126, 127 decrease smoothly from the entry passage 1 1 6 towards the exit passage 1 1 8. The expression "decreases smoothly", as used in the present description, means that the walls internal 1 14 defining the compression zones 1 26, 1 27, do not have abrupt or sharp edges defined by the interception planes, but are curved continuously. As can be seen with reference to Figure 4, the compression zones 126, 1 27 have a non-uniform thickness along the longitudinal direction of the gears 1 22, 1 23, which is greater in a centered location between the axially opposite ends of the gears of the pump 122, 123, and which is smaller at adjacent locations of each of the ends of the gears of the pump 122, 1 23. Preferably, the thickness of the compression zones decreases continuously from the location centered between the opposite ends of the pump gears 1, 22, 23 to each of the ends of the gears of the pump 122, 1 23. In addition, it is desirable that the thickness of the zones of compression 126, 1 27 decreases continuously and smoothly from the centered location between the opposite ends of the gear pumps 122, 1 23 to each end of the gear pumps 122, 123. The compression zones 126, 127 and seal areas 1 28, 129 are preferably further defined through the following criteria: the area of the compression zone is subjected in a maximized manner to the contraction which is sufficient for the areas of the sealing zones 1 26, 127 maintain a reliable seal between the teeth of the gears 1, 22, 123 and the inner walls of the gear chamber 1 20. Maximizing the surface area of the compression zone maximizes filling of the volume limited by the adjacent teeth and the internal walls of the gear chamber 120 in the areas of the sealing zones 126, 1 27, which, in turn, results in a great improvement in the efficiency of the pump. This means that higher flow rates can be achieved for a given gear pump. The superior efficiency of the pump for a pump of a certain size will result in substantial capital savings, since it will not be necessary to replace or substantially modify the associated equipment, such as a devolatilizer, in order to accommodate a pump size larger. The option of replacing a conventional gear pump with an improved gear pump, which has, in accordance with the principles of the present invention, the ability to achieve higher filling efficiency and superior performance ranges for a given size pump , will also result in reduced labor costs related to the modification or replacement of equipment associated with a pump of particular size, and a reduced period, during which a production unit is out of service. The gear pump illustrated 1 1 0, can be described as having a double compression zone where the fluid being pumped is compressed both in the direction of rotation of the gears of the pump 1 22, 1 23 and in the direction parallel to the axes of rotation of the gears of the pump 122, 1 23. The geometry of the double compression zones 126, 1 27 provides a mechanism by means of which, the fluid is induced by the rotation of the gears of the pump 1 22, 1 23 through a progressive narrowing opening, which generates an increase in pressure in the direction of rotation of the gears 122, 123 ending in a soft final contraction at the beginning of the seal areas 1 28 , 1 29. A key difference between the present invention and the prior art, is that the continuous and smooth variation of the limits of the compression zone in both the axial and radial direction, provides more time for filling the space between the teeth and, therefore, allows more fluid to be compressed over a longer path length in the gear teeth of the pump 1 22, 123, thereby providing higher ranges of production and efficiency Top filling As mentioned above, a major restriction in the area of the compression zones 1 26, 1 27, is that a reliable seal must be maintained between the teeth of the gears 1, 22, and the inner walls of the chamber. gears 1 20. This generally means, that the seal areas 128, 1 29 must be designed, formed and contoured so that the entire length of at least one tooth of each of the gears 122, 1 23, is sufficiently spaced in closed form to its associated seal zone to maintain an effective seal between the compression zone and the pump discharge. However, as illustrated in Figure 7, it is generally preferable to design, form and contour the seal areas 1 28, 1 29, so that at least two adjacent teeth in each of the gears 122, 123 are sufficiently spaced in closed form to their respective seal zones to maintain an effective seal (that is, one in which very little, if nothing, of fluid can flow between the teeth and the walls of the gear chamber in the area of the seal areas), along the entire length of the two adjacent teeth. This will prevent minor damage, such as wear or excessive abrasion, that any tooth only significantly affect the overall performance of the pump, thus ensuring a longer and more reliable service life without significantly reducing the efficiency and performance of the pump. the bomb. Because the seal regions 128, 129 are formed to follow the length of at least one tooth, and preferably two adjacent teeth of the gears 122, 123, the shape of the seal areas 1 28, 129 is determined by the tooth pattern of the gears 1 22, 1 23. In the case of the helical double gears, the teeth are wound around the gears 1 22, 123 in a helical path in a first direction (e.g., clockwise), from a first end of the gears to the entire length of the middle section of the gear and subsequently take a sharp turn and wind around the gears in a helical path in a direction opposite to the first direction (e.g., counterclockwise), from the mid-section of the gear, to a second end of the gears opposite the first end, As shown in Figure 8. Therefore, in the case of the pump 1 1 0, which has a double tunnel discharge with two unloading doors 1 30, 1 31 (Figures 5 and 6), and the which has gears do helical bles 122, 1 23, the 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 the gear chamber 120 defining the seal areas 128, 129 it results in a V-shaped seal area, as indicated in Figure 5, by the boundaries of the seal area 1 32, 1 33. It should be noted that the boundaries of the seal area 1 32, 1 33, are shown only for purposes of illustration, since there is a smooth transition from the compression zone to the seal zone, which, however, would not be easily visible. A double tunnel discharge (as shown in Figures 5 and 6) is preferred, because it provides a larger area for the compression zone in the suction part of the pump 1 1 0, without violating the requirement of that at least one tooth, and more preferably two teeth, of each of the gears 122, 1 23 seal against the portion of the walls of the gear chamber is defined in the seal form. The double tunnel discharge also allows a larger rotation angle of the gears 1 22, 1 23 before the teeth break the seal. In figures from 9 to 11, an alternative embodiment of the present invention using helical gears is shown. As with the gear pump 1 10, the gear pump 21 0 includes a housing 21 2 which defines internal walls 214, an inlet passage 6, an outlet passage 218 and a gear chamber 220 positioned between the inlet passage and the output pump. The gears 222, 223 are rotatably supported within the gear chamber 220. The gears 222, 223 have internal gear teeth which are wound helically about the entire length of the gears 222, 223. As with the pump 1 1 0, the compression zones 226, 227 and the seal areas 228, 229, are defined by the principle of providing a double compression zone wherein the fluid is compressed both in the direction of rotation of the gears 222, 223, as in the direction parallel to the rotation axes of the pump gears 222, 223, and the compression zones 226, 227 provide a mechanism by which the fluid is induced by the rotation of the gears 222, 223 to through an opening of narrowing progressively in the direction of rotation to generate an increase in pressure until the fluid reaches a soft contraction at the beginning of the zones d and seals 228, 229. By applying the same principles of the pump 1 1 0 to the pump 21 0, the thickness of each of the compression zones 226, 227 decreases continuously from the inlet passage 21 6 and up to the outlet passage 21 8, and each of the compression zones has a non-uniform thickness along the longitudinal (axial) direction of the gears 222, 223. However, as can be seen with reference to Figure 9, the thickness of the compression zone is greater at a point near one end of each of the gears 222, 223, and decreases continuously towards the opposite end. This modification is provided to adapt the principle of the present invention to the pump 210 having helical gears 222, 223, instead of helical double gears. Likewise, the seal area 228, 229 and the compression zones 226, 227 are defined by the boundaries of the seal area 232 233, which follow the contour of the helical teeth of the gears 222 223. Thus, , the seal areas 228, 229 are approximately triangular in shape. As shown in Fig. 1 2, the principles of the present invention can also be applied to the gear pump 31 0 (Figs.
12 and 1 3), which uses cylindrical gears 322, 323 having teeth which extend along straight lines parallel with the axial directions of the gears 322, 323. The pump 31 0 is similar to the pump 1 1 0 with respect to the shape of the housing 31 2, with the main difference that the seal areas 328, 329 and the compression zones 326, 327 are defined by the boundary lines of the seal area 332, 333, which they are straight lines that are parallel with the axis of rotation of the gears 322, 323 to maximize the area of the 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 inner walls of the housing 31 2 in the area of the seal area 328, 329. The present invention has been tested in the laboratory and evaluated in the manufacture of polystyrene for a given material and a differential pressure filling. n determined (between the inlet and outlet of the pump). The efficiency (ratio of the volume of the pumped product to the base of the volume of the pump defined by the volume of the tooth) as a function of the pump speed (RPM), was shown to remain relatively high (greater than 85%) in a wider range of pump speed, compared to conventional gear pumps. Those skilled in the art will appreciate that various modifications may be made to the preferred embodiment of the present invention as described therein, without departing from the scope of the present invention, as defined by the appended claims.