BACKGROUND OF THE INVENTION
The subject matter of this application relates to a pump.
BRIEF DESCRIPTION OF THE DRAWINGS
For a better understanding of the invention, and to show how the same may be carried into effect, reference will now be made, by way of example, to the accompanying drawings, in which:
FIG. 1 shows a perspective view of a pump,
FIG. 2 shows a side elevation of a pump,
FIG. 3 shows a sectional view of the pump taken on the line 3-3 in FIG. 2,
FIG. 4 shows a sectional view of the pump taken on the line 4-4 in FIG. 2, and
FIG. 5 shows an exploded perspective view of the pump.
FIG. 6 shows a first alternative embodiment of the pump of FIG. 1.
FIG. 7 shows a second alternative embodiment of the pump of FIG. 1 having a first stage and a second stage.
FIG. 8A shows an exploded view of the first stage of the pump of FIG. 7.
FIG. 8B shows an exploded view of the second stage of the pump of FIG. 8.
FIGS. 9 and 10 shows a passage that interconnects the outlet chamber of the first stage of the pump of FIG. 7 with the inlet chamber of the second stage of the pump of FIG. 7.
FIG. 11 shows a sectional view of a work rotor/sealing rotor combination that shows a novel vane structure and a novel sealing chamber.
FIG. 12 shows an enlarged view of the vane of FIG. 11.
DETAILED DESCRIPTION
Referring to FIGS. 1-3, the illustrated pump comprises a pump body 10 composed of a pump rotor housing 14, a gear housing 18 and two end caps 22, 26. The end cap 22 is formed with two recesses that accommodate respective ball bearings 27, 28 and the gear housing 18 is similarly formed with two recesses that accommodate respective ball bearings 29, 30. The outer races of the ball bearings are press fitted in the respective recesses. First and second rotor drive shafts 31, 34 are press fitted in the inner races of the bearings and extend parallel to one another through the interior space of the pump rotor housing 14. The interior space of the pump rotor housing is composed of two generally cylindrical cavities 36, 37 that intersect in a region X (FIG. 3). Two seals (only one of which, designated 35, is shown in FIG. 5) surround the periphery of the interior space and are in sealing engagement with the end cap 22 and the gear housing 18 respectively. A work rotor 38 and a sealing rotor 42 are mounted on the shafts 30, 34 respectively and are keyed for rotation with the shafts. The rotors are located in the cavities 36, 37 respectively.
The gear housing 18 is formed on the opposite side from the bearing recesses with a gear recess 46 (FIG. 4) into which the two shafts 31, 34 extend. Two spur gears 50, 54 of equal size are fitted on the shafts 31, 34 respectively and are located in the gear recess 46. The two spur gears are in meshing engagement. Each gear includes a cylindrical boss that projects into a recess 56 in the end cap 26.
An electric motor (not shown) having a drive shaft 58 is attached to the end cap 26. A drive pinion 62 is attached to the drive shaft of the motor and is in meshing engagement with the spur gear 50. Accordingly, when the motor drives the pinion 62, the two spur gears 50, 54 are driven at equal speeds in opposite directions.
The work rotor 38 is generally cylindrical and has two diametrically opposed vanes 66, 68 extending parallel to the central axis of the work rotor and projecting radially therefrom. When the work rotor rotates within the cavity 36 of the interior space, a small clearance exists between the tip of the vanes and the surface bounding the cavity 36. Thus, as the work rotor rotates, the work rotor and the pump rotor housing are in an effective sealing relationship. The cylindrical surface of the lower cavity extends at least 180 degrees about the central axis of the work rotor so that there is always at least one vane between the inlet passage and the outlet passage.
The pump rotor housing 14 is formed with an inlet passage 69 and an outlet passage 70 that communicate with the cavity 36. The upper end of each passage is internally threaded to receive a suitable hose attachment fitting.
The sealing rotor 42 is generally cylindrical and is formed with two peripheral notches 73, 74 that extend longitudinally of the rotor parallel to the axis of rotation of the rotor.
It will be appreciated from examination of FIG. 3 that the configuration of the work rotor 38 corresponds to a spur gear in which all the teeth but two have been removed and the configuration of the sealing rotor 42 corresponds to a spur gear in which all the spaces but two between the teeth have been filled.
The radius of curvature of the upper cavity 37 in the regions Y is slightly greater than the radius of the cylindrical surface of the sealing rotor. The peripheral surface of the upper cavity in each of the regions Y subtends an angle at least as great as the angle subtended by the peripheral notches 73, 74, so that during rotation of the sealing rotor the external surface of the sealing rotor remains in effective sealing relationship with the pump rotor housing with respect to flow of gas around the sealing rotor.
The radius of curvature of the cavity 37 between the regions Y is somewhat greater than in the regions Y, which facilitates manufacture of the pump rotor housing because the tolerance on the dimensions of the peripheral surface of the upper cavity between the regions Y may then be greater than in the regions Y.
As shown in FIG. 3, the vane 66 of the work rotor is positioned in the notch 73 of the sealing rotor. This position is referred to as the 12 o'clock position, having regard to the angular position of the vane 66. As the work rotor rotates in the clockwise direction (and the sealing rotor rotates in the counter clockwise direction), the trailing flank of the vane 66 rolls over the flank of the notch 73 and ultimately disengages from the notch. As the rotors continue to rotate, a very narrow clearance is defined between the cylindrical surface of the work rotor and the cylindrical surface of the sealing rotor. When the work rotor has rotated through almost 180°, the vane 68 rolls into the notch 74 and the cooperation between the surface of the vane and the surface of the notch maintains a narrow clearance between the work rotor and the sealing rotor. At all angular positions of the work rotor 38, there is a very narrow clearance between the work rotor and the sealing rotor 42. The narrow clearance provides an effective sealing relationship between the work rotor and the sealing rotor. The seal between the work rotor and the sealing rotor is referred to herein as the rotor seal. The notches in the sealing rotor accommodate the vanes when the work rotor rotates without destroying the rotor seal.
Depending on the angular position of the work rotor 38, the sealing rotor 42 and the two vanes 66, 68 define two or three chambers within the cavity 36. At the position shown in FIG. 3, there is an inlet chamber 71 and an outlet chamber 72. The inlet passage 69 opens into the inlet chamber 71 and the outlet passage 70 opens from the outlet chamber 72.
Referring again to FIG. 3, as the work rotor rotates from the 12 o'clock position to about 2 o'clock, the vane 66 reaches and passes the upper edge of the inlet passage. The inlet chamber 71 is defined between the vane 68 and the rotor seal. Thus, as the rotor rotates the volume of the inlet chamber 71 increases and tends to cause a reduction in pressure in the inlet chamber thereby inducing a flow of gas into the inlet chamber from the inlet passage 69.
When the vane 66 reaches the lower edge of the inlet passage, the inlet chamber 71 that was bounded by the trailing flank of the vane 68 becomes a transfer chamber and a new inlet chamber is created between the rotor seal and the trailing flank of the vane 66. The transfer chamber 71 between the leading flank of the vane 66 and the trailing flank of the vane 68 is isolated from the inlet passage. A quantity of gas is trapped in the transfer chamber, except for minor leakage between the tips of the vanes and the peripheral surface of the lower cavity 36. Advancing movement of the vane 66 pushes the trapped gas in the clockwise direction about the central axis of the working rotor.
As the work rotor continues to rotate, the tip of the vane 68 reaches the lower edge of the outlet passage 70. The outlet chamber and the transfer chamber are then in communication and a new outlet chamber is thereby created between the leading flank of the vane 66 and the rotor seal. The work rotor continues to rotate and the advancing of the vane 66 decreases the volume of the outlet chamber, tending to increase the pressure in the outlet chamber and expel gas from the outlet chamber through the outlet passage 40. The rotor seal and the narrow clearance between the peripheral surface of the upper cavity in the region Y and the cylindrical surface of the sealing rotor in the region Y provides a large resistance to leakage of gas from the outlet chamber. Accordingly, most gas is forced to leave the outlet chamber through the outlet passage.
The term effective sealing relationship used herein does not require a perfect seal, with the external surfaces of the work rotor and the sealing rotor, for example, continuously in sealing contact. An effective sealing relationship between two members requires that the rate at which fluid can leak between the members should be small relative to the rate at which fluid is delivered from the inlet passage to the outlet passage.
In a conventional external gear pump, the gear teeth divide the incoming flow of air into two streams, each of which is chopped by gear teeth into small volumes which are subsequently combined. This manner of operation consumes energy, resulting in heating of the gas. In the case of the pump illustrated in FIG. 1-5, all the gas proceeds from the inlet passage to the outlet passage along the same path and for each revolution of the work rotor, the flow of gas is chopped into only two volumes.
In a modification of the pump shown in FIGS. 1-5, the external surfaces of the rotors and internal surfaces of the cavities are in contact, thereby improving the rotor seal and the seals between the rotors and the pump rotor housing. In order to minimize friction between surfaces, which would result in heating of the pump components and possible bear of the pump components, the surfaces may be provided with anti-friction coatings.
One desirable advantage of the pumps disclosed herein is that several such pumps may be easily interconnected to provide a desired mass flow rate, a desired compression ratio, or both. Referring first to FIG. 6, for example, an alternative pump 100 may comprise any desired number of pump rotor housings 14 affixed to each other end-to-end. Each pump rotor housing may essentially comprise the structure shown in FIG. 3, but where the respective work rotors and sealing rotors in all of the pump rotor housings 14 share common shafts 31, 34 (shown in FIG. 4). The shafts 31, 34 may be driven using a common gear housing 18 shown in FIG. 6 by a motor (not shown) that rotates shaft 58. The entire assembly is preferably enclosed within end caps 22 and 26.
The pump rotor housings 14, gear housing 18, and end caps 22, 26 of FIG. 6 may be interconnected using dowels (not shown). Alternatively, any other technique or structure may be used to interconnect the pump rotor housings, gear housing, and end caps, such as screws, welds, rivets, a glue or other bonding agent, etc. Preferably, some or all of the outlets 70 of each pump rotor housing 14 are interconnected to each other using an exhaust manifold (not shown) to consolidate the fluid output of the assembly. In some embodiments, however, depending on the application, an exhaust manifold may be neither necessary nor desirable. Similarly, an intake manifold may be used in some applications to interconnect the inlets 69 of each pump rotor housing 14.
As can be seen in FIG. 6, a desired mass flow rate of fluid may be achieved through the pump 100 simply by selecting the number of pump rotor housings 14 to include within the assembled pump 100. It will be understood by those skilled in the art that each of the pump rotor housings 14 may be individually designed to provide differing mass flow rates by selecting the width or other dimensions of the work rotors therein.
As another alternative embodiment, FIG. 7 shows a multi-stage pump 200 that has a first stage 210 and a second stage 220. Fluid, such as air for example, enters the multi-stage pump 200 through an inlet 230 of the first stage 210 and exits through an outlet 240 of the second stage 220. The multi-stage pump 200 may include a motor 250 that rotates gears 252 and 254, which are in turn connected to respective shafts 260 and 262 (shown in FIG. 8A) that drive the work rotors and sealing rotors of each of the stages 210 and 220, as later described. The pump 200 may also include end caps 256 and 258. It should be understood that the motor 250 may be optionally omitted from the multi-stage pump 220, such that an end-user may supply a motor with the desired power or other characteristics. It should also be understood that, although the multi-stage pump 200 shows gears 252 and 254 mounted to the exterior of end cap 256, other embodiments may employ a separate gear chamber within the end caps 256 and 258, as shown in FIG. 1, for example.
Referring to FIGS. 8A and 8B, the first stage 210 may include shafts 260 and 262 that each fit within a respective one of a sealing rotor 264 and a work rotor 266. The sealing rotor 264 and work rotor 266 may each be housed in a respective one of cavities 268 and 270 that are each formed in a first stage body 272 that receives fluid from inlet 230 and is compressed by the action of the work rotor 266 of the first stage 210. The fluid then passes to the second stage 220 through an aperture 274 defined in a wall 276 of the first stage and an aperture 278 defined in a wall 280 of the second stage 220. The second stage 200 similarly includes a sealing rotor 282 and a work rotor 284 each driven by a respective one of the shafts 260 and 262 that extend through the walls 276 and 280. The sealing rotor 282 and work rotor 284 may each be housed in a respective one of cavities 286 and 288 that are each formed in a second stage body 290 that receives fluid from the aperture 278, compresses the fluid further by the action of the work rotor 284 and exhausts the fluid through the outlet 240. As can be appreciated from these figures, the work rotor of the 266 of the first stage 210 compresses fluid received from the inlet 230 by forcing the fluid through an outlet chamber, of diminishing volume as described earlier with respect to the pump of FIGS. 1-4, and which empties into the smaller inlet chamber of the second stage 220. The second stage 220 then compresses the fluid further by forcing the fluid into an outlet chamber of the second stage of a diminishing volume, again as described earlier with respect to the pump of FIGS. 1-4. Thus, the combined operation of the stages 210 and 220 can achieve a significant pressure ratio between the inlet 230 and the outlet 240. For example, where the volume of the work rotor chamber of the second stage 220 is one half that of the work rotor chamber of the first stage, the pressure ratio of the inlet 230 to the outlet 240 is approximately 5:1. Those of skill in the art will appreciate that even higher compression ratios may be achieved by adding further stages. Thus, the multi-stage pump 200 can be assembled modularly from selective pump stages of differing sizes.
Referring to FIGS. 9 and 10, the multi-stage pump 200 preferably is constructed without a pressure tube line between the stages. Specifically, the walls 276 and 280 that respectively form the interface between the first stage 210 and the second stage 220 may each define a respective half of an opening 292 that connects the outlet chamber of the first stage 210 with the inlet chamber of the second chamber 220. Each half of the opening 292 may comprise an elongate portion 294a and 294b, respectively, and an aperture 296a and 296b into a respective stage 210 or 220. The elongate portions respectively defined by the walls 276 and 270 should be generally parallel and generally aligned with each other, while the apertures 296a and 296b should be oriented on opposite sides of the aligned elongate portions, because the synchronized rotation of the work rotors of the two stages means that the inlet chamber of the second stage is laterally offset from the outlet chamber of the first stage in the multi-stage pump 200. However, it is possible that other embodiments may employ structures or configurations that have the work rotors of the respective stages rotate in opposite directions, in which case the walls 276 and 280 may simply define a straight opening between the outlet chamber of the first stage and the inlet stage of the second stage, and in that case, there may simply be one wall between the stages. Where two walls 276 and 280 are used, the walls may be preferably each be formed of a plastic material and bonded together using a bonding agent. Alternatively, the walls 276 and 280 may be formed of metal, fiberglass, etc., and other means may be used to secure the walls 276 and 280 to each other, such as screws, rivets, welds, etc.
As noted previously, a multi-stage pump 200 may be constructed from modular components to achieve a desired compression or pressure ratio. For example, such components could comprise end caps, a plurality of stages of different respective volumes, and a plurality of interfaces between stages where each interface comprises two walls together defining the opening 292. The plurality of stages could include intermediate stages that lack both an inlet 230 and an outlet 240, so that a three stage, four stage, or higher-stage pump could be assembled with intermediate stages simply receiving fluid from one stage having a higher volume flow and ejecting fluid into another stage with a lower volume flow. Furthermore, those of ordinary skill in the art will appreciate that such modularity may include the assembly depicted in FIG. 6, e.g. a multi-stage compressor shown in FIG. 7 may be combined with one or more identical multi-stage compressors as generally disclosed in FIG. 6, with the outlets of the respective multi-stage compressors connected by an exhaust manifold, etc. Stated differently, each pump rotor housing 14 shown in FIG. 6 may essentially comprise a multi-stage pump 200 (but preferably using a common gear drive system and common motor).
The multi-stage pump 200 shown in FIGS. 7-10 may also include several novel features as further described below, although it should be understood that each of these additional features may be incorporated into the pump depicted in FIGS. 1-5, or any other existing or future pump. First, referring to FIGS. 8A and 8B, the shafts 260 and 262 extend through multiple elements, including one or more of the end caps 256 and 258, as well as the walls 276 and 280. Because of machine tolerances that allow rotation of the shafts through these elements, leakage may occur not only from the high pressure chambers of the stages 210 and 220 into the atmospheric region surrounding the pump, but such leakage may occur from one stage to the next as the pressure between each stage also varies considerably. A novel sealing mechanism may take advantage of these pressure differences. Specifically, sealing rings 300 may each be aligned within respectively associated recesses 310 such that the pressure difference pushes the seal into a tighter sealing engagement with the walls and/or end caps. For example, the seals 300 shown in FIG. 8A between the end cap and the rotors 264, 266 fit very tightly around the respective shafts 260 and 262, and thus surround the small gap between these shafts and the respective apertures in the end cap 256 through which they extend. When the multi-stage pump 200 is being operated, these seals will be pushed into a very tight sealing engagement against the end cap 256 by the pressure induced by the work rotor. This same effect is duplicated between the first stage 210 and the second stage 220, as well as the second stage 220 and then end cap 258, if necessary. Preferably, the seals may be fabricated of a Teflon™ material that includes microspheres at a 30% ratio, although seals fabricated of other appropriate materials may also be used.
Second, referring to FIG. 11, the cavity 270 preferably includes opposed regions 320 in which the radius of the cavity diminishes slightly so that the tips of the vanes 330 come very close to the surface of the cavity 270 so as to effectively seal off the inlet 230 from the chamber that transfers fluid to the outlet 240, once the tips of the vanes 330 reach the region 320 closest to the inlet 230. Preferably, however, the opposed regions 230 are positioned slightly below a horizontal centerline through the work rotor 266, as the present inventor realized that positioning the regions 230 any closer to the inlet (or the outlet) would create a zone during rotation where the vanes 330 are simply pushing fluid between the inlet and the outlet, without either picking up fluid from the inlet, or pushing air through the outlet. The optimal distance below the centerline for the regions 230 depends largely on the width of the vanes 330, given that one vane needs to completely clear the region 230 at the outlet chamber before the transported fluid begins compressing, and at that point, the leading surface of the other vane will be beyond the horizontal centerline of the cavity 270. Stated differently, by limiting the tight seal between the vanes 330 and the surface of the cavity 270 to the part of the cavity slightly below the horizontal centerline, as disclosed above, the efficiency of the work rotor is improved by preventing the vanes 330 from having to move any fluid around the cavity during any portion the work rotor's rotation where it is not compressing that fluid at the outlet chamber of the stage. One of ordinary skill in the art will understand that the term “below” is of relative directionality in that, if the position of the work rotor and the sealing rotor were reversed, the opposed regions 320 would be slightly “above” the horizontal centerline, or if the pump had work and sealing rotors horizontally side-by-side, the opposed regions 320 would be slightly to the “left” or “right” of a vertical centerline of the work rotor, etc.
Third, referring to FIGS. 11 and 12, each vane 330 of a work rotor 266 or 284 may preferably engage with a respective notch 340 in the opposed sealing rotor 264 or 282 along opposed curved surfaces 342 of the vane 330 and 344 of the notch 340. At the 12-o'clock position of the vane 330, these curved surfaces approximately contact each other, within designed tolerances, preferably at an iso-speed line where the distance to the rotational axis of the work rotor is approximately equal to the distance to the rotational axis of the sealing rotor, i.e. a line where the sealing rotor and the work rotor are moving tangentially at the same speed. Preferably, as well, the radius of curvature of the surfaces 342 and 344 are substantially the same so as to both provide a tight seal while simultaneously avoiding locking of the vane 330 in the notch 340.
Fourth, referring to FIG. 12, because of design tolerances that allow the vane 330 to rotate within the notch 340, there is a small gap between the tip of the vane 330 and the surface of the notch 340 as it begins to clears the sealing rotor 264, and this small gap tends to cause undesirable leakage from the inlet 230 into the notch 340, so that fluid is undesirably transported backwards through the sealing rotor 264. To counteract this effect, the vane 330 may include a tip that is shaped to create negative feedback on this leakage. Specifically, the tip may include a groove that creates turbulence, such as a vortex, as the fluid passes backwards over the tip of the vane 330 and into the notch 340. The vortex creates backpressure that counteracts the flow of fluid into the notch 340, as shown by the arrow 346 in FIG. 12. In some embodiments, the groove may be U-shaped as depicted in FIG. 12. Alternatively the groove may be V-shaped, or still alternatively the shape of the tip may be any shape that creates negative feedback on the flow of fluid into the notch 340.
It will be appreciated that the invention is not restricted to the particular embodiment that has been described, and that variations may be made therein without departing from the scope of the invention as defined in the appended claims, as interpreted in accordance with principles of prevailing law, including the doctrine of equivalents or any other principle that enlarges the enforceable scope of a claim beyond its literal scope. For example, the invention is not restricted to the sealing rotor having the same number of notches as the number of vanes of the work rotor. With suitable adjustments in timing of rotation of the rotors, the sealing rotor may have only one notch. Moreover, the work rotor may have more than two vanes, although it will be appreciated that as the number of vanes increases, the volume of the pump available for pumping fluid will decrease. Unless the context indicates otherwise, a reference in a claim to the number of instances of an element, be it a reference to one instance or more than one instance, requires at least the stated number of instances of the element but is not intended to exclude from the scope of the claim a structure or method having more instances of that element than stated. The word “comprise” or a derivative thereof, when used in a claim, is used in a nonexclusive sense that is not intended to exclude the presence of other elements or steps in a claimed structure or method.
Patent Application
20140161655
