Enlarging the volume of the gear pump to reduce cavitation at the meshing area

Abstract

A gear having a volume expansion portion for reducing cavitation at a meshing portion of a gear pump is provided. The gear has a gear tooth profile, a body, a plurality of gear teeth with an involute tooth profile extending radially outward from the body, including adjacent first and second gear teeth each having a tip and a root, with a space formed between the adjacent first and second gear teeth, and a root pocket formed directly at the root of the gear teeth in the space, which increases the gear root volume without changing the gear tooth profile, thereby adding buffering for trapped fluid. A gear pump including the gear is also disclosed.

Description

[Technical Field]

[0001] This application claims the benefit of priority to U.S. Provisional Application No. 63 / 433,587, filed December 19, 2022, and U.S. Patent Application No. 18 / 372,480, filed September 25, 2023, the contents of which are incorporated herein by reference.

[0002] The present invention relates to gear mechanisms, and more particularly to cavitation between meshing gear teeth. [Background technology]

[0003] In his theory of hydraulic turbines, Euler predicted the phenomenon of cavitation in 1754. However, it was not until 1895 that the harmful effects of cavitation were identified. Since then, a considerable amount of research has been conducted to avoid the adverse effects on metallic parts caused by cavitation. Initially, such research focused on hydraulic turbine runners, ship propellers, valves, and piping. Subsequent research has shown that cavitation occurs between meshing gear teeth in oil-lubricated gearboxes that transmit torque (see J.B. Hunt et al., "Cavitation between meshing gear teeth," Wear, Vol. 71, Issue 1, pp. 65-78 (September 1, 1981)).

[0004] All pumps operate by creating a low pressure at the inlet, allowing atmospheric pressure (or pressure within the gear mechanism) to force the fluid through the pump. This process makes all pumps susceptible to cavitation. Cavitation is the formation of vapor cavities (gas bubbles) within a liquid when the local pressure of the liquid drops rapidly below the liquid's vapor pressure. This pressure drop causes vapor bubbles to form within the liquid, which usually collapse back into liquid form within a short time. This collapse is violent, producing a loud popping sound and shock waves that can damage nearby surfaces. These shock waves can cause wear in some mechanical components. Vapor cavities that implode near solid surfaces can cause cyclic stresses from repeated exposure to such implosions. Even strong metals can be punctured by exposure to the powerful, localized jets resulting from imploding gas bubbles. If left unchecked, the damage can eventually destroy the pump.

[0005] Reference is now made to the drawings, in which like reference numerals refer to like elements throughout the various views. FIG. 1 is a schematic side cross-sectional view of a gear pump 10. FIG. 1 illustrates a conventional gear pump 10 having two identical spur gears (gears 2 and 3) disposed within a pump housing 6. Each gear 2 and 3 has a plurality of teeth 1 disposed along its circumference (e.g., nine teeth 1 are shown for each gear 2 and 3). In the illustrated example, a first gear 2 is mounted on and integrally supported by a corresponding gear shaft 4, thereby rotating the first gear 2 in a clockwise direction 2a. A second gear 3 is mounted on and integrally supported by a corresponding gear shaft 5, thereby rotating the second gear 3 in a counterclockwise direction 3a. A motor (not shown) rotates the gears 2 and 3.

[0006] As the pair of gears 2 and 3 of the gear pump 10 rotate, the teeth 1 of the first gear 2 mesh (contact) with the teeth 1 of the second gear 3 and then disengage from the teeth 1 of the second gear 3. That is, the teeth of the gears 2 and 3 mesh with each other at meshing position 7. When the first gear (operating or driving gear) 2 rotates clockwise, causing the second gear (driven gear) 3 to rotate counterclockwise, the gear pump 10 draws in fluid F through the inlet opening 8 of the housing 6, where it is transported by the rotating gears 2 and 3 along the interior wall of the housing 6 and around the outer periphery of the gears 2 and 3, as indicated by arrow 8a, to reach the outlet opening 9 of the housing 6, where the fluid F exits the gear pump 10, as indicated by arrow 9a. In this way, the gears 2 and 3 pump the fluid F by continuously trapping a portion of the fluid F in one location and transferring it to another.

[0007] Internal cavitation can occur in one or more locations in gear pump 10. Figure 1A shows one common cavitation location in gear pump 10, where teeth 1 of first gear 2 are close to the wall of housing 6. Figure 1B shows another common cavitation location in gear pump 10, namely, near meshing location 7. Cavitation occurring at meshing location 7 is described in more detail below.

[0008] The incoming fluid F is drawn into the inlet opening 8 by the disengagement of the gears 2, 3 at a position relatively close to the inlet opening 8. Upon disengagement, the inter-tooth volume space formed between adjacent teeth 1 of each gear 2, 3 expands as the teeth 1 of the previously engaged mating gear move out of the space. Upon disengagement, the inter-tooth volume space between adjacent gear teeth 1 fills with fluid F from the inlet opening 8 and is forced to move with each gear 2, 3 between its teeth 1, along the closely adjacent inner surface of the housing wall, and to the outlet opening 9 on the discharge side of the gear pump 10. The very small clearance between the tips of the teeth 1 of the gears 2, 3 and the corresponding inner surface of the housing wall, the speed at which the gear tooth tips move along the inner surface of the housing wall, and the flat bearing surfaces closely adjacent to the sides of the gears 2, 3 keep the fluid F trapped within the inter-tooth volume and prevent it from flowing back toward the inlet opening 8.

[0009] On the discharge side of the gear pump 10, the gears 2 and 3 mesh relatively close to the outlet opening 9, causing the teeth 1 of the mating gear to enter the inter-tooth volume space formed between adjacent teeth 1 of each gear 2 and 3, resulting in the contraction of the inter-tooth volume and forcing fluid F out of the outlet opening 9. As a positive displacement gear pump 10, the fluid discharge pressure is determined primarily by the cross-sectional area of ​​the downstream conduit passage. The meshing of the teeth 1 of the gears 2 and 3 at meshing location 7, which is more or less along an axis joining the axes of symmetry of the gear shafts 4 and 5, and the presence of a flat bearing surface portion closely adjacent thereto, have the effect of isolating the fluid F at the outlet opening 9 from the fluid F at the inlet opening 8.

[0010] In the gear pump 10, cavitation can occur on the suction side in the region where the gears 2 and 3 disengage and separate. In this region, as described above, the expanding intertooth volumes between adjacent teeth 1 of each gear 2 and 3 must be filled with fluid F flowing through the inlet opening 8 under inlet fluid pressure as the teeth 1 of the mating gear leave their previous positions. When the rotational speed of the gears 2 and 3 reaches a certain threshold, the rate of increase of the expanding intertooth volumes can exceed the rate at which these volumes are filled by fluid F flowing into the inlet opening 8 under inlet fluid pressure. In this situation, the local fluid pressure drops below the vapor pressure of any gas dissolved in the fluid F or below the vapor pressure of the pumped fluid F itself, resulting in the formation of cavities (gas bubbles) due to the rupture of the continuity of the fluid F at nucleation sites on any particles or solid surfaces. Such gas, fluid F vapor, or both, evaporate from the surrounding fluid medium into the cavities.

[0011] As the interproximal volume further fills, the increasing localized fluid pressure then causes these cavities to collapse, increasing the pressure and temperature of the vapor within. This phenomenon continues until the volume of these cavities becomes a very small fraction of their original size, ultimately reaching a point where they collapse completely, generating an acoustic shock wave in a very small volume and dispersing the vapor into the surrounding fluid medium. If this collapse occurs on or near the surface of a tooth, it can erode the tooth, leaving a depression in the surface. If this cycle is repeated, it can destroy the surface of the tooth.

[0012] FIG. 2 shows a computational fluid dynamics (CFD) analysis of a gear pump, such as gear pump 10, performed by CFX Berlin Software (Berlin, Germany). Gears 2 and 3 of gear pump 10 each have, for example, nine teeth, whereas the gears of the gear pump analyzed in FIG. 2 each have eleven teeth. A computational fluid dynamics (CFD) analysis uses numerical techniques to simulate fluid motion. Thus, a computational fluid dynamics (CFD) analysis functions as a virtual fluid dynamics simulator. The results show the behavior of velocity, water vapor (cavitation), and air (bubble formation).

[0013] Various attempts have been made to address the cavitation problem in gear pumps such as gear pump 10. The cavitation effect in gear pump 10 can be addressed by using precision-machined helical gears for gears 2 and 3 that are smooth to mesh with each other. However, the internal mechanism of gear pump 10 can cause a local pressure drop of as much as 0.1 bar in water at speeds above 3000 rpm.

[0014] Another approach is to provide channels in the bearing surface adjacent to the meshing location 7 where the gears 2, 3 mesh. The channels extend in opposite directions from the meshing location 7. However, even with this measure, the rate at which the returning fluid F fills the expanding intertooth volume depends on the fluid pressure at the inlet opening 8. Therefore, above a certain rotational speed, the fluid cannot fill the intertooth volume as quickly as the expanding intertooth volume, and cavitation still occurs.

[0015] U.S. Patent No. 7,878,781, issued to Hamilton Sundstrand, Inc. (Windsor Locks, Connecticut, USA), discloses another approach to addressing the cavitation problem in gear pumps 10. In this patent, a pressurized fluid passage is provided in at least one of the bearing structures opposite the meshing region. The fluid passage extends between surface openings in the bearing surfaces of the bearing structures. These surface openings are located on either side of the mating axis of the bearing surfaces. The surface openings are separated from each other by at least the width of a tooth 1 on a pair of gears 2 and 3.

[0016] Despite these attempts, there remains a need for improved gear pumps that address the problem of cavitation. Generally, cavitation can accelerate wear and reduce pump efficiency and lifespan of gear pump components, particularly gear teeth. Accordingly, an object of the present invention is to minimize, and if possible eliminate, cavitation, thereby reducing wear and increasing the efficiency and lifespan of gear pumps. A related object is to significantly reduce or eliminate pressure differentials at the limits of the gear pump. Another object is to provide a gear pump that minimizes fuel consumption by reducing energy losses. [Prior art documents] [Patent documents]

[0017] [Patent Document 1] U.S. Patent No. 7,878,781 Summary of the Invention

[0018] To meet the above needs, achieve the above goals, and in view of its objectives, the present invention provides a gear with increased volume for reducing cavitation in meshing of a gear pump. The gear of the present invention includes a gear tooth profile, a body, and a plurality of gear teeth with an involute tooth profile extending radially outward from the body. The plurality of gear teeth includes adjacent first and second gear teeth, each having a tip and a root, with a space formed between the adjacent first and second gear teeth. The gear of the present invention has a root pocket formed directly at the root of the gear tooth in the space. The root pocket adds compliance to the trapped fluid by increasing the gear root volume without changing the gear tooth profile.

[0019] The present invention also provides a gear pump. The gear pump includes a first gear. The first gear has (i) a gear tooth profile, (ii) a first body, (iii) a plurality of gear teeth of a first involute tooth profile extending radially outward from the first body, the plurality of gear teeth including adjacent first and second gear teeth each having a tip and a root, with a space formed between the adjacent first and second gear teeth, and (iv) a root pocket formed directly at the root of the first gear tooth in the space, the root pocket increasing the gear root volume without changing the gear tooth profile, thereby adding buffering (compliance, adaptability) for the trapped fluid. The gear pump also includes a second gear. The second gear is configured to mesh with the first gear in a meshing region that forms an increased gear meshing volume due to the root pocket. The second gear has a second body and a plurality of gear teeth with a second involute tooth profile extending radially outward from the second body. The gear pump further includes a first gear bearing and a second gear bearing. The first gear bearing and the second gear bearing are configured to position the first gear and the second gear along a bearing centerline. The tooth root pocket does not extend into either the first gear bearing or the second gear bearing. The gear pump further includes a housing having a fluid inlet and a fluid outlet. The first gear, the second gear, the first gear bearing, and the second gear bearing are disposed within the housing.

[0020] It is to be understood that the foregoing summary and the following detailed description are exemplary, but are not restrictive, of the invention. [Brief explanation of the drawings]

[0021] The invention will be best understood by reference to the following detailed description in conjunction with the accompanying drawings, in which: It is emphasized that, according to common practice, the various features of the drawings are not to scale. Rather, the dimensions of the various features have been arbitrarily expanded or reduced for clarity. The drawings include the following figures:

[0022] [Figure 1] FIG. 1 is a schematic cross-sectional side view showing a conventional gear pump. [Figure 1A] An enlarged view of dashed area 1A in FIG. 1 shows one common cavitation location in the gear pump shown in FIG. 1, where the teeth of the first gear are close to the wall of the gear pump housing. [Figure 1B] An enlarged view of the area 1B enclosed by the dashed line in FIG. 1 shows another typical cavitation location (near the meshing location) in the gear pump shown in FIG. [Figure 2] A computational fluid dynamics (CFD) analysis of a gear pump such as the conventional gear pump shown in Figure 1 is shown. [Figure 3] 1 is a perspective view of a conventional high-pressure gear pump, showing the general layout of the gear pump. [Figure 4] FIG. 4 is a perspective view of one of the bearings of the high-pressure gear pump shown in FIG. 3. [Figure 5] FIG. 5 is a front view of one of the movable bearings (floating bearings) of the high-pressure gear pump shown in FIGS. 3 and 4. [Figure 6A] FIG. 1 is a front perspective view of an exemplary gear pump assembly. [Figure 6B] FIG. 6B is an enlarged view of the bearing dam half highlighted by dashed area 6B in FIG. 6A. [Figure 7] FIG. 1 is a side view of a series of exemplary gear teeth. [Figure 8] FIG. 1 is a schematic side view of a gear according to the present invention modified so that cavities (pockets) are machined directly into the roots of the gear teeth. [Figure 9] 9 is a schematic diagram illustrating an exemplary oval shape and dimensions of the pocket shown in FIG. 8. [Figure 10] 9 is a perspective view showing a part of a gear pump including a gear having the pocket shown in FIG. 8, the gear being disposed between a fixed bearing (stationary bearing) and a floating bearing (movable bearing). [Figure 11]FIG. 10 is a perspective view showing a portion of a gear having a beam disposed across the center of a pocket. [Figure 12] FIG. 1 is a schematic diagram illustrating a conventional bearing configuration used in the MFP-590 main fuel pump. [Figure 13] An analysis of trapped volume during gear meshing in the conventional bearing configuration shown in FIG. 12 is shown. [Figure 14] FIG. 1 is a schematic diagram showing the advanced bearing configuration analyzed and tested on the MFP-590 pump. [Figure 15] 15 shows an analysis of trapped volume during gear meshing in the advanced bearing configuration shown in FIG. 14. [Figure 16] FIG. 15 is a schematic diagram illustrating a trapped volume analysis for a particular fluid fuel passing through the advanced bearing configuration shown in FIG. 14. [Figure 17] 13 is a schematic diagram illustrating a trapped volume analysis for a particular fluid fuel passing through the conventional bearing configuration shown in FIG. 12. [Figure 18A] The calculated trapped volume for the advanced bearing configuration shown in Figure 14 is shown for a gear rotation angle of 10°. [Figure 18B] The calculated trapped volume for the advanced bearing configuration shown in Figure 14 is shown for a gear rotation angle of 11°. [Figure 19A] 15 is a graph of measured pressure versus time for an MFP-590 pump equipped with only the advanced bearing configuration shown in FIG. 14, illustrating discharge pressure ripple. [Figure 19B] 15 is a graph of measured pressure versus time for an MFP-590 pump equipped with the advanced bearing configuration shown in FIG. 14 and the pocket shown in FIG. 8, illustrating discharge pressure ripple. DETAILED DESCRIPTION OF THE INVENTION

[0023] In this specification and the appended claims, a number of terms shall be defined to have the following meanings: As used herein, the term "substantially" is a descriptive term indicating approximation, meaning "to a great extent" or "for the most part, but not entirely, of what is specified," and is intended to avoid precise numerical bounds on specified parameters. As used herein, directional terms (e.g., up, down, right, left, front, back, top, bottom) are used only with reference to the drawings and do not imply absolute orientations.

[0024] The term "about" means that the amount, size, composition, parameter, or other quantity or characteristic need not be exact and may, desirably, be approximate and / or larger or smaller, reflecting tolerances, conversion factors, rounding, measurement error, and other factors known to those skilled in the art. When a value is described as "about" or "approximately equal to" a particular numerical value, the value is within ±10% of that numerical value. For example, a value of "about 10" is intended to include values ​​from 9 to 11. Regardless of whether the term "about" is used at the endpoint of a numerical value or range in the specification, the endpoint of that numerical value or range is intended to include both the form modified by "about" and the form not modified by "about." It will be further understood that the endpoints of each range are significant both in relation to the other endpoint and independently of the other endpoint.

[0025] The term "about" refers to any range of values ​​unless otherwise specified. For example, "about 1, 2, or 3" is synonymous with "about 1, about 2, or about 3," and further includes "about 1 to 3," "about 1 to 2," and "about 2 to 3." Specific and preferred values ​​disclosed for components and steps, and ranges, are merely exemplary and do not exclude other defined values ​​or other values ​​within defined ranges. Components and method steps of the present invention include any recited value or any combination of values, specific values, more specific values, and preferred values.

[0026] As used herein, the indefinite article and its corresponding definite article mean "at least one" or "one or more," unless otherwise stated. The words "include," "including," "have," "having," "comprises," "comprising," and similar terms mean exhaustive rather than limiting, i.e., inclusive and not exclusive.

[0027] [Gear Pump Overview] FIG. 3 is a perspective view of a gear pump 10 that is typical of aerospace fluid pumps that operate to pump fuel, lubricant, or other fluids. An aircraft engine, for example, includes a main fuel pump that is central to its control system. This type of pump supplies fuel to the combustion chamber by pumping the required flow rate from a fuel tank. The output flow rate from such a fuel pump is also used as hydraulic fluid to operate actuators (e.g., actuators used to open airflow discharge gates from the engine's core flow to the fan flow).

[0028] FIG. 3 shows the general layout of a high-pressure gear pump 10 (described in U.S. Pat. No. 10,094,291, granted to Safran Aircraft Engines, Paris, France). The gear pump 10 includes a first gear 2 and a second gear 3. The first gear 2 and the second gear 3 mesh with each other, discharging and pumping fuel between the gear teeth 1. The first gear 2 is driven by a gear shaft 4 (see FIG. 1). As shown in FIG. 3, each gear (pinion) 2, 3 has a stub shaft 13, 14, 15, 16 on either side. The first stub shaft 13, 14, shown on the right side of FIG. 3, is supported by fixed bearings 17, 18, each having a first clearance. The second stub shaft 15, 16, shown on the left side of FIG. 3, is supported by floating bearings 19, 20, each having a second clearance. These bearings 17, 18, 19, and 20 are all plain bearings, but the fixed bearings 17 and 18 are held in the housing 6 with a smaller gap than the movable bearings 19 and 20. This allows the fixed bearings 17 and 18 to be displaced in the axial direction so as to compress (press) the gears 2 and 3, thereby reducing the gap that can cause the fed fluid to be recirculated to the low-pressure side.

[0029] Satisfactory operation of the high pressure gear pump 10 depends on adequate sealing between its various components. It is essential to limit leakage of fluid pumped out of the housing 6 and around the gears 2, 3 as it recirculates to the inlet opening 8 of the gear pump 10. Leakage due to recirculation around the gears 2, 3 is minimized by a plurality of springs 30 that press the movable bearings 19, 20 against the gears 2, 3. This is referred to as compressing the gears 2, 3.

[0030] Another criterion for satisfactory operation of the gear pump 10 is explained with reference to Figures 4 and 5. Friction between the bearings 17-20, the gears 2 and 3, and their stub shafts 13-16 is avoided by a hydrodynamic lift layer formed by the recirculation of the pumped fluid. Each bearing 17-20 has several relief patterns drilled into it. The relief patterns include a high-pressure dish 33 and a low-pressure dish 34 located on either side of a separating flange 60 at the periphery of the axially inner surface 35 (near the gears 2 and 3). The high-pressure dish 33 and the low-pressure dish 34 communicate with adjacent fluid spaces at the inlet opening 8 and the outlet opening 9 of the gear pump 10. The high-pressure dish 33 communicates with a curved high-pressure groove 36 opening on the axially inner surface 35 and with a high-pressure groove 37 opening on the radially inner surface 38 of the bearings 17-20 via a hole (not shown). A low-pressure groove 39 extends from the confluence (boundary) of the axially inner surface 35 and the radially inner surface 38. The low-pressure groove 39 is connected to the low-pressure dish 34 via a collecting groove 40. In the bearings 17-20 configured in this manner, fluid circulation from the high-pressure dish 33 to the low-pressure dish 34 is maintained by the operation of the gear pump 10. This forms a fluid dynamic pressure layer between the axially inner surface 35 and the radially inner surface 38, providing dynamic lubrication for the bearings 17-20. Therefore, the stub shafts 13-16 are supported by the fluid dynamic pressure layer of the radially inner surface 38 that fills the gap, and the fluid dynamic pressure layer of the axially inner surface 35 is formed adjacent to the side of the gears 2 and 3. This keeps the gears 2 and 3 slightly separated from the bearings 17-20, preventing the gap from being completely closed despite the presence of the spring 30.

[0031] As shown in FIG. 5, the movable bearings 19, 20 in these known structures have special features on their axially outer surfaces, away from the gears 2, 3. The axially outer surfaces of the movable bearings 19, 20 are divided into two crescent-shaped sections 42, 43, which are arranged in different planes. The two crescent-shaped sections 42, 43 are separated by a step with a seal, so that the crescent-shaped section 42 is exposed to high-pressure fluid and the crescent-shaped section 43 is exposed to low-pressure fluid. The movable bearings 19, 20 are held in place by the step separating the crescent-shaped sections 42, 43. The movable bearings 19, 20 are held in grooves reamed into the housing 6 so that they are offset from the rotational axes of the gears 2, 3. Multiple springs 30 are compressed between the movable bearings 19, 20 and one surface of the housing 6. The springs 30 are attached to corresponding sections 41. These sections 41 are provided only where the crescent-shaped portion 42 is largest (widest), which is approximately one-quarter of the circumference of the movable bearings 19, 20. As a result, the axial thrust is unbalanced, and includes a moment about the lateral axis of the movable bearings 19, 20, which cancels out the opposite moment caused by the pressure difference of the pumped fluid at the axial inner surface 35. Therefore, tilting of the movable bearings 19, 20 about the lateral axis does not occur, and the movable bearings 19, 20 maintain coaxiality with the supporting stub shafts 15, 16, despite the assembly gap of the movable bearings 19, 20 in the housing 6.

[0032] Channels on bearing surface As mentioned above, one approach to addressing the cavitation problem in gear pumps is to provide channels in the bearing surface. A specific example of this approach is provided by U.S. Pat. No. 9,932,980, issued to Woodward, Inc. (Fort Collins, Colorado). U.S. Pat. No. 9,932,980 discloses a gear pump bearing with inlet and discharge relief cuts in the end faces of the floating and stationary bearings. These relief cuts allow pumped fluid to flow from the gear mesh to the top and bottom of the gear on the discharge side, and from the top and bottom of the gear to the gear mesh on the inlet side. These relief cuts leave a portion of the bearing material near the centerline between the inlet and discharge sides, forming a bearing dam. This bearing dam substantially seals the inlet and discharge sides, thereby maintaining pump efficiency. In some embodiments, the shape of the bearing dam can significantly affect the discharge and filling of the gear, which can in turn affect the cavitation performance of the gear pump. In summary, the gear pump described in US Pat. No. 9,932,980 includes a bearing dam having a configuration that ostensibly reduces fluid cavitation and resulting damage.

[0033] More specifically, U.S. Patent No. 9,932,980 describes and illustrates an exemplary gear pump assembly 100. FIG. 6A shows an oblique front perspective view of assembly 100. As shown in FIG. 6A, drive gear bearing 104 includes drive gear bearing half 204a and drive gear bearing half 204b. Drive gear 114 includes drive gear teeth 134, a central shaft portion 234a (e.g., a journal) extending axially from drive gear teeth 134, and a central shaft portion 234b extending axially from drive gear teeth 134 opposite central shaft portion 234a. Drive gear bearing half 204a includes bore 250a, and drive gear bearing half 204b includes bore 250b. When assembly 100 is assembled, bore 250a is configured to receive and rotatably support central shaft portion 234a, and bore 250b is configured to receive and rotatably support central shaft portion 234b.

[0034] 6A , driven gear bearing 106 includes driven gear bearing half 206a and driven gear bearing half 206b. Driven gear 116 includes driven gear teeth 136, a central shaft portion 236a extending axially from driven gear teeth 136, and a central shaft portion 236b extending axially from driven gear teeth 136 opposite central shaft portion 236a. Driven gear bearing half 206a includes bore 250c, and driven gear bearing half 206b includes bore 250d. When assembly 100 is assembled, bore 250c is configured to receive and rotatably support central shaft portion 236a, and bore 250d is configured to receive and rotatably support central shaft portion 236b.

[0035] Assembly 100 includes central fluid dam half 258a in region 6B of FIG. 6A adjacent to holes 250a and 250c in bearing half 204a and bearing half 206a, respectively. FIG. 6B is an enlarged view of central fluid dam half 258a shown in region 6B of FIG. 6A. Corresponding central fluid dam halves (not shown) are also provided adjacent to holes 250b and 250d in bearing half 204b and bearing half 206b, respectively. Central fluid dam half 258a and the corresponding central fluid dam halves cooperate to form a bearing dam.

[0036] 6B, central fluid dam half 258a of bearing dam includes inlet face 260 and outlet face 261. Inlet face 260 includes slots 262 formed as relief cuts in inlet face 260. Outlet face 261 includes vents 263 formed as relief cuts in outlet face 261. In the assembled configuration of assembly 100, central fluid dam half 258a, drive gear teeth 134, and driven gear teeth 136 provide a barrier that substantially blocks fluid flow between fluid inlet cavity 160 and fluid discharge cavity 180 along a bearing split line that transverses the bearing centerline. The configuration of inlet face 260, outlet face 261, slots 262, and vents 263 is designed to address cavitation.

[0037] [Gear tooth shape] 7 is a side view illustrating a set of exemplary gear teeth 300. In some embodiments, the gear teeth 300 correspond to the drive gear teeth 134 and / or driven gear teeth 136 of the exemplary gear pump assembly 100.

[0038] Gear teeth 300 extend radially from gear 302. In some embodiments, gear 302 may be drive gear 114 or driven gear 116. Gear 302 has a root diameter 304. Root diameter 304 is the diameter at the base of intertooth spaces 306. Gear 302 also includes a pitch circle 308. In some embodiments, pitch circle 308 may be a circle derived from the number of gear teeth 300 and a predetermined diametral or circular pitch, or may be a circle around which spacing or tooth profiles are set to establish tooth ratios.

[0039] One of the fundamentals of gear design is the gear tooth profile. The tooth profile is the shape of the gear tooth curve, measured from the root to the tip of the gear tooth. The functional (working) portion of the tooth profile is the area that actually makes contact during tooth meshing. Gears generally have an involute tooth profile. This involute curve helps the gear transmit power smoothly during rolling action. Tooth thickness, diametral pitch, and pressure angle all help determine the gear tooth profile. These factors are determined by the desired contact ratio between the meshing portions of the gear. Gear tooth profile also varies depending on the number of teeth in the gear; the higher the number of teeth, the more linear the gear tooth profile, ultimately forming what is known as a rack gear.

[0040] Each gear tooth 300 includes a tip 310 and a root 312. The tip 310 is the height that the gear tooth 300 protrudes beyond the pitch circle 308, and the root 312 is the depth of the intertooth space 306 between the pitch circle 308 and the root diameter 304. Each gear tooth 300 also includes a pressure angle 320. The pressure angle 320 is the angle, at a pitch point 322 on the pitch circle 308, between a pressure line that is normal to the tooth surface at the pitch point 322 and a plane that is tangent to the pitch circle 308. In an involute tooth such as gear tooth 300, the pressure angle 320 can also be described as the angle between a line of action 324 and a line 326 that is tangent to the pitch circle 308.

[0041] The gear teeth 300 of the gear 302 shown in FIG. 7 exhibit a standard full-fillet root profile. As defined herein, a "standard full-fillet root profile" provides a constant radius that extends in a continuous arc from one tooth 300 to the next. The typical root shape of a spur gear is a full fillet that simultaneously tangents the involute tooth profile and the root diameter. The lowest point in the constant-radius fillet establishes the root diameter. In the case of a hobbed gear, the tooth shape is generated by the path the tool tip follows as the teeth 300 are cut. In the case of a form-ground tooth, the radius is formed at the tip of the grinding wheel. The adjacent flanks of the two teeth 300 and the root portion between the two teeth 300 are simultaneously formed by the grinding wheel, which matches the final finished tooth profile of the space between the teeth 300.

[0042] In contrast to the standard full-fillet root profile, an improved gear root profile (root shape) is disclosed in U.S. Patent No. 9,057,372, granted to Hamilton Sundstrand Co. (Windsor Locks, Connecticut, USA). This improved gear root profile provides a larger fluid volume to reduce fluid displacement compared to the standard full-fillet root profile. Ostensibly, the increased volumetric flow of the improved gear root profile helps reduce cavitation within the gear meshing area. However, such an approach's effectiveness is limited by the reduced volumetric efficiency, as other factors, such as overall pump leakage, may be affected by the profile change.

[0043] The following discussion will be directed to the shape of the gear teeth 300.

[0044] [Root pocket] As noted above, various attempts have been made to address the cavitation problem in gear pumps such as gear pump 10. Despite these attempts, a need exists for an improved gear pump 10 that addresses the cavitation problem. As shown in FIGS. 8, 9, and 10, this need is met by the present invention. According to the present invention, this need is met by increasing the gear root volume, thereby providing additional compliance for trapped fluid. The present invention incorporates additional features into gear pump 10 that increase the overall fluid volume within the gear meshing area over standard or conventional gear profiles. This additional feature addresses issues associated with mesh venting, which can typically result in cavitation damage and gear journal failure, thereby reducing the effects of cavitation and harmful pressure spikes.

[0045] As shown in FIG. 8, the additional volume is achieved by modifying a gear 500 with multiple teeth 502. Each tooth 502 has a contacting side 503a, a non-contacting side 503b, a tip 504, and a root 506, and extends radially outward from a body 512. A spline 508 is formed in the center of the body 512. The gear 500 is modified by machining pockets (cavities) 510 directly into the roots 506 of the teeth 502 of the gear 500. As shown, material is removed only from the non-contacting side 503b of the teeth 502 to form the pockets 510. While the example of FIG. 8 shows 18 teeth 502 and pockets 510, one skilled in the art will recognize that the number of teeth 502 and corresponding pockets 510 can be increased or decreased depending on the application. The gear 500 modified with the pocket 510 may be incorporated into the gear pump 10, the gear pump assembly 100, or other applications.

[0046] The pockets 510 are shaped to maximize the volume of liquid (e.g., fuel) occupied by each pocket 510 while taking into account structural considerations (e.g., strength requirements) of the gear 500. Thus, a design tradeoff exists. The pockets 510 must be sized and shaped to maximize their volume while minimizing adverse effects on the integrity of the gear 500. As shown in FIG. 9, a preferred shape for the pockets 510 is elliptical. An ellipse is defined by a closed, always concave curve, i.e., a closed curve that bounds a convex area. Common objects, such as footballs and eggs, have elliptical cross sections. While other shapes are possible for the pockets 510 depending on the application of the gear 500, an elliptical shape is preferred for at least one particular application.

[0047] The application includes a gear 500 configured for use in a gear pump 10. The exemplary gear 500 has 18 teeth 502, a pressure angle of approximately 30°, a root diameter of approximately 2.925 inches, a true involute form diameter (TIF diameter) of approximately 3.059 inches, a circular tooth thickness of approximately 0.280 inches, and an outer diameter of approximately 3.700 inches. In this application, the pocket 510 has a depth D of approximately 0.350 (±0.005) inches, a length L of approximately 0.450 (±0.001) inches, a width W of approximately 0.179 (±0.001) inches, and a radius of curvature R of approximately 0.060 inches. The radius of curvature helps to avoid stress concentrations.

[0048] As shown, the pocket 510 does not extend the entire length of the gear 500. That is, the gear 500 has a first length, and the root pocket (hereinafter simply referred to as "pocket") 510 has a second length, which is shorter than the first length. In the gear 500 for this application, the length between the teeth 502 in which the pocket 510 is formed is approximately 0.700 inches. Therefore, the length of the pocket 510 (approximately 0.450 inches) is approximately 65% ​​of the length of the gear 500. Preferably, the length of the pocket 510 is approximately 50% to approximately 80% of the length of the gear 500. More preferably, the length of the pocket 510 is approximately 55% to approximately 75% of the length of the gear 500. Even more preferably, the length of the pocket 510 is approximately 60% to approximately 70% of the length of the gear 500.

[0049] The purpose of the plunge cut in the gear root to create pocket 510 is to increase the overall volume of the gear meshing area while using the existing gear tooth profile. The intent is to position pocket 510 to avoid contact with the involute-shaped working portion of gear 500, thereby avoiding modification of the normal gear tooth profile. FIG. 10 shows a portion of a gear pump 520 including gear 500 with pocket 510. Gear 500 is positioned between fixed bearing 517 (located on fixed gear journal 517a) and movable bearing 519 (located on floating gear journal 519a). Because pocket 510 does not extend the entire length of gear 500, the gear tooth profile remains the same at the points where gear 500 contacts bearings 517 and 519, eliminating the need for modifications to bearings 517 and 519. This allows the flanks of the bearings 517, 519 and the gear tooth profile to remain unaffected by the modifications to the gear 500 (i.e., the incorporation of pocket 510). By cutting directly into the root area without affecting the gear tooth profile, a significant amount of volume can be added to the trapped fluid. This can provide a much more effective volume increase than previous attempts to address the cavitation problem in gear pumps, particularly for larger gear pumps 10, and can minimize the effects of changes in trapped volume without the deleterious effects of tooth profile modifications.

[0050] The design tradeoffs represented by the size and shape of the pocket 510 disclosed herein are exemplary only. Those skilled in the art will recognize that other configurations, including multiple pockets, may achieve optimal design tradeoffs for different applications. FIG. 11 illustrates an alternative shape for the pocket 510, including a beam 530 that crosses the center of one or more pockets 510. The beam 530 may extend partially or entirely through the depth of the pocket 510. If the beam 530 extends entirely, it maximizes the strength of the gear 500 while essentially dividing the pocket 510 into two separate pockets. The beam 530 may be added for structural considerations, thereby adding strength to the gear at the expense of a small amount of bulk.

[0051] [Comparative test] A significant amount of testing was conducted to identify and characterize the cavitation problem addressed herein, to apply prior approaches that have failed in their attempts to solve the problem, and to achieve and confirm success with a gear 500 modified to machine pockets (cavities) 510 directly into the roots 506 of the teeth 502. Initial testing was conducted on a MFP-590 main fuel pump configuration. This pump has a conventional bearing configuration 550 with bearing channels 552, 554, and 556 (vent cuts, flow sections) in the driven bearing 560 and the drive bearing 562, similar to the aforementioned channels disclosed in U.S. Pat. No. 9,932,980. The MFP-590 fuel pump is relatively large, has a relatively high gear tip speed, and operates at relatively high pressures.

[0052] FIG. 12 shows a conventional bearing configuration 550. The conventional bearing configuration 550 has an inlet 564 and a discharge (outlet) 566. Initial testing showed severe cavitation erosion on both the bearing bridge (breaking through the bearing in localized areas) and the bearing journal after approximately 8 hours of operation. The conventional bearing configuration 550 is claimed to operate for approximately 5,400 hours or more without overhaul or inspection. Thus, initial testing raised concerns that cavitation erosion could limit the life of the components and lead to premature failure.

[0053] FIG. 13 illustrates an analysis of trapped volume during gear meshing in the conventional bearing configuration 550 shown in FIG. 12, which helps explain why the conventional bearing configuration 550 experienced severe cavitation erosion. As the gears rotate in the MFP-590 fuel pump, a small volume of fluid is forced from the discharge port 566 through the meshing and back to the inlet port 564. At certain points during rotation, a small volume of fluid (called trapped fluid 570) becomes trapped within the gear mesh between first seal point 572 and second seal point 574. The trapped fluid 570 is under high pressure. As the gears change position, the volume of the trapped fluid 570 also changes, and the high-pressure trapped fluid 570 attempts to escape as a high-velocity jet that forces its way through seal points 572 and 574. Because the trapped fluid 570 is essentially incompressible, these rapid volume changes can cause various pressure oscillations between the two gears. These are often high-energy vibrations that lead to cavitation and subsequent damage to surrounding hardware. High-pressure vibrations can also lead to unwanted gear movement, further exacerbating the problem. This high-energy phenomenon typically manifests as physical damage due to cavitation erosion issues within journal bearings or in the surfaces adjacent to the gear tooth profile.

[0054] Conventional gear pump design practices avoid removing material from the gears. Therefore, a conventional design approach to increasing volume is to remove material from the bearings. In line with this approach, to address cavitation issues reflected in initial testing of the MFP-590 fuel pump with conventional bearing configuration 550, more aggressive bearing channels (vent cuts, flow sections) were machined into driven bearing 560 and drive bearing 562. FIG. 14 illustrates an advanced bearing configuration 580. Advanced bearing configuration 580 includes an inlet 564 and a discharge port (outlet) 566. As shown in FIG. 14, advanced bearing configuration 580 additionally includes a notch 582, a channel 584, and a discharge vent 586. Additionally, inlet dam material was removed from region 588.

[0055] FIG. 15 shows a trapped volume analysis of the advanced bearing configuration 580 shown in FIG. 14 during gear meshing, which helps explain why the advanced bearing configuration 580 was insufficient to avoid cavitation damage. Fuel flows at approximately 6.25 gallons per minute. In the gear mesh position shown in FIG. 15, fuel in "Bucket B" (under high pressure) is compressed (squeezed) into "Bucket A" (under low pressure) through the narrow gap between the non-contacting gear sides, the horizontal slot, and the orifices to the discharge and inlet ports. The fuel travels in the direction of arrow "C." This results in high compression, pressure spikes, and cavitation. Even with aggressive vent cuts, sufficient mesh venting (flow in the gear mesh area) was not achieved without causing excessive leakage across the bearing dam.

[0056] Thus, previous design approaches to reducing cavitation damage showed only marginal improvement. This result was likely due to the unique characteristics of the MFP-590 fuel pump configuration (relatively large size, relatively high gear tip speed, and relatively high pressure operation). The need for a new approach was recognized. An analysis of the confined fluid 570 was completed to help identify alternative approaches to reducing cavitation effects and harmful pressure spikes.

[0057] FIG. 16 is a schematic diagram illustrating a trapped volume analysis for a specific fluid fuel passing through the advanced bearing configuration 580 shown in FIG. 14. The analysis was performed for a fluid fuel having the following properties: specific gravity 0.818, discharge coefficient 0.75, and bulk modulus 150,000 PSIG. "PSIG" refers to gauge pressure, expressed in pounds per square inch gauge. "PSIG" is a unit of pressure relative to ambient or atmospheric pressure, measuring pressure without consideration of local atmospheric pressure. Gauge pressure applies when the pressure within the gear mechanism is greater than atmospheric pressure.

[0058] In the gear mesh position shown in FIG. 16, fuel in "Bucket B" (under high pressure) is compressed (squeezed) into "Bucket A" (under low pressure). The fuel moves from a high pressure region 592, which is relatively high at approximately 2200 PSIG, to a low pressure region 594, which is relatively low at approximately 200 PSIG (reduced by a factor of 10). FIG. 16 shows the gear teeth positioned on two bearings. For clarity, the bearings that would be visible on the gear teeth have been omitted. The line of action 324 is shown.

[0059] FIG. 17 is a schematic diagram illustrating a trapped volume analysis for a particular fluid fuel passing through the conventional bearing configuration shown in FIG. 12. In the gear mesh position shown in FIG. 17, fuel in "Bucket B" (under high pressure) is compressed (squeezed) into "Bucket A" (under low pressure) through connection 596. Connection 596 provides a small, narrow opening between high pressure region 592 and low pressure region 594. Fuel travels from high pressure region 592 to low pressure region 594. FIG. 17 shows gear teeth positioned on two bearings. For clarity, the bearings that would be visible on the gear teeth have been omitted. Line of action 324 is shown.

[0060] As described above, the modified configuration, which adds pockets 510 directly to the roots 506 of gear teeth 502, reduces cavitation effects and harmful pressure spikes by increasing the gear root volume (adding trapped fluid cushioning). This modified configuration significantly reduces the compression of "Bucket B" shown in FIGS. 16 and 17 by adding compressible fluid fuel cushioning (compliance). Test results have shown a significant reduction in damaging erosion. Specifically, testing of an MFP-590 main fuel pump configuration including gear 500 with pockets 510 showed minimal to no cavitation erosion after approximately 50 hours of operation. The amount of cavitation erosion that exists is typical of similar configurations successfully deployed and is not life-limiting.

[0061] Additional analytical testing was performed to quantify the improvement achieved by the modified configuration with the addition of pocket 510. An analysis of advanced bearing configuration 580 was performed to provide a baseline. FIGS. 18A and 18B reflect the trapped volume calculated for advanced bearing configuration 580. Considering 360° of gear rotation and 18 gear teeth, a complete passage through the mesh occurs over a 20° (=360 / 18) range. The analysis provided calculations in 1° rotational increments over the 20° range, identifying the lowest volume point within the mesh at 10°-11°.

[0062] Thus, Figure 18A reflects the results calculated for an advanced bearing configuration 580 with a gear rotation angle of 10°, and Figure 18B reflects the results calculated for an 11° gear rotation angle. Figures 18A and 18B respectively show the driven gear 561, the driving gear 563, the line of action 324, the pitch circle 308, and the five trapped fluid volume areas (engagement areas 1, 2, 3, 4, and 5). The following table shows the areas of the trapped areas (engagement areas) for the two rotation angles: [Table 1]

[0063] The width of the driven gear 561 and the drive gear 563 is approximately 0.631 inches. Therefore, the confined volume of the meshing area 1 at a rotation angle of 10° is 0.007816 in 2 ×0.631in=0.004932in 3 This confined volume is the initial volume (Vi). The confined volume of meshing area 1 at a rotation angle of 11° is 0.007595 in 2 ×0.631in=0.004792in 3 This confined volume is the final volume (Vf). The change in volume (dV = Vi - Vf) is 0.004932 in 3 to 0.004792in 3 Subtracting this gives us 0.00014in 3 The pressure difference (dP) between the two regions can then be calculated at the lowest volume point within the interlocking region. Assuming the bulk modulus of the fluid (B) is constant at 150,000 psi, dP = B(dV / V) gives dP = 150,000 psi × (0.00014 in 3 / 0.004932in 3 ) = 4258 PSID ("PSID" stands for "pounds per square inch difference" and is used to measure pressure relative to something other than atmospheric pressure).

[0064] An analysis of the improvement achieved by the modified configuration with the addition of pocket 510 was performed in comparison to the baseline of advanced bearing configuration 580. Pocket 510 was 0.0295 in 3 Therefore, the confined volume of meshing area 1 at a rotation angle of 10° is 0.004932 in 3 +0.0295in 3 =0.03443in 3 The confinement volume of meshing area 1 at a rotation angle of 11° is 0.004792 in 3 +0.0295in 3 =0.03429in 3 The change in volume (dV = Vi - Vf) is 0.03443 in 3 to 0.03429in 3Subtracting this gives us 0.00014in 3 The pressure difference (dP) between the two regions can then be calculated at the lowest volume point within the interlocking region. Assuming the bulk modulus of the fluid (B) is constant at 150,000 psi, dP = B(dV / V) gives dP = 150,000 psi × (0.00014 in 3 / 0.03443in 3 ) = 610 PSID. Thus, the addition of pocket 510 reduced the compression of the trapped volume of the gear meshing area by approximately a factor of seven (610 / 4258).

[0065] The volumetric gain of the improved configuration with the addition of pocket 510 can also be calculated compared to the baseline value of advanced bearing configuration 580. The trapped volume of meshing area 1 at a rotation angle of 10° is 0.004932 in for advanced bearing configuration 580. 3 and 0.03443 in for the improved configuration with pocket 510 added. 3 Therefore, with this improved configuration, the trapped volume has increased by approximately 600% ((0.03443-0.0049320) / 0.004932 x 100 = 598%). The trapped volume of meshing area 1 at an 11° rotation angle is 0.004792 in for advanced bearing configuration 580. 3 and 0.03429 in for the improved configuration with pocket 510 added. 3 Therefore, with this improved configuration, the confinement volume was still increased by approximately 600% ((0.03429-0.004792) / 0.004792 x 100 = 616%).

[0066] Because the volume change is fixed, the negative effects of the change can be minimized by increasing the overall volume, which makes the fixed compression a smaller percentage of the overall volume. The additional refinement of pocket 510 increases the overall volume. This refinement therefore reduces pressure spikes, which in turn reduces cavitation throughout the fluid.

[0067] The improvement achieved by the modified configuration with the addition of pocket 510 compared to the baseline of advanced bearing configuration 580 can also be measured by the effect of the improvement on both pressure ripple and gear stress. Pressure ripple can be calculated as the pressure fluctuation, or the amplitude of the pressure deflection, either upward (positive) or downward (negative), divided by the total pressure in the gearing, multiplied by 100. So, for example, if the total pressure is 200 psi and the pressure fluctuates by 10 psi, the pressure ripple will be 5%.

[0068] Pressure testing was completed for the advanced bearing configuration 580 and the modified configuration with the addition of pocket 510. The gear speed for the pressure test was 3000 rpm, the discharge pressure was 462 psi, and the gear discharge pressure ripple target was approximately 5% maximum. FIG. 19A is a graph showing the measured pressure versus time for an MFP-590 pump equipped with only the advanced bearing configuration 580, showing a discharge pressure ripple of approximately ±30 psi, or 6.5%. FIG. 19B is a graph showing the measured pressure versus time for an MFP-590 pump equipped with the advanced bearing configuration 580 and pocket 510, showing a discharge pressure ripple of approximately ±5 psi, or 1%. The addition of pocket 510 significantly reduced the discharge pressure ripple. Thus, the addition of pocket 510 transformed the tested pump from one that failed to meet its performance target to one that met its discharge pressure ripple target.

[0069] Finally, a finite element analysis (FEA) of the gear stress for the gear with pocket 510 was completed, and the factor of safety (FoS) for the gear was calculated. When designing a product, engineers strive to achieve the required factor of safety (FoS). This requirement helps engineers provide additional confidence that the part will not fail even if overloaded. The factor of safety (FoS) can be calculated in a variety of ways. However, ultimately, each calculation checks the amount of safe load beyond the designed working load. Therefore, FoS = (actual load) / (working load). If the part is unsafe, there is a significant risk of part failure. When the factor of safety (FoS) is 1, the design load is equal to the safe load. Therefore, for a safer design, the factor of safety (FoS) should always be greater than 1. If the factor of safety (FoS) is less than 1, the risk of failure is too high. A typical factor of safety (FoS) for aircraft parts is 1.5 to 2.

[0070] An MFP-590 fuel pump with a gear having a pocket 510 was analyzed. The maximum gear operating condition was selected for the test because it represents the condition where the gear experiences the highest torque. In the analyzed application, the maximum gear operating condition is defined by a torque applied to the gear teeth of approximately 500 in-lbf, a speed of approximately 6618 rpm, and a pressure of approximately 1600 psid. Jet A, one of the most commonly used fuels in commercial aircraft, constituted the fluid. Consistent with a conservative approach, a single tooth tip contact was assumed in the analysis. However, in reality, the load is distributed across multiple gear teeth.

[0071] The gears are made of high-strength, wear-resistant CPM-10V steel. The gear teeth were subjected to a nitriding surface treatment to further enhance wear resistance. Nitriding is a thermochemical treatment process that enriches the surface with nitrogen to increase surface hardness. This process relies on the low solubility of nitrogen in the ferrite crystal structure and promotes the precipitation of iron nitride or alloy nitrides. The yield strength of the gear material in the transverse direction was approximately 70.4 ksi. The yield stress is the stress limit at which the material begins to deform. The endurance limit (R = -1) of the gear material was approximately 54.0 ksi. The endurance limit is defined as the stress range at which the material exhibits infinite life under repeated stress without crack propagation.

[0072] The maximum principal stress is the maximum normal stress acting on one of the principal planes of a component (such as the gear under analysis) where the shear stress has a value of zero. The maximum principal stress was calculated to be 17.5 ksi. Therefore, the factor of safety (FoS) for the gear against high-cycle fatigue was calculated to be 54 ksi / 17.5 ksi = 3.1. High-cycle fatigue is a type of fatigue caused by small elastic strains under a large number of load cycles before failure occurs. The stress results from a combination of mean and alternating stresses. Therefore, a finite element analysis (FEA) performed on the gear with pocket 510 showed not only an acceptable factor of safety (FoS), but also a relatively high factor of safety (FoS).

[0073] The trapped fluid within the pump's gear mesh expands and contracts due to the movement of the rotating gears. This rapid expansion and contraction creates very large localized pressure fluctuations. These high-energy cycles can cause cavitation damage to surrounding hardware, such as bearings and gears. The pressure cycles can also create gear imbalances, generating large forces that can cause undesired gear movement leading to cavitation damage and potential journal failure. Gears modified to include pockets 510 reduce the pressure levels inherent in the gear mesh, thereby reducing cavitation and its adverse effects.

[0074] While the foregoing description has been made with reference to specific embodiments and examples, the present invention is not limited to the details set forth above. Rather, various changes in details may be made within the scope and equivalents of the claims without departing from the spirit of the present invention. Uses and applications of the improved gears disclosed herein include, for example, any application where a change in confined volume may occur.

Claims

1. A gear having a gear tooth profile, The main body and a plurality of gear teeth having an involute tooth profile extending radially outward from the body, the plurality of gear teeth including adjacent first and second gear teeth each having a tip and a root, with a space formed between the adjacent first and second gear teeth; a tooth root pocket formed directly at the tooth root of the gear tooth in the space, the tooth root pocket increasing the gear tooth root volume without changing the gear tooth profile, thereby adding trapped fluid cushioning; Gear with.

2. The root pocket is oval. The gear of claim 1 .

3. the root pocket has a depth of about 0.350 inches, a length of about 0.450 inches, a width of about 0.179 inches, and a radius of curvature of about 0.060 inches; 3. The gear of claim 2.

4. the gear has a first length; the root pocket has a second length that is shorter than the first length; The gear of claim 1 .

5. The second length is between about 50% and about 80% of the first length.

5. The gear of claim 4.

6. The second length is about 65% of the first length.

6. The gear of claim 5.

7. the root pocket has a center; the gear further comprises a beam portion that intersects the center of the root pocket; The gear of claim 1 .

8. The root pocket has a depth, the beam portion extends to penetrate only a portion of the depth of the root pocket; 8. The gear of claim 7.

9. The gear has strength, The root pocket has a depth, the beam portion extends to penetrate the entire depth of the root pocket, dividing the root pocket into two separate sections while maximizing the strength of the gear; 8. The gear of claim 7.

10. a first gear having (i) a gear tooth profile; (ii) a first body; (iii) a plurality of gear teeth of a first involute tooth profile extending radially outward from the first body, the plurality of gear teeth including adjacent first and second gear teeth each having a tip and a root, wherein a space is formed between the adjacent first and second gear teeth; and (iv) a root pocket formed directly at the root of the first gear tooth in the space, the root pocket increasing a gear root volume without changing the gear tooth profile, thereby adding trapped fluid cushioning; a second gear configured to mesh with the first gear at a meshing region that defines an increased gear meshing volume by the root pocket, the second gear having a second body and a plurality of gear teeth having a second involute tooth profile extending radially outward from the second body; a first gear bearing and a second gear bearing configured to position the first gear and the second gear along a bearing centerline, wherein the root pocket does not extend into either the first gear bearing or the second gear bearing; a housing having a fluid inlet and a fluid outlet, the housing having the first gear, the second gear, the first gear bearing, and the second gear bearing disposed therein; A gear pump comprising:

11. The root pocket is oval. The gear pump according to claim 10.

12. the root pocket has a depth of about 0.350 inches, a length of about 0.450 inches, a width of about 0.179 inches, and a radius of curvature of about 0.060 inches; The gear pump according to claim 11.

13. the first gear has a first length; the root pocket has a second length that is shorter than the first length; The gear pump according to claim 10.

14. The second length is between about 50% and about 80% of the first length.

14. The gear pump according to claim 13.

15. The second length is about 65% of the first length.

15. The gear pump according to claim 14.

16. the root pocket has a center; the first gear further includes a beam portion that intersects the center of the root pocket; The gear pump according to claim 10.

17. the first gear and the second gear define a compression gear meshing area; the root pocket is configured to reduce compression of the gear meshing area by approximately seven times; The gear pump according to claim 10.

18. the first gear and the second gear define a confined volume; The root pocket is configured to increase the trapped volume by approximately 600%. The gear pump according to claim 10.

19. the gear pump has a discharge pressure ripple of about 1%; The gear pump according to claim 10.

20. The gear pump has a safety factor against high cycle fatigue of greater than 3. The gear pump according to claim 10.

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