Improved double helical gear for hydraulic gear systems, having a variable helix angle and a tooth profile that does not cause locking.

Abstract

A double helical gear (1) for a hydraulic gear device (2) having a tooth profile (4) that does not cause locking, configured to be coupled to a support shaft (5) to form a drive gear or driven gear of the hydraulic device, comprising a plurality of teeth (6) extending longitudinally or axially with a variable helix angle in a compound function, wherein the tooth profile (4) maintains shape continuity in each cross section, and each tooth (6) in the longitudinal direction is divided into five zones: a starting zone (A), a proximal intermediate zone (B), a central zone (C), a distal intermediate zone (D), and a terminal zone (E), wherein the starting zone (A), central zone (C), and terminal zone (E) have a variable helix angle (ψ), and the proximal intermediate zone (B) and distal intermediate zone (D) have a constant helix angle (ψ).

Description

[Technical Field]

[0001] The present invention relates to a double helical gear having a non-entanglement tooth profile, which meshes within a hydraulic gear system.

[0002] More specifically, the present invention relates to a gear intended to mesh with similar gears in a hydraulic gear system without locking.

[0003] A typical example of a hydraulic gear system to which the gears of the present invention are best applied, and which is specifically referenced below herein, is a rotary positive displacement gear pump. However, the gears of the present invention can also be applied to all hydraulic systems that operate via a hydraulic gear motor and / or a pair of gears, and these are therefore intended to be within the scope of the present invention. [Background technology]

[0004] As is well known in the art, rotary positive displacement gear pumps generally have two gears. These gears are almost always straight-toothed, with one, called the drive gear, coupled to the control shaft and rotationally driving the other gear, called the driven gear.

[0005] In gears with straight teeth, each pair of teeth meshes and disengages simultaneously across the entire axial width of the tooth surface. This type of engagement mechanically generates vibration and noise due to fluctuations in the load on the teeth and the impacts caused by approaching and separating.

[0006] Another particularly significant drawback of the conventional type of gear pumps described above is that the fluid being pumped is either trapped, i.e., trapped and compressed within the sealed space between the teeth in the meshing region, or in any case subjected to volume changes, thereby resulting in harmful and uncontrolled localized stress peaks that cause direct hydraulic operating noise.

[0007] A known technical solution to avoid direct mechanical operating noise is to employ gears with helical teeth. The teeth of these helical gears are oriented along a cylindrical spiral, rather than parallel to the gear axis.

[0008] In gears with helical teeth, the inclination of the teeth causes each pair of teeth to gradually mesh and disengage, resulting in quieter and more regular power transmission.

[0009] These gears offer advantages in many respects and substantially satisfy the objective of reducing operating noise, but their unique structure gives rise to other problems. In fact, the inclination of the teeth divides the transmitted force into a tangential component necessary for transmitting torsional moment and an axial component that tends to displace the gear rather.

[0010] To avoid this problem, thrust bearings or two oppositely oriented helices with complementary angles are used, thereby suppressing the induced axial thrust.

[0011] The present invention aims to avoid the use of thrust bearings or any other devices to compensate for internally generated axial forces, and instead follows the trend of counter-direction helices, further intending to suppress hydraulic noise caused by fluid confinement.

[0012] Figure 1, attached to this specification, shows a known example of a gear having opposite helices, commonly referred to as a herringbone gear, where the two opposite helices are connected at a cusp.

[0013] The herringbone gear shown in Figure 1 is used as a rotor for hydraulic pumps in low-speed, high-power applications.

[0014] Despite being used for many years, this type of tooth profile creation is limited in accuracy and hardness due to manufacturing difficulties caused by machining at the tip.

[0015] In fact, the machine used to manufacture this type of gear is a vertical cutting machine, in which two opposite helices are machined simultaneously by the reciprocating motion of blades that interfere with each other at their tips.

[0016] The limitation of this process is that the machining near the cusp is extremely delicate and complex, making it impossible to manufacture large, high-hardness gears because it is usually not possible to obtain gear sets with materials that have a hardness higher than 35 on the Rockwell C scale.

[0017] To improve hardness properties, these gears can be treated, for example, with thermal nitriding after machining the teeth. However, tooth distortion after heat treatment forces designers to use wider tolerances to avoid damage to the tooth surface, resulting in reduced efficiency.

[0018] An alternative solution is shown in Figure 2, in which a gap is provided between the two helices, allowing for the use of various machine tools in the manufacture of the gear system and enabling optimal precision even at high hardness levels, such as above 58-60 on the Rockwell C scale. However, these gear systems cannot be used for pressurized applications.

[0019] For example, other alternative solutions have been proposed for manufacturing pumps equipped with high-hardness double helical gears, as disclosed in Patent Documents 1 and 2, which in particular relate to the pumping of molten plastic materials, i.e., operating at low speeds and therefore using involute tooth profiles without considering noise due to fluid entrapment, which is negligible in these applications.

[0020] To at least partially solve the above problems, in order to provide a double helical gear having a tooth profile that does not cause jamming and can be easily machined even when reaching high hardness, the above Patent Document 3 by the applicant of the present case proposes a tooth profile of these gears in which the twist angle is variable along its development. In a preferred modification shown in the attached FIG. 3, its development is divided into three zones. In the first zone of the helical development, the twist angle is constant, in the second zone the angle is variable, and in the third zone the angle is constant again.

[0021] Although the pump manufactured using the gears specified in the application recalled above substantially meets the on-site requirements, nevertheless, it has some important problems that have remained unsolved until today.

[0022] In particular, in this new type of known pump, sudden breakage of vulnerable parts or edges with typical chipping shown in FIG. 28b occurs at the ends of the gears under certain operating conditions. This phenomenon observed by the applicant in the machine manufactured according to the teachings of the above Patent Document 3 may limit the use of the device in some applications.

[0023] Furthermore, due to the discontinuity of the function of the fluid volume displaced over time due to the shape of the tooth profile, the reduction of vibration still does not seem to be optimal.

[0024] Therefore, the technical problem underlying the present invention is to prove that it is easy to manufacture by a substantially conventional numerical control machine, and while suppressing the presence of vulnerable parts or edges at the ends of the rotor itself that cause restrictions on the use of the pump, it is possible to simultaneously suppress mechanical and hydraulic operating noise and avoid the generation of axial thrust that requires any force compensation. It is to devise a new type of double helical gear for a hydraulic gear device having structural and functional characteristics.

[0025] Another object of the present invention is to manufacture gears for positive displacement pumps and other types of hydraulic devices that are completely free from jamming.

Prior Art Documents

[0026] [Patent Document 1] U.S. Patent Application Publication No. 2004 / 0031152 [Patent Document 2] U.S. Patent No. 7040870 [Patent Document 3] European Patent Application Publication No. 17181600.2 [Overview of the Initiative]

[0027] The underlying solution of the present invention is to obtain a double helical gear for a hydraulic gear system, which is coupled to a support shaft to form a drive gear or driven gear for the hydraulic system, and comprises a plurality of teeth, the plurality of teeth extending in the longitudinal or axial direction of the teeth with a variable helix angle in a continuous function, each tooth having a central zone with a variable helix angle that causes a helix transition from right-handed to left-handed, the tooth profile maintaining shape continuity in each cross section, and each tooth having at least a start zone and an end zone with variable helix angles at two opposing side ends of the gear, in which the helix angle decreases as it approaches the side ends of the gear.

[0028] A continuous function is a function that has no discontinuities, and this definition also includes, for example, a compound function composed of multiple circular arcs that can be connected by straight lines.

[0029] Preferably, the continuous function has neither corners nor peaks, and particularly preferably, the central zone has a curved transition from right-handed to left-handed helix, with the transition point having a helix angle of zero.

[0030] In a preferred modification of the present invention, each tooth further comprises a proximal intermediate zone and a distal intermediate zone having a constant twist angle, connecting the starting zone and the ending zone to the central zone.

[0031] According to a possible alternative, the start zone, end zone, and middle zone, each having a variable helix angle, are directly connected by an inflection point, resulting in a variable helix angle along the entire tooth profile. In this case, the helix angle can be unfolded, for example, according to a cosine curve.

[0032] Preferably, the helix angle is zero at the opposing side ends of the gear; that is, the tooth profile of each individual tooth joins perpendicularly to the opposing side surfaces of the gear.

[0033] The means proposed above ensure that the axial component of the force exchanged by the gear when it meshes with another identical gear during use is zero at the side end, thereby minimizing the risk of edge chipping even under unfavorable operating conditions.

[0034] Preferably, the tooth profile is a mirror image of the central plane passing through the transition point between the right-handed and left-handed helical curves.

[0035] Preferably, the tooth profile is parameterized and sized such that the torsional contact ratio parameter is between 0.6 and 1, preferably between 0.6 and 0.8, and more preferably equal to 0.65.

[0036] Preferably, the gear tooth profile is one that does not cause locking.

[0037] A tooth profile that does not result in this trapping can be defined by the apex and base of two arcs or ellipses connected by an involute waveform contained between two part-off diameters: a lower truncation diameter where the transition between the base and the involute tooth profile occurs, and an upper truncation diameter where the transition between the involute tooth profile and the apex occurs.

[0038] Preferably, the lower involute cutoff diameter is JPEG2026521096000002.jpg1179 They are selected as equal to, where, JPEG2026521096000003.jpg1114 is the pitch circle diameter, and the parameter p1 is between 9.7% and 9.9%, preferably equal to 9.8%.

[0039] Preferably, the upper involute cutoff diameter is JPEG2026521096000004.jpg1277 They are selected as equal to, where, JPEG2026521096000005.jpg1114 is the pitch circle diameter, and the parameter p2 is between 12.1% and 12.3%, preferably equal to 12.2%.

[0040] Preferably, the top of each tooth has a cutting edge that is primarily intended to move within the pump body, defined by a limited thickness that protrudes relative to the tooth profile.

[0041] The technical problems identified above are also solved by a hydraulic gear system comprising a pair of gears according to the above proposal.

[0042] In particular, the device may be a positive displacement pump or a hydraulic gear motor.

[0043] This technical challenge can also be solved by a method for manufacturing double helical gears with a tooth profile that does not cause locking for hydraulic gear systems using an automated numerically controlled machine operated by appropriate software. The process involves determining the axial rotation equations for all tooth profile cross-sections calculated on the pitch circle diameter to obtain a series of coordinates representing the helical path, A process of manufacturing a solid model by sliding the front tooth profile of the rotor along a helical path using 3D software, The process of transferring a solid model to CAD-CAM, This includes a process of finishing the interdental spaces on the processing station.

[0044] If desired, this method may include a quenching treatment.

[0045] Further features and advantages of the present invention will become more apparent from the description of its embodiments, which is given below with reference to the accompanying drawings, which are given as non-limiting examples. [Brief explanation of the drawing]

[0046] [Figure 1] A schematic perspective view of a herringbone gear manufactured according to prior art is shown. [Figure 2] A schematic perspective view of a double helical gear having spaced-apart helices, manufactured according to prior art, is shown. [Figure 3] A perspective view of a double helical gear with a variable-angle spiral, manufactured according to prior art, is shown. [Figure 4] For example, a plan view of a pair of double helical gears according to the present invention, connected to each other within a hydraulic gear system that is a positive displacement pump, is shown. [Figure 5] A perspective view of a double helical gear having teeth unfolded along a variable-angle spiral, manufactured according to the present invention, is shown. [Figure 6] A schematic side view of the cross-section of the gear according to the present invention is shown. [Figure 7] This shows the geometric structure of the involute segments of the gear tooth profile. [Figure 8a] Figure 6 shows a magnified detail of F. [Figure 8b] A schematic side view of two gears according to the present invention that mesh with each other is shown, with the pressure lines identified. [Figure 8c] Figure 8a shows a magnified detail of G. [Figure 8d] This shows a schematic geometric diagram of the contact angle of involute tooth profiles in a gear device composed of gears according to the present invention. [Figure 9a] This shows a schematic front view of a cross-section of two adjacent teeth in occlusion. [Figure 9b] Figure 9a shows a schematic side view of the gear mechanism. [Figure 10a] The following are geometric schematic diagrams related to a typical spiral. [Figure 10b] The following are geometric schematic diagrams related to a typical spiral. [Figure 11a] This shows the development of one turn of the helix in a known herringbone gear. [Figure 11b] This shows the development of one turn of the helix in a known herringbone gear. [Figure 12a] The schematic diagrams of the geometric shapes of the helical development in the gear according to the present invention are shown. [Figure 12b] The schematic diagrams of the geometric shapes of the helical development in the gear according to the present invention are shown. [Figure 12c] The schematic diagrams of the geometric shapes of the helical development in the gear according to the present invention are shown. [Figure 13] A geometric schematic diagram of the helical development in a gear according to the present invention is shown. [Figure 14] A geometric schematic diagram of the helical development in a gear according to the present invention is shown. [Figure 15a] The geometrical diagrams shown illustrate the kinematic analysis of the contact points between the helices of a gear on the pitch circle diameter as the rotational speed changes. [Figure 15b] The geometrical diagrams shown illustrate the kinematic analysis of the contact points between the helices of a gear on the pitch circle diameter as the rotational speed changes. [Figure 16a] The geometric schematic diagrams for calculating the equations of motion of the contact points between two axial helices in three different zones are shown. [Figure 16b] The geometric schematic diagrams for calculating the equations of motion of the contact points between two axial helices in three different zones are shown. [Figure 17] The geometric schematic diagrams for calculating the equations of motion of the contact points between two axial helices in three different zones are shown. [Figure 18a] The geometric schematic diagrams for calculating the equations of motion of the contact points between two axial helices in three different zones are shown. [Figure 18b]The geometric schematic diagrams for calculating the equations of motion of the contact points between two axial helices in three different zones are shown. [Figure 18c] The geometric schematic diagrams for calculating the equations of motion of the contact points between two axial helices in three different zones are shown. [Figure 19a] The geometric schematic diagrams for calculating the equations of motion of the contact points between two axial helices in three different zones are shown. [Figure 19b] The geometric schematic diagrams for calculating the equations of motion of the contact points between two axial helices in three different zones are shown. [Figure 19c] The geometric schematic diagrams for calculating the equations of motion of the contact points between two axial helices in three different zones are shown. [Figure 20a] This is a schematic side view of a fluid trapped in a chamber formed by the teeth of a gear mechanism according to the present invention. [Figure 20b] This is a schematic side view of a fluid trapped in a chamber formed by the teeth of a gear mechanism according to the present invention. [Figure 21a] The geometric schematic diagrams for analyzing the fluid volume trapped within the chamber formed by the teeth of the gear mechanism are shown. [Figure 21b] The geometric schematic diagrams for analyzing the fluid volume trapped within the chamber formed by the teeth of the gear mechanism are shown. [Figure 21c] The geometric schematic diagrams for analyzing the fluid volume trapped within the chamber formed by the teeth of the gear mechanism are shown. [Figure 22] The geometric schematic diagrams for analyzing the fluid volume trapped within the chamber formed by the teeth of the gear mechanism are shown. [Figure 23] The geometric schematic diagrams for analyzing the fluid volume trapped within the chamber formed by the teeth of the gear mechanism are shown. [Figure 24a]The geometric schematic diagrams for analyzing the fluid volume trapped within the chamber formed by the teeth of the gear mechanism are shown. [Figure 24b] The geometric schematic diagrams for analyzing the fluid volume trapped within the chamber formed by the teeth of the gear mechanism are shown. [Figure 24c] The geometric schematic diagrams for analyzing the fluid volume trapped within the chamber formed by the teeth of the gear mechanism are shown. [Figure 25a] A schematic side view of two gears according to the present invention that mesh with each other is shown, with the pressure lines identified. [Figure 25b] Figure 25a shows an enlarged view of H. [Figure 26] This shows a geometric diagram for analyzing the forces acting on the teeth of a gear mechanism. [Figure 27a] This shows the direction of the force acting along the tooth alignment of the gear in the first gear device according to prior art. [Figure 27b] This shows the direction of the force acting along the tooth alignment of the gear in the second gear mechanism according to prior art. [Figure 27c] This shows the direction of the force acting along the tooth alignment of the gear according to the present invention. [Figure 28a] A schematic perspective view of a gear according to prior art is shown, with the direction of the exchanged force identified. [Figure 28b] Figure 28a shows a typical chipping pattern in the tooth profile of a gear according to prior art. [Figure 29] A schematic perspective view of the gear according to the present invention is shown, with the direction of the exchanged force identified. [Figure 30a] The diagram shows the tooth arrangement in a gear according to prior art, and the tooth angle with respect to a plane that crosses its lateral end is identified. [Figure 30b] The diagram shows the tooth arrangement in a gear according to the present invention, with the tooth angle relative to a plane transverse to its lateral end identified. [Modes for carrying out the invention]

[0047] Referring to Figure 5, a gear of the type having a double helix tooth profile, manufactured according to the present invention, is schematically shown as 1 overall.

[0048] Gear 1 is intended, in particular, but not limited to, for hydraulic gear systems, and the following description refers to this particular application area for the sake of simplicity.

[0049] To better understand all aspects of the present invention, a "cylindrical helix" is defined as a curve drawn by a moving point that performs continuous circular motion and, simultaneously, various motions perpendicular to the plane of rotation.

[0050] Furthermore, "spiral pitch" is defined as the distance the spiral generation point moves axially during one complete rotation.

[0051] Furthermore, a “non-entrapment tooth profile” is defined as a rotor tooth profile that allows the pumped fluid to flow between the meshing teeth of the rotor without being trapped, compressed, or subjected to volume fluctuations as it slides within the pump. To achieve this effect, a non-entrapment tooth profile can perfectly match the corresponding tooth profile without defining a cavity between two meshing tooth profiles. A non-entrapment tooth profile can be manufactured as an arc and created using heads and legs connected to each other by tooth surfaces having circular involutes.

[0052] The present invention aims to manufacture a double helical gear that can be used together with the same type of gear in a gear system for a positive displacement pump using a rotor with a reverse twist. Advantageously, according to the present invention, gear 1 has a tooth profile that does not cause locking and a helical shape that suppresses the central corner point of a conventional herringbone gear manufactured according to the prior art, and suppresses the fragile edge present in a rotor manufactured according to the prior art at the end of the rotor.

[0053] This suppresses problems associated with machining rotors with such tooth profiles using machine tools at their source.

[0054] Figure 4 shows a top view of gear 1, which, when coupled with the corresponding gear 1', defines a double helical type gear 2 that does not cause trapping in a hydraulic device, such as a positive displacement pump.

[0055] Gear 1 is conventionally coupled or fitted to the support shaft 5 to form a drive gear or a driven gear, depending on its assigned role in the hydraulic system.

[0056] In exemplary embodiments described herein as non-limiting examples, the gear 1 has a front tooth profile 4 having seven teeth, but this does not preclude the use of a different number of teeth.

[0057] Advantageously, according to the present invention, the double helix unfolding portion 3 of the gear 1 changes along the axial direction of the teeth in a continuous function and arc pattern while maintaining the shape continuity of its cross-section, and its cross-section coincides with the front tooth profile 4.

[0058] In this invention, a continuous function means a function that has no discontinuities. This function can be an intrinsic function (e.g., a cosine curve) or a compound function (e.g., three circular arcs connected by a linear transition).

[0059] In other words, gear 2 has no sharp points or acute angles in its central zone. Each corresponding tooth 6 is continuous and has no undercuts. Furthermore, each corresponding tooth 6 has no sharp points in its central zone, and this geometric shape is defined in both the central zone and its ends of the gear, with the outer portion ending at a helix angle equal to 0°.

[0060] This unique helical development, described in further detail below, allows for the acquisition of a pair of rotors in which the pitch and helix angle change with mathematical regularity, and in particular, the preferred tooth profile shown in the attached figure ensures transmission continuity with a contact ratio equal to 0.65.

[0061] Essentially, this means that the other two teeth 6 begin to mesh simultaneously before the two teeth 6 separate. The contact is continuous and reversible, moving outward from the center of the rotor or vice versa, depending on the clockwise or counterclockwise rotation and the helical arrangement.

[0062] Furthermore, note that the tooth profile is conjugate along the entire length of the rotor; that is, the tangents to the tooth profile at the contact points coincide, and the common normal passes through the center of instantaneous rotation.

[0063] Referring to Figure 5, the rotor covered by the present invention is shown in detail, which comprises several segments, in particular the following segments, whose helix angle is variable or constant: - Segment A: Having a variable twist angle, - Segment B: Having a constant twist angle, - Segment C: Having a variable twist angle, - Segment D: Having a constant twist angle, - Segment E: Has a variable torsion angle.

[0064] Essentially, the longitudinal arrangement of the rotor teeth can be divided into five zones: the starting zone, the proximal intermediate zone, the central zone, the distal intermediate zone, and the terminal zone. Segment A corresponds to the starting zone, segment B to the proximal intermediate zone, segment C to the central zone, segment D to the distal intermediate zone, and segment E to the terminal zone.

[0065] Dividing the tooth into five segments or zones is not essential for carrying out the present invention. It should be noted that the starting and distal segments can be directly connected to the central segment by an inflection point without a transition section having a constant twist angle. In this case, the tooth profile of the tooth apex unfolded on a flat surface can be a cosine curve, but in a preferred embodiment, the ending and distal segments can follow a circular arc—or other planar curve—and the intermediate segment is linear.

[0066] The lengths of the various rotor segments A, B, C, D, and E are adjusted according to mechanical considerations and change as the rotor band changes, according to the geometric rules detailed below.

[0067] Next, we will describe the tooth profile pattern 4 of the single tooth 6 of gear 2.

[0068] Referring to Figure 6, the tooth profile 4 of the tooth 6 of gear 2 has two truncated diameters connecting two arcs 7 and 9 at the top and bottom of the tooth. JPEG2026521096000006.jpg914 , JPEG2026521096000007.jpg917 It is formed by segments having circular involutes 8 contained within them.

[0069] Referring to Figure 7, the involute segment is characterized by the following equation: JPEG2026521096000008.jpg52130

[0070] Referring to Figure 6 above, the involute parameters used in preferred embodiments of the present invention are shown below.

[0071] The lower involute truncation diameter is selected to be equal to the following: JPEG2026521096000009.jpg1570 However, p1 is between 9.7% and 9.9%, preferably equal to 9.8%.

[0072] The upper involute truncation diameter is selected to be equal to the following: JPEG2026521096000010.jpg1376 However, p2 is between 12.1% and 12.3%, preferably equal to 12.2%.

[0073] In either case, JPEG2026521096000011.jpg1114 This is the pitch circle diameter of gear 1.

[0074] With respect to the diameters of the top and bottom, the following relationship applies, as can be seen from Figures 8a, 8b, and 8c. JPEG2026521096000012.jpg1580 However, p3 is contained between 19.5% and 20.5%, preferably equal to 20%. JPEG2026521096000013.jpg1172 However, p4 is included in the range of 19.5% to 20.5%, preferably equal to 20%.

[0075] Referring particularly to Figure 8a, the top of each tooth 6 has a cutting edge 10 defined by a limited thickness that protrudes from the tooth profile 4, allowing it to move easily within the pump body and to narrow the dimensional tolerance with respect to the outer diameter of the gear.

[0076] Figure 8b shows the meshing between two gears. JPEG2026521096000014.jpg1114 is the pitch circle diameter, Cd is the differential circumference, and Rp represents the pressure line. Detailed G in Figure 8c shows the pressure angle φ, which is the difference between the pressure line and the pitch circle diameter. JPEG2026521096000015.jpg1114 This is the angle between the tangent line and the object.

[0077] Preferably, the parameters of the tooth profile 4 according to the present invention are such that the pressure angle φ is between 28° and 32°, and preferably equal to 30°.

[0078] From the design parameters listed above, and referring particularly to Figure 8d, the contact angle α between the two teeth engaging on the involute segment is approximately equal to 29°.

[0079] Regarding the torsional contact ratio, the tooth profile 4 of the rotor 1 is designed to avoid fluid trapping between the groove bottoms and tops of two meshing teeth, as already mentioned. However, this results in a circumferential contact ratio Rc, defined as the ratio of the involute contact angle to the angle between the circumferential pitch and the surrounding area, being less than 1. In this particular case, it is approximately equal to 0.5.

[0080] This essentially means that teeth that make contact and transmit motion separate before the next tooth meshes, which necessitates the manufacture of gears with that helical shape.

[0081] Referring specifically to Figures 9a and 9b, two adjacent teeth 6 in a cross-section perpendicular to the axis of rotation of the rotor 5 are shown as I and II. The same teeth 6 in a cross-section perpendicular to the axis of rotation at a distance L are shown as I' and II'.

[0082] In order to maintain a constant length of the contact line between the six teeth along the axial direction during the rotation of the rotors, I and II' must each be located at a distance Lf that is 2·π / (number of teeth n°).

[0083] In this case, we can see that the torsional contact ratio Re is equal to 1.

[0084] Since Hertz's contact stress always acts on the same surface, this is an ideal operating condition.

[0085] However, based on the design specifications of the pump, which is designed to operate at a pressure of 250 bar and a speed of 3600 rpm in the preferred case, and the strength calculations of the gears, it is necessary that the helix angle is not greater than 45°, and therefore equal to 0.65, and in any case greater than or equal to 0.6, and in particular a torsional contact ratio Re that falls between 0.6 and 0.8 has been established.

[0086] Here, we will explain the calculation of the coordinates of the helical development in three-dimensional space. This makes it possible to determine the inter-tooth space of the rotor using 3D software.

[0087] Refer to Figures 10a and 10b to review some geometric definitions beforehand.

[0088] A spiral is a curve in three-dimensional space, drawn by lines at a constant angle wound around a cylinder.

[0089] The unfolding of one turn of a spiral consists of straight-line segments corresponding to the hypotenuse of a right triangle, where the pitch P and the circumference π·d of the spiral are the sides forming the right angle. The slope is determined by the angle α contained between the hypotenuse and the sides forming the right angle corresponding to the circumference of the spiral. Therefore, tan(α) = P / (π·d).

[0090] Referring to Figures 11a and 11b, it can be seen that the unfolding of one turn of the helix in the conventional version of the herringbone gear is a straight segment with an inclination α. ​​Considering the half-gear, a torsional contact ratio equal to 1 is obtained.

[0091] Figures 12a and 12b suggest a similar configuration for the helical development proposed by the present invention. Naturally, in this case, the hypotenuse is replaced by a curve with a variable slope α. Figure 12c shows the pattern of this curve in space.

[0092] For simplicity, only gear half with a torsional contact ratio equal to 1 will be considered below, for the sake of simplifying the formulas and explanations. The right triangle on which the helical development is drawn will be used as the basis for the following explanation.

[0093] In the following schematic diagram, in depicting the spiral unfolded triangle, in order to obtain a contact ratio of 1, the variables p and (π·dp) corresponding to the sides enclosing the horizontal and vertical right angles, respectively, are replaced with new variables P / n° (number of teeth) and π·dp / n° (number of teeth), using the spiral pitch P and pitch circle diameter dp used in the calculation of the average twist angle.

[0094] Using these new variables, we proceed from the graphs in Figures 11a and 11b to the graphs in Figures 12a and 12c.

[0095] Given the band length that determines the pump's discharge rate, it can be seen that the torsion angle derived from a given geometric configuration does not exceed the set design parameter.

[0096] In the structure according to the present invention, all cross-sections perpendicular to the axis of rotation are equal to one another.

[0097] The choice to use two identical circular arcs for the unfolding of the rotor at its central and terminal ends in order to construct the helix is ​​primarily related to simplifying all the mathematical relationships derived from it. However, any other curve can also be used. Furthermore, the straight-line segments can be replaced with curves connecting the two ends. In this particular case, the helix is ​​formed by a straight line tangent to two circular arcs of the same radius, which are contained within an angle equal to the design angle.

[0098] A variation in which straight segments are replaced with curved segments can be defined, for example, by a single cosine curve. In this case, the cosine curve defines both the start and end zones and the middle zone, keeping the helix angle constantly variable throughout the entire length of the tooth.

[0099] Referring to Figure 13, the torsion angle is obtained by the following relationship. JPEG2026521096000016.jpg2599

[0100] In this specific example, R e =1, and the number of teeth n°=7 are used. JPEG2026521096000017.jpg2466

[0101] Referring to Figures 15a and 15b, the kinematic analysis of the contact points between the gear helices on the pitch circle diameter as the rotational speed changes will be explained here.

[0102] We assume that the contact is a point contact at the pitch circle diameter.

[0103] The spiral expansion in the plane is shown by dividing the graph into the following three distinct zones:

[0104] First zone z1 having a variable torsion angle, Second zone z2 having a constant twist angle, Third zone z3 having a variable torsional angle.

[0105] Replace the X and Y axes as follows:

[0106] X → Arc [C] where the spiral generation point has moved. Y → Distance traveled in the axial direction by the spiral generation point [F] When rotating, the two gears maintain contact at a single point until they have traveled a distance F equal to the band length.

[0107] The equation of motion in the axial direction of the contact point between the two helices is described as a compound function of three distinct zones z1, z2, and z3.

[0108] Referring to Figures 16a and 16b, the following relationship applies to the first zone z1. JPEG2026521096000018.jpg16136

[0109] JPEG2026521096000019.jpg1828 Using this, JPEG2026521096000020.jpg2337 C1=θ1 r p

[0110] This can be written as follows: JPEG2026521096000021.jpg39120 C=θ r p

[0111] By replacing it in item 2 above, θ r p =[kr p -kr p cosφ] θ r p =kr p (1-cosφ) 3.1 θ=k(1-cosφ)→θ=kk cosφ JPEG2026521096000022.jpg4962

[0112] By substituting in Equation 1, JPEG2026521096000023.jpg20125 Functions that are valid for 0 < φ < α as θmax → φ = α

[0113] Recall 3.1: In θ=kk cosφ, if we substitute φ=αv, θmax = kk cosα 3.3 θmax=k(1-cosα)

[0114] The time law for the motion of the contact point between two interlocking helices can be obtained here by substituting in Equation 3.2.

[0115] θ = ωt JPEG2026521096000024.jpg19137

[0116] Another important relationship can be obtained from 3.3 by substituting θmax=ω*tmax into the following equation. ω tmax = k(1-cosα) JPEG2026521096000025.jpg2290

[0117] Referring to Figure 17, we can derive 3.6 by deriving 3.4. This is the velocity at which the point is translated parallel to the rotor axis. JPEG2026521096000026.jpg17134

[0118] By deriving 3.6, we obtain 3.7, which is the equation for the acceleration experienced by the fluid along its axial movement toward the center or end of the gear, depending on the direction of rotation. JPEG2026521096000027.jpg20136

[0119] With respect to zone z2, referring to Figures 18a to 18c, it can be written as follows: JPEG2026521096000028.jpg4046

[0120] From Equation 1, By replacing it with item 2 above, JPEG2026521096000029.jpg1735 JPEG2026521096000030.jpg151133

[0121] From 3.8, JPEG2026521096000031.jpg4192

[0122] By deriving 3.7, JPEG2026521096000032.jpg26139

[0123] Regarding Zone 3, referring to Figures 19a to 19c, the same relationship as in Zone 1 applies, and in particular, it can be written as follows: JPEG2026521096000033.jpg86137

[0124] 2. JPEG2026521096000034.jpg22136

[0125] By replacing it in item 1 above, JPEG2026521096000035.jpg28136

[0126] Here, by deriving 4.3, we obtain 4.4. JPEG2026521096000036.jpg17136

[0127] The equation of acceleration is completed in section 4.5. JPEG2026521096000037.jpg23136

[0128] Next, to demonstrate the efficiency of the solutions described for damping vibrations, we will discuss the fluid dynamics of the pump.

[0129] The geometric considerations described above illustrate how the modified torsion angle development reduces fluid vibrations compared to the prior art.

[0130] The fluid volume trapped within the chamber formed by the gear teeth is delivered from the intake zone to the discharge zone. Upon reaching the high-pressure zone, these volumes are subjected to compression caused by the mutual penetration between the solid teeth and the fluid itself, and are discharged. Because the teeth have a helical shape, what is happening is a kind of fluid "helical extrusion."

[0131] Referring to Figure 20, to explain this phenomenon, we can consider a cross section that moves axially before being discharged radially from the pump. Cross section S moves axially at the same velocity as the contact point of the helices of two gears that lie on the pitch circle diameter. Assuming that the contact is always a point contact and only on the pitch circle diameter (a theoretical assumption to explain the phenomenon), the contact point on the gear helices can be considered to move continuously along the axial direction over infinite time. This occurs because as soon as the helical contact of a pair of meshing teeth ends, the next tooth immediately becomes conjugate.

[0132] Referring to Figures 21a-21c and 22, which graphically show the function of fluid cross-sectional displacement in the axial direction and the associated velocity, it is shown how the described solution improves the management of vibrations caused by the pumped fluid.

[0133] Consider a tank through which a liquid is pumped via a pipe of constant cross-section. The formula for the pump flow rate can be written as follows: JPEG2026521096000038.jpg2025 and JPEG2026521096000039.jpg1625 Therefore, maximum flow rate JPEG2026521096000040.jpg1721 You can obtain this.

[0134] From the relationship obtained above, the velocity of the cross-section S is not constant, and therefore the flow rate changes over time.

[0135] To understand the phenomena related to vibrations induced in a fluid by a pump, we can consider a small liquid mass m1 in a tank, as shown in Figure 22, being struck by a mass m0 exiting the pump. The mass m1 changes its velocity in two ways: either by receiving a small force over a long period of time, or by receiving a large force over a short period of time.

[0136] In its simplest form, Newton's second law can be described as follows: F=ma Multiplying both sides by t, we get F t=mat → F t=m V → I=p.

[0137] In the equation, I = impulse and p = momentum.

[0138] Referring to two theorems in physics, namely the impulse theorem (applicable to a single object), which states that the change in momentum is equal to the impulse of the force acting on an object, and the law of conservation of momentum, which states that the total momentum of a system consisting of two interacting particles remains constant, we can write the following:

[0139] p = p0 + p1 = m0V0 + m1V1

[0140] At the moment t’ below, p’ = p’0 + p’1 = m0V’0 + m1V’1 is obtained.

[0141] According to the law of conservation of momentum, p ’ = p = cost is obtained.

[0142] These considerations lead to the finding that, since the interaction between two particles causes an exchange of momentum, the momentum lost by one of the two particles is equal to the momentum gained by the other particle. JPEG2026521096000041.jpg1451

[0143] By performing side-by-side subtraction, Δp0 = -Δp1 is obtained.

[0144] Here, JPEG2026521096000042.jpg49114

[0145] When two particles interact, the force acting on one particle is equal in magnitude and opposite in direction to the force acting on the other particle (this is consistent with Newton's third law of motion).

[0146] For the prior art in which a leading particle undergoes rapid acceleration by colliding with a subsequent particle having a low and constant speed, it can be seen from FIGS. 21a to 21c and the above equations that a single particle undergoes rapid acceleration by colliding with a subsequent particle having the same acceleration.

[0147] Here, referring to FIG. 23, assuming point contact on the pitch circle diameter and considering that contact occurs on a line formed obliquely across the entire involute arc with respect to the gear band, the contact line formed between the rotating gears will be described.

[0148] Here, a constant helix will be described for the calculation of the length of the contact line of a helical gear.

[0149] The symbols are shown below. Rp: The length of the line drawn when the involute of a rotating gear touches the pressure line (the direction in which pressure is exchanged). Rs: Contact initiation radius on the involute Rf: Radius of contact on the involute α s :Twist angle calculated for Rs θ ev : Contact angle on an involute JPEG2026521096000043.jpg3152

[0150] As shown in Figures 24a to 24c, by analyzing the contact line of a gear with a constant helix and a torsional contact ratio equal to 1, it can be seen that the overall length is always kept the same. In fact, as a pair of teeth gradually disengage, the contact line shortens by the same amount as the subsequent pair of teeth engage.

[0151] The above explanation applies to gears with a torsional contact ratio equal to 1. The case where this value is less than 1 is analyzed below.

[0152] The contact line shortens in proportion to the offset at the opposite end of the gear band of the tooth following the meshing tooth.

[0153] Consider the selected value of Re for this invention: Re = 0.65.

[0154] This means that the front of the gear rotates along the band by an angle equal to the following: JPEG2026521096000044.jpg2558

[0155] Therefore, the offset of the teeth adjacent to the opposing ends of the band is equal to the following: JPEG2026521096000045.jpg5881

[0156] Next, we analyze how the contact line length changes when Re=1. JPEG2026521096000046.jpg2562

[0157] In order to ensure the continuity of occlusion and to always have teeth that are in contact, JPEG2026521096000047.jpg1238 It must be that way.

[0158] In this invention, JPEG2026521096000048.jpg1233 Since it was selected, JPEG2026521096000049.jpg1245 This is the result.

[0159] The above formula can be used to illustrate the significant advantages brought about by the following inventions compared to the prior art.

[0160] By recalling Equation 3.4 for a segment with a variable twist angle and Equation 3.7 for a segment with a constant twist angle, JPEG2026521096000050.jpg2991

[0161] By substituting the value of 11° in equations 3.4 and 3.7, respectively, the following can be obtained. JPEG2026521096000051.jpg1292 Variable torsion segment JPEG2026521096000052.jpg1089 Constant torsion segment

[0162] As can be seen from the obtained values, the contact line length in segments with a variable twist angle is longer than the contact line length in segments with a constant twist angle.

[0163] Alternatively, it can be written that the following relationship applies to segments with a variable torsion angle.

[0164] 3.3 θmax=k(1-cosα)

[0165] Sections 3.4 and 3.7 can be rewritten as follows: JPEG2026521096000053.jpg22169

[0166] JPEG2026521096000054.jpg18118 By simplifying, JPEG2026521096000055.jpg22127 JPEG2026521096000056.jpg53137

[0167] therefore, JPEG2026521096000057.jpg29145

[0168] Equation 4.8 demonstrates that for a gear having a specific helix angle α and a selected geometric configuration, the segment with a variable helix angle has a longer contact line.

[0169] This makes it possible to improve the stress distribution when the contact line takes its minimum length value (because the torsional contact ratio is less than 1) compared to prior art.

[0170] The forces acting on the teeth of a gear are analyzed below.

[0171] Starting with the equation that determines the power transmitted to the fluid, from a hydraulic perspective, it can be written as follows: JPEG2026521096000058.jpg2149

[0172] From a mechanical standpoint, BP=C ω

[0173] Focusing on the power supplied to the fluid shown in A, the torsional torque transmitted to the shaft can be obtained using B. JPEG2026521096000059.jpg2444

[0174] The torque transmitted from the drive gear to the driven gear is JPEG2026521096000060.jpg1830 It becomes equal to.

[0175] Referring to Figures 25a and 25b, the resultant force F, directed according to the pressure angle φt, is schematically shown. In these figures, only the components of the resultant force F are shown because its three-dimensional projection, which is inclined perpendicular to the torsion angle ψ, cannot be seen. JPEG2026521096000061.jpg2392

[0176] Referring to Figure 26, a complete three-dimensional diagram of the forces acting on the teeth of a gear with helical teeth can be seen. Force F acts perpendicular to the point of contact.

[0177] The symbols are shown below. F:Resultant force Ft: Tangential component (force component useful for power transmission) Fr: Radial component Fa: Axial component

[0178] The following relationship can be obtained between the resultant force and its components. 5.0 F t =F cosφ n cosψ 5.1 F r =F senφ n 5.2 F a =F cosφ n sen

[0179] Ft can usually be obtained directly from the design data from 4.9, and the other forces can be obtained by the following equations. JPEG2026521096000062.jpg64120

[0180] Referring to Figures 27a to 27c, the direction of the resultant force at various points on the gear is clear in the case of a conventional herringbone rotor, a prior art rotor with variable helix angles in three zones, and the rotor according to the present invention.

[0181] Referring to Figures 28a and 28b, assuming point contact on the pitch circle diameter, it becomes clear how the direction of the resultant force at the contact point at the end of the gear has an axial component, and how this zone is exposed to periodic stresses and forces that can cause chipping of this edge under certain operating conditions.

[0182] Therefore, all components of the resultant force F are present at point a, and their modules are as follows: JPEG2026521096000063.jpg2158

[0183] Referring to Figure 29, this edge b is a smaller module JPEG2026521096000064.jpg2242 In addition to having an axial component F a =F t Instead, the focus shifts to how the resultant force F, which does not even have tan0=0, is received.

[0184] Furthermore, as shown in the comparison diagrams Figures 30a and 30b, it is clear that the cross-section of the end of tooth 6 is at a right angle, whereas in the prior art it is at an acute angle.

[0185] Furthermore, recall the kinematic equations for the velocity (at the ends) at the contact points (always assuming point contact) of the gear's helical on the pitch circle diameter. JPEG2026521096000065.jpg51137

[0186] Dealing with velocity, approaching infinity is paradoxical, as the gear has a helical development in that segment where the angle is equal to 0°. Therefore, in the initial segment, the rotor can be considered "straight" teeth in all respects, and the contact in that zone no longer appears as point contact, but its length can be calculated and it is distributed along a line determined by the mechanical properties of the material.

[0187] The manufacturing method for the gear 1 described above includes the following steps.

[0188] Refer to Figures 12a to 12c.

[0189] Process 1 Determine the axial rotation equations for all tooth profile cross-sections calculated on the pitch circle diameter. A series of coordinates (x) representing the helical path. i ,y i ,z i ) can be obtained.

[0190] Process 2 Using 3D software, a solid model is manufactured by sliding the rotor's front tooth profile along the helical path shown in Figures 8b and 8c.

[0191] Process 3 Transfer the solid model to CAD / CAM.

[0192] Process 4 For example, heat treatment and inter-tooth space formation using numerically controlled machining stations such as 5-axis machining centers.

[0193] The present invention brilliantly solves the technical problems and achieves several advantages. Its first advantage comes from the fact that it is possible to manufacture gears having a reverse helix with a partially or entirely variable helix angle, a tooth profile that does not cause locking, and a shape that suppresses the apex at the center of the rotor.

[0194] Furthermore, the precise and continuous opposite twist angle of the teeth does not generate any axial force, and therefore, it is less likely to displace the gear or damage the tooth ends, allowing it to be incorporated into gears without axial compensation.

[0195] In short, the present invention makes it possible to manufacture a rotor having a reverse twist, a tooth profile that does not cause trapping, and a helical shape that suppresses the corner point at the center of the rotor itself, and thus all problems related to the machining of rotors by machine tools are solved.

[0196] Furthermore, the present invention makes it possible to manufacture gears for hydraulic systems having a reverse helix with a partially or entirely variable helix angle.

[0197] Obviously, a person skilled in the art could make several modifications and variations to the above invention for incidental reasons and to satisfy specific requirements, but all of them fall within the scope of protection of the present invention as defined in the following claims.

Claims

1. A double helical gear (1) for a hydraulic gear system (2), configured to be coupled to a support shaft (5) to form a drive gear or driven gear of the hydraulic system, comprising a plurality of teeth (6), wherein the plurality of teeth (6) extend in the longitudinal or axial direction of the teeth (6) with a variable helix angle (ψ) in a continuous function, and each tooth (6) has a central zone (C) having a variable helix angle (ψ) that causes a helix transition from right-handed to left-handed, and the tooth profile (4) of the teeth maintains shape continuity in each cross-section, wherein each tooth (6) has at least a start zone (A) and an end zone (E) at two opposing side ends of the double helical gear (1), each having a variable helix angle (ψ), and in these start zone (A) and end zone (E), the helix angle decreases as it approaches the side ends of the double helical gear (1).

2. The double helical gear (1) according to claim 1, wherein each tooth (6) further comprises a proximal intermediate zone (B) and a distal intermediate zone (D) having a constant helix angle (ψ) that connect the starting zone (A) and the ending zone (E) to the central zone (C).

3. The double helical gear (1) according to claim 1, wherein the starting zone (A), the ending zone (E), and the central zone (C), each having a variable helix angle, are directly connected by an inflection point.

4. The helix angle (ψ) is zero at the opposing side ends of the gear (1), i.e., the tooth profile (4) of each individual tooth is joined perpendicularly to the opposing side surfaces of the gear (1), according to any one of claims 1 to 3.

5. The double helical gear (1) according to claim 4, wherein the axial component (Fa) of the force (F) exchanged by the gear that meshes with another identical gear during use is zero at the side end.

6. The double helical gear (1) according to any one of claims 1 to 5, wherein the tooth profile (4) of the tooth (6) is a mirror image of the central plane passing through the transition point between right-handed and left-handed helical.

7. A double helical gear (1) according to any one of claims 1 to 6, wherein the torsional contact ratio parameter (Re) is between 0.6 and 1, preferably between 0.6 and 0.8, and more preferably equal to 0.

65.

8. The double helical gear (1) according to any one of claims 1 to 7, wherein the tooth profile (4) of the teeth of the gear (1) is a tooth profile that does not cause locking.

9. The tooth profile (4) that does not result in the aforementioned trapping has two truncated diameters A double helical gear (1) according to claim 8, defined by the apex and base (7, 9) of two arcs or ellipses connected by an involute tooth profile (8) contained between them.

10. The lower involute cutoff diameter is They are selected equally, and here, The double helical gear (1) according to claim 9, wherein p1 is the pitch circle diameter, and the parameter p1 is between 9.7% and 9.9%, preferably equal to 9.8%.

11. The upper involute cutoff diameter is They are selected equally, and here, The double helical gear (1) according to claim 9 or claim 10, wherein p2 is the pitch circle diameter and the parameter p2 is between 12.1% and 12.3%, preferably equal to 12.2%.

12. A double helical gear (1) having a cutting edge (10) defined by a limited thickness that protrudes from the tooth profile (4) at the top of each tooth (6).

13. A hydraulic gear device comprising a pair of gears according to any one of claims 1 to 12.

14. The hydraulic gear apparatus according to claim 13, wherein the device is a positive displacement pump or a hydraulic gear motor.

Images