Multi-gear small-tooth-difference transmission capable of eliminating jitter

Abstract

The multi-gear small-tooth-difference transmission comprises at least two axis revolution gears and an eccentric balance weight part, through the design of angular momentum of the axis revolution gears B1 and B2 and the cooperation of the installation position of the eccentric balance weight part T1, the problem of axial shaking of a multi-gear small-tooth-difference transmission in the prior art is solved, and the axial shaking speed of the multi-gear small-tooth-difference transmission is improved through the design of the angular momentum of the axis revolution gears B1 and B2 and the cooperation of the installation position of the eccentric balance weight part T1. And the operation is more stable.

Description

Technical Field

[0001] This invention application relates to a transmission device for a low-tooth-difference gearbox, belonging to the field of mechanical technology. Background Technology

[0002] Existing multi-gear transmissions with small tooth differences, such as cycloidal pinwheel transmissions, typically consist of two axially offset planetary gears mounted at 180°, or three axially offset planetary gears mounted at 120°. Because the centers of mass of these planetary gears are not on the same plane perpendicular to the shafts, the transmission generates an axial overturning moment during operation, causing the entire transmission to vibrate. This vibration is particularly noticeable during startup or shutdown.

[0003] Therefore, cycloidal pinwheel speed reducers are difficult to apply to the fourth, fifth, or later joints of a robot. Even when used in the third joint, noticeable vibrations still occur during startup or shutdown, necessitating the use of the second or fourth joint's motor idle to balance angular momentum. In our actual experiments, we found that this balance becomes increasingly difficult to achieve the further back the joint. This vibration is particularly difficult to suppress when the weights of the pivot gears are unequal. This is a common problem in industrial robots or multi-joint robots. To address these issues, this invention is hereby disclosed. Summary of the Invention

[0004] A multi-gear transmission with few tooth differences to eliminate vibration is characterized by comprising at least two axially pivoting gears and at least one eccentric counterweight component, wherein the axially pivoting gears are axially offset, the weight of the eccentric counterweight component is less than the weight of the axially pivoting gears, and the angular momentum of the axially pivoting gears can be equivalent to and symmetrical with the angular momentum of the eccentric counterweight component in the same plane. That is, the angular momentum of the eccentric counterweight component and the axially pivoting gears are in a stable geometric relationship in the angular momentum triaxial coordinate system, which can simultaneously cancel out the axial and radial angular momentum of the multi-gear transmission with few tooth differences.

[0005] The implementation methods include, but are not limited to: the first method is that the absolute values ​​of the angular momentum of the axial revolution gears are not equal, and in the axial direction, the eccentric counterweight is set on the outside of the axial revolution gear with the larger angular momentum; the second method is that the axial revolution gears are fixed together, and in the axial direction, the eccentric counterweight is set on both sides of the axial revolution gears; the third method is that the axial revolution gears are fixed together, and in the axial direction, the eccentric counterweight is set between the axial revolution gears; the fourth method is that each axial revolution gear has at least one corresponding eccentric counterweight, wherein the eccentric counterweight is not another axial revolution gear, including one of the above methods, or a combination of the above methods.

[0006] The obvious advantage of the above design is that, due to the addition of the eccentric counterweight, the angular momentum of the central rotating gear can be equivalently placed on the same plane, and the angular momentum of the eccentric counterweight, the central rotating gear, and the central rotating gear are always in a triangular stable relationship. This eliminates the problem of machine vibration during startup and shutdown in existing technologies, and this elimination is reliable and does not change with rotational speed or angle of motion.

[0007] The second advantage of this design is that the central axis rotation gear can be fixedly connected, achieving stable angular momentum balance through eccentric counterweights. This eliminates the need for components such as pins, pin discs, or cross rails. The added eccentric counterweights can reduce their own weight by increasing the eccentricity; in fact, the weight of the eccentric counterweights can be even less than that of the pins, pin discs, or cross rails, thereby improving the torque-to-weight ratio. When applied to the fourth, fifth, or later joints of a robot, the reduced moment of inertia significantly increases the robot's flexibility.

[0008] As an effective application of this invention, the transmission includes at least four gears. Gears A1 and A2 are fixed-axis gears, and gears B1 and B2 are axial revolving gears. The axial revolving gears mesh with the fixed-axis gears, driven by an eccentric bearing. Gear A1 meshes with gear B1, and gear A2 meshes with gear B2. The number of teeth of gear B1 is N, the number of teeth of gear A1 is N+P, the number of teeth of gear B2 is N+M, and the number of teeth of gear A2 is N+M+Q, where P≥1, M≥1, and Q≥1. The gears are combined in one of the following two ways: the first way is that the fixed-axis gears are fixedly connected, and one of the axial revolving gears can be fixed and does not rotate, while the unfixed axial revolving gear can drive the output shaft; the second way is that the axial revolving gears are fixedly connected in the direction of rotation, and one of the fixed-axis gears can be fixed and does not rotate, while the unfixed fixed-axis gear can drive the output shaft.

[0009] The purpose of the above design is to achieve a large reduction ratio within a relatively small space. The reduction ratio is at its maximum when P = M = Q = 1. Due to the large reduction ratio, the reverse drive resistance is also large, especially when N is greater than 20, in which case reverse drive may experience self-locking. When reverse drive is required, the reduction ratio can be reduced by increasing the values ​​of P, M, and Q, thereby reducing the reverse drive resistance.

[0010] As an effective improvement of this invention application, the eccentric counterweight part includes a structure that can force the surrounding fluid to flow, the surrounding fluid including a lubricant or air, to play a role in lubrication and heat dissipation, or the installation position of the eccentric counterweight part has an axial adjustment function to adapt to the different angular momentum of the shaft revolution gears of different batches, having one or both of the above structures.

[0011] One of the purposes of the above design is to use the structure and power of the eccentric counterweight itself to force the lubricant or air to flow within the transmission chamber, thereby lubricating and cooling the various friction parts. The eccentric counterweight can be adjusted and positioned axially during installation, allowing for faster transmission assembly and saving time.

[0012] As an alternative, at least two more gears can be added to the four gears: a fixed-axis gear A3 and a central rotating gear B3. The central rotating gear B3 is driven by an eccentric shaft to mesh with the fixed-axis gear A3. The number of teeth of the central rotating gear B3 is N+P, and the number of teeth of the fixed-axis gear A3 is N+P+1, where P≥1. Alternatively, multiple sets of gears can be added in this manner.

[0013] The purpose of the above design is to increase the number of reduction ratios and provide more options for reverse drive of the transmission.

[0014] As an alternative to the gear design in this patent application, the gears of the transmission may include a pin gear, the pin gear including a pin gear housing and pin teeth, the pin gear housing containing tooth grooves, and the pin teeth being rolling bodies capable of rolling along the tooth surfaces of the tooth grooves in the pin gear housing. The transmission also includes a device for preventing the pin teeth from falling out of the tooth grooves of the pin gear, the device for preventing the pin teeth from falling out may be, but is not limited to, one or a combination of more than one of the following: baffles, flanges, and annular guards.

[0015] The advantages of the above design are that the needle teeth can use solid rolling elements, reducing the difficulty of machining, and all contact surfaces are rolling contacts, thus resulting in higher torque output capability.

[0016] As a design improvement, the design includes a pin gear in which the pin teeth can roll in the tooth grooves of the pin gear housing, and the eccentric counterweight includes a movable counterweight block that can move relative to the eccentric counterweight body, or the movable counterweight block is restricted to a certain range of movement, or the movable counterweight block is connected to the eccentric counterweight body by an elastic member.

[0017] The purpose of this design is that when using pin gears, and each pin gear meshes with only one axial revolving gear, the pin teeth in the pin gear can roll within the tooth grooves of the pin gear housing. Typically, about half of the pin teeth are not in close contact with the cycloidal gear, and the tooth grooves in the pin gear housing do not tightly constrain these pin teeth that are not in contact with the cycloidal gear. Therefore, when the machine reverses or suddenly stops, the pin teeth will move due to inertia. The inertial movement of the pin teeth will cause their angular momentum to be delayed and act on the transmission. The purpose of setting a movable counterweight block in the eccentric counterweight component, which can move relative to the main body of the eccentric counterweight component, is to counteract this angular momentum delay effect of the pin teeth.

[0018] The design schemes for the tooth profile curve of the tooth groove or the tooth profile curve of the gear in this patent application include, but are not limited to, the following schemes: at least one tooth profile curve contains a circular arc curve, or at least one tooth profile curve contains a cycloid, or at least one tooth profile curve contains an elliptical curve, or at least one tooth profile curve contains an elliptic cycloid, or at least one tooth profile curve contains a straight line, or at least one tooth profile curve contains an involute, or at least one tooth profile curve contains a conic curve; or a combination of one or more of the above schemes.

[0019] As an alternative to this patent application, the transmission also includes a clutch C, the design method of which includes, but is not limited to: the first method is to select and fix the rotation of one of the gears by switching the clutch C, and output by the other gear; the second method is to switch the output gear ratio according to the external load or machine command, including one of the above two clutch design methods, or a combination of the two methods.

[0020] The purpose of the above design is to provide multiple gears in a single transmission. When there are only two pivoting gears and M = P = Q = 1, two gears with a multiplicative relationship and the same direction can be achieved, i.e., n. 2 The gear ratios are either 1 and n:1, or [n×(n+2)]:1 and n:1. Of course, more gears are possible with more rotating shafts. A significant advantage is its applicability to scenarios with constantly changing loads, such as the joints of reciprocating transport machinery, the joints of heavy-duty running machinery, or vehicles that need to travel on various road conditions.

[0021] This invention applies for a multi-gear transmission with a small tooth difference that eliminates vibration. It can be used with various engines to form a power output module. The engine may include, but is not limited to, an axial flux motor, an external rotor motor, an internal rotor motor, or a turbine.

[0022] This invention discloses a multi-gear transmission with low tooth difference that eliminates vibration. It can be applied in transmission systems, including but not limited to propeller drives, hub drives, or mechanical actuators. The mechanical actuators can be applied to rotary joints of industrial machinery, wearable robotic arms, reciprocating transport machines, embodied intelligent machines, vehicle steering systems, aircraft steering rudder transmission systems, ship steering rudder transmission systems, or the transmission can be applied to electric vehicles or electric construction machinery.

[0023] The technical advantages and scope of protection of this invention will become clearer below with reference to the accompanying drawings. The drawings and their descriptions are for illustrative purposes only and the specific examples herein do not limit the scope of protection of this invention. Attached Figure Description

[0024] Symbol explanations in the diagram: A1—Fixed-axis gear #1; A11—Pin tooth in the pin gear; A12—Pin tooth housing in pin tooth v; A13—Pin pin shaft; A2—Fixed-axis gear #2; A3—Fixed-axis gear #3; B1—Pin gear #1; B2—Pin gear #2; B12—Pin shaft; B121—Pin shaft disc; B13—Conical step of the pin shaft; B131—Conical opening of the pin hole on the pin gear; B21—Gear groove on the pin tooth housing; B221—Annular guard #1; B222—Annular guard #2; B223—Annular guard #3; B224—Annular guard #4; B225—Annular guard #5; B226—Annular guard #6; B3v—Pin gear #3; C—Clutch; D—Fixed disc connecting the housing; E1—Eccentric shaft #1 E2vv2 eccentric shaft; Input shaft v—G1; Output shaft—G2; F1—First bearing; F2—Second bearing; F3—Third bearing; F4—Fourth bearing; F5—Fifth bearing; F6—v—Sixth bearing; H1—First section position; H2—Second section position; O—Shaft center position; J1—No.1 cross slide rail; J11—No.1 cross slide rail disc; J2—No.2 cross slide rail; J21—No.2 cross slide rail disc; K1—Positioning bolt; K2—Bus sleeve; T1—No.1 eccentric counterweight; T2—No.2 eccentric counterweight; T3—Blade in the eccentric counterweight; T11—Eccentric counterweight body; T12—Modible counterweight; T13—Spring; T21—Fluid inlet; T22—Fluid outlet; T31—Adjustable position buckle; S—Centrifugal force direction.

[0025] Figure 1 This is a coordinate analysis diagram of the axial overturning force of a multi-gear transmission with a small tooth difference in the existing technology.

[0026] Figure 2 This is a coordinate diagram of the first embodiment of the present invention.

[0027] Figure 3 This is a coordinate diagram of the second embodiment of the present invention.

[0028] Figure 4 This is a coordinate diagram of the third embodiment of the present invention.

[0029] Figure 5 This is a schematic diagram of the axial cross-section of an application example based on the first solution of this invention.

[0030] Figure 6 yes Figure 5The radial section diagram of the second section position H2 on the right side of the middle section.

[0031] Figure 7 yes Figure 5 Radial cross-sectional view of the first cross-section position H1 on the left.

[0032] Figure 8 yes Figure 5 A radial cross-sectional view of another gear tooth arrangement at the first cross-section position H1 on the left.

[0033] Figure 9 This is a partial schematic diagram of a pin gear according to this patent application.

[0034] Figure 10 This is a schematic diagram of the axial cross-section of an application example of the second solution according to the present invention.

[0035] Figure 11 This is a schematic diagram of the axial cross-section of an application example of the third solution according to the present invention.

[0036] Figure 12 This is a schematic diagram of an eccentric counterweight component.

[0037] Figure 13 This is a schematic diagram of the second type of eccentric counterweight component. Detailed Implementation

[0038] The advantages and preferred embodiments of the present invention are illustrated with reference to the accompanying drawings. Specific examples are provided to illustrate the advantages of the invention and should not be construed as limiting the scope of protection of the invention.

[0039] Figure 1 This is a coordinate analysis diagram of the axial overturning force in a multi-gear, low-tooth-difference transmission in the prior art. The upper part of the diagram shows the analysis of two axially offset planetary gears installed at 180°. The middle of the diagram is the input shaft G1, which is on the y-axis. It contains planetary gear B1, whose center of mass is O1; therefore, its centrifugal force is in the S1 direction. Planetary gear B2, whose center of mass is O2, has a centrifugal force in the S2 direction. Because the centrifugal forces S1 and S2 are offset on the y-axis and cannot be equivalently represented on the same plane or symmetrically, an overturning force, as indicated by the arc-shaped arrow in the diagram, is generated. This overturning force is particularly noticeable during shutdown or startup.

[0040] Figure 1The lower half of the diagram shows an analysis of three axially offset planetary gears installed at 120° angles. The input shaft G1 is in the center of the diagram, located on the y-axis. Gear B1 is the first axially offset planetary gear, with its center of mass O1; therefore, its centrifugal force is in the direction of S1. Gear B2 has its center of mass O2, so its centrifugal force is in the direction of S2. Gear B3 has its center of mass O3, so its centrifugal force is in the direction of S3. Because centrifugal forces S1, S2, and S3 are offset on the y-axis and cannot be equivalently represented on the same plane or be symmetrical, an overturning force around the x-axis, as shown by the arc-shaped arrow in the diagram, is generated.

[0041] Figure 2 This is a coordinate schematic diagram according to the first embodiment of this invention. The two axially rotating gears B1 and B2 are offset axially and installed at 180°. The angular momentum of gear B1 is greater than that of gear B2, possibly due to different eccentricities or because gear B1 is heavier. The center of mass of gear B1 is O1, therefore its centrifugal force direction is S1. The center of mass of gear B2 is O2, therefore its centrifugal force direction is S2. The eccentric counterweight T1 is above the y-axis, that is, outside the axially rotating gear B1 with the larger angular momentum. The center of mass of counterweight T1 is O3, therefore its centrifugal force direction is S3. As shown in the diagram, centrifugal forces S2 and S3 can be equivalent to S4, which is in the same plane as S1, and centrifugal force S4 can be symmetrical to S1. Therefore, the first solution of the present invention can eliminate overturning force, and compared with the prior art, there will be no shaking caused by angular momentum whether it is starting, stopping suddenly or accelerating.

[0042] Figure 3 This is a coordinate schematic diagram of the second embodiment of the present invention. Two axially rotating gears, B1 and B2, are fixedly connected. The angular momentum of gear B1 and gear B2 can be equal or unequal. The center of mass of gear B1 is O1, therefore its centrifugal force direction is S1. The center of mass of gear B2 is O2, therefore its centrifugal force direction is S2. Axially, two eccentric counterweights T1 and T2 are located on opposite sides of the axially rotating gears. The center of mass of counterweight T1 is O3, therefore its centrifugal force direction is S3. The center of mass of counterweight T2 is O4, therefore its centrifugal force direction is S4. As shown in the diagram, the centrifugal forces S1 and S2 of the axially rotating gears can be equivalent to and symmetrical with the centrifugal forces S3 and S4 of the eccentric counterweights in the same plane. Therefore, the second embodiment of the present invention can eliminate overturning force.

[0043] Figure 4This is a coordinate schematic diagram of the third embodiment of the present invention. Two axially rotating gears, B1 and B2, are fixedly connected. The angular momentum of gear B1 and gear B2 can be equal or unequal. The center of mass of gear B1 is O1, therefore its centrifugal force direction is S1. The center of mass of gear B2 is O2, therefore its centrifugal force direction is S2. Axially, an eccentric counterweight T1 is positioned between the two axially rotating gears. The center of mass of eccentric counterweight T1 is O3, therefore its centrifugal force direction is S3. As shown in the diagram, the centrifugal forces S1 and S2 of the axially rotating gears can be equivalent to and symmetrical to the centrifugal force S3 of the eccentric counterweight in the same plane. Therefore, the third embodiment of the present invention can eliminate overturning force.

[0044] Figure 5 This is an axial cross-sectional schematic diagram of an application example according to the first solution of this invention application. The input shaft G1 is in the middle of the diagram, with its axis at O2. A pin disk B121 is mounted on the input shaft G1 via a first bearing F1. Pin disk B121 has a pin B12. A second fixed-axis gear A2 is mounted on the outer side of pin disk B121 via a second bearing F2. A clutch C is mounted on the left side of the fixed disk D connecting to the outer casing. Clutch C can connect to the outer fixed-axis gear A2 or to the inner pin disk B121. In the diagram, clutch C is connected to the second fixed-axis gear A2, so the second fixed-axis gear A2 is fixed and does not rotate, while the pin disk B121 can rotate.

[0045] Figure 5 The input shaft G1 is also equipped with a second eccentric shaft E2, and the second eccentric shaft E2 is mounted on the second axial orbital gear B2 via the third bearing F3. Its axis is O1. The second axial orbital gear B2 meshes with the pin tooth A11 on the second fixed shaft gear A2.

[0046] Figure 5 The left side of the No. 2 fixed-axis gear A2 is connected to the No. 1 fixed-axis gear A1 via the fourth bearing F4. The No. 1 fixed-axis gear A1 is connected to the input shaft G1 via the sixth bearing F6, and the No. 1 fixed-axis gear A1 is fixedly connected to the output shaft G2. The input shaft G1 also has the No. 1 eccentric shaft E1 mounted on it, which is connected to the No. 1 central rotating gear B1 via the fifth bearing F5. The axis of the central rotating gear B1 is O3. The No. 1 central rotating gear B1 meshes with the No. 1 fixed-axis gear A1 via the pin tooth A11.

[0047] Figure 5As can be seen, pin B12 passes through both the second-axis planetary gear B2 and the first-axis planetary gear B1, with the portion on the right passing through the second-axis planetary gear having a larger diameter. The first function of this design is that the different diameters of the holes through which pin B12 passes between the first-axis planetary gear B1 and the second-axis planetary gear B2 allow for adjustment of the weight of the two gears and a reduction in the difference in angular momentum between them. The second function of this design is to create a step or ramp between the pin and the gear, which can withstand axial thrust. For example... Figure 5 In the transmission, the tapered opening B131 of the pin hole on the No. 1 shaft planetary gear B1 cooperates with the tapered step B13 of the pin. This cooperation can withstand the axial thrust from left to right from the shaft planetary gear. The third bearing F3 and the fifth bearing F5 can withstand the axial thrust from right to left from the shaft planetary gear. At the same time, the first bearing F1 can withstand the thrust from left to right from the pin disc B121. Therefore, the entire transmission can withstand bidirectional axial inertial force or vibration. It has a simple and compact structure, and each bearing can be an angular contact bearing or a tapered bearing.

[0048] Figure 6 yes Figure 5 The diagram shows a radial section at position H2 on the right. The center of the diagram is the input shaft G1, with its axis at O2. The input shaft G1 is fixedly connected to eccentric shaft E2 (position 2). Eccentric shaft E2 is connected to a third bearing F3, whose axis is O1. Axis O1 revolves around axis O2 or performs a oscillating motion. The outer ring of bearing F3 is fitted with a planetary gear B2 (position 2) that meshes with the outermost fixed gear A2 (position 2). Fixed gear A2 is a pin gear with pin teeth A11. Pin B12 passes through a hole in planetary gear B2.

[0049] Figure 7 yes Figure 5 The diagram shows a radial section at position H1 on the left. The center of the diagram is the input shaft G1, with its axis at O2. The input shaft G1 is fixedly connected to an eccentric shaft E1. The outer layer of eccentric shaft E1 is connected to a fifth bearing F5, whose axis is O3. Axis O3 revolves around axis O2 or performs a pendulum motion. The outer ring of the fifth bearing F5 is fitted with a planetary gear B1 (axis 1) that meshes with the outermost fixed gear A1 (axis 1). Fixed gear A1 is a pin gear with pin teeth A11. Pin B12 passes through a hole in planetary gear B1.

[0050] The following is combined Figure 5 , Figure 6 , Figure 7This will illustrate the operation and advantages of this case. Since the number of teeth on the axial orbital gear B2 is greater than that on the axial orbital gear B1, under the same module, the angular momentum of the axial orbital gear B2 is greater. According to the first technical solution of this invention, "the absolute values ​​of the angular momentum of the axial orbital gears are not equal, and in the axial direction, the eccentric counterweight is located on the outside of the axial orbital gear with the larger angular momentum." Therefore, in Figure 5 As can be seen, the input shaft G1 is also equipped with an eccentric counterweight T1, which is mounted on the outside of the axial orbital gear B2, which has a larger angular momentum. Therefore, designing the transmission according to the structure in this case can eliminate axial overturning force.

[0051] Combination Figure 5 , Figure 6 , Figure 7 It can be seen that the third bearing F3 and the fifth bearing F5 are symmetrically arranged at 180 degrees. When the input shaft G1 rotates clockwise 1 turn, the second shaft planetary gear B2 meshes with the second fixed shaft gear A2. Since the second fixed shaft gear A2 is fixed on the fixed disk D connecting the outer shell by the clutch C and cannot rotate, the pin disk B121 will push the second shaft planetary gear to rotate counterclockwise 1 / 15 turn. The pin B12 on the pin disk B121 also passes through the first shaft planetary gear B1, and the first shaft planetary gear B1 is also forced to rotate counterclockwise 1 / 15 turn. Since the first shaft planetary gear B1 is simultaneously meshing clockwise with the first fixed shaft gear A1, the first fixed shaft gear A1 rotates clockwise 1 / 225 turn, and the reduction ratio is 225:1.

[0052] When clutch C disengages from fixed-axis gear A2 and switches to fixed-pin disc B121, the circumferentially fixed-axis gear B1 is also fixed and cannot rotate. Therefore, input shaft G1 rotates clockwise one revolution, and fixed-axis gear A1 rotates clockwise 1 / 15 revolution, resulting in a reduction ratio of 15:1. The direction of rotation is the same as when clutch C fixes fixed-axis gear A2. Similarly, when fixed-axis gears A1 and A2 are fixed together, and circumferentially fixed-axis gears B1 and B2 each use different pins and pin discs to restrict circumferential rotation, and when circumferential gear B2 is fixed by clutch C, the pin disc of circumferential gear B1 can output a reduction ratio of 224:1.

[0053] Therefore, when M = p = Q = 1, two gears with a multiplicative relationship and the same direction can be achieved, i.e., n 2 The gear ratios are either :1 and n:1, or [n×(n+2)]:1 and n∶1.

[0054] Figure 8 yes Figure 5 A radial cross-sectional view of another gear tooth arrangement at the first cross-section position H1 on the left. Figure 8 and Figure 7 The structures are identical. The center of the diagram is the input shaft G1, with its axis O2. An eccentric shaft E1 is fixedly connected to the periphery of input shaft G1. The outer layer of eccentric shaft E1 is connected to the fifth bearing F5, whose axis O3 revolves around axis O2 or performs a pendulum motion. The outer ring of the fifth bearing F5 is fitted with the first axis revolving gear B1, which meshes with the outermost fixed-axis gear A1. The first fixed-axis gear A1 is a pin gear, containing pin teeth A11. A pin B12 passes through a hole in the first axis revolving gear B1.

[0055] Figure 8 and Figure 7 The difference is, Figure 8 number of teeth ratio Figure 7 Middle-aged, when Figure 5 , Figure 6 , Figure 8 The components are combined into a gearbox, where P=1, M=2, and Q=1. Therefore, the reduction ratios are different. The third bearing F3 and the fifth bearing F5 are symmetrically arranged at 180 degrees. When the input shaft G1 rotates clockwise one revolution, the second-axis planetary gear B2 meshes with the second-axis fixed gear A2. Since the second-axis fixed gear A2 is fixed to the fixed disk D connecting the housing by the clutch C and cannot rotate, the pin disk B121 will push the second-axis planetary gear to rotate counterclockwise by 1 / 15 revolution. The pin B12 on the pin disk B121 also passes through the first-axis planetary gear B1, forcing the first-axis planetary gear B1 to also rotate counterclockwise by 1 / 15 revolution. Because the first-axis planetary gear B1 is simultaneously meshing clockwise with the first-axis fixed gear A1, the first-axis fixed gear A1 rotates clockwise by 1 / 105 revolution, resulting in a reduction ratio of 105:1.

[0056] Figure 5 , Figure 6 , Figure 8 Combination and Figure 5 , Figure 6 , Figure 7 Compared to the maximum reduction ratio, the combined reduction ratio is reduced by about half. Its effect, in this case, is that when the N value is large, the resistance of reverse drive is smaller, especially when the P value is greater than 1 and the M value is greater than 2, the advantage of reverse drive is more obvious. Alternatively, it can better match the engine's maximum and rated speeds. Those skilled in the art can make corresponding adjustments during the design process based on the technology disclosed in this patent application; specific examples are not listed here.

[0057] Figure 9This is a partial schematic diagram of a pin gear according to this patent application. The pin gear housing A12 contains a tooth groove B21, and the pin tooth A11 can roll on the tooth surface of the tooth groove B21. An annular guard B221 is also installed on the pin gear housing. The annular guard B221 is above the tooth groove B21 and has a hole directly opposite the tooth groove. The pin tooth A11 contains a pin A13, which passes through the hole in the annular guard B221. The inner diameter of the hole is larger than the outer diameter of the pin A13 but smaller than the outer diameter of the pin tooth A11. The purpose of this design is to allow the pin tooth A11 to roll freely on the tooth surface of the tooth groove B21 without falling off.

[0058] Figure 10 This is an axial cross-sectional schematic diagram of an application example of the second scheme of this invention. The input shaft G1 is located in the center of the diagram, with its axis at O2. The input shaft G1 is mounted on a fixed disk D connecting to the outer casing via a first bearing F1. The fixed disk D connecting to the outer casing is fixedly connected to a second fixed-axis gear A2 via a positioning bolt K1. The second fixed-axis gear A2 contains pin teeth A11, and pins A13 are located at both ends of the pin teeth A11. The pins A13 pass through holes in the first annular guard B221 and the second annular guard B222, and both annular guards at both ends are fixed to the fixed-axis gear.

[0059] Figure 10 An eccentric shaft E1 is fixedly connected to the middle of the input shaft G1. A bushing K2 is mounted on the eccentric shaft E1 via a second bearing F2. The outer layer of the bushing K2 is fixedly connected to the axial revolving gears B2 and B1. The axial revolving gear B2 meshes with the fixed-axis gear A2 via pin teeth A11, and the axial revolving gear B1 meshes with the fixed-axis gear A1. The fixed-axis gear A1 is connected to the axial gear A2 via bearing F3 and to the input shaft G1 via bearing F4. The pin teeth of the fixed-axis gear A1 have annular guards B223 (No. 3) and B224 (No. 4) at both ends, similar in structure to those of the fixed-axis gear B2. The annular guards, pin teeth, and tooth grooves in the fixed-axis gears are also similar in structure. Figure 9 Similar to the case study.

[0060] Figure 10 According to the second embodiment of this invention, the input shaft G1 is further equipped with eccentric counterweight T1 (number 1) and eccentric counterweight T2 (number 2). When the input shaft G1 drives the axial rotation gears B1 and B2 via the eccentric shaft E1, it also drives the eccentric counterweights T1 and T2 to rotate. Figure 10 As can be seen, the center of gravity of the planetary gears B1 and B2 is located at the axis O1, which is relatively close to the axis O2 of the input shaft. The center of gravity of the eccentric counterweights T1 and T2 is farther from the axis O2 of the input shaft. Furthermore, when the gear thickness is much smaller than the gear radius, the eccentricity of the counterweights can be larger. Therefore, the weight of the eccentric counterweights T1 and T2 can be less than the weight of the planetary gears B1 and B2. Figure 5 In the case comparison, the weight of the eccentric counterweight can even be less than... Figure 5 The weight of pin B12 and pin disc B121 in the case allows for a larger overall torque-to-weight ratio in the transmission.

[0061] Figure 11 This is a schematic diagram of the axial cross-section of an application example of the third solution according to the present invention. Figure 11 The case is in Figure 10 Based on the existing A1-B1 and A2-B2 gear sets, a third gear set has been added: a fixed-axis gear A3 and a central revolving gear B3. Among the added gears, the fixed-axis gear A3 is also a pin gear. Figure 10 The pin gear structure is similar, and the pin teeth of the A3 gear are also equipped with ring-shaped guards B225 and B226 at both ends.

[0062] Figure 11 In the example, when fixed-axis gear A2 is fixed and fixed-axis gear A1 drives the output shaft, it can be used as the first gear. When fixed-axis gear A3 is fixed and fixed-axis gear A1 drives the output shaft, it can be used as the second gear. When fixed-axis gear A2 is fixed and fixed-axis gear A3 drives the output shaft, the reduction ratio is greater, and it can be used as the third gear. Of course, the gear fixing order and gear positions can also be used in other combinations as needed by professionals in this industry.

[0063] Figure 12 This is a schematic diagram of an eccentric counterweight component. The eccentric counterweight component body T11 is fixedly mounted on the eccentric shaft E1, which is mounted on the input shaft G1 via an adjustable latch T31 ​​and rotated by the input shaft G1. The adjustable latch T31 ​​allows for adjustment and positioning of the eccentric counterweight component's axial position during assembly, and can also be used to adjust the shock absorption effect of each product. At both ends of the eccentric counterweight component body T11 are blades T3. When the blades T3 rotate with T11, centrifugal force forces the surrounding fluid to flow radially, thereby cooling and lubricating the various friction surfaces in the transmission.

[0064] Figure 13 This is a schematic diagram of the second type of eccentric counterweight component. The eccentric counterweight component body T11 is fixedly mounted on the input shaft G1. At both ends of the eccentric counterweight component body T11 are fluid inlets T21 and fluid outlets T22, which are interconnected. Their function is that when T11 rotates, lubricating oil or air enters the fluid inlet T21 and exits through the fluid outlet T22. The fluid outlet T22 faces the gears, thereby cooling and lubricating the various friction surfaces in the transmission.

[0065] Figure 13 In the eccentric counterweight body T11, a movable counterweight block T12 is also installed. A spring T13 connects the movable counterweight block T12 and the eccentric counterweight body T11. The purpose of this structure is to allow the movable counterweight block T12 to swing within a certain range. When using pin gears, and each pin gear meshes with only one axis-centric revolving gear, the pin teeth in the pin gear can roll in the tooth grooves of the pin gear housing. The inertial motion of the pin teeth will cause their angular momentum to be delayed and act on the transmission. The movable counterweight block T12 is to counteract this angular momentum delay effect of the pin teeth.

[0066] Other specific examples can be designed based on the principle of the multi-gear low-tooth-difference transmission for eliminating vibration disclosed in this application of the present invention, which will not be listed hereafter.

Claims

1. A multi-gear transmission with a small tooth difference that eliminates vibration, characterized in that: It comprises at least two axially rotating gears and at least one eccentric counterweight, wherein the axially rotating gears are offset in the axial direction, and the angular momentum of the axially rotating gears can be equivalent to and symmetrical with the angular momentum of the eccentric counterweight in the same plane; the implementation methods include: a first method in which the absolute values ​​of the angular momentum of the axially rotating gears are not equal, and the eccentric counterweight is located on the outside of the axially rotating gear with the larger angular momentum; a second method in which the axially rotating gears are fixedly connected, and the eccentric counterweight is located on both sides of the axially rotating gears; a third method in which the axially rotating gears are fixedly connected, and the eccentric counterweight is located between the axially rotating gears; including one of the above methods, or a combination of the above methods.

2. The transmission according to claim 1, characterized in that: It contains at least four gears. Gears A1 and A2 are fixed-axis gears, and gears B1 and B2 are axial revolving gears. The axial revolving gears mesh with the fixed-axis gears, driven by eccentric bearings. Gear A1 meshes with gear B1, and gear A2 meshes with gear B2. Gear B1 has N teeth, gear A1 has N+P teeth, gear B2 has N+M teeth, and gear A2 has N+M+Q teeth, where P≥1, M≥1, and Q≥1. The gears can be combined in one of two ways: the first way is that the fixed-axis gears are fixed together, and one of the axial revolving gears can be fixed and not rotate, while the unfixed axial revolving gear can drive the output shaft; the second way is that the axial revolving gears are fixed together in the direction of rotation, and one of the fixed-axis gears can be fixed and not rotate, while the unfixed fixed-axis gear can drive the output shaft.

3. The transmission according to claim 1, characterized in that: The eccentric counterweight component includes a structure that forces the flow of surrounding fluid, including lubricant or air, to serve for lubrication and heat dissipation.

4. The transmission according to claim 2, characterized in that: Based on the four gears, add at least two more gears: a fixed-axis gear A3 and a central rotating gear B3. The central rotating gear B3 is driven by an eccentric shaft to mesh with the fixed-axis gear A3. The number of teeth of the central rotating gear B3 is N+P, and the number of teeth of the fixed-axis gear A3 is N+P+1, where P≥1. Alternatively, multiple sets of gears can be added in this way.

5. The transmission according to claim 1, characterized in that: The gears in the transmission include a pin gear, which includes a pin gear housing and pin teeth. The pin gear housing contains tooth grooves, and the pin teeth are rolling bodies that can roll along the tooth surface of the tooth grooves in the pin gear housing. The transmission also includes a device to prevent the pin teeth from falling out of the tooth grooves of the pin gear. The device to prevent the pin teeth from falling out may include one or a combination of more than one of baffles, flanges, and annular guards.

6. The transmission according to any one of claims 1, 2, 3, 4, or 5, characterized in that: The eccentric counterweight component can be adjusted and positioned axially during installation to accommodate the different angular momentum of the rotating gears of different batches; or the gearbox includes a pin gear, the pin teeth of which can roll in the tooth grooves of the pin gear housing; the eccentric counterweight component includes a movable counterweight block that can move relative to the eccentric counterweight component body; or the movable counterweight block is restricted to a certain range of movement; or the movable counterweight block is connected to the eccentric counterweight component by an elastic component; having one or a combination of the above structures.

7. The transmission according to claim 1, 2, 3, 4, or 5, characterized in that: It also includes a clutch C, the design methods of which include: the first method is to select and fix the rotation of one of the gears by switching the clutch C, and output by the other gear; the second method is to switch the output gear ratio according to the external load or machine needs, including one of the above two clutch design methods, or a combination of the two methods.

8. The transmission according to any one of claims 1, 2, 3, 4, or 5, characterized in that: The design schemes for the tooth profile curve of the tooth groove or the tooth profile curve of the gear include the following schemes: at least one tooth profile curve contains a circular arc curve, or at least one tooth profile curve contains a cycloid, or at least one tooth profile curve contains an elliptical curve, or at least one tooth profile curve contains an elliptic cycloid, or at least one tooth profile curve contains a straight line, or at least one tooth profile curve contains an involute, or at least one tooth profile curve contains a conic curve; or a combination of one or more of the above schemes.

9. The transmission according to any one of claims 1, 2, 3, 4, or 5, characterized in that: The transmission works in conjunction with the engine to form a power output module. The engine may include an axial flux motor, an external rotor motor, an internal rotor motor, or a turbine.

10. The transmission according to any one of claims 1, 2, 3, 4, or 5, characterized in that: The transmission is used in a transmission system, and its application scenarios include, but are not limited to, propeller drive, hub drive, or mechanical actuators. The mechanical actuators can be used in, but are not limited to, rotary joints of industrial machinery, wearable robotic arms, reciprocating transport machines, embodied intelligent machines, vehicle steering systems, aircraft steering rudder transmission systems, ship steering rudder transmission systems, or the transmission can be used in electric vehicles or electric construction machinery.

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