Transmission with two pendulum shaft gears generating two multiplier transmission ratios
Through the design of four gears and clutch switching, the swing shaft gear transmission achieves multiple gear ratios and torque carrying capacity in a small space, solving the problems of limited space and torque in traditional transmissions, and is suitable for mechanical systems with varying loads.
Patent Information
- Authority / Receiving Office
- CN · China
- Patent Type
- Utility models(China)
- Current Assignee / Owner
- 王踊
- Filing Date
- 2025-07-04
- Publication Date
- 2026-05-01
AI Technical Summary
Existing swing shaft gear transmissions have limited transmission ratios in small spaces and insufficient torque carrying capacity, which cannot meet the application requirements of load changes.
It adopts a four-gear structure, in which two fixed-axis gears and two sway-axis gears are driven to mesh through eccentric bearings. Combined with clutch switching, two multiplier transmission ratios are achieved, which increases the torque carrying capacity. The gear combination is improved by needle gear design and retainer to balance angular momentum.
It achieves two gear ratios within a relatively small space, improving torque carrying capacity, adapting to load changes, and making it suitable for complex road conditions and heavy machinery.
Smart Images
Figure CN224187964U_ABST
Abstract
Description
A transmission with two sway shaft gears producing two multiplier gear ratios Technical Field
[0001] This utility model relates to a transmission device for a gearbox and belongs to the field of mechanical technology. Background Technology
[0002] Currently known cycloidal gearboxes typically consist of a cycloidal gear meshing with a pin gear, or a pin gear meshing with a cycloidal gear. Two gears with the same number of teeth rotate 180 degrees around the axis of a third gear. The transmission is achieved by the different number of teeth on the cycloidal and fixed-axis gears, usually with only one fixed gear ratio. Because each tooth of the pin gear or pin gear has a necessary radial structure, each tooth requires a large radial space. Therefore, the number of teeth is limited for the same diameter, thus limiting the reduction ratio. Furthermore, since the necessary path for torque transmission in a pin gear or pin gear is the pin tooth or pin, and the pin tooth or pin is only fixed to the body at both ends—meaning it is in a suspended state and needs to transmit the main torque—the torque that each tooth can withstand is relatively small.
[0003] To address the aforementioned problems, this utility model is proposed. The purpose of this utility model is to provide a transmission that achieves a large transmission ratio within a relatively small space, and can output two gears with different transmission ratios, while simultaneously improving the transmission's torque-carrying capacity, thus providing a wider range of applications for mechanical transmissions. Summary of the Invention
[0004] A transmission with two sway-shaft gears generating two multiplier transmission ratios is characterized by comprising at least four gears, where gears A1 and A2 are fixed-axis gears, and gears B1 and B2 are sway-shaft gears. The sway-shaft gears mesh with the fixed-axis gears, driven by an eccentric bearing. Gears A1 mesh with gear B1, and gears A2 mesh with gear B2. Gear B1 has N teeth, gear A1 has N+1 teeth, gear B2 has N+1 teeth, and gear A2 has N+2 teeth. The gears are arranged in two ways: First, gears A1 and A2 are fixedly connected, and one of gears B1 and B2 can be fixed in the circumferential direction without rotation, while the other can drive the output shaft. Second, gears B1 and B2 are fixedly connected in the rotational direction, and one of gears A1 and A2 can be fixed without rotation, while the other can drive the output shaft.
[0005] The obvious advantage of the above design is that it achieves a large transmission ratio within a relatively small space.
[0006] Furthermore, the so-called transmission with two swing shaft gears generating two multiplier transmission ratios also includes a clutch C. The clutch C has two installation methods. The first installation method is to install it on the fixed end of the housing, and the switching of the clutch C selects between the fixed swing shaft gear and the fixed shaft gear. The second installation method is to install it on the output shaft end, and the switching of the clutch C selects between the output from the swing shaft gear and the output from the fixed shaft gear. This includes a combination of the above two gear combinations and the two clutch installation methods.
[0007] The obvious advantage of the above design is that it achieves two gear ratios within a relatively small space.
[0008] Furthermore, it includes a clutch C; the clutch C has two installation methods. The first installation method is to install it on the fixed end of the housing, and to select the fixed shaft gear A2 or the swing shaft gear B2 by switching the clutch C; the second installation method is to install it on the output shaft end, and to select whether the output is from the fixed shaft gear A1 or the swing shaft gear B1 by switching the clutch C. The clutch C can switch the output gear ratio of the transmission according to the external load or machine needs.
[0009] The purpose of the above design is to realize two gear positions that are multiplicative and have the same direction, i.e., n 2 The transmission ratios of 1 and n:1, or [n×(n+2)]:1 and n:1, have the obvious advantage of being applicable 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 different road conditions.
[0010] When a large transmission ratio is achieved with a smaller number of teeth, the torque carrying capacity of each tooth must be improved. Therefore, as a necessary technical solution, at least two of the four gears are pin gears. Preferably, gears A1 and A2 are pin gears. The space between the pin gear sleeve and the pin gear housing has teeth fixed on the pin gear housing. The minimum radius of curvature of the tooth profile curve of the tooth on the pin gear housing is R1, and the outer radius of the pin gear sleeve in contact with it is R2. Then the inequality R1≥1.06×R2 is satisfied.
[0011] The obvious advantage of the above design is that torque can be directly transmitted between the pin sleeve and the pin housing, without needing to pass through suspended pins, thus increasing the torque-carrying capacity of each tooth. The main function of the pins is to position the pin sleeve, so the pins can be subjected to very little force. Furthermore, the pin design does not need to worry about both transmitting torque and being suspended, so the diameter of the pins can be reduced, the thickness of the pin sleeve can be increased, and the torque-carrying capacity of each tooth can be further increased.
[0012] As an improvement, a retainer is installed between the needle tooth sleeve cavity and the needle tooth in the pin gear. The retainer is made of copper-containing materials or plastic. The function of the retainer is to maintain the relative position of the needle tooth and the needle tooth sleeve, preventing collision between them. The reason for this design is that, due to the increase in torque, the deformation of parts such as gears, crankshafts, and needle tooth sleeves increases, and the risk of deformation and collision between the needle tooth and the needle tooth sleeve increases. Adding a retainer can extend the service life of these parts.
[0013] Alternatively, in this invention, at least two gears may have a number of teeth that is not a prime number, or all gears may have a number of teeth that is not a prime number. The purpose of having a non-prime number of teeth is to allow the gears to be processed quickly using a multi-head grinding machine, thereby reducing production costs.
[0014] According to the transmission disclosed in this utility model, there is a further improvement: the angular momentum balance of the transmission is maintained by setting different counterweights or different center-of-gravity distances between the swing shaft gear B1 and the swing shaft gear B2.
[0015] As an improvement, the design of the pendulum gears B1 and B2 includes one of the following two design methods: The first design method involves pin holes on both pendulum gears B1 and B2, with pins on the pin disk fixing the rotation of the pendulum gears circumferentially. The contact surface between the pin hole and the pin includes a conical surface or a step, capable of withstanding a certain axial force. Furthermore, the diameter of the pin hole on pendulum gear B2 is larger than the diameter of the pin hole on pendulum gear B1. The second design method involves cross slide rails on both pendulum gears B1 and B2, with a cross slide rail disk fixing the rotation of the pendulum gears circumferentially. The contact surface between the cross slide rail and the cross slide rail disk includes a conical surface or a step, capable of withstanding a certain axial force. The angular momentum of pendulum gears B1 and B2 is balanced by the size and configuration of the cross slide rails or the holes through which they pass on the gears.
[0016] One advantage of the above design is that it achieves angular momentum balance without the need for additional components. Another advantage of the pin and pin hole or the cross slide rails being able to withstand axial forces is that the direction of the axial force can be matched with the bearings in the transmission. The bearings in the transmission can be tapered bearings or angular contact bearings. The purpose of this design is to enable the entire transmission to withstand axial impact forces with fewer parts.
[0017] As a design improvement, the design schemes for the tooth profile curves of the teeth meshing with the needle sleeve on the needle tooth housing of the needle gear and the tooth profile curves of the sway shaft gear include the following schemes: at least one gear tooth profile curve contains a circular arc curve, or at least one gear tooth profile curve contains a cycloid, or at least one gear tooth profile curve contains an elliptical curve, including one or more of the above schemes.
[0018] The transmission of this utility model 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.
[0019] The transmission of this invention can be applied in various mechanical transmission systems, including but not limited to applications in driving propellers, driving vehicles, or mechanical joints. The mechanical joints include but are not limited to rotary joints of industrial machinery, wearable robotic arms, intelligent mechanical devices, robotic hands, vehicle steering systems, aircraft steering rudder transmission systems, or ship steering rudder transmission systems.
[0020] The technical advantages and scope of protection of this utility model will become clearer with reference to the accompanying drawings. The drawings and their descriptions are for illustrative purposes only and the specific examples therein do not limit the scope of protection of this utility model. Attached Figure Description
[0021] Symbol explanations in the diagram: A1—Fixed-axis gear No. 1; A11—Pin tooth sleeve; A12—Pin tooth; A121—Retainer; A2—Fixed-axis gear No. 2; A3—Teeth on the pin tooth housing; B1—Swing shaft gear No. 1; B2—Swing shaft gear No. 2; B12—Pin shaft; B121—Pin shaft disc; B13—Conical step of the pin shaft; B131—Conical opening of the pin hole on the swing shaft gear; B21—Cross slide rail No. 1; B211—Cross slide rail disc No. 1; B22—Cross slide rail No. 2; B221— 2nd cross slide rail disc; C—clutch; D—fixed disc connecting the housing; E1—1st eccentric shaft; E2—2nd eccentric shaft; Input shaft—G1; Output shaft—G2; F1—first bearing; F2—second bearing; F3—third bearing; F4—fourth bearing; F5—fifth bearing; F6—sixth bearing; H1—first section position; H2—second section position; O—axis center position; K—connecting ring; R1—tooth profile curve radius of tooth A3 on the needle tooth housing; R2—needle tooth sleeve radius.
[0022] Figure 1 is an axial cross-sectional view of the first type of transmission with two swing shaft gears generating two multiplier transmission ratios according to the structural design disclosed in this utility model.
[0023] Figure 2 is a radial cross-sectional view of the second cross-section position H2 on the right in Figure 1.
[0024] Figure 3 is a radial cross-sectional view of the first cross-section position H1 on the left in Figure 1.
[0025] Figure 4 is an axial cross-sectional view of a second type of transmission with two swing shaft gears generating two multiplier transmission ratios, based on the structural design disclosed in this utility model.
[0026] Figure 5 is a radial cross-sectional view of the first cross-section position H1 on the left in Figure 4.
[0027] Figure 6 is a radial cross-sectional view of the second cross-section position H2 on the right in Figure 4.
[0028] Figure 7 is a radial cross-sectional view of the part where the needle sleeve meshes with the gear.
[0029] Figure 8 is a radial cross-sectional view of a needle sleeve including a retainer according to the present invention.
[0030] Figure 9 is a radial cross-sectional view of another type of needle sleeve including a retainer in this utility model. Detailed Implementation
[0031] The advantages and preferred embodiments of this utility model are described with reference to the accompanying drawings. Specific examples are provided to illustrate the advantages of this utility model and should not be construed as limiting the scope of protection of this utility model.
[0032] Figure 1 is an axial cross-sectional view of the first type of transmission with two swing shaft gears generating two multiplier transmission ratios according to the structural design disclosed in this utility model. In the middle of the figure is the input shaft G1, with its axis at O2. The input shaft G1 is mounted on a pin disk B121 via a first bearing F1. Pin disk B121 has a pin shaft B12. A second fixed shaft 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 shaft gear A2 or the inner pin disk B121. In the figure, clutch C is connected to the second fixed shaft gear A2, so the second fixed shaft gear A2 is fixed and does not rotate, while the pin disk B121 can rotate.
[0033] In Figure 1, the input shaft G1 is also equipped with a second eccentric shaft E2, and the second eccentric shaft E2 is mounted with a second pendulum gear B2 via a third bearing F3, with its axis being O1. The second pendulum gear B2 meshes with the pinion sleeve on the second fixed shaft gear A2.
[0034] In Figure 1, the left side of fixed-axis gear A2 (No. 2) is connected to fixed-axis gear A1 (No. 1) via bearing F4. Fixed-axis gear A1 (No. 1) is connected to input shaft G1 via bearing F6, and output shaft G2 is fixedly connected to it. Input shaft G1 also has eccentric shaft E1 (No. 1) mounted on it. E1 is connected to sway gear B1 (No. 1) via bearing F5, with its axis being O3. Sway gear B1 (No. 1) meshes with fixed-axis gear A1 (No. 1) via pin sleeve A11 and pin teeth A12.
[0035] As shown in Figure 1, pin B12 passes through the No. 2 swing gear B2 and the No. 1 swing gear B1, with the portion on the right passing through the No. 2 swing gear having a larger diameter. The first function of this design is that the different diameters of the holes through which pin B12 passes through the No. 1 and No. 2 swing gears allow for adjustment of the weight of the two gears and balance of their angular momentum. The second function is to create a step or ramp between the pin and the gears to withstand axial thrust. For example, in Figure 1, the tapered opening B131 of the pin hole on the No. 1 swing gear B1 engages with the tapered step B13 of the pin. This engagement can withstand the axial thrust from left to right from the swing gear. The third bearing F3 and the fifth bearing F5 can withstand the axial thrust from right to left from the swing gear, while 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 forces or vibrations, and the structure is simple and compact. Each bearing can be an angular contact bearing or a tapered bearing.
[0036] Figure 2 is a radial cross-sectional view of position H2 on the right side of Figure 1. The center of the figure is the input shaft G1, with its axis 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 O1 revolves around axis O2 or performs a pendulum motion. A pendulum gear B2 (position 2) is mounted on the outer ring of bearing F3 and meshes with the outermost fixed-axis gear A2 (position 2). Fixed-axis gear A2 is a pin gear, containing pin teeth A12 and pin tooth sleeves A11. Pin B12 passes through a hole in pendulum gear B2.
[0037] Figure 3 is a radial cross-sectional view of the first section H1 on the left side of Figure 1. The center of the figure is the input shaft G1, whose axis is O2. The input shaft G1 is fixedly connected to the outer periphery of eccentric shaft E1. The outer layer of eccentric shaft E1 is connected to the 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 pendulum gear B1, which meshes with the outermost fixed-axis gear A1. Fixed-axis gear A1 is a pin gear, containing pin teeth A12 and pin tooth sleeves A11. Pin B12 passes through a hole in pendulum gear B1.
[0038] The operation and advantages of this utility model will be explained below with reference to Figures 1, 2, and 3. 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 swing shaft gear B2 meshes with the second fixed shaft gear A2. Since the second fixed shaft gear A2 is fixed to the fixed disk D connecting the outer casing by the clutch C and cannot rotate, the pin disk B121 will push the second swing shaft gear to rotate counterclockwise by 1 / 15 revolution. The pin B12 on the pin disk B121 also passes through the first swing shaft gear B1, forcing the first swing shaft gear B1 to rotate counterclockwise by 1 / 15 revolution as well. Because the first swing shaft gear B1 is simultaneously meshing clockwise with the first fixed shaft gear A1, the first fixed shaft gear A1 rotates clockwise by 1 / 225 revolution, resulting in a reduction ratio of 225:1.
[0039] When clutch C releases the second fixed shaft gear A2 and switches to the fixed pin disk B121, the first swing shaft gear B1 is also fixed in the circumferential direction and cannot rotate. Then the input shaft G1 rotates clockwise 1 turn, and the first fixed shaft gear A1 rotates clockwise 1 / 15 turn. The reduction ratio is 15:1, and the direction of rotation is the same as when clutch C fixes the second fixed shaft gear A2.
[0040] Figure 4 is an axial cross-sectional view of the second type of transmission with two swing shaft gears generating two multiplier transmission ratios, based on the structural design disclosed in this utility model. In the middle of the figure is the input shaft G1, with its axis at O2. The input shaft G1 is mounted with the second cross slide rail plate B221 via the first bearing F1. The left side of the second cross slide rail plate B221 is mounted with the second cross slide rail B22, which is connected to the second swing shaft gear B2 on the left side. The outer side of the second cross slide rail plate B221 is mounted with the second fixed shaft gear A2 via the second bearing F2. A clutch C is mounted on the left side of the fixed disk D connecting to the outer casing. The clutch C can connect to the outer second fixed shaft gear A2 or the inner second cross slide rail plate B221. In the figure, the clutch C is connected to the second cross slide rail plate B221, so the second swing shaft gear B2 cannot rotate, while the second fixed shaft gear A2 can rotate.
[0041] In Figure 4, the input shaft G1 is also equipped with a second eccentric shaft E2, and the second eccentric shaft E2 is mounted with a second pendulum gear B2 via a third bearing F3, with its axis being O1. When the input shaft G1 rotates, the second pendulum gear B2 meshes with it through the pin tooth sleeve on the second fixed shaft gear A2.
[0042] In Figure 4, the left side of the fixed-axis gear A2 (No. 2) is fixedly connected to the fixed-axis gear A1 (No. 1) via a connecting ring K. The fixed-axis gear A1 (No. 1) is connected to the output shaft G2 via the fifth bearing F5. The output shaft G2 is connected to the input shaft G1 via the sixth bearing F6. The input shaft G1 also houses the eccentric shaft E1 (No. 1), which is connected to the sway gear B1 (No. 1) via the fourth bearing F4. The sway gear B1 (No. 1) meshes with the fixed-axis gear A1 (No. 1) via a pin sleeve A11 and pin teeth A12.
[0043] In Figure 4, a cross slide rail B21 is installed to the left of the first swing shaft gear B1, and a cross slide rail disc B211 is installed to the left of the first cross slide rail B21. The first cross slide rail disc B211 is fixedly connected to the output shaft G2. The angular momentum of the two swing shaft gears is balanced by adjusting the mass distribution of the first and second cross slide rails and their accessories. In addition, there can be a step or ramp between the cross slide rails and the gears to withstand the axial thrust.
[0044] Figure 5 is a radial cross-sectional view of the first section H1 on the left side of Figure 4. The center of the figure is the input shaft G1, whose axis is O2. The input shaft G1 is fixedly connected to the outer periphery of eccentric shaft E1. The outer layer of eccentric shaft E1 is connected to the fourth bearing F4, whose axis is O3. Axis O3 revolves around axis O2 or performs a pendulum motion. The outer ring of the fourth bearing F4 is fitted with pendulum gear B1, which meshes with the outermost fixed-axis gear A1. Fixed-axis gear A1 is a pin gear, containing pin teeth A12 and pin tooth sleeves A11. Cross slide rail B21 passes through a hole in pendulum gear B1.
[0045] As shown in Figure 5, the structure at cross slide rail B21 (number 1) allows the first pendulum gear B1 to slide up and down when cross slide rail B21 is stationary. Cross slide rail B21 and cross slide rail disk B211 share a 90° rotational connection, allowing the first pendulum gear B1 to move laterally along cross slide rail B211. Furthermore, the axis O3 of the first pendulum gear B1 can revolve around the axis O2 of the first fixed-axis gear A1 or perform a pendulum motion. However, when the cross slide rail disk is not rotating, the first pendulum gear B1 also cannot rotate. This design allows for greater torque capacity and minimal backlash.
[0046] Figure 6 is a radial cross-sectional view of position H2 on the right side of Figure 4. The center of the figure is the input shaft G1, with its axis 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 O1 revolves around axis O2 or performs a pendulum motion. The outer ring of the third bearing F3 is fitted with pendulum gear B2 (position 2), which meshes with the outermost fixed-axis gear A2 (position 2). Fixed-axis gear A2 is a pin gear, containing pin teeth A12 and pin tooth sleeves A11. Cross slide rail B22 (position 2) passes through a hole in pendulum gear B2 (position 2).
[0047] The operation and advantages of the second embodiment of this utility model are explained below with reference to Figures 4, 5, and 6. The third bearing F3 and the fourth bearing F4 are symmetrically arranged at 180 degrees. When the input shaft G1 rotates clockwise one revolution, the second swing shaft gear B2 meshes with the second fixed shaft gear A2. Because the second cross slide rail plate B221 is fixed to the fixed plate D connecting the outer casing by the clutch C and cannot rotate, the second swing shaft gear B2 cannot rotate around the axis O1. The second fixed shaft gear A2 will rotate clockwise 1 / 16 revolution. The connecting ring K connects the second fixed shaft gear and the first fixed shaft gear, and the first fixed shaft gear A1 is also forced to rotate clockwise 1 / 16 revolution. Since the first swing shaft gear B1 is simultaneously meshing clockwise with the first fixed shaft gear A1, the first swing shaft gear B1 rotates counterclockwise 1 / 224 revolution, resulting in a reduction ratio of 224:1.
[0048] When clutch C releases the No. 2 cross slide rail plate B221 and switches to fix the No. 2 fixed shaft gear A2, the No. 1 fixed shaft gear A1 is also fixed. Then the input shaft G1 rotates clockwise 1 turn and the No. 1 swing shaft gear B1 rotates counterclockwise 1 / 14 turn. The reduction ratio is 14:1, and the direction of rotation is the same as when clutch C fixes the No. 2 fixed cross slide rail plate B221.
[0049] Figure 7 is a radial cross-sectional view of the meshing part between the pin sleeve and the gear. The left side of the figure shows the sway gear B1, and the right side shows the fixed-axis gear A1. The fixed-axis gear A1 is a pin gear, containing pin teeth A12, which are surrounded by a pin sleeve A11. The fixed-axis gear A1 also contains teeth A3 on the pin housing. When the pin sleeve A11 meshes with the sway gear B1, it also meshes with teeth A3 on the pin housing. In the example shown in the figure, the tooth profile curve of tooth A3 on the pin housing is a circular arc curve, where R1 is the radius of the circular arc curve; R2 is the radius of the pin sleeve. In this case, R1 = R2 × 1.082, which meets the requirement of R1 ≥ 1.06 × R2 in this utility model. The first advantage of this design is that it can improve the torque carrying capacity of each tooth; the second advantage is that it ensures rolling contact of the meshing surface under impact force; and the third advantage is that, compared with traditional pin gears where the inner side of the pin sleeve rubs against the pin teeth, this design is beneficial for chip removal during long-term use.
[0050] Figure 8 is a radial cross-sectional view of a needle tooth sleeve including a retainer according to this utility model. The needle tooth A12 is in the center of the figure. The needle tooth A12 does not contact the inner ring of the needle tooth sleeve A11. A retainer A121 is located within the inner ring of the needle tooth sleeve A11. The function of the retainer A121 is to prevent the needle tooth A12 from colliding with the needle tooth sleeve A1. The retainer A121 can be made of a copper alloy or plastic.
[0051] Figure 9 is a radial cross-sectional view of another type of needle sleeve including a retainer in this invention. The needle tooth A12 is in the center of the figure. Inside the inner ring of the needle sleeve A11 is a retainer A121. The function of the retainer A121 is to prevent the needle tooth A12 from colliding with the needle sleeve A1. The retainer A121 is not closed in the circumferential direction. The advantage of this design is that when the thermal expansion coefficient of the retainer material differs significantly from that of the needle sleeve, and the operating temperature is high, the retainer can still function.
[0052] Based on the transmission structure proposed in this utility model, which generates two multiplier transmission ratios using two swing shaft gears, other specific cases can also be designed. Application scenarios include, but are not limited to, combining with a turbine shaft engine to form a power module, or combining with an electric motor to form a power module. The electric motor can be an axial flux motor, an external rotor motor, or an internal rotor motor.
[0053] As can be seen from the above examples, the transmission of this utility model, which uses two swing shaft gears to generate two multiplier transmission ratios, can achieve n within a relatively small space. 2 The transmission ratio can be either n:1 or n:1, or [n×(n+2)]:1 or n:1, with the two gears rotating in the same direction. This allows for a wider range of applications, such as driving vehicles to adapt to various complex road conditions, driving robots to achieve explosive power and endurance for running and jumping, or reciprocating transport robots.
[0054] This utility model is applied in mechanical transmission devices, and its application scenarios include, but are not limited to, propeller drives, hub drives, or mechanical joints. Among them, mechanical joints include, but are not limited to, rotary joints of industrial machinery, wearable robotic arms, human-machine collaborative mechanical devices, robotic hands, vehicle steering systems, aircraft steering rudder transmission systems, ship steering rudder transmission systems, or many other application scenarios, or other mechanical actuators, which will not be listed further.
Claims
1. A transmission with two sway shaft gears generating two multiplier transmission ratios, characterized in that: It contains at least four gears. Gears A1 and A2 are fixed-axis gears, and gears B1 and B2 are sway-axis gears. The sway-axis gears are driven by an eccentric bearing to mesh with the fixed-axis gears. Gear A1 meshes with gear B1, and gear A2 meshes with gear B2. Gear B1 has N teeth, gear A1 has N+1 teeth, gear B2 has N+1 teeth, and gear A2 has N+2 teeth. The gears can be combined in one of the following two ways: The first way is that gears A1 and A2 are fixedly connected, and one of gears B1 and B2 can be fixed and not rotate, while the other can drive the output shaft. The second way is that gears B1 and B2 are fixedly connected in the direction of rotation, and one of gears A1 and A2 can be fixed and not rotate, while the other can drive the output shaft.
2. The transmission according to claim 1, characterized in that: At least of them There are two pin gears. The space between the pin gear's pin sleeve and pin gear shell has teeth fixed on the pin gear shell. The minimum radius of curvature of the tooth profile curve of the tooth on the pin gear shell is R1, and the outer radius of the pin sleeve that contacts it is R2. Then the inequality R1≥1.06×R2 is satisfied.
3. The transmission according to claim 1, characterized in that: The four gears include a pin gear. A retainer is installed between the inner cavity of the pin gear sleeve and the pin teeth. The retainer is used to maintain the relative position of the pin teeth and the pin gear sleeve. The retainer is made of copper-containing materials or plastic.
4. The transmission according to claim 1, characterized in that: There are at least two gears whose number of teeth is not a prime number, or there are no prime numbers among the gears.
5. The transmission according to claim 1, characterized in that: Among them, It includes a clutch C, which has two installation methods. The first installation method is to install it on the fixed end of the housing, and the rotation of the fixed shaft gear or the swing shaft gear is selected by switching the clutch C. The second installation method is to install it on the output shaft end, and the output is selected by the fixed shaft gear or the swing shaft gear by switching the clutch C. The clutch C can switch the output gear ratio according to the external load or machine needs, and includes one of the above two clutch installation methods.
6. The transmission according to claim 1, characterized in that: The angular momentum balance of the transmission is maintained by setting different counterweights or different center-of-gravity distances between the swing shaft gear B1 and the swing shaft gear B2.
7. The transmission according to claim 1, characterized in that: The design of the pendulum gears B1 and B2 includes one of the following two design methods: The first design method involves pin holes on both pendulum gears B1 and B2, with pins on the pin disk fixing the rotation of the pendulum gears circumferentially. The contact surface between the pin hole and the pin includes a conical surface or a step, capable of withstanding a certain axial force. Furthermore, the diameter of the pin hole on pendulum gear B1 is larger than the diameter of the pin hole on pendulum gear B2. The second design method involves cross slide rails on both pendulum gears B1 and B2, with a cross slide rail disk fixing the rotation of the pendulum gears circumferentially. The contact surface between the cross slide rail and the cross slide rail disk includes a conical surface or a step, capable of withstanding a certain axial force. The angular momentum of pendulum gears B1 and B2 is balanced by the size and configuration of the cross slide rails or the holes through which they pass on the gears.
8. The transmission according to claim 1, characterized in that: wherein the gears The design schemes for the tooth profile curves of gears A1 and A2, and gears B1 and B2 include the following schemes: at least one gear tooth profile curve contains a circular arc curve, or at least one gear tooth profile curve contains a cycloid, or at least one gear tooth profile curve contains an elliptical curve, including one or more of the above schemes.
9. The transmission according to claim 1, characterized in that: The two swing shaft gears generate two multiplier transmission ratios, which work together 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, 5, 6, 7, 8, or 9, characterized in that: The two sway shaft gears that generate two multiplier transmission ratios are used in transmission systems. Their application scenarios include, but are not limited to, propeller drives, hub drives, or mechanical actuators. Among them, mechanical actuators include, 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, or ship steering rudder transmission systems.