Ultra-high reduction speed changer of double differential reduction gear

Abstract

A housing (18) and a first input shaft (I A1 At least a first input (20) rotatable about a first output shaft (O A The transmission has at least a first output (26) rotatable about a first output shaft (11). The transmission further includes a first planetary gear set including a first outer gear (14), a first inner planetary gear (11), and a first outer planetary gear (15) connected to and drivable by the first input, the first outer planetary gear being rigidly connected to and axially aligned with one another. The first planetary gear set is rotatable through the first outer gear (22, 24) and is also rotatable about a first output shaft (23). The transmission further includes a second outer gear (16) meshing with the first outer planetary gear, and a second inner gear (12) connected to the first output and rotatable about the first output shaft through the first inner planetary gear.

Description

Technical Field

[0001] The present invention is directed to a high reduction transmission, specifically a high reduction transmission of a double differential transmission.

Background Art

[0002] The basic function of a high-low reduction transmission is to reduce a high input RPM to a low RPM, for example, to drive the wheels of a vehicle or the rotor of a helicopter. The output RPM of such a transmission is usually in the range of 0 to 1,000 RPM. For example, when the prime mover is an electric motor or a jet engine, the input RPM can be 20,000 RPM or more.

[0003] The state-of-the-art transmissions capable of achieving high reduction include the following. 1. A multi-stage transmission employing cylindrical gears 2. A bevel worm gear reducer with a ratio of 20 in one stage 3. A planetary transmission equipped with an oscillating bevel gear 4. A cycloid transmission

[0004] A multi-stage transmission with cylindrical gears requires a number of shafts with bearings and gears. When the reduction ratio is 20, at least four stages are required. Four reduction stages require four shafts, eight bearings, and four gear meshes. If the efficiency of one single stage is 99.4% (0.994 4 = 0.976), observing only the four gear meshes shows an overall efficiency of 97.6%. A four-stage cylindrical transmission requires a fairly large transmission housing envelope.

[0005] A bevel worm gear drive can be referred to as, for example, a high reduction hypoid (HRH) or a super reduction hypoid (SRH). The worm-shaped pinion typically has 1 to 5 teeth, and the ring gear typically has 27 to 75 teeth. The maximum achievable ratio ranges up to 75. Ratios above about 15 have a reduced backdrive capability. A gear set without backdrive capability is self-locking. A self-locking gear set cannot be used in a vehicle's driveline or in a helicopter's main rotor drive. A bevel worm gear drive also creates high sliding speeds due to large components in the face width direction. For example, an SRH pinion with 5 teeth, when meshed with a 60-tooth ring gear, creates a relative slide of 617 m / min between the sides at a pinion speed of 10,000 RPM (equal to the transmission input speed). This is higher than the maximum slide expected in a hypoid axle drive of a sports car when driving faster than 125 MPH (pinion speed = 4,000 RPM). This example illustrates that doubling the transmission input not only reduces efficiency but also has a risk of surface damage and premature failure.

[0006] A planetary gearbox (e.g., PCT / US2020 / 059810) can achieve very high reduction in the range of about 20 to 100 without generating high relative surface sliding. As the shaft angle between two bevel gears approaches 180°, the relative sliding speed decreases to 0. Due to shaft angles higher than 160° in the most common planetary gearboxes, the relative sliding speed is not significant even when the input speed is 20,000 RPM or higher. The drawbacks of a planetary gearbox are the angular bearing seat of the nutating member and the very large forces applied to the bearings by the angular seat. Another possible cause of problems in a planetary gearbox is the axial inertial force generated by the nutating member. High-speed planetary gearboxes require a mirror image arrangement of an even number of nutating members, as well as precise gear timing and precise balance adjustment.

[0007] A cycloid gearbox is a two-dimensional analog of a peri-cyclic gearbox. One rotation of the eccentric input shaft rotates the output shaft by 1 to 2 tooth pitches. The radial inertial force of the cycloid gearbox cannot be corrected by the arranged configuration of the second cycloid disk. As a result, a high reduction cycloid gearbox is only used when the low input speed is decelerated at a very low output speed.

[0008] If a high ratio of 10 to 100 is to be achieved, the designer may prefer a multi-stage cylindrical gearbox, which is often combined with a planetary gearbox. Multi-stage gearboxes are well known in the art and provide a reasonable power density.

[0009] There is still a need to create a compact high reduction gearbox that is easy to manufacture and has predictable operating conditions. Preferably, if the individual parts involved in the gearbox are themselves known as standard mechanical design parts, the prediction of durability and service life is enhanced by applying calculation algorithms provided by standards such as the American Gear Manufacturers Association (AGMA), the International Standardization Organization (ISO), and / or other national standards. These algorithms rely on hundreds of thousands of fatigue life tests and many application factors that have been evaluated over decades. In safety engineering, such proven algorithms and application factors are valuable tools for developers.

Summary of the Invention

[0010] The present invention is directed to a gearbox comprising a housing, wherein at least a first input is rotatable about a first input shaft and at least a first output is rotatable about a first output shaft. The gearbox further comprises a bevel gear and a first outer gear connected to the first input and drivable by the first input.

[0011] The transmission further includes a first planetary gear set having a first internal planetary gear and a first external planetary gear, where the first internal planetary gear and the first external planetary gear are each bevel gears, the first internal planetary gear is rigidly connected to the first external planetary gear and is axially aligned with the first external planetary gear. The first planetary gear set is rotatable via a first outer gear. The first internal planetary gear is rotatable about a first planetary gear axis, and the first external planetary gear is rotatable about the first planetary gear axis, such that the first planetary gear set is rotatable about the first planetary gear axis. Further, the first planetary gear set is also rotatable about a first output axis.

[0012] The transmission of the present invention further includes a second outer gear that is a bevel gear and meshes with the first external planetary gear, and a second inner gear that is a bevel gear, where the second inner gear is rotatable about the first output axis, the second inner gear is rotatable via the first internal planetary gear, and the second inner gear is connected to the first output.

[0013] In the transmission of the present invention, for a predetermined rotational speed of the first input, the rotation of the first planetary gear set about the first output axis results in a first rotational differential component of the second inner gear, and the rotation of the first planetary gear set about the first planetary gear axis results in a second rotational differential component of the second inner gear. The first rotational differential component and the second rotational differential component produce a second inner gear speed and a first output rotational speed that are lower than the predetermined input rotational speed.

Brief Description of the Drawings

[0014]

Figure 1

Figure 2

Figure 3

Figure 4

Embodiments for Carrying Out the Invention

[0015] As used herein, the terms "invention", "the invention", and "the present invention" are intended to broadly refer to all of the subject matter of this specification and any of the following claims. Descriptions containing these terms should not be understood to limit the subject matter described in this specification or the meaning or scope of any of the following claims. Further, this specification does not attempt to explain or limit the subject matter covered by any claim in any particular part, paragraph, description, or drawing of this application. The subject matter should be understood by reference to the entire specification, all the drawings, and any of the following claims. The present invention can have other configurations and can be practiced or carried out in various ways. Also, it is understood that the expressions and terms used herein are for illustrative purposes and should not be regarded as limiting.

[0016] Here, the details of the present invention will be considered by way of example only, with reference to the accompanying drawings that illustrate the invention. In the drawings, like features or components are referred to by like reference numerals. The size and relative size of a particular aspect or element may be exaggerated for clarity or for purposes of detailed explanation. Doors, casings, housings, inner or outer guards, etc. may be omitted from the drawings for a better understanding and clearer view of the present invention.

[0017] As used herein, the terms "comprising," "having," "including," and variations thereof are meant to include the recited elements and their equivalents, as well as additional elements. The use of a letter or number to identify an element of a method or process is for identification purposes only and is not meant to indicate that the element should be performed in a particular order. As used herein, the singular forms "a," "an," and "the" are intended to include the plural forms as well, unless the context clearly dictates otherwise, and the term "and / or" includes any and all combinations of one or more of the associated listed items.

[0018] In the following description of the drawings, reference may be made to directions such as upper, lower, above, below, rear, bottom, top, front, back, etc., which are for convenience in reference to the drawings (as normally viewed). These directions are not intended to be construed literally or to limit the invention in any way. Additionally, terms such as "first," "second," "third," etc. are used herein for purposes of description and are not intended to indicate importance or significance, or to imply any such importance or significance, unless explicitly stated.

[0019] The differential device can be regarded as an extension of a two-dimensional planetary speed reducer into three dimensions. In a planetary gear, a specific ratio can be achieved, for example, by connecting an internal gear to a sun gear or by connecting an internal gear to a housing.

[0020] Generally, a standard differential divides the input engine torque in two ways and allows each output of the differential to rotate at a different speed. The function of a standard differential can be described with reference to FIG. 1, which shows a two-dimensional view of an operating unit such as is used in an automobile. The input rotation is transmitted from the final drive gear 1, via the carrier 8, to two planetary gears (i.e., planets) 2 and 3. Planets 2 and 3 transmit rotation to side gears 4 and 5, which are each connected to output shafts 6 and 7, respectively. When output shafts 6 and 7 are connected to the vehicle's drive wheels, in the case of straight-ahead and equal two-wheel traction, ω out1 and ω out2 are equal to each other and to the input speed ω in .

[0021] When traveling a curve, the wheel going towards the outside of the curve (e.g., connected to shaft 6) must travel a longer distance than the wheel going towards the inside of the curve (e.g., connected to shaft 7). To prevent wheel slip when traveling a curve, the wheel going towards the outside of the curve must rotate faster and the wheel going towards the inside of the curve must rotate slower. The differential automatically accommodates this requirement. At a constant drive speed ω in , shaft 7 cannot rotate at ω in while maintaining the same traction torque as shaft 6, and requires a lower speed ω out2 = ω in - Δω. Next, to maintain the same traction torque as shaft 7, shaft 6 requires a higher speed ω out1 = ω in + Δω. Planets 2 and 3 automatically start rotating by plus or minus Δω to maintain torque equilibrium between shafts 6 and 7. A differential transmission is regarded as a three-dimensional version of a planetary transmission.

[0022] In the most common application of an automobile, side gears 4 and 5 are connected to the drive wheels via output shafts 6 and 7. When both drive wheels have the same traction when the vehicle is going straight, and when both wheels have the same diameter, there is no relative movement between the four gears 2, 3, 4, and 5 (Δω = 0), and, the input rotation ωin is transmitted to the two output shafts 6 and 7 at a ratio of 1 (ω out1 = ω out2 = ω in ). When traveling on a curve, the wheels facing the outside of the curve (connected to shaft 6 for example) must travel a longer distance than the wheels facing the inside of the curve (connected to shaft 7 for example). The differential enables this requirement by the rotation of the planet (in this example, gear 2 rotates at +Δω and gear 3 rotates at -Δω). As a result of such rotation, the output speed of the wheels facing the outside of the curve becomes ω + Δω, and the output speed of the wheels facing the inside of the curve becomes ω - Δω, which corresponds to the curve running condition while maintaining the vehicle speed (corresponding to ω) without wheel slip or loss of tractive force. Wherein ω ··· Carrier input speed Δω ··· Delta rotation of the side gear

[0023] If the number of teeth of all four differential gears is the same, the rotation of the planet is exactly Δω (for example, the upper part rotates in the clockwise (CW) direction and the lower part rotates in the counterclockwise (CCW) direction).

[0024] An innovative solution for high output density low, medium, and high reduction speed changers is the double differential shown in FIG. 2. The double differential transmission is symmetric and has a high output density. The double differential transmission preferably has a symmetric orientation of eight gears including a first inner planet gear 11, a second inner planet gear 13, a first inner gear 10, a second inner gear 12, a first outer planet gear 15, a second outer planet gear 17, a first outer gear 14, and a second outer gear 16, and all the gears are each located around a carrier 19 including a central sleeve to which diametrically opposed struts 36 having threaded end portions to which nuts can be fixed are attached. The input rotation φ1 (centered on axis I A1 from shaft 20 to gear 14) rotates gears 15 and 17 by φ3 and -φ3 respectively around their respective struts 36. Since gear 16 is non-rotatably connected to the housing 18 (for example, via pin 35 as shown in the figure), (axis S A1Rotation φ3 about (centered on) and rotation -φ3 about (centered on) axis S A2 Rotation -φ3 about (centered on) and rotation φ3 about (centered on) axis S actuate the rotation -φ2(23) of the carrier 19 with a magnitude of rotation corresponding to the gear tooth ratio z2 / z1. In the embodiment of FIG. 2 (and FIG. 3), the first inner gear 10 is optional but may be included for purposes of balance and / or stability.

[0025] The first inner planet gear 11 is rigidly connected to the first outer planet gear 15, such as via a spline connection, and the second inner planet gear 13 is rigidly connected to the second outer planet gear 17, such as via a spline connection. The connections between 11 and 15, and between 13 and 17, directly transmit the rotation of the carrier 19 as a first rotation component (first differential) to the second inner gear 12 that functions as an output gear. Rotations φ3 and -φ3 (rotation 22 about axis S A1 and rotation 24 about axis S A2 also are transmitted to gears 11 and 13 that represent a second rotation component (second differential) to the second inner gear 12. The magnitude of this second rotation depends on the gear tooth ratio z3 / z4. If the tooth numbers of all eight gears are equal, the ratio is infinite, which means that, as a result of the input rotation 21, the output shaft 26 (rotatable about axis O A1 does not rotate (rotation 25 = φ4 = 0). For example, if the tooth number z2 differs from all other tooth numbers z1, z3, z4 by only one tooth, the overall transmission ratio φ1 / φ4 becomes very high. Preferably, axes S A1 and S A2 coincide (i.e., S A1 = S A2 ).

[0026] The input rotation 21 from the shaft 20 is transmitted to the gears 15 and 17, causing the rotation 22 of gear 15 and the rotation 24 of gear 17. Both gears 15 and 17 mesh with gear 16. Gear 16 is rigidly connected to the housing 18. The fact that gear 16 cannot rotate causes the rotation 23 of the carrier 19. Gears 15 and 11, and gears 17 and 13 are rotationally constrained to each other. The carrier rotation 23 gives a first rotational component (first differential) to the second inner gear 12. Rotations 22 and 24 add a second rotational component (second differential) to the second inner gear 12. If all eight relevant bevel gears have the same number of teeth, the output rotation 25 is zero. This explanation means, for example, that a 90° rotation φ2 of the carrier 19 rotates gears 15 and 17 90° in the directions of 22 and 24. Thus, the output gear 12 receives a 90° rotation φ2 from the carrier and a 90° rotation φ3 from gears 11 and 13 (in the opposite direction), and as a result, does not rotate independently of the input rotation 21. This embodiment does not seem to have a practical benefit, but this example is used to demonstrate the function of the double differential transmission. In this example, the ratio is φ1 / φ4 = ∞.

[0027] Using the individual tooth numbers to derive the ratio formula, various possible ratios result from the change of the tooth numbers of the gears from 14 / 16 to 15 / 17 and from 10 / 12 to 11 / 13.

[0028]

Number

[0029] Or,

[0030]

Number

[0031]

Number

[0032] [Number]

[0033] Plug (2) into (4)

[0034] [Number]

[0035] Or,

[0036] [Number]

[0037] Plug (6) into (3)

[0038] [Number]

[0039] Plug (6) into (2)

[0040] [Number]

[0041] Plug (8) into (7)

[0042] [Number]

[0043] Rearrangement

[0044] [Number]

[0045] Wherein, z1 ··· the number of teeth of gear 14 and gear 16 z2 ··· The number of teeth of gear 15 and gear 17 z3 ··· The number of teeth of gear 10 and gear 12 z4 ··· The number of teeth of gear 11 and gear 13 φ1 ··· Rotation angle of gear 14 φ2 ··· Rotation angle of carrier 19 φ3 ··· Rotation angle of gear 15 (when gear 17 is in the negative φ3 direction) φ4 ··· Rotation angle of gear 12 (and output shaft 26) R ··· The ratio of the input speed divided by the output speed

[0046] In the following four examples, various combinations of the number of teeth are used to demonstrate a very high range of ratios that can be achieved with a double differential device without significantly changing the size of the transmission.

[0047] Example 1: z1 = 40; z2 = 39; z3 = 40; z4 = 40; ratio R = -78.000

[0048] Example 2: z1 = 40; z2 = 41; z3 = 40; z4 = 40; ratio R = 82.000

[0049] Example 3: z1 = 45; z2 = 50; z3 = 40; z4 = 40; ratio R = 20.000

[0050] Example 4: z1 = 30; z2 = 50; z3 = 40; z4 = 40; ratio R = 5.000

[0051] The functional expansion of a double differential transmission consisting of a two - dimensional view of an extended double differential device with two input shafts is shown in Figure 3. Compared with Figure 2, in Figure 3, gears 30, 31 and shaft 32 are added. In the extended arrangement, the second outer gear 16 is not rotationally constrained by the housing 18, but gear 16 is here connected to the first cylindrical gear 30, and the first cylindrical gear 30 is rotatably arranged with respect to the housing 18, input shaft I A2It meshes with a second cylindrical gear 31 (i.e., pinion) connected to a second input shaft 32 that is rotatable around it. This second input possibility allows for various combinations of input speeds with two different prime movers, for example, electric motors having different speed and torque characteristics. For example, one motor can be a high-torque and low-speed motor that operates on a constant-speed signal without speed adjustment. For example, if an output RPM of 0 is required, the second motor rotates in the reverse direction. For example, in the case of a rapid acceleration up to a vehicle's cruising speed, first the second motor is turned off, and the accumulated kinetic energy of the differential gear and carrier is used for the vehicle's acceleration. After a few seconds, for example when the vehicle reaches half of its cruising speed, the second motor is operated in the positive rotation direction. In the first stage of acceleration, a large amount of energy is drawn from the battery of a conventional electric vehicle. The extended double differential device allows for storing kinetic energy during smooth operation periods and during deceleration and braking operations.

[0052] The second outer gear 16 is connected to a first cylindrical gear 30 rotatably arranged with respect to the housing 18 and meshes with a second cylindrical gear 31 connected to the second input shaft 32. In the case of two input shafts, the numerical value of the ratio that brings about the following relationship between the output rotation for two input rotations is not one.

[0053]

Number

[0054] Or

[0055]

Number

[0056]

Number

[0057]

Number

[0058] Plug (12) into (14)

[0059]

Number

[0060] Or

[0061]

Number

[0062] Plug (16) into (13)

[0063]

Number

[0064] Plug (16) into (12)

[0065]

Number

[0066] Plug (18) into (17)

[0067]

Number

[0068] Second input rotation

[0069]

Number

[0070] Wherein z5 ··· Number of teeth of gear 30 z6 ··· Number of teeth of gear 31 φ5 ··· Rotation angle of gears 16 and 30

[0071] Applying equation (19) to different input rotations φ6 can result in two special cases. In case 1, the output speed (rotation angle φ4) is equal to the speed of gear 16 (rotation angle φ5). In this case, the output rotation φ4 is equal to the input rotation φ1, resulting in a ratio of R = 1.00. (Plugged into (19))

[0072]

Number

[0073] The solved (21) for φ4

[0074]

Number

[0075] Or simplified

[0076]

Number

[0077] As a result

[0078]

Number

[0079] In case 2, the input rotation φ5 is 0, which simplifies equation (19) and makes it equal to equation (9). (Plugging φ5 = 0 into (19))

[0080]

Number

[0081] Deletion of the 0 term

[0082] [Number]

[0083] Equation (9) is based on the fact that gear 16 is firmly connected to the transmission housing, which presents the case of φ5 = 0 and proves that equation (19) is decisive.

[0084] The gears of the double differential device can be straight bevel gears, spiral bevel gears, or face gears with cylindrical-shaped gears. In the case of high input speeds, the spiral bevel gears on the ground achieve the highest efficiency and the lowest noise emission in relation to the high load capacity. The axial force of the double differential device is the same as that of the differential device of an automobile equipped with straight bevel gears.

[0085] Due to the fact that hypoid offset gears are preferably not used, the relative surface slip has no component in the face width direction and consists only of the slip of the profile. For a spiral bevel gear set with a ratio close to 1.0, an outer diameter of 120 mm (common in automotive double differential transmissions), and a speed of 1,000 RPM, the relative profile slip is at most approximately 84 m / min. The relative speed between the two fastest gears (14 and 15) of a double differential transmission is only about 50% of the input speed. Equation 8, φ3 = φ1 / 2·z1 / z2, gives the speed of gear 15 as only 48.8% of the input speed for z1 = 40 and z2 = 41 (φ3 = φ1 / 2·40 / 41 = 0.488·φ1). Thus, the relative speed between gear 14 and gear 15 is, in this case, φ1 - φ3 = 0.512·φ1. This means that the relative speed between the fastest gears of a double differential transmission is usually only about half of the input speed. When the input speed is 10,000 RPM, the double differential has only 10·84 m / min·0.512 = 430.08 m / min. Compared to a standard spiral bevel gear transmission, the slip speed of the double differential transmission is only 51.2% in this case. An overview of the slip speeds and efficiencies of the various types of transmissions described is shown in Table 1. The calculations of the slip speeds and efficiencies based on Table 1 can be determined by commercially available software such as Gleason UNICAL (trademark) bevel gear analysis and optimization software.

[0086]

Table 1

[0087] The comparison in Table 1 clearly shows the advantages of the double differential reducer over other types of reducers. The low relative surface slip indicates less friction and results in higher transmission efficiency. Less friction results in less heat generated within the transmission. The calculated gear efficiencies are shown in the last column of Table 1. A high gear efficiency value of 98.8% at a ratio of 80 and a transmission input speed of 10,000 RPM has not been reported in state-of-the-art transmissions.

[0088] The extended double differential device allows for a variety of applications with a second input (Input 2). For example, if Input 2 is connected to a non-variable speed, low-speed, high-torque motor rotating at 1,500 RPM in the clockwise (CW) direction, and Input 1 is connected to a variable, high-speed, low-torque motor rotating in the counterclockwise (CCW) direction, the speed of Input 1 (e.g., -9,500 RPM) can be selected such that the output speed is 0 RPM in the counterclockwise (CCW) direction. This example is based on the following number of teeth. z1 = 45 z2 = 50 z3 = 40 z4 = 40 z5 = 60 z6 = 20 The speed of Input 2 (shaft 32) is n6 = 1,500 RPM CW (positive equal), and for the first reduction z6 / z5 = 20 / 60, the speed of gear 30 is n5 = 500 RPM. The speed of the output shaft is n4 = 0.

[0089] Equation (19) is also valid when the rotational speed n in RPM is used instead of the angle φ. n4 = (n1 + n5) / 2·[1 - z1 / z2·z4 / z3] + n5·z1 / z2·z4 / z3 0 = (n1 + 500) / 2·[1 - 45 / 50·40 / 40] + 500·45 / 50·40 / 40, or 0 = (n1 / 2 + 250)·0.1 + 450 The result is n1 = -9,500 RPM

[0090] The actual application of this example could be a vehicle idling in front of an intersection traffic signal. When the red signal changes to green, n1 can be decreased from -9,500 RPM to 0 in order to accelerate the vehicle from 0 MPH to 35 MPH. During the acceleration period, most of the acceleration energy is supplied by utilizing the kinetic energy of a double differential assembly with gears 10, 11, 12, 13, 14, 15, 16, and 17, as well as carrier 19, and a motor connected to input 1. Driving faster than 35 MPH simply requires rotating the input in the opposite direction. At a vehicle speed of 70 MPH, the speed of input 1 reaches n1 = +9,500 RPM. Depending on the vehicle's duty cycle (highway or city driving), the low-speed motor can be turned off and a clutch (not shown) can be applied to lock input 2. In this case, a variable-speed motor connected to input 1 supplies all the energy required, for example, for light-load city driving.

[0091] Constantly trying to reverse charge the battery with bursts of recovery energy results in very low electrical efficiency and limits the battery's chemical capacity to accept large amounts of energy within just a few seconds. For example, a mid-size sedan driving at 35 MPH has about 0.4 kWh of kinetic energy. Significantly reducing speed in front of a just turned red traffic signal requires recovering 0.4 kWh within about 2 - 3 seconds. As a result, either in the brake disk or the electronic vehicle control module, up to 0.10 - 0.15 kWh can be reverse charged to the battery and 0.25 kWh may be converted to heat. The double differential including the motor at input 1 can store about 0.24 kWh with an efficiency of about 96%, which means that when the vehicle comes to a complete stop in front of a red signal, 0.23 kWh becomes available in the form of the rotation of the double differential. This energy is used just a few minutes later to accelerate the vehicle after the signal turns green. With today's battery technology, short-term energy storage cannot be done efficiently. The double differential concept allows reducing the battery size by maintaining the same driving range capability.

[0092] The combination of two input speeds allows for a variety of possibilities for employing a compound planetary gearset in various operating conditions by achieving optimal motor and transmission efficiency. An additional aspect of simple energy storage in the high-speed rotating differential carrier unit aids the vehicle's battery, especially when a high energy burst is required to accelerate a large truck from 0 MPH to 30 MPH. In contrast to an internal combustion engine, an electric motor requires little energy while operating in an idle state without external resistance.

[0093] It is also possible to use a compound planetary gearset with two inputs to collect and transfer energy from an electric motor and a combustion engine to the drive wheels of a hybrid vehicle. Such an arrangement allows for finding an optimal speed combination for each of the two prime movers, which permits the elimination of any additional transmissions in a hybrid vehicle.

[0094] As previously described, the function of the compound planetary gearset does not require the first inner gear 10. The first inner gear 10 was used to make the transmission symmetric and was expected to help keep the torques of gears 11 and 13 equal in the case of large teeth and deformation of the transmission housing (under high loads). If symmetry and balance are not a concern, the compound planetary gearset may be simplified to consist of seven, six, or even five gears, thereby reducing manufacturing costs, by deleting gear 10 and further gears 13 and / or 17.

[0095] To allow for placement of the compound planetary gearset of the present invention between the wheels of a vehicle's drive shaft, an additional embodiment is shown in FIG. 4, which is a two-dimensional view of a modified version of the compound planetary gearset from FIG. 2. The transmission of FIG. 4 has an additional differential function between two output shafts 26 and 41. Output shaft 26 remains on the right side of the transmission housing, output shaft O A2The additional output shaft 41 rotatable about the center exits the transmission housing on the left side. The gear 10 not required for the correct functioning of the double differential is removed, and the shaft 41 functions as the main transmission shaft that was the function of the shaft 26 in FIG. 2. The first inner gear 12 in FIG. 2 is replaced by the gear 40 in FIG. 4. The gear 40 is hollow inside to create a space for the arrangement of the four differential gears 42, 43, 44, and 45. The gears 42 and 43 are planetary gears held in a predetermined position relative to the gear 40 by pins 46. The pins 46 are connected to the gear 40 which is the gear for the final output speed. The gears 44 and 45 are side gears. The output shaft 26 is connected to the side gear 44, and the output shaft 41 is connected to the side gear 45. The design of FIG. 4 realizes the same differential function between the two output shafts 26 and 41 as described for the output shafts 7 and 6 in FIG. 1. The end cap 47 closes the differential inside the gear 40 and functions as a radial sleeve bearing for the shaft 26 and as a thrust sleeve bearing for the gear 44. The wall of the hollow space of the gear 40 is utilized as a thrust sleeve bearing for the gears 42 and 43. Preferably, the input shaft I A1 and the output shaft O A2 are parallel, and more preferably, coincide (i.e., I A1 =O A2 ).

[0096] The transmission of FIG. 4 includes an additional differential device that, for example, accommodates different wheel speeds when the vehicle is traveling over a bend. A differential device similar to that shown in FIG. 1 is incorporated into gear 40. The transmission in FIG. 4 has an output shaft 26 that can be connected to the right wheel and an output shaft 41 that can be connected to the left wheel. The input shaft 20 still lies on the left side of the transmission. When the input shaft 20 is connected to an electric motor with a hollow shaft, the transmission and the drive electric motor of the invention can be arranged in line with the drive axle of the vehicle. This means that the output shaft 26 can be connected to the right drive wheel via a first drive shaft and a constant velocity (CV) joint, and that the output shaft 41 can be connected to the left drive wheel via a second drive shaft and a CV joint.

[0097] Although the invention has been described with reference to preferred embodiments, it is to be understood that the invention is not limited to these particular embodiments. The invention is intended to cover modifications that would be apparent to a person skilled in the art in the subject matter to which the invention pertains without departing from the spirit and scope of the appended claims.

Claims

1. A transmission, comprising a housing, at least a first input and at least a first output, and further comprising a double differential device, wherein the transmission is: A first outer gear, which is a bevel gear, is connected to the first input and is drivable by the first input; A first planetary gear set, which comprises a first inner planetary gear and a first outer planetary gear. The first inner planetary gear and the first outer planetary gear are each bevel gears. The first inner planetary gear is rigidly connected to the first outer planetary gear and is axially aligned with the first outer planetary gear. The first planetary gear set is rotatable via the first outer gear; A second outer gear, which is a bevel gear and meshes with the first outer planetary gear; A second inner gear, which is a bevel gear, is connected to the first output and is rotatable via the first inner planetary gear; The first input is rotatable about a first input shaft; The first output is rotatable about a first output shaft; The first inner planetary gear is rotatable about a first planetary gear shaft, and the first outer planetary gear is rotatable about the first planetary gear shaft. As a result, the first planetary gear set is rotatable about the first planetary gear shaft; The first planetary gear set is further rotatable about the first output shaft; The second inner gear is rotatable about the first output shaft; The double differential device has a first inner gear that is a bevel gear and has symmetry by being rotatable via the first inner planetary gear; In an axial sectional view, the pitch lines of the bevel gear connections between the first outer gear and the first outer planetary gear, the first outer planetary gear and the second outer gear, and the first inner planetary gear and the second inner gear intersect at the intersection of the rotation axes of the first outer gear and the first inner planetary gear. A transmission.

2. The transmission according to claim 1, wherein the first input shaft and the first output shaft are parallel to each other.

3. The transmission according to claim 2, wherein the first input shaft and the first output shaft coincide with each other.

4. The transmission according to claim 1, wherein the second outer gear is non-rotatable.

5. For a predetermined rotational speed of the first input, rotation of the first planetary gear set around the first output shaft results in a first rotational differential component of the second inner gear, rotation of the first planetary gear set around the first planetary gear shaft results in a second rotational differential component of the second inner gear, the transmission according to claim 1, wherein the first rotational differential component and the second rotational differential component result in a second inner gear speed and a first output rotational speed that are lower than the predetermined input rotational speed.

6. The transmission further comprises a second planetary gear set disposed on the side opposite to the first planetary gear set, and the second planetary gear set comprises a second inner planetary gear and a second outer planetary gear, the second inner planetary gear and the second outer planetary gear are each bevel gears, the second inner planetary gear is rigidly connected to the second outer planetary gear and is axially aligned with the second outer planetary gear, and the second planetary gear set is rotatable via the first outer gear, the second inner planetary gear is rotatable around a second planetary gear shaft, the second outer planetary gear is rotatable around the second planetary gear shaft, and as a result, the second planetary gear set is rotatable around the second planetary gear shaft, the transmission according to claim 1, wherein the second planetary gear set is further rotatable around the first output shaft.

7. The transmission according to claim 6, wherein the first planetary gear shaft and the second planetary gear shaft coincide with each other.

8. The transmission further comprises a second input rotatable around a second input shaft, the second input including a first cylindrical gear rotatable around the second input shaft. The transmission further includes a second cylindrical gear rotatable around the first output shaft and meshing with the first cylindrical gear, and the second cylindrical gear is rigidly connected to the second outer gear.

9. The transmission according to claim 1, further comprising the second inner gear including a differential mechanism having an internal space and a plurality of bevel gears located within the internal space.

10. The transmission according to claim 9, further comprising a second output rotatable about a second output shaft, wherein the first output and the second output are connected to the second inner gear and the differential mechanism.

11. The transmission according to claim 10, wherein the first output and the second output extend in opposite directions.

12. The transmission according to claim 10, wherein the second output extends within and through the first input.

13. A method for reducing the rotational speed of an input shaft in a transmission according to any one of claims 1 to 12, the method comprising: rotating the input shaft at a predetermined rotational speed, whereby rotating a first planetary gear set about a first output shaft, thereby resulting in a first rotational differential component of a second inner gear connected to an output shaft; and rotating the first planetary gear set about a first planetary gear shaft, thereby resulting in a second rotational differential component of the second inner gear connected to the output shaft, wherein the first rotational differential component and the second rotational differential component result in a rotational speed of the second inner gear and the output shaft that is lower than the predetermined rotational speed of the input shaft.

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