WO2013053981A1

WO2013053981A1 - Power transmission arrangement, method and use of power transmission arrangement

Info

Publication number: WO2013053981A1
Application number: PCT/FI2011/050886
Authority: WO
Inventors: Taimo Majalahti; Pekka Majalahti
Original assignee: Individual
Current assignee: Individual
Priority date: 2011-10-12
Filing date: 2011-10-12
Publication date: 2013-04-18
Anticipated expiration: 2014-04-12
Also published as: EP2802791A4; WO2013053988A2; EP2802791A2; WO2013053988A3

Abstract

This invention relates to a power transmission arrangement, method and use comprising at least a gear assembly (30) with a rotary first axle (1) and a rotary second axle (14), and two or more pinion gear assemblies (5) equipped with two or more gear wheels (6-10, 39a, 39b) on an axle (12), and two or more annuli (15-19, 41) meshing with the gear wheels (6-10, 39a, 39b), and a gearbox housing (27). The arrangement comprises at least a stationary or stationary lockable annulus (15- 18, 41) meshing with the coaxial non-equal diameter gear wheels (6-9, 39b) one set of the gear wheels at a time, and a rotary annulus (19) meshing with gear wheels (10, 39a) coaxial with the gear wheels (6-9, 39b) and having a different diameter than the gear wheels (6-9, 39b), and all the gear wheels (6-10, 39a, 39b) on the same axle (12) are fixedly coupled to each other and/or to the same axle (12).

Description

POWER TRANSMISSION ARRANGEMENT, METHOD AND USE OF POWER TRANSMISSION ARRANGEMENT

The present invention relates to a power transmission arrangement as defined in the preamble of claim 1 , to a power transmission method as defined in the preamble of claim 1 1 and to a use of the power transmission arrangement as defined in the preamble of claim 20.

The power transmission arrangement according to the invention is suited very well for instance to be used as either a reducing or accelerating gearing in connection with small and/or large vehicles and also for instance for wind turbine applications which require large accelerating gear ratio to speed-up the rotor of the electrical generator.

According to the prior art there are various types of reducing or accelerating gearings from simple cylindrical gear pairs or spur gear systems to complicated epicyclic gearings or planetary gearings with one or more planet gears rotating about a sun gear, and comprising also an outer ring gear or annulus meshing with the planet gears. In the spur gear systems small gearwheels are typically used and in planetary gearings the sun gear is usually relatively small. As a problem with these structures is a disadvantageous pres- sure angle in a tooth contact with relatively small gear wheels. Because of the extremely high pressure in the tooth contact the lubricating oil does not work properly and the teeth wear too fast, particularly in power transmissions with large torques.

Another disadvantage, particularly in connection with conventional epicyclic gearings is a heavy bearing load of the planet gears. For that reason epicyclic gearings for heavy-duty power transmission are usually very massive that increases costs. For instance the power transmission in a wind turbine converts the slow, high-torque rotation of the turbine blades into a much faster rotation needed by the electrical generator. These gear arrangements are huge and may weigh several tons, and typically contain more than one stage of gearing to achieve a needed gear ratio from 50:1 to 120:1 to speed-up the generator's rotor for effective electricity generation. The first stage of the gearing is usually a planetary gear, but a sufficient durability of these gearings has been a serious problem for a long time. Hybrid vehicles according to the prior art use various gear structures in their power transmission arrangements. However, these power transmission arrangements have generally only a limited start-out torque, which is a problem in case of large and heavy motor vehicles and machineries. Therefore the power transmission arrangements pur- posed for hybrid vehicle use are usually made for small and light vehicles only. For example the electro-mechanical transmission used in Toyota Prius and many other Toyota motor vehicles is suitable primarily for light vehicles only because of its limited start-out torque. In order to have a powerful start-out torque needed in very heavy vehicles said Toyota-Prius system would have excessively large and heavy electric motor. Also the electric motor of the Toyota-Prius system ought to rotate at extremely high rpm at full speeds of the vehicle. For this reason hybrid vehicles use primarily especially designed electric motors that have a greater start-out torque than in ordinary electric motors. These kind of special electric motors are, however, very expensive. One speed change mechanism according to prior art is presented in the US patent No. US3705522 (Ogava). The patent shows a speed change mechanism with an input shaft and an output shaft with a gear system with two pinion gears on an eccentric shaft and two annuli meshing with the pinion gears. In addition the speed change mechanism comprises small diameter long pinion gears meshing with the two pinion gears on the eccentric shaft. A troublesome problem here is the small diameter of the long pinion gears, which causes a disadvantageous pressure angle in the tooth contact with the two pinion gears on the eccentric shaft, likewise in connection with small sun gears mentioned above. The object of the present invention is to eliminate the drawbacks described above and to achieve a reliable, cost effective and efficient power transmission arrangement that is suitable for several different purpose of use, for instance to be used in light or heavy motor vehicles, in wind turbines to speed-up the rotor of the electrical generator and in other machinery as a reducing or accelerating gearing. Likewise the object of the pre- sent invention is to achieve a light, compact and durable power transmission arrangement that gives a long service life with large reducing or accelerating gear ratio. The power transmission arrangement according to the invention is characterized by what is presented in the characterization part of claim 1 . Correspondingly the power transmission method according to the invention is characterized by what is presented in the characterization part of claim 1 1 , and the use of the power transmission arrangement according to the invention is characterized by what is presented in the characterization part of claim 20. Other embodiments of the invention are characterized by what is presented in the other claims.

The solution of the invention has the advantage that thanks to its ingenious structure the power transmission arrangement according to the invention is compact and extremely durable because the pressure in tooth contacts is smaller than in conventional solutions and the bearing loads are small. These features make it possible to use the power transmission arrangement according to the invention for several purposes both as a reducing gearing and accelerating gearing as well as in applications requiring large gear ratio. The solution is for example well adaptable as an accelerating gearing for wind turbine applications, which require large accelerating gear ratio between 50-120:1 to speed-up the generator's rotor for effective electricity generation. For instance in a 3 MW wind turbine application the solution of the invention is clearly more economical and it can be designed and manufactured about 50% lighter than a comparable conventional planetary gearings with sun gears. Yet, the transmissions according to the invention never go to a hub level as in a conventional planetary gear-set with a sun gear. Therefore the gear power transmissions according to the invention are tangential and circularly translational and taking place at outer rim-level meshing, only. Because of their large diameter gear meshing contacts the solutions according to the invention transfer larger torques and can without damage withstand exceptionally large torques and powerful rotational shock loads coming via their power-in or power-out axles. Mentioned 50% less weight means that the whole wind turbine construction can be made much lighter and more feasible. Also the available transmission ratio between 50-120:1 is much higher than the gear ratio available with conventional planetary gearings with sun gears.

Yet further advantages are in a motor vehicle use. When the power transmission arrangement according to the invention is used as a power transmission in motorized vehi- cles, its regenerative braking features can be used to slow down the vehicle and generated electricity can be stored in a battery or super condenser packet and used at the next start-out or acceleration. This way especially at city driving lots of gas or diesel fuel can be saved and pollution greatly diminished. In addition to the said regenerative braking and start-out use the one and the same electric machine can be used in many ways. Whenever the engine is running, the same electric machine works as an alternator for the vehicle's electricity requirements and in driving as a hybrid booster power for quick passing also. In traffic jams the primary motor can be temporarily stopped and only electric mode can be used for driving reducing pollution greatly.

In the following, the invention will be described in detail by the aid of examples by referring to the attached simplified and schematic drawings, wherein presents in a cross section and in a simplified and schematic way a basic transmission arrangement according to the invention,

presents in a cross section a pinion gear assembly according to the invention,

presents in a cross section and in a simplified and schematic way one advantageous embodiment according to the invention,

presents in a cross section and in a simplified and schematic way another advantageous embodiment according to the invention,

presents in a oblique top and side view and in a simplified and schematic way yet another advantageous embodiment according to the invention, presents in a cross section and in a simplified and schematic way the same kind of structure as presented in Fig. 5,

presents in a partial cross section a combined carrier power-out unit used in the structures of Fig. 5 and Fig. 6,

presents a partial transversal section of the embodiment of Fig. 5 along the line A-A,

presents in a cross section and in a simplified and schematic way yet another advantageous embodiment according to the invention and presents in a cross section and in a simplified and schematic way yet another advantageous embodiment according to the invention. The gear assembly or gearing according to the power transmission arrangement of the invention is a further evolutionary modification of an ordinary epicyclic or planetary gear and therefore it is later called a trans-planetary gearing or trans-planetary gear assembly, or even shorter as a trans-planetic gear. The trans-planetary gearing according to the invention does not include a sun gear at all and the pinion gears that act as planet gears are meshing with internal gears only. The internal gears are also called ring gears or annuli. The power transmission arrangement according to the invention can be a complete stepless electro-mechanical transmission with two electric machines, but it can be constructed also without an electric machine and used as either a reduction or accel- erating gearing. However, the trans-planetary and gear-changing functions of the trans- planetary gearing are greatly enhanced when an electric machine is combined to the trans-planetary gear assembly to co-operate with the primary rotator means, such as a motor, wind turbine or another power unit producing rotating movement to power this tangential and circularly translational, without a sun gear constructed, mechanical power transmission assembly with one or several gear ratio levels taking place only at a large diameter or outer-rim-level meshing between the planet gears and the annuli of the trans-planetary gearing.

As a result of the large meshing diameter and exceptionally large effective radius of cur- vature of the tooth flanks, the tooth contact area is doubled and correspondingly the Hertzian contact pressure at every meshing of gears compared to a conventional planetary gear is halved. Consequently the trans-planetary gearing according to the invention can thus without damage withstand exceptionally large torques and powerful rotational shock loads to its tooth contacts of large diameter.

In addition the trans-planetary power transmission assembly is advantageously arranged to transmit torque through an elastic joint with a small rotational twist angle between the non-equal diameter planet gears on each planet axle, for compensating machining tolerance errors of their tooth meshing. Further the trans-planetary gearing according to the invention transmits rotating power through one or several circumferential differences between two to five or even more together-joined non-equal diameter planetary gears on each planetary axle to obtain designed transmission ratios by a stepless gear changing method for forward gears and, when needed, at least one reverse gear. The diameter in this description can be understood as the pitch diameter or of the gears and annuli. The circumferential difference means also the difference of the pitch diameters of the gears and annuli. The trans-planetary gearing according to the invention operates one of said circumferential differences at a time through two together-joined non-equal diameter planetary gears on each planet axle to load against each other via their matching annuli to transfer designed reduction gear ratios to power-out axle. Correspondingly when said power-out axle is serving as power-in axle and is directly rotated by the primary rotator means, said non-equal diameter planetary gears, when loading against each other through their matching annuli, transfer designed accelerated rotation or rotations to the carrier-power-out axle unit, while said power transfers are loading axle bearings of said non-equal-diameter planet gears, only.

In Fig. 1 a basic transmission arrangement according to the invention is shown in a simplified and schematic way. Later it is called also the basic trans-gear, or for short the BTG. The shown construction is a trans-planetary automatic transmission comprising at least a gear assembly 30 with a rotary first axle 1 and a rotary second axle 14. In some advantageous embodiments the shown construction comprises also at least one electric machine 40 on the same first axle 1 . The first axle 1 with grooves 2 at its first end and acting as a power-in axle is placed in a gear housing or gearbox, later gearbox housing 27, so that only the first end of the first axle 1 is seen outside of a gearbox housing 27. Inside the gearbox housing 27 two carrier flanges 3 and 4 with an axial distance to each other are coupled by a non-rotary connection to the first axle 1 . The first carrier flange 3 is connected to the first axle 1 through grooves 13 and the second carrier flange 4 is integrated to the first axle 1 being the uniform part with the first axle 1 . However, the carrier flange 4 could be also groove connected like the carrier flange 3. Thus these three inter-connected components 1 , 3, and 4 have no relative rotation to each other but revolve coaxially together at the same speed of rotation around the central axis of the first axle 1 .

The gear assembly 30 comprises a group of pinion gear assemblies 5 functioning like a planet gear assemblies without a sun gear, the group of pinion gear assemblies 5 being arranged between the carrier flanges 3 and 4 at even radial intervals. A number of the groups may be two, three, four or even more. Each pinion gear assembly 5 comprises a number of interconnected and coaxially arranged pinion gears or planet gears 6, 7, 8, 9 of which each having a different diameter, and arranged so that the planet gear 6 having the smallest diameter is the closest to the first end of the pinion gear assembly 5, then is the next larger planet gear 7, then the next larger planet gear 8 and finally the largest planet gear 9 is the closest to the second end of the pinion gear assembly 5. Preferably the non-equal diameter planet gears 6, 7, 8, 9 are machined on the same material, such as steel so that they all are integrated together to the pinion gear assembly 5 that forms also a support shaft or planet axle 12 for the planet gears 6, 7, 8, 9. The planet axle 12 of each pinion gear assembly 5 is rotatory mounted in bearings 12a on the carrier- flanges 3 and 4. Yet at least one more pinion gear or planet gear 10 is connected to each pinion gear assembly 5 coaxially with the other planet gears 6, 7, 8, 9. The diameter of the planet gear 10 is only a little smaller than the diameter of the largest planet gear 9 but larger than the diameter of the other planet gears 6-8. The planet gear 10 has been coupled to rotate together with the pinion gear assembly 5 through an elastic joint 1 1 transmitting torque with a small rotational twist angle between the non-equal diameter planet gears 6-9 on the same planet axle 12 for compensating machining tolerance errors of their tooth meshing. Each pinion gear assembly 5 with the planet gears 6-10 rotates as one unit around the central axis of the planet axle 12. Thus the pinion gear assembly 5 with the inter-connected planet gears 6-10 forms a kind of stepped cone.

The basic transmission arrangement according to the invention or the BTG has normally four essentially identical pinion gear assemblies 5, which are revolving on the central axis of the first axle 1 . The gear assembly 30 of the BTG comprises also a group of matching ring gears or annuli 15-18 that are carried and rotated by the meshing with their planet gears 6-9. All the planet gears 6 on the separate planet axles 12 form a first gear wheel set, likewise all the planet gears 7 on the separate planet axles 12 form a second gear wheel set, and all the planet gears 8 on the separate planet axles 12 form a third gear wheel set, and further all the planet gears 9 on the separate planet axles 12 form a fourth gear wheel set, and finally all the planet gears 10 on the separate planet axles 12 form one more gear wheel set. In general all the planet gears 6-10 of the same diameter on the separate planet axles 12 form together a gear wheel set that is meshing with its corresponding annulus.

Each planet gear 6, or each set of planet gears 6, is meshing with the first annulus 15 and each planet gear 7 is meshing with the second annulus 16. In addition the annulus

15 is thread-bolt connected to flange 28, which is mounted in bearings 28a by the first axle 1 . Further each planet gear 8 is meshing with the third annulus 17 and each planet gear 9 is meshing with fourth annulus 18, and further each planet gear 10 is meshing with the rotary annulus 19, which is the uniform material, such as steel construction with the second axle 14 acting as a power-out axle, the second end of which equipped with grooves 2 and protruding out from the gearbox housing 27. Later these uniform material components 14 and 19 are named as a power-out unit 20. The all of said annuli 15-19 and the second axle 14 rotate around the central axis of the first axle 1 .

At the first end the first axle 1 is carried in a bearing 1 a by the first end flange 21 of the gearbox housing 27 and the other end of the first axle 1 is carried in bearings 1 b by the second axle 14 that is further carried in a bearing 14a by the second end flange 22 of the gearbox housing 27. Said end flanges 21 and 22 belong to and are parts of the gearbox housing 27. In addition, if needed the gearbox housing 27 is equipped with an internal support flange 23 that is stationary connected inside the gearbox housing 27, and the first axle 1 is also carried in a bearing 23a by the support flange 23.

The electric machine 40 comprising at least a stator coils 24, rotor 25 and suitable con- trol means is placed inside the gearbox housing 27. The stator coils 24 are mounted to the inner side of the first end flange 21 , and the rotor 25 is connected through grooves 26 to the first axle 1 close to the stator coils 24. Thus the rotor 25 is carried by and rotating together with the first axle 1 at the same rotational speed as the first axle 1 . The BTG comprises also a gear changing means 29 such as a brake actuator unit for each annulus 15-18 to stop the rotational movement of the annuli 15-18 one at a time. The gear changing means 29 are mounted inside the stationary gearbox housing 27 and operated by a suitable power unit, such as a hydraulic, mechanical or electrical power unit, and by a suitable control system. The power unit and the control system are not shown in Fig. 1 .

Figs. 1 as well as Figs 3, 4, 8 and 9 are schematic and simplified drawings, for instance for clarity reasons only one pinion gear assembly 5 is presented and some lines of the annuli 15-19 and the gear changing means 29 have been left off.

Next the function of the BGT is described. With this configuration of gears when the first axle 1 is rotated clockwise then the first, second and third gear rotate the second axle 14 counter-clockwise and correspondingly the reverse gear rotates the second axle 14 clockwise. When the first axle 1 is being rotated clockwise by the primary rotator means the carrier flanges 3 and 4 are rotated clockwise along with it at the same revolutions per minute (rpm). Since the pinion gear assemblies 5 are mounted in bearings by said carrier flanges 3 and 4 the pinion gear assemblies 5 are rotated clockwise around the central axis of the first axle 1 by the primary rotator means as well. An end load is resisting the rotation of the power-out unit 20, whose annulus 19 is meshing with the planet gears 10 having the second largest diameter of all the planet gears 6-10. Therefore the pinion gear assemblies 5, while being rotated clockwise around the central axis of the first axle 1 , are simultaneously forced to rotate counter-clockwise around the central axis of their own planet axle 12. The simultaneous counter-clockwise rotation of the pinion gear assemblies 5 forces each of the annuli 15-18 to rotate to the direction and at the rpm dictated by the circumferential difference between the planet gear 10 and each of the planet gears 6-9.

When the annuli 15-18 are allowed to rotate freely, the annuli 15-17, which are of smaller diameter than the annulus 19 meshing with the planet gears 10, will revolve clockwise. Whereas the annulus 18, which is of a little larger diameter than the annulus 19, will at the same time revolve counter-clockwise. As said before the circumferential difference between said planet gears 6-9 compared to the planet gears 10 dictates the rpm and the direction of the rotation of each of the annuli 15-18. One of the annuli 15-18 at a time is stopped by the gear changing means 29, such as a brake actuator unit 29e mounted inside the stationary gearbox housing 27. When the first axle 1 is rotated by the primary rotator means while all of the axial-friction brake actuator units 29e are released, then the annuli 15-18 are freely revolving and there is no driving connection through the first axle 1 to the second axle 14. This is a clutch-free state of the BTG arrangement.

A gear change order is explained next beginning with the first gear, then second, third and finally the reverse gear.

When the brake actuator unit 29e of the annulus 17 is activated to stop the clockwise rotation of the annulus 17 the small circumferential difference between the planet gears 8 and 10 is slowly and powerfully with high reducing gear ratio rotating the power-out unit 20 counter-clockwise. Thus there is a great trans-planetary reducing gear ratio through the first axle 1 to the second axle 14 and the torque of the primary rotator means is greatly multiplied to rotate the second axle 14 counter-clockwise. This is the first gear of the BTG because the diameter of the planet gears 8 is next smaller than the diameter of the planet gears 10.

When the second gear is needed the brake actuator unit 29e of the annulus 17 is released and the brake-actuator unit 29e of the annulus 16 is simultaneously activated to stop the faster clockwise revolving annulus 16. Now the noticeably greater circumferential difference between the planet gears 7 and 10 is counter-clockwise rotating the sec- ond axle 14. Thus there is less gear ratio through the first axle 1 to the second axle 14, which is therefore rotating at faster rpm with less torque. This is the second gear of the BTG.

When the third gear is needed the brake actuator unit 29e of the annulus 16 is released and the brake actuator unit 29e of the annulus 15 is simultaneously activated to stop the clockwise revolving annulus 15. When the annulus 15 is fully stopped by the axial brake actuator unit 29e the greatest circumferential difference between the planet gears 6 and 10 is through the trans-planetary action rotating the second axle 14 counter-clockwise and the gear ratio from the first axle 1 to the second axle 14 is the smallest. Thus the rpm of the power out axle 14 is accelerated step by step to the highest possible rotational speed. This is the third gear of the BTG.

When the reverse gear is needed the brake actuator units 29e are released and the second axle 14 is stopped. If the BTG is serving as a power transmission in a motorized vehicle, the stopping of said second axle 14 means that by using regular wheel brakes the motor vehicle is stopped and therefore the rotation of the second axle 14 is also stopped. During all the gear changes and transmission ratios and also when the reverse gear is engaged and used the first axle 1 can continuously be rotated clockwise by the primary rotator means. Now the axial brake actuator unit 29e of the annulus 18 is acti- vated to stop the annulus 18. The planet gears 9 are of the larger diameter than the planet gears 10. While the annulus 18 is held stationary by the brake actuator unit 29e the planet gears 9 and 10 are through their circumferential difference rotating the second axle 14 to the reverse direction, which here means clockwise. This is the reverse gear of the BTG. Since all the brake actuator units 29e are of the friction type the gear change from the forward direction to the reverse direction can actually be done even while the second axle 14 is slowly rotating counter-clockwise. Yet such an action may quite heavily load the brake actuator unit 29e of the annulus 18 and the planet gears 9 and 10 and therefore it is not recommended but only if the brake actuator unit 29e of the annulus 18 and the planet gears 9 and 10 are dimensioned to take such extra heavy loads.

To obtain reliable function of gears the goal is to obtain inexpensive manufacturing with fairly large tolerances and yet long service life without internal overloading at tooth meshing and load axles. A satisfactory solution is obtained with the innovative clutch with high stiffness radially and optimally designed torsional flexibility to compensate machining tolerance errors of the connection of the gears on the same axle. The clutch of this kind is located on the planet axle 12. It can be the elastic joint 1 1 in the planet gear 10 mentioned above or an elastic arrangement in the planet axle 12. Two set of planet gears, one set of the planet gears 6-9 at a time, and the planet gears 10 are loaded by the carrier force. The problems may occur that tolerance errors cause large tooth forces at the two set of planet gears involved. The goal is to obtain more even dis- tribution of loading on the gear tooth even with rather large manufacturing tolerances of the teeth. Large tolerances also lower manufacturing costs. Design goals of the clutch are: the clutch should have a maximal radial stiffness and a slight torsional flexibility to compensate machining tolerance variations in tangential direction. Said torsional flexibility makes flank pressure distributions much more even and lowers pressure peaks. These results effectively increase the tooth flank endurance and tooth bending life.

In Fig. 2 one pinion gear assembly 5 according to the invention is shown in a cross section. The pinion gear assembly 5 of Fig. 2 is of the same type as shown in a simplified and schematic way in Fig. 1 but their diameter order is reversed. In this embodiment the planet gears 6-9 are integrated to the planet axle 12 of the pinion gear assembly 5 being the same material as the planet axle 12, for example steel. A cylindrical area near the second end the planet axle 12 is equipped with grooves 10a by which the planet gear 10 is coupled to the planet axle 12 to rotate around the central axis of the planet axle 12 together with the planet gears 6-9 on the same planet axle. The planet gear 10 has been coupled to rotate together with the pinion gear assembly 5 through an elastic joint 1 1 transmitting torque with a small rotational twist angle between the planet gears 6-9 on the same planet axle 12 for compensating machining tolerance errors of their tooth meshing, as mentioned earlier. The elastic joint 1 1 can also be of different type than shown in Fig. 2.

In Fig. 3 one advantageous embodiment of the power transmission arrangement according to the invention is shown in a simplified and schematic way. This embodiment is also called embodiment 1 or for short the EM1 . The structure of the EM1 is essentially similar to the BTG but differs from the BTG mainly in some constructional details. Namely the structure of the gear changing means 29 is now different from the structure of the BTG, but it could also be similar. Another difference relates to the mutual sizes of the planet gears 6-9 compared to the size of the planet gear 10. In this case the planet gears 6-8 for forward motion have a larger diameter than the planet gear 10, and correspondingly the planet gear 9 for reverse or backing up motion has a smaller diameter than the planet gear 10, which is totally opposite to the corresponding structure of the BTG arrangement and means that the rotation directions are also opposite to the BTG. Yet another minor difference is that now the first carrier flange 3 is integrated to the first axle 1 acting as a power-in axle and being the uniform part with the first axle 1 and the second carrier flange 4 is connected to the first axle 1 through grooves 13. However, the carrier flange 3 could be also groove connected like the carrier flange 4.

The structure of this kind makes it possible to achieve a very effective start-out torque. If the end load is not too high, the start-out can be done using the electric machine 40 only, acting as an electric motor, to rotate the first axle 1 and the start-out takes place softly from zero. If a heavy start-out torque is needed, the primary motor can be coupled to the first axle 1 . Together with the electric motor 40 the primary motor is rotating the first axle 1 clockwise. The outcome is extremely powerful torque rotating the second axle 14 acting as a power-out axle. A motor vehicle with the power transmission according to the EM1 and equipped with one electric motor 40 can have a considerable smaller gasoline or diesel engine and yet have the same start-out and passing power as a vehicle with a larger engine would have. A group of pinion gear assemblies 5 functioning like a planet gear assemblies without a sun gear are arranged between the carrier flanges 3 and 4 at even radial intervals. A number of the groups may be two, three, four or even more. Each pinion gear assembly 5 comprises a number of interconnected and coaxially arranged pinion gears or planet gears 6, 7, 8, 9 of which each having a different diameter, and arranged so that the planet gear 6 having the largest diameter is the closest to the first end of the pinion gear assembly 5, then is the next smaller planet gear 7, then the next smaller planet gear 8 and finally the smallest planet gear 9 is the closest to the second end of the pinion gear assembly 5. Preferably the non-equal diameter planet gears 6, 7, 8, 9 are machined on the same material, such as steel so that they all are integrated together to the pinion gear assembly 5 that forms also a support shaft or planet axle 12 for the planet gears 6, 7, 8, 9. The planet axle 12 of each pinion gear assembly 5 is mounted in bearings 12a on the carrier-flanges 3 and 4. Yet at least one more pinion gear or planet gear 10 is connected to each pinion gear assembly 5 coaxially with the other planet gears 6-9. The diameter of the planet gear 10 is only a little larger than the diameter of the smallest planet gear 9 but smaller than the diameter of the other planet gears 6-8. The planet gear 10 has been coupled to rotate together with the pinion gear assembly 5 through an elastic joint 1 1 transmitting torque with a small rotational twist angle between the non-equal diameter planet gears 6-9 on the same planet axle 12 for compensating machining tolerance errors of their tooth meshing. Each pinion gear assembly 5 with the planet gears 6-10 rotates as one unit around the central axis of the planet axle 12. Thus the pinion gear assembly 5 with the inter-connected planet gears 6-10 forms a kind of stepped cone.

The EM1 according to the invention has normally four essentially identical pinion gear assemblies 5 which are revolving on the central axis of the first axle 1 . The EM1 comprises also a group of matching annuli 15-18 that are carried and rotated by the meshing with their planet gears 6-9. The planet gears 6-9 and the planet gears 10 are meshing with their own annuli 15-19, and all of the annuli 15-19 and the second axle 14 rotate around the central axis of the first axle 1 the structure and function corresponding the structure and function of the BTG. The gear changing means 29 of the EM1 differs from the gear changing means 29 of the BTG though it could also be similar. The gear changing means 29 of the EM1 comprise a gear change ring 29a having grooves around its external rim and internal rim. The grooves can be like a toothing of the gear rim. The gear change ring 29a is mounted onto the inner surface of the stationary gearbox housing 27 and guided to move axially along the toothing-like grooves 29b inside the gearbox housing 27 so that the gear change ring 29a cannot rotate. The EM1 comprises also means for moving the gear change ring 29a axially in order to stop the rotational movement of the annuli 15-18 one at a time, which annuli 15-18 have external toothing corresponding the grooves of the internal rim of the gear change ring 29a. The means for moving the gear change ring 29a axially contains at least a power unit, such as a hydraulic, mechanical or electrical actuator, and a suitable control system. The power unit and the control system are not shown in Fig. 3. Next the function of the EM1 is described. Though it is not shown in Fig. 3 a conventional friction clutch is recommended to be arranged between the output axle of the primary motor and the first axle 1 of the EM1 . With this reversed configuration of gears when the first axle 1 is rotated clockwise then the first, second and third gear rotate the second axle 14 clockwise and correspondingly the reverse gear rotates the second axle 14 counter-clockwise. This is totally opposite to arrangement of the BTG described above.

When the first axle 1 is being rotated clockwise by the primary rotator means the carrier flanges 3 and 4 are rotated clockwise along with it at the same revolutions per minute (rpm). Since the pinion gear assemblies 5 are mounted in bearings 12a by said carrier flanges 3 and 4 the pinion gear assemblies 5 are rotated clockwise around the central axis of the first axle 1 by the primary rotator means as well. An end load is resisting the rotation of the power-out unit 20, whose annulus 19 is meshing with the planet gears 10 having the second smallest diameter of all the planet gears 6-10. Therefore the pinion gear assemblies 5, while being rotated clockwise around the central axis of the first axle

1 , are simultaneously forced to rotate counter-clockwise around the central axis of their own planet axle 12. The simultaneous counter-clockwise rotation of the pinion gear assemblies 5 forces each of the annuli 15-18 to rotate to the direction and at the rpm die- tated by the circumferential difference between the planet gears 10 and each set of the planet gears 6-9.

When the annuli 15-18 are allowed to rotate freely the annuli 15-17, which are of larger diameter than the annulus 19 meshing with the planet gears 10, will revolve counterclockwise. Whereas the annulus 18, which is of a slightly smaller diameter than the annulus 19, will at the same time revolve clockwise. As said before the circumferential difference between the planet gears 6-9 compared to the planet gears 10 dictates the rpm and the direction of the rotation of each of the annuli 15-18. One of the annuli 15-18 at a time is stopped by the axially moving gear change ring 29a mounted inside the stationary gearbox housing 27. When the first axle 1 is rotated by the primary rotator means while none of the annuli 15-18 is locked stationary by the gear change ring 29a all the annuli 15-18 are freely revolving and there is no driving connection through the first axle 1 to the second axle 14. This is a clutch-free state of the EM1 arrangement.

A gear change order of the EM1 is explained next beginning with the first gear, then second, third and finally the reverse gear.

The recommended friction clutch between the output axle of the primary motor and the first axle 1 of the EM1 is released and the electric machine 40 is used as a regenerative brake to stop the first axle 1 . The first axle 1 with the flanges 3 and 4 and all the rotating gearings are quickly stopped and the annulus 17 is locked to the gearbox housing 27 by the gear change ring 29a. If the end load is not too high, start-out can be done using only the electric motor 40 to rotate the first axle 1 and start-out takes place softly from zero. If a heavy start-out torque is needed, the primary motor can be re-clutched to the first axle 1 . Together with the electric motor 40 the primary motor is rotating the first axle 1 clockwise. The outcome is an extremely powerful torque rotating the second axle 14. This is the first gear of the EM1 . When the second gear is needed the primary motor is throttled down and again the electric machine 40 is used as a regenerative brake and the revolutions of the first axle 1 decrease quickly and the freely rotating annuli 15-16 come to a stop. At that split-second the gear change ring 29a is axially moved to lock the annulus 16 to the stationary gearbox housing 27 and the second gear rotating the second axle 14 clockwise is on. Accel- eration by the primary motor can immediately continue. This is the second gear of the EM1 .

When the third gear is needed the primary motor is throttled down and again the electric machine 40 is used as a regenerative brake and the revolutions of the first axle 1 decrease quickly and the freely rotating annulus15 comes to a stop. At that split-second the gear change ring 29a is axially moved to lock the annulus 15 to the gearbox housing 27. Acceleration by the primary motor can immediately continue. Thus the rpm of the power out axle 14 is accelerated step by step to the highest possible clockwise rotation. This is the third gear of the EM1 .

When the reverse gear is needed, the primary motor is un-clutched from the first axle 1 and the motor vehicle is stopped by using its regular brakes. When the second axle 14 and the first axle 1 are not rotating anymore the gear change ring 29a can be axially moved to lock the reverse gear annulus 18 to the stationary gearbox housing 27. When the primary motor is re-clutched to the first axle 1 , the second axle 14 starts to rotate counter-clockwise in a reverse direction. This is the reverse gear of the EM1 .

At a regenerative braking both the BTG and EM1 effectively transmit the rotating powers and torques from the second axle 14 to the rotor 25 of the electric machine 40 forcing the rotor 25 to revolve at higher revolutions and thus enabling the electric machine 40 effectively generate electricity. With automated gear selections the regenerative braking is effective in lower speeds of the vehicle, too. Thus the regenerative braking can be used to slow down the vehicle and generated electricity can be stored in a battery, bat- teries or a super condensers packet and can be used at the next start-out or acceleration. This way, especially at city driving, lots of gas or diesel fuel can be saved and pollution can be greatly diminished. In addition to the said regenerative braking and start-out use the one and the same electric machine 40 with the rotor 25 revolving with the first axle 1 can be used in many ways. Whenever the engine, acting as a rotator means, is running, the same electric machine 40 acts as an alternator for various electricity requirements of the motor vehicle and in driving as a hybrid booster power for quick passing, too. When the electric machine 40 is used in traffic jams, the primary motor can be temporarily stalled and driving goes on using only electric mode and thus pollution is greatly reduced. When the brake actuator units 29e in the BTG or the gear change ring 29a in the EM1 are released from locking the annuli 15-18 there is no driving connection between the first axle 1 and the second axle 14. Therefore at times of electric blackouts the same electric machine 40 can be used also as an auxiliary electric power source. A motor vehicle having the BTG or the EM1 transmission equipped with one electric machine 40 can have a considerable smaller gasoline or diesel engine and yet have the same start- out and passing power as a vehicle with a larger engine would have. In Fig. 4 another advantageous embodiment of the power transmission arrangement according to the invention is shown in a simplified and schematic way. This embodiment is also called embodiment 2 or for short the EM2. The structure of the EM2 is essentially similar to the BTG but differs from the BTG mainly in some constructional details. Namely the EM2 has only three forward gears but no reverse gear. And in addition all the annuli 15-17 and 19 of the EM2 are mounted in the bearings. Further difference is that the annuli 15-17 of the EM2 are not stopped by brakes but by the gear changing means 29 having at least the inside and outside grooved gear change ring 29a, which cannot rotate but can be axially moved by an electric or hydraulic actuator to effect gear changes. The gear changing means 29 of the EM2 correspond to the gear changing means 29 of the EM1 . Otherwise the EM2 gear functions are based on the same trans- planetary principles of the circumferential differences of its planet gears and their matching annuli as in the BTG.

A group of pinion gear assemblies 5 functioning like a planet gear assemblies without a sun gear are arranged between the carrier flanges 3 and 4 at even radial intervals. A number of the groups may be two, three, four or even more. Each pinion gear assembly 5 comprises a number of interconnected and coaxially arranged pinion gears or planet gears 6, 7, 8 of which each having a different diameter, and arranged so that the planet gear 6 having the smallest diameter is the closest to the first end of the pinion gear as- sembly 5, then is the next larger planet gear 7 and finally the largest planet gear 8 is the closest to the second end of the pinion gear assembly 5. Preferably the non-equal diameter planet gears 6, 7, 8 are machined on the same material, such as steel so that they all are integrated together to the pinion gear assembly 5 that forms also a support shaft or planet axle 12 for the planet gears 6, 7, 8. The planet axle 12 of each pinion gear assembly 5 is mounted in bearings 12a on the carrier-flanges 3 and 4. Yet at least one more pinion gear or planet gear 10 is connected to each pinion gear assembly 5 coaxially with the other planet gears 6, 7, 8. The diameter of the planet gear 10 is only a slightly smaller than the diameter of the largest planet gear 8 but larger than the diame- ter of the other planet gears 6, 7. The planet gear 10 has been coupled to rotate together with the pinion gear assembly 5 through an elastic joint 1 1 like in the BTG solution.

The transmission arrangement according to the EM2 has normally three or four essen- tially identical pinion gear assemblies 5, which are revolving on the central axis of the first axle 1 acting as a power-in axle. The EM2 comprises also a group of matching an- nuli 15-17 that all are carried by bearings and rotated by the meshing with their planet gears 6-8. Each planet gear 6 is meshing with the annulus 15 and each planet gear 7 is meshing with the annulus 16. Further each planet gear 8 is meshing with the annulus 17, and each planet gear 10 is meshing with the annulus 19, which structure is similar to corresponding structure of the BTG or the EM1 .

In addition the annulus 15 is thread-bolt connected to a flange 28, which is mounted in bearings 28a by the first axle 1 , and a gear coupling part 31 of the annulus 15 is fixedly connected to the flange 28 and groove-connected to the gear change ring 29a in order to lock the annulus 15 stationary. Further the annulus 16 is threaded-bolt connected to a flange 32, which is mounted in bearings 32a to the flange 28. The flange 32 is groove- connected to the gear change ring 29a in order to lock the annulus 16 stationary. And finally the annulus 17 is thread-bolt connected to flange 33, which is mounted in bear- ings 33a to the second axle 14 acting as a power-out axle, and groove-connected to the gear change ring 29a in order to lock the annulus 17 stationary. All the annuli 15-17 and 19 and the second axle 14 rotate around the central axis of the first axle 1 . The locking grooves 29i of the flanges 31 -33 to lock the annuli 15-17 stationary by the gear change ring 29a are on the outer periphery of the flanges 31 -33 and are for instance similar to a normal toothing of a gear wheel.

Next the function of the EM2 is described. Thus, with this configuration of gears when the first axle 1 is rotated clockwise then the first, second and third gear rotates the second axle 14 counter-clockwise. When the first axle 1 is being rotated clockwise by the primary rotator means the carrier flanges 3 and 4 are rotated clockwise along with it at the same revolutions per minute (rpm). Since the pinion gear assemblies 5 are mounted in bearings 12a by said carrier flanges 3 and 4 the pinion gear assemblies 5 are rotated clockwise around the central axis of the first axle 1 by the primary rotator means as well. An end load is resisting the rotation of the power-out unit 20, whose annulus 19 is meshing with the planet gears 10 having the largest diameter of all the planet gears 6-8 and 10. Therefore the pinion gear assemblies 5, while being rotated clockwise around the central axis of the first axle 1 , are simultaneously forced to rotate counter-clockwise around the central axis of their own planet axle 12. The simultaneous counter-clockwise rotation of the pinion gear assemblies 5 forces each of the annuli 15-17 to rotate to the direction and at the rpm dictated by the circumferential difference between the planet gears 10 and each set of the planet gears 6-8.

When the annuli 15-17, being of smaller diameter than the annulus 19 meshing with the planet gears 10, are allowed to rotate freely, they will revolve clockwise. As said before the circumferential difference between said planet gears 6-8 compared to the planet gears 10 dictates the rpm and the direction of the rotation of each of the annuli 15-17. One of the annuli 15-17 at a time is locked stationary by the gear change ring 29a, though the brakes according to the BTG can be used also. When the first axle 1 is rotated by the primary rotator means while none of the annuli 15-17 is locked stationary by the gear change ring 29a there is no driving connection through the first axle 1 to the second axle 14. This is a clutch-free state of the EM2 arrangement.

A gear change order is explained next beginning with the first gear, then second and finally the third gear.

The recommended friction clutch between the output axle of the primary motor and the first axle 1 of the EM2 is released and the electric machine 40 is used as a regenerative brake to stop the first axle 1 . The first axle 1 with the flanges 3 and 4 and all the rotating gearings are quickly stopped and the annulus 17 is locked to the gearbox housing 27 by the gear change ring 29a. If the end load is not too high, start-out can be done using only the electric motor 40 to rotate the first axle 1 and a stepless start-out takes place softly from zero rpm. If a heavy start-out torque and acceleration is needed, the primary motor can be re-clutched to the first axle 1 . Together with the electric motor 40 the primary motor is rotating the first axle 1 clockwise. The outcome is an extremely powerful torque rotating the second axle 14 counter-clockwise. This is the first gear of the EM2.

When the second gear is needed the primary motor is throttled down and the largest annulus 17 is released by the gear change ring 29a, and again the electric machine 40 is used as a regenerative brake and the revolutions of the first axle 1 decrease quickly and the freely rotating annulus 16 comes to a stop. At that split-second the gear change ring 29a is axially moved to lock the annulus 1 6 to the stationary gearbox housing 27 and the second gear rotating the second axle 14 counter-clockwise is on. Acceleration by the primary motor can immediately continue. This is the second gear of the EM2.

When the third gear is needed the primary motor is throttled down and the second larg- est annulus 16 is released by the gear change ring 29a, and again the electric machine 40 is used as a regenerative brake and the revolutions of the first axle 1 decrease quickly and the freely rotating annulus16 comes to a stop. At that split-second the gear change ring 29a is axially moved to lock the annulus 15 to the gearbox housing 27. Acceleration by the primary motor can immediately continue. Thus the rpm of the power out axle 14 is accelerated to the highest possible counter-clockwise rotation. This is the third gear of the EM2.

When the reverse gear is needed, the primary motor is un-clutched from the first axle 1 and the motor vehicle is stopped by using its regular brakes. The gear change ring 29a is axially moved to lock the largest annulus 17 to the stationary gearbox housing 27, and the electric machine 40 is used as the power source to rotate the first axle 1 counterclockwise. The clockwise reverse rotation of the second axle 14 starts from zero softly accelerating. This is the reverse gear of the EM2. In Fig. 5 yet another advantageous embodiment of the power transmission arrangement according to the invention is shown. This embodiment is also called embodiment 3 or for short the EM3. Fig. 5 shows the arrangement in a schematic way and Fig. 6 shows a cross section of one advantageous embodiment of the arrangement according to Fig. 5. Further Fig. 7 shows a partial cross section of the first axle 1 of the embodiment of Fig. 6, the first axle 1 serving as a power-out unit 35 of the EM3, and Fig. 7a shows a partial transversal section of the power transmission arrangement of Fig. 6 along the line A-A.

The annulus 19 and the second axle 14, acting now as a power-in axle, belong to and are now parts of a one-steel power-in unit 34, which is by a cast-steel construction united to rotate as one unit. Correspondingly a power-out unit 35 shown alone in Fig. 7 is formed by the first axle 1 acting now as a power-out axle and also as a supporting axle comprising carrier flanges 36 and 37 at an axial distance from each other, and a supporting flange 38 axially between the carrier flanges 36 and 37. Further the power-out unit 35 comprises a group of axial carrier bridges 38a, for instance four carrier bridges 38a fastened on the periphery of the carrier flanges 36 and 37 connecting the carrier flanges 36, 37 and the supporting flange 38 together. The power-out unit 35 is advantageously a one cast-steel construction also, and therefore the first axle 1 , the carrier bridges 38a, the carrier flanges 36 and 37 and the supporting flange 38 rotate simulta- neously together at the same rpm.

The construction of the EM3 includes two annuli 19 and 41 that are radially and axially wide enough to accommodate the group of planet gears 39a, 39b meshing with them. Grooved planet axles 12 are mounted in bearings 12a to the carrier flanges 36, 37 and to the supporting flange 38. Smaller diameter planet gears 39a are groove-connected to the power-in-ends of the grooved planet axles 12 and larger diameter planet gears 39b are groove-connected to the power-out-ends of the same grooved planet axles 12. Thus each planet axle 12 is groove-connected to one smaller planet gear 39a and to one larger planet gear 39b to rotate as a one together-connected unit 12, 39a, 39b and forming now the pinion gear assembly 5 mentioned earlier.

This embodiment of the EM3 comprises four planet axles 12 with two planet gears 39a, 39b on each planet axle 12. Fig. 7 shows spaces 36a and 37a between the flanges 36 and 37 for the planet gears 39a and 39b. The space 36a for the larger diameter planet gears 39b is between the first carrier flange 36 and the supporting flange 38, and the space 37a for the smaller diameter planet gears 39a is between the supporting flange 38 and the second carrier flange 37. Each space 36a, 37b is axially wide enough for two axially one after the other placed planet gears, either two larger planet gears 39b or two smaller planet gears 39a. The spaces 36a, 37a are needed because the diameter of the planet gears 39a, 39b is so large that they could not rotate in the same vertical plane. This is easily seen in Fig. 7a where the smaller diameter planet gears 39a are seen in front of the carrier flange 37. The two planet gears 39a one upon the other are rotating in a vertical plane that is closer to the first end the gearbox housing 27 than the vertical plane where the two side by side being two planet gears 39a are rotating. In this way the size of the transmission arrangement can be made small and still relatively large and durable gear wheels with lower teeth can be used. It is also easily seen in Fig. 7a that there is no room for the sun gear between the planet gears 39a or 39b. The smaller diameter planet gears 39a are meshing with the annulus 19, which belongs to the power-in assembly 34. Whereas the larger diameter planet gears 39b are meshing with the fulcrum annulus 41 , which is stationary groove-connected to the non-rotating gearbox housing 27. It can be said also that each set of planet gears 39a is meshing with the annulus 19 and each set of planet gears 39b are meshing with the fulcrum annulus 41 . Parts 41 a on the outer periphery of the annulus 41 in Fig. 5 represent the locking groove connections. End flanges 21 and 22 are thread-bolt-connected to non- rotating stationary gearbox housing 27. The power-out unit 35 is mounted in bearings 1 b inside the power-in unit 34, which is carried in a bearing 14a by the second end flange 22. The power-out unit 35 is carried in a bearing 1 a by the first end flange 21 , too.

In Fig. 5 the EM3 comprises a structure where the elastic joints 1 1 are used with a small rotational twist angle between non-equal diameter planet gears 39a and 39b on each planet axle 1 2 for compensating machining tolerance errors of their tooth meshing to distribute even flank pressure to every tooth contact of the embodiment EM3. Corre- spondingly In Fig. 6 the EM3 comprises another structure where the planet axles 1 2 themselves are designed to optimally twist under the heavy opposite directional torsion loads directed to the structure of the power transmission arrangement of the invention. These both solutions and even they both together can be used to achieve a reliable way for compensating machining tolerance errors of the tooth meshing in the power trans- mission arrangement of the invention.

The embodiment of the EM3 can be constructed for example using the grooved axle 1 2 with an optimal twist angle and/or said elastic joints 1 1 between each planet axle 12 and the larger planet gear 39b to transmit torque with a small rotational twist angle and with eight maximally large sideways or radially overlapping planet gears 39a, 39b as shown in Figs. 5, 6 and 7a. As well the solution of the EM3 can be constructed using two, three or more optimally twisting planet axles 12 and/or said elastic joints 1 1 and accordingly with four, six, eight or more planet gears as the particular applications and their known peak torques require. The elastic joints 1 1 are designed to twist a bit easier than the stiffer but optimally twisting planet axles 12 resulting in larger, gradually stiffening twist-angle between the non-equal diameter planet gears 39a and 39b.

Next the function of the EM3 is described. The rotational directions either clockwise or counter-clockwise are looked as axial view from the primary rotator towards said power- in unit 34.

The power-in unit 34 with the annulus 1 9 is rotated clockwise by a primary rotator, for example a wind turbine rotor. Thus the smaller diameter planet gears 39a, which are meshing with the annulus 19, are under heavy clockwise torque. By a non-rotary connection through the grooved planet axles 12 the larger diameter planet gears 39b are under the same clockwise torque also. The fixedly to the stationary gearbox housing 27 connected and therefore non-rotating fulcrum annulus 41 , which is meshing with the larger diameter planet gears 39b, will not give in but remains stationary.

The heavy opposite directional torsion loads are transferred by these planet axles 1 2, which are designed to optimally twist under said opposite directional torsion loads and thus these optimally twisting axles 12 distribute even flank pressure to every tooth contact of the EM3. Also, the optimally twisting planet axles 12 together with the elastic joints 1 1 transmit torque with a larger, gradually stiffening rotational twist-angle between the non-equal diameter planet gears 39a, 39b on each planet axle 1 2 and thus guaranteeing even flank pressure to every tooth contact of the EM3 at all torque loads.

Therefore the planet gears 39a, 39b are forced to rotate counter-clockwise at a high rpm. The planet gears 39a, 39b are mounted in bearings 1 2a through their planet axles

12 by the carrier flanges 36 and 37. The heavy opposite directional torques force simultaneously the non-equal diameter planet gears 39a, 39b with their groove connecting axles 12 to circulate clockwise in the matching annuli 1 9 and 41 . Thus the whole carrier- power-out unit 35 including the first axle 1 is rotated at a high gear ratio clockwise by the power-in unit 34.

The following comparison and explanation is given for easier understanding of the de- scribed trans-planetary accelerating function: The hypothesis for a transmission ratio is 100. When a deduction trans-planetary gearing is functioning, the power-in axle must rotate the two, differing, non-equal diameter planet gears a hundred (1 00) revolutions inside their matching annuli, until the power-out axle rotates one revolution. Next the reversed or rather the accelerating function as with the EM3 is explained. In this case the rotation of the carrier-power-out unit 35 is accelerated through the power-in unit 34, which is, in this example, rotated one revolution by the primary rotator, like a wind turbine rotor. Now the two planet gears 39a, 39b by the opposite direction torsion loads through their planet axle 1 2 are rotated one hundred (1 00) revolutions inside their matching annuli 19, 41 until the power-out unit 35 has also rotated one hundred revolutions. Thus, when the second axle 14 and the power-in unit 34 rotate one revolution the first axle 1 and the whole power-out unit 35 rotate one hundred (100) revolutions.

The EM3 can be used either as a reducing or accelerating gearing in applications requir- ing large gear ratio but it is especially suitable as an accelerating gearing for wind turbine applications, which require large accelerating gear ratio of 50-1 20:1 to speed-up the generator's rotor for the effective electricity generation.

The embodiment EM3 according to the invention is characterized especially in that the optimally twisting axles 1 2 distribute even flank pressure to every tooth contact of the

EM3. The planet gears 39a and 39b are firmly groove-connected to their axles 12. The opposite direction torsion loads are transferred by the axles 12, which are designed to optimally twist under said torsion loads and thus these optimally twisting axles 12 distribute even flank pressure to every tooth contact of the EM3 and in spite of the small un- avoidable machining defects or slight deformations caused by heat treating of its gearwheels the all of its meshing contacts are evenly loaded. The result is a much longer service life. Further the embodiment EM3 according to the invention is characterized in that the exceptionally large effective radius of curvature of its tooth flanks doubles the tooth contact area and accordingly halves the Hertzian contact pressure at every meshing of gears of the EM3 compared to the tooth contact area and said Hertzian contact pressure at the meshing of a sun gear and the planet gears of a conventional planetary gear set, thus resulting in a lighter and more economical machinery with a greatly extended service life.

The embodiment EM3 according to the invention is also characterized in that the relative slippage at its gear meshing is minimized due to the large diameter of its planet gears 39a, 39b in relation to their annuli 19, 41 . The relative slippage in the meshing of the annuli 19, 41 and all of the planet gears 39a, 39b is minimal compared to the slippage that takes place in the meshing of a sun gear and the planet gears of a conventional planetary gear. The very small slippage in its gear meshing improves also greatly the service life the EM3.

The embodiment EM3 according to the invention is yet characterized in that the pressure angle at all of its gear meshings or tooth contacts is uniquely optimal and is corollary of the advantageous diameter ratio of the large diameter planet gears 39a, 39b in relation to their annuli 19, 41 resulting in a smooth and quiet running of the machinery without vibrations.

Transmissions never go to a hub level as in a conventional planetary gear-set with a sun gear. Therefore one can conclude the gear power transmissions of the embodiment EM3 are tangential and circularly translational and taking place at outer rim-level meshing, only. Because of their large diameter gear meshing contacts the solutions according to the invention can without damage withstand exceptionally large torques and powerful rotational shock loads coming via their power-in or power-out axles.

The embodiment EM3 according to the invention has only one gear ratio. Thus the gear changing means 29 are not needed. However, the operation principle and the basic structure are essentially similar to the gear assembly 30 of the other embodiments of the invention. The annulus 19 is connected to the second axle 14 to rotate together with the second axle 14 and in addition the gear assembly 30 comprises the annulus 41 that is connected to the stationary gearbox housing 27 so that the annulus 41 cannot rotate but acts as a fulcrum. The annulus 41 corresponds the annuli 1 5-18 when they are stopped one at a time to serve as a non-rotating fulcrum. The rotary annulus 19 is meshing with the planet gears 39a, one on each planet axle 12. Planet gears 39a correspond to planet gears 1 0 in the other embodiments and have here a slightly smaller diameter than the planet gears 39b on the same planet axles 12. The slightly larger planet gears 39a correspond to planet gears 6-9 one set at a time, and are meshing with the non-rotating annulus 41 .

In Fig. 8 yet another advantageous embodiment of the power transmission arrangement according to the invention is shown in a simplified and schematic way. This embodiment is also called embodiment 4 or for short the EM4. The subassembly with the gear assembly 30 including the pinion gear assembly 5, annuli 1 5-19, and the electric machine 40 of the EM4 structure is essentially similar to the BTG but differs from the BTG mainly in some other constructional details. Namely the EM4 comprises an additional electric machine 64, a power-split arrangement 65 and a clutch arrangement 66, all of which are carried in bearings by a common rotating third axle 42 acting as a power-through axle. The power-split arrangement 65 comprises a group of planet gears 46, 47 mounted to a group of axles 50. In addition the power-split arrangement 65 comprises annuli 44 and 45, a flange 43, a planetary carrier 49 and the first axle 1 acting as a power-in axle.

The first axle 1 acting as the power-in axle is seen outside of a stationary gearbox housing 27. Inside the gearbox housing 27 the first axle 1 expands radially to the flange 43. The first axle 1 with the flange 43 is mounted in bearings 1 a to the first end flange 21 . The flange 43 is thread-bolt-connected to an annulus 44. Thus the first axle 1 , the flange 43 and the annulus 44 form a together-connected or integrated power-in assembly 63, which is directly rotated by a primary power source, such as a vehicle motor or another rotator. Mutually joined or integrated planet gears 46 and 47 are of one material or one steel construction and revolve therefore always together as one unit. Axles 50 of the planet gears 46, 47 are mounted to the planetary carrier 49. Axles 50 can advanta- geously be two, three, four or more.

The mutually joined planet gears 46, 47 are of differing or non-equal diameters as are their matching annuli 45 and 44, also. The annulus 45 is connected through grooves 45a to the third axle 42 acting as the power-through axle. The mutually joined planet gears 46, 47 are mounted in bearings 48 to their circulating axles 50. Mounted in bearings 49a and carried by the third axle 42 the planetary carrier 49 revolves around the central axis of the third axle 42. The integrated power-in assembly 63 is rotated by the primary rotator or motor. The third axle 42 is carried in a bearing 42a by the first axle 1 , which is fur- ther mounted in bearings 1 a to the first end flange 21 as mentioned earlier.

The joined planet gears 46 are meshing with the annulus 45, which is through grooves 45a connected to the power-through axle or the third axle 42 and thus the annulus 45 and the third axle 42 are forced to rotate together as a power-through unit 67. A flange 53 is thread-bolt connected to the stationary gearbox housing 27. A rotor 54 of the additional electric machine 64 is thread-bolt connected to the planetary carrier 49, which with the rotor 54 is mounted in bearings 49a to the third axle 42. Stator coils 55 are mounted to a flange 56. The third axle 42 is supported and carried in a bearing 56a by the flange 56. The stator coils 55 are stationary since the flange 56 is thread-bolt connected to the stationary gearbox housing 27.

The clutch arrangement 66 is used when it is advantageous for the power transfer. Between the rotor 54 and the power-in assembly 63 there is a clutch disc-pack 51 for connecting and locking trans-planetary power-split arrangement 65 with the third axle 42 and the rotor 54 to rotate as one locked-in unit.

A ring-like hydraulic or electric actuator 53a is mounted to the flange 53 to actuate locking in and unlocking of the power-split arrangement 65 to the third axle 42 and to the rotor 54 whenever it is advantageous for the power transfer. Through a clutch pack compactor 52 the actuator 53a compacts the clutch disc-pack 51 to lock the rotor 54 axially to the annulus 44.

When the trans-planetary power-split arrangement 65 is clutched to the rotor 54 then the power-in axle or the first axle 1 , the rotor 54 and the power-through axle or the third axle 42 rotate at the same rpm and there is no relative rotation between the gear wheels 44,

47 and 46, 45. Thus the locking in the power-split arrangement 65 enables the primary rotator or motor directly with its full rpm and torque to rotate the power-through axle or the third axle 42. Electricity, generated by the additional electric machine 64, is fed by wiring 57 to the frequency converter 59 and from the frequency converter 59 by wiring 58 to the stator coils 24 to power the electric machine 40 to rotate the third axle 42, also. An electric wiring 57 from the stator coils 55 of the additional electric machine 64 to a frequency con- verter 59 and from the frequency converter 59 by a wiring 58 to the stator coils 24 of the electric machine 40 are wound through the cover of the gearbox housing 27. The frequency converter 59 and an accumulator/super condensers packet 61 are situated outside the gearbox housing 27. A wiring 60 connects the frequency converter 59 and the accumulator/super condensers packet 61 together. A wiring 62 from the frequency con- verter 59 leads to electrical uses/sources being outside the gearbox housing 27.

The stator coils 24 of the electric machine 40 are mounted to a flange 21 a, which is thread-bolt connected to the stationary gearbox housing 27. The rotor 25 of the electric machine 40 is connected through grooves 26 to the third axle 42. The annuli 45 and 44, the carrier 49 and the rotors 54 and 25 rotate around their common central axis that is the central axis of the third axle 42, which is also coaxial with the first axle 1 and with the second axle 14.

The power-through axle or the third 42 is through grooves 13 connected to the two car- rier flanges 3 and 4. Like in the BTG the group-gearings or the group of the pinion gear assemblies 5 are mounted in bearings 1 2a to the carrier flanges 3 and 4. The pinion gear assemblies 5 are comprised of five non-equal diameter planet gears 6, 7, 8, 9 and 10. The planet gear 10 is connected to each pinion gear assemblies 5 through an elastic joint 1 1 transmitting torque with a small rotational twist angle between the non-equal diameter planet gears on the same axle 1 2 for compensating machining tolerance errors of their tooth meshing.

The non-equal diameter annuli 1 5-19 are meshing with the planet gears 6-1 0. The annulus 15 is tread-bolt connected to the flange 28, which is mounted in bearings 28a to the third axle 42 and thus the annulus 15 is rotating on the bearings 28a. The planet gear 6 is meshing with the annulus 1 5 and the planet gear 7 is meshing with the annulus 1 6. The planet gear 8 is meshing with the annulus 17 and the planet gear 9 is meshing with the annulus 1 8 like in the BTG. The planet gear 10 is meshing with the power-out annulus 19, which is by the same- steel construction connected to the power-out axle or the second axle 14. These uniform material components 14 and 19 form the power-out unit 20 as mentioned earlier. The power-through axle or the third axle 42 is carried in a bearing 23a by the flange 23. The third axle 42 is mounted in bearings 1 b to the power-out axle or the second axle 14, which is carried in bearings 14a by the second end flange 22 of the stationary gearbox housing 27.

Inside the gearbox housing 27 there are gear changing means 29 like four brake actua- tors as in the BTG or like a gear change ring with the groove guiding as in the EM1 or in the EM2. The gear changing means 29 are provided to stop one of the fulcrum annuli 15-18 at a time. Like for example in the BTG and in the EM1 or in the EM2.

Next the function of the EM4 is described. With this configuration of gears when the first axle 1 is rotated clockwise then the first, second and third gear rotates the second axle 14 counter-clockwise and the reverse gear rotates the second axle 14 clockwise.

The power-in assembly 63 is rotated clockwise by the primary rotator/motor and thus the rotated power-in assembly 63 exerts clockwise torque to the smaller diameter planet gears 47 of the power-split arrangement 65. Through their same material or the same- steel construction the larger diameter gears 46 are under the same clockwise torque, too. The joined larger planet gears 46 are meshing with the annulus 45, which is through the grooves 45a connected to the power-through axle or the third axle 42. The end load is resisting the rotation of the power-out axle 14 and therefore the third axle 42 will not readily give in and rotate. Opposite direction torsion loads between the joined planet gears 46 and 47 are thus caused by the resisting fulcrum annulus 45.

Therefore the mutually integrated or joined planet gears 46 and 47 are together forced to rotate counter-clockwise around central axis of their own axles 50. Simultaneously, the opposite direction torques caused by the circumferential difference of the joined planet gears 46 and 47 force the planetary carrier 49 to accelerate its clockwise rotation. Thus the unit comprising the planetary carrier 49 and the rotor 54 is rotated clockwise at a high gear ratio by the power-in assembly 63. The electricity generating magnetic drag in the air gap between the rotor 54 and the sta- tor coils 55 is dragging and resisting the rotation of the planetary carrier 49 and the rotor 54. Simultaneously said magnetic drag is transferred through the high gear ratio trans- planetary power-split arrangement 65 as highly multiplied clockwise torque of the pri- mary motor to the rotation of said power-through axle or the third axle 42. Thus the combined actions of the primary power source/motor, said drag of the electricity- generation, magnetising flux in the air gap between the magnetised generator elements 54, 56 and the power-split gear arrangement 65 with the power-through unit 67 are functioning together to efficiently transmit the rotational torque of the primary power source/motor to the power-through axle or the third axle 42.

It is necessary to observe that from the primary rotator/motor to the rotor 54 the unlocked trans-planetary power-split arrangement 65 functions as an accelerating gearing, but when the fulcrum annulus 45 begins to rotate and is gaining rpm, the high gear ratio is simultaneously lessening. When the power-through unit 67 rotates at the same rpm with the power-in axle or the first axle 1 then the gear ratio is 1 . Transferring said magnetic drag from the rotor 54 to the power-through unit 67 the trans-planetary power- split arrangement 65 functions as a high ratio reducing gearing, which keeps its high gear ratio constantly in spite of fluctuations of the rpm of the power-through unit 67.

When the additional electric machine 64 is without electricity the rotor 54 can freely revolve and there is only a minimal friction-caused torque to the power-through unit 67. Thus the rotating power-through unit 67 serves as a rotating fulcrum against which the joined one-steel planetary gears 46, 47 base when they transfer accelerated clockwise rotation to the planetary carrier 49 and to the rotor 54 fixedly coupled to the planetary carrier 49.

Thus, when the primary rotator/motor rotates the first axle 1 clockwise the rotation of the first axle 1 is conveyed to rotate the third axle 42 clockwise through the power-split ar- rangement 65 and through the grooves 45a always when the trans-planetary power-split arrangement 65 is locked. When the power-split arrangement 65 is unlocked the first axle 1 is conveyed to rotate the third axle 42 only when electric power is taken from the additional electric machine 64. When the power-split arrangement 65 is unlocked and the additional electric machine 64 is without electricity the rotor 54 can revolve freely and no rotating power is conveyed to the third axle 42.

The gear assembly 30 of the EM4 comprises essentially the same kind of components and means than the BGT; namely the carrier flanges 3 and 4, the group of pinion gear assemblies 5, the annuli 15-19, the second axle 14, the gear changing means 29 and flanges 23 and 28. Thus the gear assembly 30 of the EM4 is essentially similar to the corresponding gear assembly of the BTG and works essentially in the same way. Thus the function of the gear assembly 30 of the EM4 has already been explained in connec- tion with the BTG. The only main difference between the gear assemblies 30 is that the EM4 has an additional rotation axle, the third axle 42 acting as the power-through axle, between the first axle 1 and the second axle 14, and the carrier flanges 3 and 4 are connected to the third axle 42 instead of the first axle 1 . The trans-planetary power-split arrangement 65 shown in Fig. 8 has only one-step gear ratio but it could have also a multi-step gear ratio, for instance a two-step or three-step gear ratio or even larger than the three-step gear ratio.

In Fig. 9 yet another advantageous embodiment of the power transmission arrangement according to the invention is shown in a simplified and schematic way. This embodiment is also called embodiment 5 or for short the EM5. The structure and main functions of the EM5 are essentially similar to the EM1 but differ from the EM1 mainly in some constructional details. Namely the EM5 comprises inside the gearbox housing 27 an additional rotary flange 69 with a gear changing means 29 with a rotary gear change ring 29c and a circular brake-actuator unit 70, and a synchronizer ring 71 relating to the locking of the annuli 15-18. The similar features and functions between the EM1 and EM5 are not explained here again because they have been described in connection with the EM1 .

In the embodiment EM5 the gear change ring 29c of the gear changing means 29 is a rotary component. The rotary gear change ring 29c is axially movable through grooves

29d in relation to the rotary flange 69, which is mounted in bearings 69a to the power-in axle or the first axle 1 . Since the rotary gear change ring 29c is grooved to the rotary flange 69 they have no relative rotation to each other. Stationary grooves 29b are machined into the inner periphery of the gearbox housing 27 and the non-rotary gear lock- ing ring 29a is axially movable in the grooves 29b like in the EM2 or EM3. The synchronizer ring 71 is mounted to the rotary gear change ring 29c. The electric or hydraulic brake-actuator unit 70 is mounted to the flange 23 so that when activated the brake- actuator unit 70 presses against the rotary flange 69 slowing down the rotation of the circular rotary flange 69 and the rotary gear change ring 29c rotating with the circular rotary flange 69.

Next the function of the EM5 is explained. Though it is not shown in Fig. 9 a conventional friction clutch is recommended to be arranged between the output axle of the primary motor and the first axle 1 of the EM5. With this reversed configuration of gears when the first axle 1 is rotated clockwise then the first, second and third gear rotate the second axle 14 clockwise and correspondingly the reverse gear rotates the second axle 14 counter-clockwise. When changing gears one of the annuli 1 5-18 at a time is locked stationary like for example in the EM1 . Here the locking is made with different gear changing means 29 than in the EM1 . Now one of the annuli 1 5-18 at a time is stopped by united actions of the synchronizer ring 71 and the rotary gear change ring 29c stopped by the brake-actuator unit 70, which is mounted to the stationary flange 23 and powered through a wiring or hydraulic pressure 68. When the first axle 1 is rotated by the primary rotator/motor while none of the annuli 15-1 8 is slowed down or stopped through the rotary gear change ring 29c or is locked to the gearbox housing 27 by the gear locking ring 29a, there is no driving connection through the first axle 1 to the second axle 14. This is the clutch-free state of the EM5.

Thus through the electrically or hydraulically movable rotary gear change ring 29c, the brake-actuator unit 70 is effectively stopping one of the annuli 15-1 7 at a time to effect gear changes of the EM5. During each of said gear changes there is a small, momentary torque loading the primary motor. The gear change order is explained beginning of the first gear then second, third and last the reverse gear.

In the beginning both the rotary gear change ring 29c and the non-rotary gear locking ring 29a are axially at a free space between the annuli 17 and 1 8. The activated synchronizing ring 71 causes friction between the slowly counter-clockwise rotating annulus 17 and the rotary gear change ring 29c coupled together with the rotary flange 69 by grooves, and quickly said rotary gear change ring 29c rotates the same, fully synchronized rpm with the annulus 17. Then the grooved, rotary gear change ring 29c is axially moved to lock into the periphery-grooved annulus 17. The brake-actuator unit 70 is activated to slow and stop the counter-clockwise rotation of the annulus 17 and so clockwise rotation of the second axle 14 starts softly with a high torque. The rotary gear change ring 29c and the nonrotary gear locking ring 29a are axially moved in tandem. Therefore the grooved, non- rotary gear locking ring 29a is at the same action axially moved into the stopped periphery-grooved annulus 17, also, pushing the rotary gear change ring 29c to free axial space between the annuli 16 and 17. Thus there is a great trans-planetary step-down gear ratio through the first axle 1 to the second axle 14, and the torque of the primary rotator and/or the electric machine 40 is greatly multiplied to rotate the power-out axle or the second axle 20 clockwise. This is the first gear of the EM5.

The gear change to the second gear repeats many of the same functions as of engaging the first gear. While the annulus 17 is kept stationary by the non-rotary gear locking ring 29a, with drag caused by the synchronizer ring 71 the rotary gear change ring 29c is quickly dragged to the same rpm with the annulus 16. Then the grooved, rotary gear change ring 29c is axially moved to lock into the periphery-grooved annulus 16. When the braking by the brake-actuator unit 70 starts, the non-rotary gear locking ring 29a is moved to a free space between the annuli 16 and 17. When the annulus 16 is fully stopped, then immediately the grooved, non-rotary gear locking ring 29a is at one action axially moved into the stopped periphery-grooved annulus 16, also, pushing the rotary gear change ring 29c to a free space between the annuli 15 and 16. The second gear of EM5 is on.

When the third gear is needed, the annulus 16 is kept stationary by the non-rotary gear locking ring 29a and by the synchronizer ring 71 caused friction the rotary gear change ring 29c is quickly dragged to the same rpm with the annulus 15. Then the grooved rotary gear change ring 29c is axially moved to lock into the periphery-grooved annulus 15 and the brake-actuator unit 70 is activated to slow down the annulus 15 and at the same action the non-rotary gear locking ring 29a is moved into the free space between the annuli 15 and 16. The annulus 15 is fully stopped and kept stationary by the brake- actuator unit 70. During the third gear is on, the non-rotary gear locking ring 29a can be kept in the free space between said annuli 15 and16. This is the third gear of the EM5. Thanks to the structure of the power transmission arrangement according to the invention the regenerative braking for reducing the vehicle speed can be used while the third gear is on. Immediately, while the third gear is on, the regenerative braking is started by the electric machine 40 thus somewhat reducing the rpm of the second axle 14. When changing to the second gear to speed up the rotation of the rotor 25 of said electric machine 40 for more effective regenerative braking, the holding friction of the brake- actuator unit 70 is lessened to let the annulus 15 to start a regulated counter-clockwise rotation, which makes the first axle 1 to rotate faster and the annulus 16 to stop its clockwise rotation. At that split-second the non-rotary, grooved gear locking ring 29a and the rotary gear change ring 29c are in tandem axially moved into the stopped, periphery- grooved annulus 16 and the brake-actuator unit 70 is released. The second gear of the EM5 is on and the vehicle is effectively regenerative braked by the electric machine 40.

To change from the second gear to the first gear to speed up the rotation of said rotor 25 of the electric machine 40 for yet more effective regenerative braking, the holding friction of the brake-actuator unit 70 is lessened to let the annulus 16 to start a regulated counter-clockwise rotation, which makes the first axle 1 to rotate still faster and the annulus 17 to stop its clockwise rotation. At that split-second the non-rotary, grooved gear locking ring 29a and the rotary gear change ring 29c are in tandem axially moved into the stopped, periphery-grooved annulus 17 and the brake-actuator unit 70 is released.

The first gear of the EM5 is on and more effective regenerative braking by the electric machine 40 transforms all braking forces into electricity, which is stored for example in the battery 61 to be used at the next start-out or acceleration. The recommended clutch between the primary rotator/motor and the first axle 1 is released if the primary rotator/motor is at whatever phase of a regenerative braking going to slow down the rotation of the first axle 1 , thus wasting the braking torque. Also, for engaging the reverse gear, the first axle 1 is un-clutched from the primary motor and stopped. The annulus 18 stops and is locked stationary by the non-rotary gear locking ring 29a while the rotary gear change ring 29c stays at a free space between the annuli 17 and 1 8. The primary rotator/motor is re-clutched to the first axle 1 . The reverse gear of the EM5 is on. The reverse gear can also be engaged even when the second axle 14 is still rotating clockwise. Then the first axle 1 is first un-clutched from the primary rotator/motor and both the rotary gear change ring 29c and the non-rotary gear locking ring 29a are moved to the free space between the annuli 1 7 and 18. The electric machine 40 is used as a regenerative brake to stop the first axle 1 and then used as an electric motor to counter- clockwise rotate the second axle 14 until the annulus 18 stops. Then the non-rotary gear locking ring 29c is moved to lock the annulus 18 to the stationary gearbox housing 27. The counter rotating electric machine 40 is now used as a regenerative brake, which manoeuvre will cause strong counter-clockwise torque to the power-out axle or the second axle 14. When the power-in axle or the first axle 1 is closer to stop then the primary motor is re-clutched and with strong hybrid torque the second axle 14 is counterclockwise rotated.

The battery capacity can be used to power the electric machine 40 and thus the electric machine 40 can be used at any transmission ratio level as a booster power to assist the primary motor to rotate said first axle 1 . The prescribed hybrid function greatly adds the out-coming torque of the second axle 14 at all the gears. All of the prescribed gear changes take steplessly place during the normal power transfer, therefore neither at the accelerating nor at regenerative phases the rotating power flows are interrupted by gear changes.

Fig. 9 and its explanation are not an exact presentation explaining exact parts and their function, but they are rather pointing to a new type of a gear change method adaptable a multi-gear trans-planetary gear arrangements such as the basic trans-gear arrangement BTG and the embodiments EM1 , EM2, EM4 and EM5 that can be arrayed with this gear change method for smooth, uninterrupted and stepless gear changes.

At regenerative braking the EM5 effectively transfers the rotating powers and torques from the second axle 14 to the rotor 25 of the electric machine 40 forcing it to revolve at higher revolutions and thus enabling the electric machine 40 effectively generate electricity. With automated gear selections the regenerative braking is effective in the tricity. With automated gear selections the regenerative braking is effective in the lower speeds of the vehicle, too.

The regenerative braking can be used to slow down the vehicle and generated electricity can be stored in battery/batteries or in super condensers packets and used at the next start-out or acceleration. This way especially at city driving lots of gas or diesel fuel can be saved and pollution greatly diminished.

In addition to the said regenerative braking and start-out use, the one and the same electric machine 40 can be used in many ways. Whenever the engine is running, the same electric machine 40 works as an alternator for the vehicle's electricity requirements and in driving as a hybrid booster power for quick passing also. In traffic jams the primary motor can be temporarily stopped and only electric mode can be used for driving reducing pollution greatly. When all of the annuli 15-18 can rotate freely there is no driv- ing connection between the first axle 1 and the second axle 14. Therefore at times of electric black-outs the same electric machine 40 can be used as an auxiliary electric power source, also. A motor vehicle with the EM5 transmission, which is equipped with one electric machine 40 can have a considerable smaller gasoline or diesel engine and yet have the same start-out and passing power as a vehicle with a larger engine would have.

Instead of being integrated into the pinion gear assembly 5 each of the planet gears 6-9 can as well be groove-connected to their common planet axle 12 and the planet gear 10 can be joined through said elastic clutch 1 1 and groove connection to said planet axle 12. Thus the planet gears 6-10, groove connected to each other by their common planet axle 12, rotate around the central axis of the planet axle 12 as one unit, also.

As a summary, the trans-planetary power split arrangement (65) without a sun gear is designed in the unlocked state:

- to transfer a high speed rotation to the rotor 54 of the additional electric machine

64 for the effective electricity generation

- to simultaneously transfer the electromagnetic drag as high multiplied torque to the power-through axle 42 - to enable the additional electric machine 64 to function as a primary engine's starting motor

- to enable the additional electric machine 64 to function as an alternator for vehicle's other electricity needs

- to enable the additional electric machine 64 to function as an auxiliary electric power source during electrical black-outs when or wherever auxiliary electric power is needed

- and when locked by the functional clutch arrangement 66 to the rotor 24 of the additional electric machine 64, to enable the primary engine to directly rotate the power-through axle 42 whenever it is beneficial in power transfer and the power transmission arrangement according to the invention is known of:

- the planetary-like construction without a sun gear

- the differential gear construction using two to three or more together-connected planet gears 6-10 on each planet axle 12 meshed to matching the annuli 15-19

- the construction having two or more annuli 15-18 arranged to rotate individually

- the construction where the rotational speed of every chosen annulus 15-18 can be changed individually

- the feature that by operating the gear changing means 29 like the gear locking brake actuators 29e or the non-rotary gear change ring 29a or the rotary gear change ring 29c one of the annuli 15-18 at a time acts as a fulcrum for a designed transmission ratio, and in which construction the braking or locking is done by using a mechanical friction principle, cogged locking, hydraulic principle or electromagnetic principle

- the construction where planet gears 6-10, 39a, 39b on the same planet axle 12 have no relative rotation to each other

- the construction where using small circumferential difference on the planet gears 6-10, 39b, 39a and corresponding annuli 15-19, 41 huge gear ratios are achieved

- the construction where a mechanical reverse gear is created by the diameter difference of the fulcrum annulus 15-8, 41 and the rotary annulus 19

- the construction where some of the annuli 15-18 are carried on their planet gears 6-9 only - the planet axles 12 which are designed to optimally twist under opposite directional torsional loads to distribute even flank pressure to all gears meshing under load and serve as a vibration damper in gearing

- the construction of the special elastic clutch 1 1 with slight torsional flexibility to compensate machining tolerance variations in tangential direction on the same axle 12 and serve as a vibration damper in gearing

- the construction of other different methods for torsional flexibility such as hydraulic, slightly elastic materials

- the construction where by combining trans-planetary gearing assembly 30 to the electro-mechanical parts 40, 59, 61 , 64 new efficient features are achieved, such as effective regenerative braking, opportunity to use electric motors 30, 64 as a hybrid booster power and starting motor

- the construction where by combining trans-planetary assembly 30 parts to the electro-mechanical parts 40, 59, 61 , 64 new efficient features are achieved, such as to use complete construction as a hybrid application, as a totally electric application and totally mechanical application and the transmission mode can be selected and changed while driving

- the construction whereby trans-planetary transmission assembly 30 connected to one electric machine 30, 64 will enable feasible hybrid construction

- the construction whereby normal electric motors 30, 64 can be used in hybrid solutions due to the high obtainable gear ratio

- the construction whereby using minimal gearings high accelerating or reducing gear ratio of 50-120:1 can be obtained

- the construction whereby 40-60% lighter construction of the power transmission arrangement can be obtained

- the construction whereby the exceptionally large effective radius of curvature of its tooth flanks doubles the tooth contact area and accordingly halves the Hertzian contact pressure at every meshing of gears, compared to said tooth contact area and said Hertzian contact pressure at the meshing of a sun gear and the planet gears of a conventional planetary gear set, thus resulting in a lighter and more economical machinery with greatly extended service life.

It is obvious to the person skilled in the art that the invention is not restricted to the example and embodiments described above but that it may be varied within the scope of the claims presented below. Thus, for example, the structure of the arrangement, its functions and its components can vary.

It is also obvious to the person skilled in the art that the structure of the elastic clutch on the planet axles can differ from what is shown above.

It is further obvious to the person skilled in the art that the uses of the power transmission arrangement according to the invention may be also other uses than previously mentioned.

Claims

1 . Power transmission arrangement comprising at least a gear assembly (30) with a rotary first axle (1 ) and a rotary second axle (14), and two or more pinion gear assemblies (5) equipped with two or more gear wheels (6-10, 39a, 39b) on an axle (12), and two or more annuli (15-19, 41 ) meshing with the gear wheels (6-10, 39a, 39b); and a gearbox housing (27), characterized in that the arrangement comprises at least a stationary or stationary lockable annulus (15-18, 41 ) meshing with the coaxial non-equal diameter gear wheels (6-9, 39b) one set of the gear wheels at a time, and a rotary annulus (19) meshing with gear wheels (10, 39a) coaxial with the gear wheels (6-9, 39b) and having a different diameter than the gear wheels (6-9, 39b), and all the gear wheels (6-10, 39a, 39b) on the same axle (12) are fixedly coupled to each other and/or to the same axle (12).

2. Power transmission arrangement according to claim 1 , characterized in that all the meshing tooth contacts of the gear assembly (30) have been arranged to take place between the gear wheels (6-10, 39a, 39b) and the annuli (15-19, 41 ).

3. Power transmission arrangement according to claim 1 or 2, characterized in that the pinion gear assembly (5) comprises at least two or more coaxial gear wheels (6-10, 39a,

39b) on the same axle (12) and each of the gear wheels (6-10, 39a, 39b) has a different pitch diameter, and that each of the annuli (15-19, 41 ) has a different pitch diameter.

4. Power transmission arrangement according to claim 1 , 2 or 3, characterized in that the pinion gear assembly (5) comprises means to transmit torque with a small rotational twist angle between the gear wheels (6-9, 39a) and (10, 39a) on the same planet axle (12) for compensating machining tolerance errors of their tooth meshing.

5. Power transmission arrangement according to claim 4, characterized in that the gear wheel (10, 39b) closest to the power-out end of the gearbox housing (27) comprises an elastic joint (1 1 ) transmitting torque with a small rotational twist angle between the gear wheels (6-9, 39a) on the same planet axle (12) for compensating machining tolerance errors of their tooth meshing and simultaneously serve as a vibration damper in the gearing.

6. Power transmission arrangement according to claim 4, characterized in that for compensating machining tolerance errors of the tooth meshing the axles (12) are designed to optimally twist under the heavy opposite directional torsion loads directed to the struc- ture and serve as a vibration damper in the gearing.

7. Power transmission arrangement according to any of the claims above, characterized in that the arrangement comprises an electric machine (40) containing at least a stator (24) and a rotor (25) fixedly coupled to the first axle (1 ).

8. Power transmission arrangement according to any of the claims above, characterized in that the arrangement comprises a gear changing means (29) to lock one of the annuli (15-18) at a time stationary to the gearbox housing (27).

9. Power transmission arrangement according to any of the claims above, characterized in that the arrangement comprises an additional electric machine (64), a power-split arrangement (65) and a clutch arrangement (66), all of which are carried in bearings by a common rotary third axle (42) between the first axle (1 ) and the second axle (14).

10. Power transmission arrangement according to any of the claims above, characterized in that the gear change ratio is based on the diameter difference of the parallel gears wheels (6-10, 39a, 39b) rotating around the axis of the same rotation axle (12) while circulating around the central axis of the first axle (1 ).

1 1 . Power transmission method comprising at least a gear assembly (30) with a rotary first axle (1 ) and a rotary second axle (14), and one or more pinion gear assemblies (5) equipped with a two or more gear wheels (6-10, 39a, 39b), and a two or more annuli (15- 19, 41 ) meshing with the gear wheels (6-10, 39a, 39b), and a gearbox housing (27), characterized in that the rotating torque is transmitted from the power-in axle to the power-out axle through at least a stationary or stationary lockable annulus (15-18, 41 ) and a rotary annulus (19), both the annuli (15-18, 41 ) and (19) meshing with the coaxial non-equal diameter gear wheels (6-9, 39b) and (10, 39a), which are having no relative rotation to each other.

12. Power transmission method according to claim 1 1 , characterized in that the rotating torque is transmitted from the first axle (1 ) to the second axle (14) through at least the stationary lockable annulus (15-18) one at a time and the rotary annulus (19) both the annuli (15-18) and (19) meshing with the coaxial non-equal diameter gear wheels (6-9) and (10), which are having no relative rotation to each other.

13. Power transmission method according to claim 1 1 , characterized in that the rotating torque is transmitted from the second axle (14) to the first axle (1 ) through at least a stationary annulus (41 ) and the rotary annulus (19) both the annuli (41 ) and (19) mesh- ing with the coaxial non-equal diameter gear wheels (39b) and (39a) having no relative rotation to each other.

14. Power transmission method according to any of the claims 1 1 -13 above, characterized in that all the meshing tooth contacts of the gear assembly (30) are arranged to take place between the gear wheels (6-10, 39a, 39b) on the axles (12) and the annuli (15-19, 41 ).

15. Power transmission method according to any of the claims 1 1 -14 above, characterized in that machining tolerance errors of the tooth meshing in the pinion gear assembly (5) are compensated by transmitting the torque between the first axle (1 ) and the second axle (14) with a small rotational twist angle between the gear wheels (6-9, 39a) and (10, 39a) on the same planet axle (12).

16. Power transmission method according to any of the claims 1 1 -15 above, character- ized in that the rotor (25) of the electric machine (40) is fixedly coupled to the first axle

(1 ) or the third axle (42) and used to assist a primary rotator/motor or alone to rotate the first axle (1 ) or the third axle (42), or slow down the rotation of the first axle (1 ) or the third axle (42 t) in synchronizing the gear changes and in regenerative braking and assisting the primary motor at the start-out and acceleration of the vehicle and as a boost- ing power at passing.

17. Power transmission method according to any of the claims 1 1 -16 above, characterized in that changing from one transmission level to another is effected by a non- motorized yet rotary gear changing method by using the free rotation of chosen, periph- ery-grooved annulus through the synchronizer-ring (71 ), which is rotary with the gear change ring (29c) to first drag said rotary gear change ring (29c) to the fully synchronized even speed with said chosen annulus and locked to it to be stopped by the brake- actuator unit (70).

18. Power transmission method according to any of the claims 1 1 -17 above, characterized in that changing a transmission ratio said non-motorized gear changing method is operated together with the gear changing means (29) to lock one of the annuli (16-18) at a time to the stationary gearbox housing (27).

19. Power transmission method according to any of the claims 1 1 -18 above, characterized in that the additional electric machine (64), the power-split arrangement (65) and the clutch arrangement (66), all of which are carried in bearings by a common rotary third axle (42) between the first axle (1 ) and the second axle (14) are used to assist the primary rotator/motor to rotate the third axle (42), or to stop the rotation of the third axle (42), or to generate electricity.

20. Use of the power transmission arrangement according to claim 1 for power transmission for vehicles or for other machinery as a reducing or accelerating gearing.

21 . Use of the power transmission arrangement according to claim 1 for accelerating gearing in wind turbines.