Epicyclic gear system and gas turbine engine

An epicyclic gear system for a gas turbine engine includes a planet carrier with at least one structural member, on which a planet gear is pivot-mounted by a bearing that is radially arranged between the planet gear and the structural member. Furthermore, a roller bearing device is arranged radially between the planet gear and the bearing. A relative movement between the planet gear and the bearing is prevented by the roller bearing device, if a friction torque in the region of the bearing is less than or equal to a threshold value.

This application claims priority to German Patent Application DE102018123220.9 filed Sep. 20, 2018, the entirety of which is incorporated by reference herein.

The present disclosure relates to an epicyclic gear system, and to a gas turbine engine.

A typical gas turbine engine includes a fan section, a compressor section, a combustor section, and a turbine section. The air entering the compressor section is compressed, and delivered into the combustion section where it is mixed with fuel, and ignited to generate a high-speed exhaust gas flow. The high-speed exhaust gas flow expands through the turbine section to drive the compressor, and the fan section. The compressor section typically includes low-pressure and high pressure compressors, and the turbine section includes low-pressure and high-pressure turbines.

Epicyclic gearboxes with planetary or star gear trains may be used in gas turbine engines for their compact designs, and efficient high gear reduction capabilities. Planetary and star gear trains generally include three gear train elements: a central sun gear, an outer ring gear with internal gear teeth, and a plurality of planet gears supported by a planet carrier between, and in meshed engagement with both the sun gear and the ring gear. The gear train elements share a common longitudinal central axis, about which at least two rotate. An advantage of epicyclic gear trains is that a rotary input can be connected to any one of the three elements. One of the other two elements is then held stationary with respect to the other two to permit the third to serve as an output. In gas turbine engine applications, where a speed reduction transmission is required, the central sun gear generally receives rotary input from the powerplant, the outer ring gear is generally held stationary, and the planet gear carrier rotates in the same direction as the sun gear to provide torque output at a reduced rotational speed. In star gear trains, the planet carrier is held stationary, and the output shaft is driven by the ring gear in a direction opposite that of the sun gear.

Existing gas turbine engines for aircrafts incorporate gearboxes to drive the fan section such that the fan section may rotate at a speed different than the turbine section so as to increase the overall propulsive efficiency of the engine.

In engine architectures incorporating a gearbox, a shaft driven by one of the turbine sections provides an input to the epicyclical gear assembly that drives the fan section at a reduced speed such that both the turbine section and the fan section can rotate closer to optimal speeds.

Such a gearbox incorporates roller or journal bearings to reduce friction losses at the interface between planet gears and carrier. Roller bearings have the disadvantage to be complex and sensible to the high centrifugal forces, while journal bearings are simpler and cope well with high centrifugal forces, but are less robust to noise factors, and have more serious effects on the entire system, in case of seizure.

It is the object of the present disclosure to provide a robust epicyclic gear system as well as a fail-safe gas turbine engine.

This object is achieved through an epicyclic gear system, and with a gas turbine engine with features as disclosed herein.

According to a first aspect there is provided an epicyclic gear system comprising a planet carrier with at least one structural member, on which a planet gear is pivot-mounted by a bearing. The bearing is arranged radially between the planet gear and the structural member. Furthermore, the epicyclic gear system is comprising a roller bearing device that is arranged radially between the planet gear and the bearing. A relative movement between the planet gear and the bearing is prevented by the roller bearing device if a friction torque in the region of the bearing is less than or equal to a threshold value.

The bearing of the epicyclic gear system in accordance with the present disclosure is relieved if the bearing friction increases inadmissibly. This avoids unacceptably high loads on the bearing. For this purpose, the bearing function of the rolling bearing unit is enabled if the friction torque of the bearing is greater than the threshold value.

The bearing may be designed as a journal bearing. Then, the epicyclic gear system is characterized by a simple construction, and copes well with high centrifugal forces.

Alternatively, the bearing may be designed as a roller bearing. This leads to an epicyclic gear system, which is in contrast to the last mentioned epicyclic gear system with a journal bearing more robust to noise factors, and causes less serious effects on the functionality of the entire system of a gas turbine engine in case of seizure. The splitting of the planetary gear ensures the function of the planetary gear in a structurally simple manner even if a failure of the bearing occurs.

The disclosed epicyclic gear system requires only little space if the roller bearing device comprises a radially inner ring which is part of the bearing, and comprises in the area of its radially outer surface several roller ramps, in which roller elements of the roller bearing device are arranged between the inner ring and an outer ring, which outer ring is part of the planet gear.

According to a further aspect, the roller ramps are designed to release a relative movement between the inner ring and the outer ring if a friction torque in the region of the bearing is greater than a threshold value. Then, the epicyclic gear system is operable with low control and loop control effort.

According to a further aspect of the present disclosure, the roller bearing device comprises further roller elements which are arranged radially between the inner ring and the outer ring. By means of these further roller elements the load which is effective during a relative movement between the inner ring and the outer ring, is transmittable between the outer ring and the inner ring if the friction torque in the region of the bearing is greater than the threshold value.

Each roller element of the roller bearing device may be forced by a spring force of at least one spring in a direction of an area of the respective roller ramp, in which the roller elements are locking the relative movement between the inner ring and the outer ring. Then, the epicyclic gear system is also operable with low control and loop control effort.

The threshold value of an epicyclic gear system claimed herein is adaptable by modifying the spring force of the spring.

Furthermore, the threshold of an epicyclic gear system claimed herein is adaptable by modifying the value of the tangent of a clamping angle of the roller ramp.

In a further embodiment of the epicyclic gear device according to the present disclosure the roller bearing device comprises a plurality of roller ramp units. The roller ramp units may be arranged over the circumference of the inner ring and each roller ramp unit having two at least approximately mirror image-like arranged roller ramps and a plateau area arranged there between. The plateau areas each are extending radially further than the associated two roller ramps.

Moreover, in a further embodiment of the epicyclic gear system according to the present disclosure each roller ramp unit is cooperating with at least one roller element. The at least one roller element maybe arranged between two springs. The springs are arranging the roller element in the plateau area of the roller ramp unit, if the friction moment is smaller than or equal to the threshold value.

The epicyclic gear system allows a cost-effective manufacture, and can be operated with low control and low loop control effort if the roller bearing device comprises a freewheel.

The epicyclic gear system can be operated even in the case of seizure of the bearing with high efficiency, if the freewheel is designed as a roller freewheel.

The disclosed epicyclic gear system requires only little space, if the freewheel comprises a radially inner ring, which is part of the bearing, and comprises in the area of its radially outer surface several roller ramps, in which roller elements of the freewheel are arranged between the inner ring and an outer ring, which outer ring is part of the planet gear.

According to a further aspect, the roller ramps are designed to release a relative movement between the inner ring and the outer ring if a friction torque in the region of the bearing is greater than a threshold value. Then, the epicyclic gear system is operable with low control and loop control effort.

Each roller element of the freewheel may be forced by a spring force of a spring in a direction of a first area of the respective roller ramp, in which the roller elements are locking the relative movement between the inner ring and the outer ring. Then, the epicyclic gear system is also operable with low control and loop control effort.

Depending on the respective application, it is possible to design the roller elements as balls, cylinders, barrels or the like.

As noted elsewhere herein, the present disclosure relates to a gas turbine engine. Such a gas turbine engine may include an engine core comprising a turbine, a combustor, a compressor, and a core shaft connecting the turbine to the compressor. Such a gas turbine engine may comprise a fan (having fan blades) located upstream of the engine core.

The unit may be designed as a gearbox, especially as a before mentioned epicyclic gear system that receives an input from the shaft and outputs drive to the fan so as to drive the fan at a lower rotational speed than the shaft. The input to the gearbox may be directly from the core shaft, or indirectly from the core shaft, for example via a spur shaft and/or gear or from a separate turbine. The core shaft may rigidly connect the turbine and the compressor, such that the turbine and compressor rotate at the same speed (with the fan rotating at a lower speed).

The radius of the fan may be measured between the engine centreline and the tip of a fan blade at its leading edge. The fan diameter (which may simply be twice the radius of the fan) may be greater than (or on the order of) any of: 250 cm (around 100 inches), 260 cm, 270 cm (around 105 inches), 280 cm (around 110 inches), 290 cm (around 115 inches), 300 cm (around 120 inches), 310 cm, 320 cm (around 125 inches), 330 cm (around 130 inches), 340 cm (around 135 inches), 350 cm, 360 cm (around 140 inches), 370 cm (around 145 inches), 380 (around 150 inches) cm or 390 cm (around 155 inches). The fan diameter may be in an inclusive range bounded by any two of the values in the previous sentence (i.e. the values may form upper or lower bounds).

The overall pressure ratio of a gas turbine engine as described and/or claimed herein may be defined as the ratio of the stagnation pressure upstream of the fan to the stagnation pressure at the exit of the highest pressure compressor (before entry into the combustor). By way of non-limitative example, the overall pressure ratio of a gas turbine engine as described and/or claimed herein at cruise may be greater than (or on the order of) any of the following: 35, 40, 45, 50, 55, 60, 65, 70, 75. The overall pressure ratio may be in an inclusive range bounded by any two of the values in the previous sentence (i.e. the values may form upper or lower bounds).

Specific thrust of an engine may be defined as the net thrust of the engine divided by the total mass flow through the engine. At cruise conditions, the specific thrust of an engine described and/or claimed herein may be less than (or on the order of) any of the following: 110 Nkg−1s, 105 Nkg−1s, 100 Nkg−1s, 95 Nkg−1s, 90 Nkg−1s, 85 Nkg−1s or 80 Nkg−1s. The specific thrust may be in an inclusive range bounded by any two of the values in the previous sentence (i.e. the values may form upper or lower bounds). Such engines may be particularly efficient in comparison with conventional gas turbine engines.

A gas turbine engine as described and claimed herein may have any desired maximum thrust. Purely by way of non-limitative example, a gas turbine as described and/or claimed herein may be capable of producing a maximum thrust of at least (or on the order of) any of the following: 160 kN, 170 kN, 180 kN, 190 kN, 200 kN, 250 kN, 300 kN, 350 kN, 400 kN, 450 kN, 500 kN, or 550 kN. The maximum thrust may be in an inclusive range bounded by any two of the values in the previous sentence (i.e. the values may form upper or lower bounds). The thrust referred to above may be the maximum net thrust at standard atmospheric conditions at sea level plus 15° C. (ambient pressure 101.3 kPa, temperature 30° C.), with the engine static.

In use, the temperature of the flow at the entry to the high pressure turbine may be particularly high. This temperature, which may be referred to as TET, may be measured at the exit to the combustor, for example immediately upstream of the first turbine vane, which itself may be referred to as a nozzle guide vane. At cruise, the TET may be at least (or on the order of) any of the following: 1400K, 1450K, 1500K, 1550K, 1600K or 1650K. The TET at cruise may be in an inclusive range bounded by any two of the values in the previous sentence (i.e. the values may form upper or lower bounds). The maximum TET in use of the engine may be, for example, at least (or on the order of) any of the following: 1700K, 1750K, 1800K, 1850K, 1900K, 1950K or 2000K. The maximum TET may be in an inclusive range bounded by any two of the values in the previous sentence (i.e. the values may form upper or lower bounds). The maximum TET may occur, for example, at a high thrust condition, for example at a maximum take-off (MTO) condition.

The gas turbine engines described and claimed herein may or may not be provided with a variable area nozzle (VAN). Such a variable area nozzle may allow the exit area of the bypass duct to be varied in use. The general principles of the present disclosure may apply to engines with or without a VAN.

The fan of a gas turbine as described and claimed herein may have any desired number of fan blades, for example 16, 18, 20, or 22 fan blades.

As used herein, cruise conditions may mean cruise conditions of an aircraft to which the gas turbine engine is attached. Such cruise conditions may be conventionally defined as the conditions at mid-cruise, for example the conditions experienced by the aircraft and/or engine at the midpoint (in terms of time and/or distance) between top of climb and start of decent.

Purely by way of example, the cruise conditions may correspond to: a forward Mach number of 0.8; a pressure of 23000 Pa; and a temperature of −55° C.

As used anywhere herein, “cruise” or “cruise conditions” may mean the aerodynamic design point. Such an aerodynamic design point (or ADP) may correspond to the conditions (comprising, for example, one or more of the Mach Number, environmental conditions and thrust requirement) for which the fan is designed to operate. This may mean, for example, the conditions at which the fan (or gas turbine engine) is designed to have optimum efficiency.

In use, a gas turbine engine described and claimed herein may operate at the cruise conditions defined elsewhere herein. Such cruise conditions may be determined by the cruise conditions (for example the mid-cruise conditions) of an aircraft to which at least one (for example 2 or 4) gas turbine engine may be mounted in order to provide propulsive thrust.

Other gas turbine engines to which the present disclosure may be applied may have alternative configurations. For example, such engines may have an alternative number of compressors and/or turbines and/or an alternative number of interconnecting shafts. By way of further example, the gas turbine engine shown inFIG. 1has a split flow nozzle20,22meaning that the flow through the bypass duct22has its own nozzle that is separate to and radially outside the core engine nozzle20.

However, this is not limiting, and any aspect of the present disclosure may also apply to engines in which the flow through the bypass duct22and the flow through the core11are mixed, or combined, before (or upstream of) a single nozzle, which may be referred to as a mixed flow nozzle. One or both nozzles (whether mixed or split flow) may have a fixed or variable area. Whilst the described example relates to a turbofan engine, the disclosure may apply, for example, to any type of gas turbine engine, such as an open rotor (in which the fan stage is not surrounded by a nacelle) or turboprop engine, for example.

FIG. 4shows a three-dimensional view of one of the planet gears32and a part of the planet carrier34of the epicyclic gear system30according toFIG. 3. The planet carrier34has structural members29, which are shown schematically inFIG. 3and on which the planet gears32are pivot-mounted by bearings41. The bearings41are arranged radially between the planet gears32and the structural members29, which are designed as planet carrier pins. Furthermore, the bearings41are designed as journal bearings. Furthermore, roller bearing devices42are arranged radially between the planet gears32and the bearings41. In the present embodiment each roller bearing device42comprises a freewheel42A. The freewheels42A are designed as roller freewheels. The planet gears32have only a low friction torque due to the viscous friction of the medium of the journal bearings41.

This friction torque is balanced by a different distribution of forces between teeth55,56of the planet gears32, which are in contact with teeth57of the sun gear28and with teeth58of the ring gear38. Under normal conditions, with the torque coming from the sun gear28and exiting through the planet carrier34, the overloaded tooth55is the one toward the sun gear28. Under windmilling conditions, it is the tooth56towards the ring gear38which carries more load.

As shown inFIG. 5every roller bearing device42comprises a radially inner ring43, which is part of the bearing41. Each of the radial inner rings43of the roller bearing devices42comprises several roller ramps44in the area of its radially outer surface45, in which roller elements46of the freewheels42A are rotatable arranged between the inner rings43and outer rings47, which are part of the planet gears32. The roller ramps44are designed to unblock a relative movement between the inner ring43and the outer ring47when a friction torque in the region of the bearing41is greater than a threshold value.

In case of degradated functionality, failure or seizure of the journal bearing41the first effect is an increase of the friction torque and a subsequent increase of unbalanced distribution of forces between the teeth55,56,57,58. This could lead to different kinds of major failures of the entire system. To avoid that and to limit the increase of the friction it is proposed to back-up the main function of the journal bearings41and to pass it to the roller bearing devices42with their roller bearings59automatically if the friction torque exceeds a predefined threshold value. As described below the threshold value can be tuned by the design of the roller bearing devices42.

With reference toFIG. 6, bearing cages48are shown whose movements are limited radially by a circular raceway61of the inner ring43and tangentially by springs49shown inFIG. 7andFIG. 8. The springs49are arranged between parts of the cages48and castellations62of the inner ring43which are illustrated inFIG. 6. As shown simplified inFIG. 7andFIG. 8, each roller element46of the freewheels42A is forced in the direction of an area50of the respective roller ramp44by a spring force F of the springs49, in which direction the roller elements46are blocking the relative movement between the radial inner rings43and the outer rings47.

In the operating state shown inFIG. 7the outer rings47can be turned freely clockwise (free-wheeling operation), if the radial inner rings43are at a standstill, are turned counter clockwise or are turned clockwise slower than the outer rings47. The load is transferred from the inner ring43to the outer ring47via further roller elements63of the freewheels42A which are arranged radially and rotatable between an outer raceway64of the inner ring43and an inner raceway65of the outer ring47.

If the outer rings47—e.g. with stationary radial inner rings43—are turned in the opposite direction, the clamping of the roller elements46in the roller ramps44of the freewheels42A becomes effective. The roller elements46clamp without slipping between the radial inner rings43and the outer rings47. In this direction of rotation high torque can be transmitted.

The embodiment of the freewheel42A illustrated inFIG. 7andFIG. 8also enables a freewheeling operation while the radial inner rings43are turned counter-clockwise and a driving operation when turning clockwise.

On the line51of influence which links the points52,53of contact of the roller elements46to the roller ramps44and the roller elements46to the outer tracks54of the outer rings47, in driving operation the clamping generates forces FI and FA. Because of the equilibrium of forces, these are equal. The forces FI and FA can be divided into normal forces FNI and FNA as well as into the circumferential forces FTI and FTA. The line51of influence forms against the force FNI or FNA a clamping angle ε. To achieve self-locking, the tangent of the clamping angle ε must be less than the friction value μ. E.g. for the contact point53of the roller46to the outer track54this means:
tan ε=FTA/FNA≥μ

Because of the relationship:
M=z·RA·FTA=z·RA·FNA·tan ε

z: number of roller elements46

RA: radius of the outer track54

The normal forces FNI and FNA as well as the clamping angle ε adapt automatically to the acting torque M.

According to the above mentioned explanation the threshold value may be adaptable by modifying the spring force F of the spring49and/or by modifying the value of the tangent of the clamping angle ε of the roller ramp44.

With reference toFIG. 5, the outer ring47rotates clockwise with respect to pin29and the freewheels42A are locked when the friction torque is lower than the predefined treshold value. In the last mentioned operational state of the freewheels42A all the roller elements46are pressed against the outer rings47and the radial inner rings43, which both spin at the same rotating speed. This operational state of the freewheels42A is shown inFIG. 9andFIG. 10. A part of the planet gear32, which comprises the teeth56, can be seen inFIG. 9, whereasFIG. 10shows a different part of the planet gear32including the teeth55.

The low torque load paths go through the roller elements46, which also carry the rotating force vector of the centrifugal force and an additional force originated by the planet carrier34torque coming from the pin29. The last-mentioned torque acts radially on the planet gear32and rotates with respect to it. Under this operational state, roller elements63of the roller bearings59are not in contact with the outer rings47and the radial inner rings43, so they are not affected by the load during normal operative conditions.

If the friction torque exceeds the threshold value the roller elements46of the freewheels42A run downwards over the roller ramps44and get out of contact with the outer rings47and with the radial inner rings43. Hereafter the outer rings47are pushed outwards radially by the centrifugal force and tangentially with respect to the planet carrier reference system by the force generated by the gearbox torque until the roller elements63of the roller bearings59are in contact with the outer rings47and the radial inner rings43. At that point, the load paths change, and the outer rings47, which are then carried by the roller bearings59start spinning.

If the friction torque of the journal bearings41again falls below the threshold value, the relative movement between the outer rings47and the radial inner rings43will be blocked by the freewheels42A, which will then be pushed by the revolving friction of the roller elements46. At this operational state the journal bearings41will start to operate normally.

FIG. 11shows a further embodiment of the epicyclic gear system30in an illustration according toFIG. 4. Each of the roller bearing devices42of the epicyclic gear system30shown inFIG. 11comprises a plurality of roller ramp units66which are arranged over the circumference of the inner ring43. Furthermore, each of the roller ramp units66has two at least approximately mirror image-like arranged roller ramps67,68, and a plateau area69arranged there between which is shown in more detail inFIG. 12. The plateau area69extends radially further than surfaces70,71of the two roller ramps67,68.

Moreover, each roller ramp unit66is cooperating with one roller element72. The roller elements72are arranged radially between the inner ring43and the outer ring47and are also arranged in a bearing cage73. The bearing cage73comprises castellations74which extend in axial direction of the planet gears32. The inner ring43also provides castellations75which extend in radial direction of the planet gears32. In circumferential direction of the inner ring43springs76,77are arranged between the castellations74of the bearing cage73and the castellations75of the inner ring43. The forces F76and F77of the springs76,77are directed opposite to each other, so that the springs76,77are arranging the roller elements72in the plateau areas69of the roller ramp units66, if the friction force of the bearing41is smaller than or equal to the threshold value.

In this so-called normal condition of the epicyclic gear system30only a small torque is applied to the roller bearing devices42. The roller elements72are moving in this operational state in circumferential direction of the inner ring43slightly left or right. With force equilibrium of the springs76,77the roller elements72are remaining as shown inFIG. 13andFIG. 14between the inner ring43and the outer ring47on the radially higher surface areas of the plateau areas69and close the load path between the inner ring43and the outer ring47. In this condition no relative movements happen in circumferential direction of the rings43and47between the inner ring43and the outer ring47.

If the friction force of the bearing41exceeds the threshold value, the torque applied to the roller bearing devices affects movement of the roller elements72and the bearing cage73as shown inFIG. 15toFIG. 18in circumferential direction of the rings43and47from the plateau areas69to the lower surface areas of the roller ramps67or68. In this so-called failure condition, the load path is closed by further roller elements78, and a relative movement in circumferential direction of the rings43and47between the rings43and48is released.

In the embodiment shown inFIG. 11toFIG. 18, the contact between the planet gear32and the ring gear38or sun gear28respectively depends on the operating condition of the gear box30. The gearbox30maybe subject to only torque, to only rotational speed or to torque and rotational speed.

If the gearbox30is subject to only torque, the direction of the axis passing through the gap closure and the maximum clearance is oriented tangentially with respect to the reference system centered with the main axis of the carrier34.

If the gearbox30is subject to only rotational speed, the above mentioned axis is oriented radially with respect to the carrier reference system. Hence in this case, tooth56is the tooth toward the ring gear38and tooth55is the one toward the sun gear28. The orientation of the axis will be determined by the prevailing of the centrifugal vs. the tangential forces and it will be angularly in the range comprised by the extreme conditions with only torque or only rotational speed.

The positive effect of the described solutions is that the centrifugal force and the force generated by the gearbox torque are not acting anymore on the journal bearings41if the friction in the area of the journal bearing exceeds a threshold value. This configuration of the epicyclic gear system offers the possibility of not causing a further increase of friction torque that would probably lead to an entire system failure. The journal bearings are now supported by the roller bearings at the sides of the roller bearing devices. The revolving friction in the area of the roller bearing devices is higher than the friction force which arises in the area of a journal bearing under normal conditions. On the other hand is the revolving friction lower than the friction which is generated by a journal bearing with a malfunction. This leads to a very robust system. The sliding contacts during the rotation of the roller ramp devices, e.g. between the springs, the roller ramps and the roller elements develop a dynamic friction where the abovementioned centrifugal force and gearbox torque generated force have only negligible influence.

PARTS LIST