METHOD OF POSITIONING TWO PARTS RELATIVE TO EACH OTHER IN A FORMLOCKING CONNECTION, A FORMLOCKING DEVICE AND GAS TURBINE ENGINE

Abstract

A method for positioning two parts relative to each other in a formlocking connection, the two parts having a plurality of contact faces o transmit contact forces between the two parts, computing for each possible relative position of the two parts the contact forces between the two parts and determining the relative position of the two parts with the minimal contact force among a plurality of relative positions or all possible relative positions and assembling the two parts relative to each other in the position with the minimal contact force. A formlocking device and a gas turbine engine.

Claims

1. A method for positioning two parts with existing, fixed geometries relative to each other in a formlocking connection, the two parts having a plurality of contact faces o transmit contact forces between the two parts, a) computing for each possible relative position of the two parts the contact forces between the two parts and b) determining the relative position of the two parts with the minimal contact force among a plurality of relative positions or all possible relative positions and c) assembling the two parts relative to each other in the position with the minimal contact force.

2. The method according to claim 1, wherein the minimal contact force is defined as the minimal absolute value of the contact forces at the determined relative positions or the average minimal value of the contact forces at the determined relative positions.

3. The method according to claim 1, wherein the alternating contact force is minimized.

4. The method according to claim 1, wherein at least one geometric property is measured in the area of the contact forces.

5. The method according to claim 4, wherein the at least one geometric property is the positional tolerance between the two parts.

6. The method according to claim 1, wherein the formlocking connection comprises a cylindrical spline connection, in particular with a polygonal connection, square splines, serrated splines, straight side splines, involute shaped splines or tapered tooth splines.

7. The method according to claim 1, wherein the formlocking connection comprises a face spline connection, in particular a radial serration spline or a curvic coupling.

8. The method according to claim 1, wherein the first part is a sun gear and the second part is a sun gear of a planetary gearbox, in particular in a geared turbofan engine in an aircraft.

9. A formlocking device with two parts having a plurality of contact faces to transmit contact forces between the two parts, the two parts positionable by a method of claim 1.

10. A gas turbine engine for an aircraft comprising: an engine core comprising a turbine, a compressor, and a core shaft connecting the turbine to the compressor; a fan located upstream of the engine core, the fan comprising a plurality of fan blades; and a gearbox that receives an input from the core shaft and outputs drive to the fan so as to drive the fan at a lower rotational speed than the core shaft, wherein the gearbox comprises a sun gear and a sun shaft assembled by a method of claim 1.

Description

[0052] Embodiments will now be described by way of example only, with reference to the Figures, in which:

[0053] FIG. 1 is a sectional side view of a gas turbine engine;

[0054] FIG. 2 is a close up sectional side view of an upstream portion of a gas turbine engine;

[0055] FIG. 3 is a partially cut-away view of a gearbox for a gas turbine engine;

[0056] FIG. 4A is a frontal view of a sun shaft connected to a sun gear;

[0057] FIG. 4B is sectional sideview of the view of FIG. 4A;

[0058] FIG. 4C shows a form locking connection between the sun gear and the sun shaft shown in FIGS. 4A-4B;

[0059] FIG. 5A is a graph showing the absolute positional tolerances of an external spline;

[0060] FIG. 5B is a graph showing the absolute positional tolerances of an internal spline;

[0061] FIG. 6A is a graph showing the cumulative positional tolerances of the external spline in FIG. 5A;

[0062] FIG. 6B is a graph showing the cumulative positional tolerances of the internal spline in FIG. 5B;

[0063] FIG. 7A is a graph indicating the minimized gaps between the teeth (clocking #1);

[0064] FIG. 7B is a graph indicating a further gap configuration (clocking #2);

[0065] FIG. 8 is a graph indicating the spline contact force distribution with equal load share and in an idealized spline;

[0066] FIG. 9 is a graph showing the spline contact force distribution with a non-uniform gap, clocking position 1#, planet gear position 1;

[0067] FIG. 10 is a graph showing the spline contact force distribution with a non-uniform gap, clocking position 1#, planet gear position 2;

[0068] FIG. 11 is a graph showing the spline contact force distribution with a non-uniform gap, clocking position 2#, planet gear position 1;

[0069] FIG. 12 is a graph showing the spline contact force distribution with a non-uniform gap, clocking position 2#, planet gear position 2;

[0070] FIG. 13 is a graph showing the maximal spline contact force for each clocking position;

[0071] FIG. 14 is a graph showing the spline contact forces for a worst case and a best case.

[0072] FIG. 1 illustrates a gas turbine engine 10 having a principal rotational axis 9. The engine 10 comprises an air intake 12 and a propulsive fan 23 that generates two airflows: a core airflow A and a bypass airflow B. The gas turbine engine 10 comprises a core 11 that receives the core airflow A. The engine core 11 comprises, in axial flow series, a low pressure compressor 14, a high-pressure compressor 15, combustion equipment 16, a high-pressure turbine 17, a low pressure turbine 19 and a core exhaust nozzle 20. A nacelle 21 surrounds the gas turbine engine 10 and defines a bypass duct 22 and a bypass exhaust nozzle 18. The bypass airflow B flows through the bypass duct 22. The fan 23 is attached to and driven by the low pressure turbine 19 via a shaft 26 and an epicyclic gearbox 30.

[0073] In use, the core airflow A is accelerated and compressed by the low pressure compressor 14 and directed into the high pressure compressor 15 where further compression takes place. The compressed air exhausted from the high pressure compressor 15 is directed into the combustion equipment 16 where it is mixed with fuel and the mixture is combusted. The resultant hot combustion products then expand through, and thereby drive, the high pressure and low pressure turbines 17, 19 before being exhausted through the nozzle 20 to provide some propulsive thrust. The high pressure turbine 17 drives the high pressure compressor 15 by a suitable interconnecting shaft 27. The fan 23 generally provides the majority of the propulsive thrust. The epicyclic gearbox 30 is a reduction gearbox.

[0074] An exemplary arrangement for a geared fan gas turbine engine 10 is shown in FIG. 2. The low pressure turbine 19 (see FIG. 1) drives the shaft 26, which is coupled to a sun wheel, or sun gear, 28 of the epicyclic gear arrangement 30. Radially outwardly of the sun gear 28 and intermeshing therewith is a plurality of planet gears 32 that are coupled together by a planet carrier 34. The planet carrier 34 constrains the planet gears 32 to precess around the sun gear 28 in synchronicity whilst enabling each planet gear 32 to rotate about its own axis. The planet carrier 34 is coupled via linkages 36 to the fan 23 in order to drive its rotation about the engine axis 9. Radially outwardly of the planet gears 32 and intermeshing therewith is an annulus or ring gear 38 that is coupled, via linkages 40, to a stationary supporting structure 24.

[0075] Note that the terms low pressure turbine and low pressure compressor as used herein may be taken to mean the lowest pressure turbine stages and lowest pressure compressor stages (i.e. not including the fan 23) respectively and/or the turbine and compressor stages that are connected together by the interconnecting shaft 26 with the lowest rotational speed in the engine (i.e. not including the gearbox output shaft that drives the fan 23). In some literature, the low pressure turbine and low pressure compressor referred to herein may alternatively be known as the intermediate pressure turbine and intermediate pressure compressor. Where such alternative nomenclature is used, the fan 23 may be referred to as a first, or lowest pressure, compression stage.

[0076] The epicyclic gearbox 30 is shown by way of example in greater detail in FIG. 3. Each of the sun gear 28, planet gears 32 and ring gear 38 comprise teeth about their periphery to intermesh with the other gears. However, for clarity only exemplary portions of the teeth are illustrated in FIG. 3. There are four planet gears 32 illustrated, although it will be apparent to the skilled reader that more or fewer planet gears 32 may be provided within the scope of the claimed invention. Practical applications of a planetary epicyclic gearbox 30 generally comprise at least three planet gears 32.

[0077] The epicyclic gearbox 30 illustrated by way of example in FIGS. 2 and 3 is of the planetary type, in that the planet carrier 34 is coupled to an output shaft via linkages 36, with the ring gear 38 fixed. However, any other suitable type of epicyclic gearbox 30 may be used. By way of further example, the epicyclic gearbox 30 may be a star arrangement, in which the planet carrier 34 is held fixed, with the ring (or annulus) gear 38 allowed to rotate. In such an arrangement the fan 23 is driven by the ring gear 38. By way of further alternative example, the gearbox 30 may be a differential gearbox in which the ring gear 38 and the planet carrier 34 are both allowed to rotate.

[0078] It will be appreciated that the arrangement shown in FIGS. 2 and 3 is by way of example only, and various alternatives are within the scope of the present disclosure. Purely by way of example, any suitable arrangement may be used for locating the gearbox 30 in the engine 10 and/or for connecting the gearbox 30 to the engine 10. By way of further example, the connections (such as the linkages 36, 40 in the FIG. 2 example) between the gearbox 30 and other parts of the engine 10 (such as the input shaft 26, the output shaft and the fixed structure 24) may have any desired degree of stiffness or flexibility. By way of further example, any suitable arrangement of the bearings between rotating and stationary parts of the engine (for example between the input and output shafts from the gearbox and the fixed structures, such as the gearbox casing) may be used, and the disclosure is not limited to the exemplary arrangement of FIG. 2. For example, where the gearbox 30 has a star arrangement (described above), the skilled person would readily understand that the arrangement of output and support linkages and bearing locations would typically be different to that shown by way of example in FIG. 2.

[0079] Accordingly, the present disclosure extends to a gas turbine engine having any arrangement of gearbox styles (for example star or planetary), support structures, input and output shaft arrangement, and bearing locations.

[0080] Optionally, the gearbox may drive additional and/or alternative components (e.g. the intermediate pressure compressor and/or a booster compressor).

[0081] 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 in FIG. 1 has a split flow nozzle 20, 22 meaning that the flow through the bypass duct 22 has its own nozzle that is separate to and radially outside the core engine nozzle 20. However, this is not limiting, and any aspect of the present disclosure may also apply to engines in which the flow through the bypass duct 22 and the flow through the core 11 are 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. In some arrangements, the gas turbine engine 10 may not comprise a gearbox 30.

[0082] The geometry of the gas turbine engine 10, and components thereof, is defined by a conventional axis system, comprising an axial direction (which is aligned with the rotational axis 9), a radial direction (in the bottom-to-top direction in FIG. 1), and a circumferential direction (perpendicular to the page in the FIG. 1 view). The axial, radial and circumferential directions are mutually perpendicular.

[0083] In the following, an embodiment of a method for positioning (or assembling) two parts in a formlocking connection is described in the context of an epicyclic gearbox 30 in a geared turbofan engine.

[0084] The person skilled in the art will realize that other embodiments can be used in the context of other gearboxes, e.g. in stationary turbo engines. Other embodiments of the method can e.g. be used for spline connections outside a gearbox 30 context. This can e.g. be in context of a shaft connection with cylindrical splines (square splines, serrated splines, straight side splines, involute shaped splines or tapered tooth splines). Alternatively, an embodiment can be used for face splines, such as a radial serration spline or a curvic coupling.

[0085] With an epicyclic gearbox 30 in the context of a geared turbofan engine, a spline connection 35 is used to connect the sun shaft 31 and sun gear 28. Here, the sun gear 28 can be considered as the first part, the sun shaft 31 as the second part in the formlocking spline connection.

[0086] The teeth of the planet gears 32 mesh with the outer teeth of the sun gear 28. There exists a mechanical interaction between the planet teeth meshing with the sun gear 28 and the sun gear 29 to sun shaft spline 35 with contact forces F. This interaction is especially pronounced in the gearboxes 30 in geared turbofan engines due to its relative thin cross-section. The thin cross-section is necessary to accommodate the fan shaft, as well as to minimize weight.

[0087] In use, this interaction inflicts an alternating contact force on each spline tooth as a planet gear 32 passes. The number of alternating cycles equals the number of planet gears 32. In the gearbox 30 design shown in FIGS. 3, 4A and 4B, five planet gears 32 are used, thus each spline tooth would go through five alternating cycles per one revolution of the sun gear 28. As a result, this alternating cycle can have a dramatic effect on the HCF (High Cycle Fatigue) life of the spline connection 35. It should be noted that five planet gears 32 are just chosen here as an example. Other embodiments might have more than five planet gears 32 or fewer.

[0088] In FIG. 4C the formlocking between the two parts 28, 31 is shown in detail to indicate the contact forces F when e.g. the second part 31 rotates relative to the first part 28.

[0089] The spline connection 35 comprises two parts: The internal spline (first part, on the sun gear 28) and the external spline (second part, on the sun shaft 31).

[0090] The embodiment described herein comprises spline teeth measurement data for every part to determine a distinct clocking configuration to assemble every pair of sun shaft 31 and sun gear 28 such that alternating contact force is minimized and hence the HCF strength of the parts are maximized.

Input Data

[0091] The teeth of the spline connection, i.e. the two parts 28, 31 are considered as parts having fixed geometries. This means, the geometries as such are not changed during the following procedure. The procedure is about assembling the two parts in an optimal way.

[0092] The spline connection 35 in the exemplary embodiment is specified as a Tolerance Class of 5. This is a standard that specifies the tolerance associated with the spline connection. The embodiments described herein can be applied to other tolerance classes. In essence, embodiments are applicable to any spline that is manufactured with a method that produces a distribution of resulting geometries:

between teeth 180 apart: 173 microns for both splines together (or 86.5 microns each spline).
between two adjacent teeth: 43 microns. (which is of the sum of internal and external accumulative pitch deviations.

[0093] Based on the Tolerance Class 5, the internal and external spline set of positional tolerances are calculated or measured, i.e. a deviation from a design value (FIG. 5A, 5B). FIG. 5A shows the positional tolerances on the external splines on the sun shaft 31. FIG. 5B shows the positional tolerances on internal splines on the sun gear 28. As expected, the distribution of the tolerances data over the 84 teeth of the spline connection is random, but they not necessarily follow a normal distribution. They will follow a distribution affected by the method of manufacture.

[0094] The cumulative positional tolerances are shown in FIG. 6A, 6B of the external spline (FIG. 6A) and the internal spline (FIG. 6B), as a verification of the adherence to the Tolerance Class. From FIGS. 6A, B it can be seen that no teeth 180 apart have a tolerance of more than 173 microns. From FIG. 6A, 6B it can also be seen that no two adjacent teeth have a tolerance of more than 43 microns.

[0095] As mentioned above, this can be applied to any tolerance class.

[0096] Based on the external and internal positional tolerances (see FIGS. 5A, 5B), a gap is then calculated for each spline tooth pair with existing, fixed geometries, assuming the external and internal splines are fitted together. The gap is the difference between the positional tolerances shown in FIG. 5A and 5B.

[0097] The resulting gap between the two splines can be calculated. The position of each spline contact surface is measured relative to the design (i.e. drawing) specifications. For example, if the spline contact surface was manufactured exactly according to the specification the resulting measurement would be 0 mm.

[0098] In the following, a measuring convention is established. Consider the embodiment shown in FIG. 4C and a torque application in a clockwise (CW) rotation as shown. If the contact surface is measured to be in the direction of rotation (CW in this case) compared to the nominal position, then the measured value is positive. If the measured position is in the direction opposite to the rotation (CCW in this case) then it is recorded with a negative value. If row 1 of the table below is considered, the contact surface of the external spline (left surface of 31) is measured to be more CCW than the drawing specification and thus takes a negative value of 0.0098 mm. The position of the contact surface of the internal spline (Right surface of 28) is more CW than the drawing specification and thus takes a positive value of 0.0032 mm. In this case, both contact surfaces are contributing to a gap compared to their nominal position (both contact surfaced moved in opposite directions) and thus a theoretical gap of 0.0130 mm exists for this tooth pair.

[0099] Another example with interference between teeth is considered. External tooth 21 was measured to be 0.0115 mm CW from nominal position and internal tooth 21 was measured to be 0.0108 CCW from nominal position. This means that both teeth are positioned closer together from nominal position. This pair of teeth has a theoretical interference of 0.0223 mm.

[0100] This gap calculation is performed for all the tooth pairs of the external and internal spline. After gap calculation has been made, a new adjusted gap needs to be calculated. The tooth pair with the maximum calculated interference indicates which pair of teeth will come into contact first as torque is applied to create a CW rotation. Imagine the two parts are fitted together, then the shaft 31 transmits a torque and is rotated CW until contact occurs. In this case, the contact will first occur at tooth pair 21. Because tooth pair 21 has an interference of 0.0223 mm compared to the nominal position, this means that all other pairs must be adjusted by the interference of tooth pair 21. Therefore, the external spline (31) needs to be rotated CCW by 0.0223 mm (negative value because CCW). Therefore tooth pair 1 gap becomes 0.01300.0223=0.0353 mm.

[0101] These adjusted gaps are then used in the FEM to calculate the contact forces.

[0102] The adjusted gap values change for other clocking positions. For example, if external tooth 1 is matched with internal tooth 2 the resulting gap will be 0.00980.0042=0.0140 mm. Then the new max interference tooth pair needs to be determined and the adjusted gaps can be calculated. This shows on how the measurements are used to calculate the gaps that are then fed into the FEM analysis.

TABLE-US-00001 External Internal gap adj gaps 1 0.0098 0.0032 0.0130 0.0353 2 0.0022 0.0042 0.0064 0.0287 3 0.0040 0.0027 0.0067 0.0290 4 0.0110 0.0021 0.0131 0.0354 5 0.0010 0.0062 0.0072 0.0151 6 0.0001 0.0083 0.0084 0.0307 7 0.0049 0.0115 0.0164 0.0059 8 0.0084 0.0020 0.0064 0.0159 9 0.0090 0.0109 0.0199 0.0422 10 0.0114 0.0054 0.0060 0.0283 11 0.0039 0.0020 0.0059 0.0164 12 0.0040 0.0010 0.0030 0.0193 13 0.0049 0.0068 0.0019 0.0242 14 0.0045 0.0050 0.0005 0.0218 15 0.0016 0.0038 0.0054 0.0277 16 0.0067 0.0092 0.0159 0.0382 17 0.0000 0.0120 0.0120 0.0343 18 0.0088 0.0056 0.0144 0.0367 19 0.0050 0.0027 0.0023 0.0200 20 0.0020 0.0074 0.0054 0.0169 21 0.0115 0.0108 0.0223 0.0000

[0103] Then the minimum resulting gap of all of the tooth pairs is calculated and all the spline pair tooth gaps are offset by this amount.

[0104] In effect, the minimum tooth pair gap becomes zero as the first tooth pair comes into contact. The end result is the gap of each tooth pair as shown in FIG. 7A termed clocking #1. The cumulative gap is smallest in this configuration

[0105] The internal and external spline can be assembled in many ways as the number of spline teeth (clocking). Anothernon-optimalgap distribution is shown in FIG. 7B, i.e. at clocking #2.

[0106] In the following, these gaps are inputs for a FEM analysis that is used to calculate the contact force on each spline pair when torque is applied.

Analysis

[0107] To simplify the analysis, a FEM model was created including the sun gear 28 and a partial sun shaft 31. A moment is then applied at the sun shaft 31, which in turn transfers it via the spline connection 35 to the sun gear 28 and finally the load is reacted at the helical teeth representing the five planet gears 32 (see FIG. 4A). The planet gears 32 are in this particular case not modelled but represented as normal to tooth constraints. The spline connection 35 is modelled with frictional non-linear contact.

Idealized Spline

[0108] The contact forces between the teeth of the spline 35 are calculated when considering an idealised spline (equal load share amongst spine teeth, all gaps=0 mm) as shown in FIG. 8. As the planet gear 32 rotates around the sun shaft 31, the contact force varies from roughly 7.6 kN to 11 kN.

[0109] In this idealized calculation the actually measured tolerances of the teeth are not considered.

Spline Contact Force Including Positional Tolerances

[0110] The same FEM program can be used to capture the variation in spline contact force when the positional tolerances as described above are considered.

[0111] The gap of each tooth pair 28, 31 having fixed geometries is simulated by offsetting the spline contact surfaces by the specified amount. Using the gap distribution of each clocking position the resulting spline contact force of each tooth pair is then calculated. To capture completely the variation in contact force the following is considered:

84 clocking positions (84 tooth spline)
For each clocking position, for each tooth, two planet gear 32 positions are required to capture the max and min contact force

[0112] When the spline tooth is directly below the planet gear 32, that tooth sees the maximum possible contact force (shortest load path).

[0113] When the spline tooth is between two planet gears 32, that tooth sees the minimum possible contact force (longest load path).

[0114] This yields a total of 84*2=168 analyses to determine which clocking position results in the better distribution of contact force, if the tooth that generates the maximum contact force is known.

[0115] FIG. 9 shows the contact force distribution for the clocking position #1 (see FIG. 7A) taking into account the positional tolerances. Clearly, the spline contact force distribution is no longer a smooth pattern as in FIG. 8. The variation in force has significantly increased when taking into account the positional tolerances. The contact forces vary between 0 and 38 kN.

[0116] FIG. 10 shows what happens to the contact force distribution when another planet gear 32 position is analyzed.

[0117] FIG. 11 shows how the contact force distribution changes when the parts are assembled in clocking position #2 (see FIG. 7B). The contact forces vary between 0 and 36 kN.

[0118] FIG. 12 shows what happens to the contact force distribution when another planet position is analyzed for clocking #2.

[0119] By comparing the contact force data for the clocking position #1 with that of clocking position #2 it is evident that the maximal contact forces are significantly reduced in the clocking #2 case compared to clocking #1. These clocking positions were randomly selected and are only 2 of 84 possible to illustrate the effect.

Optimal Clocking Position

[0120] As stated, there are as many possible clocking positions as there are spline teeth. Using the FEM, the spline contact force of every tooth pair for every clocking position can be calculated.

[0121] FIG. 13 shows the resulting maximum spline contact force of each of the clocking positions. Clearly, there is large variation ranging from 52.7 kN to 35.3 kN for clocking position #14 and #68 respectively.

[0122] For clocking position #14 (maximum contact force) and #68 (minimum contact force), FIG. 4 shows in detail what happens to the spline contact force of one tooth as a planet gear 32 passes by.

[0123] The optimal clocking position in the case shown is clocking position #68. If the parts are assembled in clocking position #68, this leads to a reduction of the maximum contact force by 33% and contact force range by 27% (FIG. 4) compared against the worst case (see Table 1).

[0124] Overloading one tooth can lead to the entire spline failing and should be avoided.

[0125] The embodiments of a method presented comprise spline teeth measurement data of every part to determine a distinct clocking configuration to assemble every pair of sun shaft 31 and sun gear 28 such that the alternating contact force is minimized and hence the HCF strength of the parts are maximized.

[0126] With an epicyclic gearbox 30 utilising a spline connection between sun shaft 31 and sun gear 28, an interaction exists between the teeth of planet gears 32 meshing with the sun gear 28 and the sun gear 28 to sun shaft spline contact forces. This interaction is especially pronounced in a gearbox 30 of a geared turbofan engine due to its relative thin cross-section. The thin cross-section is necessary to accommodate the fan shaft, as well as to minimize weight. In effect, this interaction inflicts an alternating contact force on each spline tooth as the planet gear 32 passes.

[0127] The embodiment comprises obtaining spline teeth measurement data for every part to determine a distinct clocking configuration to assemble every pair of sun shaft and sun gear such that alternating contact force is minimized and hence the HCF strength of the parts are maximized.

[0128] With any epicyclic gearbox 30 an interaction exists between the planet teeth meshing with the sun gear 28 and the sun gear 28 to sun shaft spline contact forces. This interaction is especially pronounced in a weight-optimised design (such as the PGB) leading to the sun gear having a relative thin cross-section. In effect, this interaction inflicts an alternating contact force on each spline tooth as the planet passes. The number of alternating cycles equals the number of planet gears. In one embodiment five planet gears 32 are used, thus each spline tooth would go through five alternating cycles per one revolution of the sun gear 28. As a result, this alternating cycle can have a dramatic effect on the HCF life of the spline 35.

[0129] It will be understood that the invention is not limited to the embodiments described above and various modifications and improvements can be made without departing from the concepts described herein. Except where mutually exclusive, any of the features may be employed separately or in combination with any other features and the disclosure extends to and includes all combinations and sub-combinations of one or more features described herein.