CA2240361A1 - Method for fabricating a split path transmission providing equal torque splitting - Google Patents

Abstract

The method includes the steps of identifying the torque distribution curves of the forward and aft split load paths of one gear train of the split path transmission system, and depicting the identified forward and aft torque distribution curves in a graphical format wherein the abscissa of the graph represents the torque input Tin coupled into the one gear train and the ordinate of the graph represents the torque transmitted through the forward and aft split load paths, respectively. The method further includes a step of selecting a predefined operating point for the split path transmission system wherein equal torque splitting between the forward and aft split load paths of the one gear train is desired. Next, the forward and aft torque distribution curves are analytically modified as required to pass through the predefined operating point. Finally, the timing relationship of the gears and pinions of the split load path represented by the analytically-modified torque distribution curve having the steeper slope is altered to provide an intentional timing difference between the gears and pinions of the forward and aft split load paths, respectively, of the one gear train such that during operation of the split path transmission system the other split load path transmits all torque input Tin to the main rotor shaft over the input torque range. The foregoing steps are repeated for each remaining gear train of the split path transmission system to provide an intentional timing difference between the gears and pinions of the forward and aft split load paths, respectively, of each remaining gear train such that equal torque splitting in the forward and aft split load paths of each remaining gear train is achieved at the predefined operating point.

Description

Description hrETH~D FOR ~ABRICA~TNG A sP~rr PATH TRANSMnSSION PROVnDING EQUAL TORQUE
SPL~NG
l.

Technical Field The present invention relates to transmission systems, and more particularly, to a method for fabricating a split path transmission system, particularly a split path transmission system for a helicopter, that provides equal 5 torque splitting in the split load paths of each engine gear train or branch at a predefined operating point and the split path transmission system fabricated thereby.

Back~round of the Invention A transmission system comprises one or more independent gear trains or 1 û branches composed of intermeshing gears and is operative to couple the power (torque) developed by a powerplant system to an output member. in those applications where the powerplant system comprises two or more engines, the transmission system includes an independent gear train or branch for coupling the torque developed by each engine to the output member, e.g., the 15 transmission system for a two-engine powerplant system would comprise two independent gear trains or branches. In such transmission systems, and in parficular, helicopter transmission systems, it may be desirable to split the power output from each engine of the powerplant system so that each associated gear train or branch includes redundant, i.e., split, load paths for coupling the 20 power from the corresponding engine to a common output member, e.g., the main rotor shaft of a helicopter. Such split path transmission system conFigurations reduce the tooth loading of the intermeshing gears, i.e., gear train assembiies, cornprising each redundant load path and result in lighter weight gear train assemblies. In addition, split path transmission systems are 25 inherently more reliable from the perspective that if one gear assembly, i.e., load W O 97/22817 PCT~B96/01390 path, becomes inoperative, the total torque from the respective engine will be transmitted through the remaining gear assembly, i.e., the redundant load path, thereby ensuring short-term emergency operation of the transmission system.
A schematic illustration of an exemplary embodiment of a split path transmission system for helicopters is iliustrated in Figure 1. Large helicopters typically have a powerplant system composed of two or three gas turbine engines, depending upon the gross weight, size, and power and redundancy requirements of the helicopter. Figure 1 illustrates the configuration of a split path transmission system 1 ûû for a powerplant system composed of two engines (conventionally identified as the left and right engines from an aft looking-forward perspective). The split path transmission system 1 ûO utilizes independent gear trains or branches to transmit the power developed by the left and right engines (not shown) to the main rotor shaft 102 of the main rotor assembly whichis operative to provide the motive power for the helicopter. The split path 1~ transmission system 100 utilizes reduction gearing assemblies to convert engine power at high RPMs (i.e., low torque) to high torque at low RPMs for rotation ofthe helicopter main and tail rotor blades.
Each gear train or branch of illustrated embodiment of the split path transmission system lûO utilizes three stages of reduction gearing assemblies toreduce the RPM of each engine output, e.g., for the illustrated embodiment an engine input to the transmission system 100 of about 23,000 RPM, to the design output of the main rotor shaft 102, e.g., about 355 RPM for the described embodiment. Each engine (not shown) provides an output to the respective branch of the split path transmission system 100 via an engine output shaft (the2~ respective engine oùtput shafts are identified by reference characters 1 04L and 104R) that is normally coupled through a spring overrunning clutch (not shown for purposes of simplification) to the associated gear train or branch 1 û6L / 1 û6R.
Each branch 1 06L / 1 06R is operative to provide torque transmission and splitting as described in the following paragraphs. A central bull gear 108 recombines the split power coupled through each branch 106L/ 106R to effect rotation of the main rotor shaft 102 (the main rotor shaft 102 is mechanically integrated in W O 97/2~8~7 PCT~B96~01390 combination with the central bull gear lû8 so that the main rotor shaft 102 rotates at the same speed as the central bull gear 108).
The first reduction stage of each branch 106L / 106R illustrated in Figure 1 is a bevel gear set that comprises a bevel pinion I lOL/ 1 lOR and a bevel gear 1 1 2L / 1 1 2R ( note that bevel gear 1 1 2R is obscured in Figure 1 ) in t intermeshing combination. Each bevel pinion 1 1 OL / 1 i OR, bevel gear 112L/ 112R combination of the described embodiment provides a reduction ratio of about 2.03/1 (from about 23,00û RPM to about 11,317 RPMJ for the respective branch lû6L/ 106R. The shaft angles of the bevel gears 1 1 2L / 1 1 2R are positioned so that the centerlines thereof are parallel to the centerline of the main rotor shaft 102.
The second stage of each branch 106L / 106R is a simple spur gear set or a high contact rafio spur gear set that comprises a spur pinion 1 1 4L / I 1 4R (note that spur pinion 1 1 4R is obscured in Figure 1 ) and a pair of spur gears 1 1 6LFwd, 1 1 6LAft / 1 1 6RFwd, 1 1 6RA~ (note that spur gear 1 1 6RFwd is parfially obscured in Figure 1) in inten~eshing combination (the "forward" and "aft" descriptors are based upon an aft looking-forward perspective). Each spur pinion 1 14L / 1 1 4R,spur gears 1 1 6kwd, 1 1 6LAft / 1 1 6RFwd, 1 1 6RAft combination of the described embodiment provides a reduction ratio of about 2.88/1 (from about 1 1,317 RPM
2û to about 3,931 RPM) for the respective branch 1 06L / 1 06R.
The third or final reduction stage of each branch 106L / 106R is a double helical output gear set that comprises a pair of double helical bull pinions 1 1 8LFwd, 1 1 8LAft / 1 1 8RFwd, 1 1 8RAft fhat are intermeshed in combination with the central bull gear 108. Each central bull gear 108, bull pinions 1 18LFWd, 1 1 8LAft / 1 1 8RFwd, 1 1 8RAft combination of the described embodiment provides a reduction ratio of about 11.07/1 (from about 3,931 RPM to about 355 RPM) for the respective branch 106L / 106R. Also illustrated in Figure 1 is a take-off bull pinion 120 fhat concurrentiy: (i) couples torque aftwardly to effect rotation of the tail rotor system (nof shown) via a shaft 122; and tii) couples torque to an oil-cooler blower unit (not shown) for providing air to an air/oil heat exchanger via a shaft 124.

W O 97/22817 PCT~B96/Q1390 The gear trains or branches 106L/ 106R of the split path transmission system 100 described 7n the preceding paragraphs function as independent means for transmitting torque from the left and right engines, respectively, to the main rotor shaft 102. The left engine torque-transmitting means is defined by the intermeshing gears and pinions comprising the left gear train or branch 106L, i.e., the left bevel pinion 110L, the left bevel gear 112L, the left spur pinion 114L, the forward and aft left spur gears 116LFwd, 116LAft, the forward and aft left double helical bull pinions 1 18LFwd, 1 18LAft, and the central bull gear 108. Similarly, the right engine torque-transmitting means is defined by the intermeshing gears and pinions comprising the right gear train or branch 106R, i.e., the right bevel pinion 110R, the right bevel gear 112R, the right spur pinion 114R, the forward and aft right spur gears 116RFwd, 116RAft, the forward and aft right double helical bull pinions 11 8RFwd, 1 18RAf~, and the central bull gear 108. From a gross structural perspective, the left and right gear trains or branches 106L / 106R are identical, i.e., equivalent physical dimensions for the constituent gears, pinions, and shafts of each engine branch 106L/ 106R. Moreover, the geometrical relationship between the left and right engine branches 106L/ 106R with respect to a longitudinal plane through the center of rotation of the central bull gear 108 is a mirror-image relationship.
In the described embodiment of the split path transmission system 100, torque splitting is effected in the second-third stages of each branch 106L / 1 06R
such that forward and aft split ioad paths are defined in the left and right engine torque-transmitting branches, respectively. Torque from eactl engine drive gear,i.e., the spur pinion 114L/114R, is split between the pair of spur gears 1 16LFwd, 116LAft/116Rrwd, 116RAft of each gear train 106L/106R. Each spur gear 116LFwd, 116LAf~, 116RFwd, 116RAf~, in turn, drives the corresponding integral, coaxial double helical bull pinion 1 18LFwd, 1 18LAf~ 8RFwd, 1 18RAf~ (integral being used herein in the sense that the respective spur gears 116 and bull pinions 1 18 are mounted on a common compound shaft (preferably fabricated as a single piece)- see Figure 1). The central bull gear 108 recombines the torque from the double CA 02240361 1998-06-ll WO 97~2817 PCTnB96J0139~

helical bull pinions 1 1 8LFwd, 1 1 8LAft/1 1 8RFwd, 1 1 3RAft of each gear train 1 06L/1 06R
to effect rotation of the main rotor shaft 102.
For the left engine torque-transmitting branch 106L, therefore, forward and aft split load paths are defined by the spur pinion 1 1 4L, the left forward spur gear 116LFwd, the left forward double helical bull pinion 118LFwd, central bull gear 108 combination and the spur pinion 1 1 4L, the left aft spur gear 1 1 6LAft, the left aft double helical pinion 118LAft, central bull gear 108 combination, respectively. Similarly, for the right engine torque-transmitting branch 106R, forward and aft split load paths are defined by the spur pinion 114R, the right 1 û forward spur gear 1 1 6RFwd, the right forward double helical bull pinion 1 1 8RFwd, central bull gear 108 combination and the spur pinion 114R, the right aft spur gear 116RA~s, the right aft double helical pinion 118RAft, central bull gear 108combination, respectively.
Ideaily, split path transmission system configurations should be designed to ensure that torque is split in equal proportions between the forward and aft split load paths of each primary torque transmitting branch, e.g., the respective gear trains 1 06L / 1 06R described hereinabove. Figure 2 illustrates the left split load path 1û6L of the split path transmission system 1ûû described in the preceding paragraphs. Further, Figure 2 illustrates the condition that spur pinion 1 1 4L is simultaneously in contact with both the left aft spur gear 1 1 6LAft and the left forward spur gear 1 1 6LFwd and the central bull gear 108 is simultaneously in contact with both the left aft double helical bull pinion 1 18LAft and the left forward double helical bull pinion 1 18LFwd. One skilled in the art will recognize that the condition described in the preceding sentence, as illustratedin Figure 2, is a necessary and sufficient condition to ensure that torque will be distributed in some manner between the left forward and aft split load paths.
However, one skilled in the art wili also recognize that such condition does not by itself ensure that the torque will be equally distributed in the ideal rnanner between the left forward and aft split load paths. The torque split, i.e., load 3û sharing, between the left forward and aft split load paths of the respective gear trains 1 06L ~ 1 06R of a split path transmission will be a natural result of the relative W O 97122817 PCT~B96/01390 flexibilities of the forward and aft split load paths and of the arc mesh path created by the simultaneously contacting pinions and gears. Figure 2 illustratesthe arc mesh path (the heavy line identified by reference characters AMP~ for the left split load path transmission system 100 described in the preceding 5 paragraphs. The length of the arc mesh path AMP is affected by the actual geometries of the elements of the system which may vary from the blueprint geometries as a result of manufacturing errors and/or tolerances. In addition, the length of the arc mesh path AMP will be affected by deflections induced in the elements of the system as a result of gear meshing, e.g., gear tooth Hertzian 10 deflections, gear tooth bending deflections, gear rim deflections, torsion and bowing of gearshafts, bearing deflections, and by housing deflections due to loading/thermal effects. These factors, individually or in combination, can cause torque loading differentials between the forward and aft split load paths if thefactors are not accommodated for properly in the design of a split path 15 transmission system.
In an attempt to minimize torque loading differences between the split load paths of split path transmission systems, the prior art has interposed a torque adjusting device within the torque load path between the engine and the central bull gear. One prior art torque adjusting device for split path transmission 20 systems is a quill shaft as exemplarily illustrated in Figure 3 of U.S. Patent No.5,113,713. Quill shafts provide a means for minimizing the torque loading differences between the split load paths by reducing the torsional spring rates of the forward and aft split load paths, which reduces the net effects of the factors that produce torque loading differentials. While the use of quill shafts to reduce 25 torsional spring rates is a relatively effective method, the method does not completely compensate for the factors causing the torque loading differences, but instead acts to minimize the net effect of such factors. Therefore, the quill shaft method does not guarantee, and rarely achieves, the ideal condition of an equal distribution of torque between the forward and aft split load paths.
3û Furthermore, incorporating a quill shaft in each gear train assembly increases the overall complexity and weight of the split path transmission system. This, in turn, -wo s7n2sl7 PCT~96~0139a increases the costs and time required for initial assemblage and subsequent maintenance of the transmission system. In addition, incorporation of quill shafts into the transmission system reduces the reliability of the system such that inspection and maintenance is required on a more frequent basis.
A need exists to provide a split path transmission system that is operative to provide substantially equal torque distribution between the forward and aft split load paths of each gear train assembly. Such a split path transmission system should achieve equal torque distribution without incorporating additionalcomponents that would increase the overall complexity or weight of the split path transmission system.

Disclosure of the Invention One object of the present invention is to provide a method for fabricating a split path transmission system wherein equal torque splitting is provided between the forward and aft split load paths of each engine gear train of the split path transmission system at a predefined operating point.
Another object of the present invention is to provide a method for fabricating a split path transmission system having equal torque splitting between the forward and aft split load paths of each engine gear train at a predefined operating point by providing an intentional timing difference in the forward and aft split load paths of each engine gear train.
These and other objects of the present invention are achieved by a method for fabricating a split path transmission system that achieves equal torque splitting between the forward and aft split load pctths of each gear train of the split path transmission system at a predefined operating point by providing an intentional timing difference between gears and pinions of the forward and aft split load paths of each gear train of the split path transmission system. The method includes the steps of identifying the torque distribution curves of the forward and aft split load paths of one gear train of the spiit path transmission system and of depicting the identified forward and aft torque 3û distribution curves in a graphical format wherein the abscissa of the graph W O 97/22817 PC~B96/01390 represents the torque input Tin coupled into the one gear train and the ordinateof the graph represents the torque transmitted through the forward and aft splitload paths, respectively.
The method further includes a step of seiecting a predefined operating 5 point for the split path transmission system wherein equal torque splitting between the forward and aft split load paths of the one gear train is desired.
Next, the forward and aft toraue distribution curves are analytically modified as required to pass through the predefined operating point. The interception point with the abscissa of the analytically-modified forward or aft torque distrii~ution 10 curve having the steeper slope is then identified to quantify a compensating input torque range. Finally, the timing relationship of the gears and pinions ofthe split load path represented by the analytically-modified torque distributioncurve having the steeper slope is altered to provide an intentional timing difference between the gears and pinions of the forward and aft split load 1~ paths, respectively, of the one gear train such that during operation of the split path transrnission system the other split load path transmits all torque input Tin to the main rotor shaft over the compensating input torque range.
The foregoing steps are repeated for each remaining gear train of the split path transmission system to provide an intentional timing difference 20 between the gears and pinions of the forward and aft split load paths, respectively, of each remaining gear train such that equal torque splitting in the forward and aft split load paths of each remaining gear train is achieved at thepredefined operating point.

Brief Descripfion otthe D,~ s 2~ A more complete understanding of the present invention and the attendant features and advantages thereof may be had by reference to the following detailed description when considered in conjunction with the accompanying drawings wherein:

W O 97~22~t7 PCT~B96J~139D

Figure 1 is a perspective view of an exemplary embodiment of a helicopter split path transmission system configured for use in combination with a powerplant system composed of two engines.
Figure 2 is a top plan schematic view illustrating the arc mesh path of the 5 forward and aft split load paths of the left engine gear train of the helicopter split path transmission system of Figure 1.
Figure 3 is a graph illustrating the torque distribution c~lrves for the forwardand aft split load paths of the ieft engine gear train of the helicopter split path transmission system of Figure 1.
1 û Figure 4 is a graph illustrating the torque distribution curves for the forward and aft split load paths of the right engine gear train of the helicopter split path transmission system of Figure 1.
Figure 5A is a top plan view of the forward spur gear, double helical bull pinion combination of the right gear train of Figure 1.
Figure 5B is an enlarged partial plan view of Figure 5A taken along loop B
thereof depicting a timing relationship of 0~ for the forward spur gear; double helical bull pinion combination.
Figure 6 is a top plan view schematic illustrating the torque and gear load vectors acting on the gear shafts of the forward and aft split load paths of the2~ left and right gear trains of the helicopter split path tr~nsmission system of Figure 1.
Figure 7 is a graph illustrating the analytically-modified torque distribution curves for the forward and aft split load paths of the right engine gear train of the helicopter split path transmission system of Figure 1.
2~ Figure 8 is a partial top plan view of the forward spur gear, double helical bull pinion combination of Figure ~A illustrating the timing relationship as modified according to the method of the present invention wherein there is equal torque splitting in the forward and aft split load paths of the right geartrain at 1 ûO% operating power.
3û Figure 9 is a schematic representation of the steps of the method according to lhe present invention for fabricating a split path transmission WO 97/22817 PCT~B96/01390 system that provides equal torque splitting between the split load paths of eachgear train thereof at a predefined operating point.

Best Modes for Carrying Out the Invention A method for fabricating a split path transmission system that provides ea,ual torque splitting between the split load paths of each engine gear train or branch at a predefined operating point, i.e., design point, and the split path transmission system fabricated by such method is described herein in terms of the helicopter split path transmission system 100 described in the "Background of the Invention" hereinabove. Such a split path transmission system 100 is similar to the transmission system incorporated in the RAH-66 Comanche helicopter being developed by the Sikorsky Aircraft Corporation. One s~illed in the art will appreciate that the method of the present invention has utility in fabricating split path transmission systems for helicopters having other powerplant system configurations, e.g., a powerplant system composed of one engine or three engines, as well as for applications other than helicopter transmission systems.Therefore, it is to be understood that the following description of the method according to the present invention is not intended to be limiting, but merely illustrative of the teachings of the method according to the present invention.
A prototype of the split path transmission system 100 described hereinabove was fabricated for gear pattern development tests and was subjected to empirical testing. During testing, the torque was measured in each of the forward and aft split load paths of each engine gear train or branch 1 06L / 1 06R over the torque output range of the engines of a powerplantsystem having predefined parameters. For exampl~ for a dual-engine powerplant system: a contingency rated power input to the transmission system from each enaine of about 1,066 kilowatts (1,430 horsepower), a 100% rated pcwer input to the transmission system from each engine of about 820 kilowatts 11,100 horsepower), and a cruise rated power input to the transmission system from each engine of about 559 kilowatts ~750 horsepower) based upon a trcnsmission inpuf speed of about 23,000 RPM from each engine. Torque - time W Og7~228~7 PCT~B96~fl~3~

traces were recorded and measured during the empirical testing to ascertain discrete torque loads for the respective forward and aft split load paths of the1eft and right gear trains or branches 1 û6L / 1 06R. The discrete torque loads were plotted as points on a graph, and a straight line best fit curve was calculated for 5 the plotted points by linear regression and then drawn on the graph to providethe respective torque distribution curves illustrated in the graphs of Figures 3, 4.
Figure 3 is a graph illustrating the individual torque distribution curves for the forward and aft split load paths of the left engine gear train 106L. The abscissa of the graph of Figure3 represents the total torque, i.e., Tin, being 10 coupled into the left engine gear train 106L by means of the left engine output shaft lû4L and the ordinate of the graph of Figure 3 represents the torque distribution, i.e., split, between the forward and aft split load paths, i.e., the left spur pinion 114L, the left forward spur gear 116LFWd~ the left forward double helical bull pinion 118LFWd~ central bull gear 108 combination and the left spurpinion 1 1 4L, the left aft spur gear 1 1 6LAft, the left aft double helical pinion 1 1 8LAf~, central bull gear 108 combination, respectively, for a given input torqueTin coupled into the left engine gear train 1 06L.
Reference numeral 5û identifies the torque distribution curve for the forward split load path, i.e., the torque being coupled through the forward split load path of the left engine gear train 1 06L for a given input torque Tin from the left engine, and reference numeral 52 identifies the torque distribution curve for the aft split load path, i.e., the torque being coupled through the aft split load path of the left engine gear train lU6L for the given input torque Tin from the left engine. The forward torque distribution curve 5û is defined by the equation TLFWd = .4621 (Tin) + 98 ~Equation 1 ) and the aft torque distribution curve 52 is defined by the equation T~A~t = .5379 (Tin) - 98 ( Equation 2) The slopes of the forward and aft torque distribution curves 50, 52, i.e., 0.4621 and 0.5379, respectively, correspond to the relative magnitudes of the net torsional spring rates of the forward and aft split load paths, respectively, of the left engine gear train 1 06L.

WO 97/22817 PCT~B96/01390 Similarly, Figure 4 is a graph illustrating the individual torque distribution curves for the forward and aft split load paths of the right engine gear train 1 06R.
The abscissa of the graph of Figure 4 represents the total torque, i.e., Tin, being coupled into the right engine gear train 106R by means of the right engine 5 output shaft 1 04R and the ordinate of the graph of Figure 4 represents the torque distribution, i.e., split, between the forward and aft split paths, i.e., the right spur pinion 1 1 4R, the right forward spur gear 1 1 6RFwd, the right forward double helical bull pinion 118RFwd, central bull gear 108 combination and the right spur pinion 114R, the right aft spur gear 116RAft, the right aft double helical pinion 118RA~t, central bull gear 108 combination, respectively, for a given input torque Tin coupled into the right engine gear train 1 06R.
Reference numeral 60 identifies the torque distribution curve for the forward split load path, i.e., the torque being coupled through the forward split load path of the right engine gear train 1 06R for a given input torque Tin from the 15 right engine, and reference numeral 62 identifies the torque distribution curve for the aft split load path, i.e., the torque being coupled through the aft split load path of the right engine gear train 1 06R for a given input torque Tin from the right engine. The forward torque distribution curve 60 is defined by the equation TRFwd = .5902 (Tin) - 179 (Equation 3)~0 and the aft torque distribution curve 62 is defined by the equation TRA~ = .4098 (Tin) + 1 79 ( Equation 4) The siopes of the forward and aft torque distribution curv~s 60, 62, i.e., 0.5902 and 0.4u98, respectively, correspond to the relative magnitudes of the net torsional spring rates of the forward and aft split load paths, respectively, of the 25 right engine gear train 1 06R.
An examination of Figures 3, 4 shows that for any given input torque Tin, the sum of the torques being coupled through the forward and aft split load paths of either gear train 106L/ lû6R equals the given torque inputTin. For example, with reference to Figure 3, for the input torque of 8,ûûO in-lb (904 joule) 30 from the left engine, the forward torque distribution curve 50 indicates thatabout 3,80û in-lb (429 joule) of torque is being coupled through the forward split -W 09~28~7 PCTnB96~1390 load path and the aft torque distribution curve 52 indicates that about 4,2ûû in-lb 1475 joule) of tora,ue is being coupled through the aft split load path.
Figures 3, 4 further show that for any given input torque Tin (except for the input torques Tin of about 2,586 in-lb and 1,9$û in-lb for the left and right gear trains 106L / 106R, respectively) there is an unequai distribution or spiit of input ~ torque Tin between the forward and aft split load paths of the respective gear trains 106L / 106R. For example, with reference to Figure 4, for the input torque Tin of 8,0ûû in-lb (904 joule) from the right engine, the forward torque distribution curve 60 indicates that about 4,600 in-lb (520 joule) of torque is being coupledthrough the forward split load path and the aft torque distribution curve62 indicates that about 3,400 in-lb (384 joule) of torque is being coupled through the aft split load path.
Further examination of Figures 3, 4 reveals several resultant phenomenon with respect to the prototype of the split path transmission system 100 that wassubjected to empirical testing. First, there is no correspondence between the individual torques being coupled through the forward and aft split load paths, respectively, of the left and right gear trains 106L/ lû6R for any given input torque Tin. That is, the magnitudes of the torques being coupled through the forward split load paths of the left and right gear trains 106L / 106R are dissimilar ~likewise for the aft split load paths of the left and right gear trains 106L/ 106R).
For example, for the input torque Tin of 8,00û in-lb (904 joule), the forward torque distribution curve 50 indicates that about 3,800 in-lb (429 joule) is being coupled through the forward split load path of the left gear train 106L while the forward torque distribution curve 60 indicates that about 4,600 in-lb (520 joule) is being 2~ coupled through the forward split load path of the right gear train 106R.Secondly, it should also be noted that, for higher input torques Tin, i.e., Tin > 4,000 in-lb (452 joule), the relationship between the forward and aft torque distribution curves 50, 52 indicates that more torque is being coupled through the aft split load path of the left gear train 106L than through the forward split load path. Conversely, however, in the right gear train 106R, the relationship between the forward and a~t torque distribution curves 60, 62 indicates that W O 97/22817 PCT~B96/01390 more torque is being coupled through the forward split load path than the aft split load path.
Figures 3, 4 also show two further resultant phenomena with respect to the torque distribution curves 5û, 52/60, 62 for the split load paths of the left and right branches 106L/ 106R, respectively. First, that the forward and aft torque distribution curves 50, 52/6û, 62 do not intersect at the origin, i.e., the condition of no load or zero torque input. Secondly, that the forward and aft torque distribution curves 5û, 52 of Figure 3 diverge with increasing input torque Tin for Tin > 2,586 in-lb (292 joule), with the divergence increasing with increasing input torque Tin. A similar conclusion may be drawn from an examination of the forward and aft torque distribution curves 6û, 62 of Figure 4 Inote, however, that that the divergence of the torque distribution curves 6û, 62 with increasing torque is with respect to Tin > 1,980 in-lb (224 joule), and that the divergencebetween the torque distribution curves 60, 62 is more pronounced than the divergence between the torque distribution curves 50, 52).
The inventors recognized that these phenomenon were indicative of underlying causative factors inherent in the split path transmission system lûO
that prevented achievement of equal load splitting in the forward and aft split load paths of the left and right gear trains 106L/ 106R, respectively. The inventors undertook analytical and empirical analyses of these phenomenon for the purpose of identifying and understanding the underlying causative factors and interactions with the end of attaining the objective of designing a split path transmission system to achieve equal torque splitting in the forward and aft split load paths of each engine gear train a predefined operating point.
With respect to the no-load phenomenon, the forward and aft torque distribution curves 5û, 52 of the left gear train 1 û6L (or the forward and aft torque distribution curves 6û, 62 of the right gear train 106~) should intersect at the origin since the prototype transmission that was tested was designed so that every gear member should have a tooth in contact at the no-load condition. This 3û condition is referred to as a gear train timing of zero degrees. An empirical examination of the gears and pinions comprising the split load paths of the W O 97/2~817 PCTnB96/~1390 prototype of the split path transmission system 100 showed that the tolerances of such gears and pinions were well within design specification tolerances, i.e., no built-in timing error due to the manufacturing process. To facilitate a more complete understanding of the method according to the present invention, 5 Figures 5A, 5B are presented to illustrate the concept of gear tlming and tooth contact at the no-load condition.
Figure 5A illustrates the forward spur gear 1 1 6RFwd and the forward double helical bull pinion 118RFwd of the right gear train 106R which are disposed in coaxial combination on a compound shaft ~see also Figure 1). Gear timing is 1 û defined in terms of the angular relationship between the pitch point on the drive side of the index tooth of the right forward spur gear 11 6RFwd and the pitch point on the drive side of the index tooth of the right forward double helical bull pinion 118RFwd. Referring to Figure ~B, the pitch diameters of the spur gear 1 1 6RFwd and the doubie helical buil pinion 1 1 8RFwd are defined by reference 15 characters PDl16 and PDlls, respectively, and the index teeth of the spur gear 116RFwd and the double helical bull pinion 118RFwd are identified by reference characters ITI16 and ITlls, respectively.
The pitch points of the index teeth ITl16, ITIls are defined by the intersection of the respective pitch diameters PDl16, PDll~ with the loaded side of 2û the corresponding index tooth ITl16, ITlls. A first timing lineTLl is extended outwardly from the coaxial center of rotation CCR of the compound shaft to pass through the pitch point of the index tooth ITlls of the right forward double helical bull pinion 118RFwd. For the iliustrated example, the timing lineTLl also passes through the pitch point of the index tooth ITI16 of the right forward spur 2~ gear spur gear 116RFwd such that the angular relationship or gear timing between the right forward spur gear spur gear 1 1 6RFwd and the right forward double helical bull pinion 118RFwd is û~. The prototype transmission system thatwas empirically tested was designed such that, theoretically, at the no-load condition with a û~ timing relationship as illustrated, the driven side of the right 3û forward spur gear 1 1 6RFwd should be in mechanical contact with the drive side of the right spur pinion 1 1 4R and the drive side of the right forward double helical CA 0224036l l998-06-ll W O 97/22817 PCT~B96/0139Q

bull pinion 1 18RFwd should be in mechanical contact with the drive side of the central bull gear 108.
The inventors determined that several factors account for the fact that the distribution curves 5û, 52/60, 62 do not converge at the no-load condition, 5 including tolerances associated with the bearing bore locations (mounting sites for the bearings supporting the gears and pinions cornprising the left and righttorque~ r~nlission branches 106L/ 106R of the split path transmission system 100) and the topological modifications of the teeth of the respective gears and pinions comprising the left and right split load paths. Topological 10 modifications of the teeth of such gears and/or pinions are effected to ensure proper mesh at the predefined operating point, i.e., to counterbalance the effect of teeth deflections at the predefined operating point - see, e.g., the discussion in the specification of U.S. Patent No. 5,315,790 entitled "Gear Tooth Topological Modification". These topological modifications, however, cause the 15 teeth to contact at different timing positions at the no-load condition (rather than, for example, at the theoretical 0~ timing position described hereinabove) With respect to the other phenomenon described hereinabove, i.e., divergence of the forward and aft torque distribution curves 50, 52 /6û, 62 of the left and right gear trains 106L/ lO~R, respectively, the disparity in torque load 20 distributions between the forward and aft split load paths, respectively, of the left and right gear trains 106L/ 106R, and the disparate divergence rates between the left-branch torque distribution curves 50, 52 and the right-branch torque distribution curves 60, 62, the inventors determined that such phenomenon are due to the cumulative effect of the disparate deflections 25 affecting the intermeshing gears and pinions comprising the respective forward and aft split load paths of the prototype of the split path transmission system l oo described hereinabove. Such disparate deflections include gear tooth Hertzian deflections, gear tooth bending deflections, gear tooth rim effects, spur gear~double helical pinion shaft torsion and deflections, spur pinion/bevel gear30 shaft deflections, bearing deflections, and housing deflections due to loading and/or thermal effects. Such disparate deflections are the result of the mesh CA 0224036l l998-06-ll W O 97f22817 PCTnB96~13g~

forces between the interacting gears and pinions of the forward and aft split load paths of the left and right gear trains 1 06L / 1 06R.
A computerized analytical calculation of the gear mesh forces and resulting deflections of the split load paths of the split path transmission system 100 was conducted. The spiral bevel gear mesh forces were calculated by standard equations to define the tangential, separating, and thrust forces acting on the spiral bevel gear. The identified tangential and separating forceswere combined as a single vector in the local X-Y coordinate system (a local Cartesian coordinate system was defined for each split load path wherein the local Z axis was coincident with the central bull gear centerline and Z-0 was located at the imaginary apex of the double helical mesh). A transverse plane gear force analysis was conducted on the spur pinion, spur gears, and double helical bull pinion meshes using input torque, the transverse plane base radii, and the X-Y positions of the gear centers as input parameters. Output parameters from the analytical calculation included output torque, the mesh force as a vector acting on the gear base radius, and the operating pressure angle. The axial forces of the helical meshes were calculated once the transverse plane results were obtained.
The next step was a computerized calculation of the deflections produced by the calculated gear mesh forces. The mean of the time-varying gear mesh stiffness was used to calculate the gear teeth deflections. The time-varying gear mesh stiffness was determined using Cornell's method to determine a single tooth pair stiffness, and then the gear contact ratio was considered todetermine mesh stiffness. Compound shaft torsion was calculated using a spring 2~ constant. For the described embodiment, a spring constant value of ~.93X lûEû4in-lb/deg was used based upon the known material composition and geometry of the compound shaft. The bull gear support deflection was calculated using a spring constant along the centerline of the aircraft and a spring constant perpendicular to the centerline lfor the described embodiment, 3û spring constants having values of 8.96X lOE06 Ib/in and 3.46X lûE06 Ib/in,respectively) as calculated by finite element analysis. The input shaft and CA 02240361 1998-06-ll WO97/22817 PCT~B96/01390 compound shaft deflections were calculated by means of the classical Euler equation for the elastic curve of a beam. The complex shapes of the respective beams were approximated as a series of sections having constant or linearly varying moments of inertia, and the necessary mathematical boundary S conditions were imposed by assuming that the shaft supports were "semi-fixed"
(average of pinned support and fixed supports). The resultant set of differential equations for the beam approximations were integrated and then solved by matrix algebra. ~eam deflections in both the X-Z and Y-Z planes were calculated, and the solutions were added as vectors. Bearing deflections were 10 calculated using bearing reaction forces (determined based upon the calculated beam deflections) using methods know to those skilled in the art.
These calcuiations revealed that the deflections arising as a result of the mesh forces between the interacting gears and pinions of the split path transmission system 100 are dependent upon the "dynamic geometr,v" of the 1~ split path transmission system lûû, i.e., the magnitude of the individual mesh forces (deflections) and the direction of such applied loads (deflections) with respect to individual gears, pinions, and gear shafts and with respect to the gross structural configuration of the split path transmission system. For example, Figure 6 is a schematic representation of the torque and reaction vectors acting20 on the gear shafts of the forward and aft split load paths, respectively, of the left and right gear trains lû6L/ 106R, i.e., the mesh forces (deflections) identified in terms of direction as well as magnitude. The vectors are identified generally byreference characters "VRl-VR8" and "VLl-VL8", respectively. Also identified in Figure 6 by means of reference characters "R" and appropriate subscripts are 2~ the directions of rotation of the gears and pinions compr,sing the for~vard and aft split load paths of the left and right gear trains 1 06L / 1 06R, respectively.
An examination of Figure 6 shows that the net force (or deflection) acting on the upper end of the left-branch compound shaft due to the vector addition of VLl and VL2 is dissimilar to the net force ~or deflection) acting on the upper 30 end of the right-branch compound shaft due to the vector addition of VRl and VR2. With respect to the lower end of the compound shaft, the net force (or CA 0224036l l998-06-ll W O 97/2~17 PCTnB96/01390 deflection) in the left branch due to the vector addition of VL3 and VL4 is dissimilar to the net force tor deflection) in the right branch due to the vector addition of VR3 and VR4.
Similar results follow for the disparate mesh forces (or deflections) affecting the individual gears and pinions of the left and right gear trains 106L/ lû6R, respectively. For example, it was determined that the net force acting on the central bull gear 108 causes a lateral deflection of the central bull gear 108 center in the direction of arrow NFlos (identified by the dashed arrow and reference characters NFlos in Figure 6). This lateral deflection is approximately coincident with a line joining the central bull gear 108 centerand the left spur pinion 114L center. Considering such a lateral deflection, theloaded windups (as used herein, the total "windup" is the total effect of lateral movement and i-ooth deflections in the helical bull pinions 118, torsional twistinduced in the compound shaft, and lateral movement and tooth deflections iri the spur gears 1 16) in the forward and aft split load paths, respectively, of the left gear train 1 û6L are approximately equal, but the loaded windups in the forward and aft split load paths, respectively, of the right gear train 106R are dissimilar.
Or, for example, the mesh force exerted by the bevel pinion 1 1 0L of the left gear train 106L tends to deflect the left spur pinion 114L into meshing engagement with the forward and aft left spur gears 1 1 6LFwd, 1 1 6LAft. Conversely, the mesh force exerted by the bevel pinion 110R of the right gear train 106R tends to deflect the right spur pinion 1 1 4R out of meshing engagement with the forward and aft left spur gears 1 1 6RFwd, 1 1 6RAf~.
As a resuit of an analytical evaluation based upon the directional aspects 2~ of the mesh forces between the interacting gears and pinions of the split path transmission system 100, and concomitantly, the deflections produced by such mesh forces, the inventors determined that: (i) the net effect of the vector addition of such mesh forces (or deflections) on the left and right gear trains 106L/lû6R, respectively, is dissimiiar; and (ii) the net effect of the vector 3û addition of such mesh forces (or deflections) on the forward and aft split load ~ paths of the left and right gear trains 1 06L/lU6R, respectively, is dissimilar. That is, W O 97/22817 PCTnB96101390 since the magnitude of the total loaded windups will, as a rule, be different ineach of the split load paths of each gear train lû6L / 106R, the slope of the line represented by a graph of Tin versus Tindividual will be different in each split load path of each gear train 106L / 106R.
The inventors concluded that the disparity in the fotal loaded windups produced an unequal distribution of torque between the forward and aft split load paths of each of the gear trains 106L / 106R. That is, the dynamic geometryof the split path transmission system 100 results in torqve distribution curves having disparate slopes that are unique to each set of splil load paths, e.g., the forward torque distribution curve 50 (or 60) versus the aft torque distribution curve 52 (or 62), as well as between split load paths of the left and right geartrains 106L / 106R, e.g., the left forward (or aft) torque distribution curve ~0 (or 52) versus the right forward (or aft) torque distribution curve 60 (or 62). Furthermore, the inventors concluded that the unequal torque distribution between the forward and aft split load paths of the gear trains 106L / 106R could be adjusted by changing the gear train timing las illustrated in Figures 5A, 5B and previously described) appropriately such that the load would be shared equally between the forward and aft split load paths at a predefined operating point.
To accomplish the appropriate gear train timing adjustment, first, the resultant torque distribution curves 50, 52 / 60, 62 of the forward and aft split load paths of the left and right gear trains 106L / 106R must have been ascertained by empirical or analytical means. Then, the forward and aft torque distributions curves 50, 52 / 60, 62 of each gear train 106L / 106R are analytically manipulated to identify the requisite timing dimension for the forward and aft split load paths of each gear train 106L / 106R so that an equal torque distribution between the forward and aft split load paths at a predefined operating point is achieved.
The first step in such analytical manipulation is to identify the predefined operating or design point for the split path transmission system 100, i.e., the input torque Tin wherein an equal load splitting between the forward and aft split load paths of each gear train 106L / 106R is desired.

WO g7J22817 PCTnB96/013gO

The predefined operating point is quantified in terms of the torque output from the associated engine at some specified flight operating condition (see, e.g., Figure 7 wherein reference numeral 63 identifies the predefined operating ~ or design point for a specific application). For military helicopters, for example, the predefined operating point may be quantified with respect to the contingency rated power (a short duration, very high engine power rating generally intended for one-engine inoperative (OEI) or emergency operation), the intermediate rated power (a high engine power rating limited to thirty minutes duration - also identified as military power), the maximum continuous 1 û power rating (the maximum power rating at which the engine can be operatedcontinuously - also identified as "normal" or "100%" power), or the cruise powerrating (the power rating that optimizes the flight operating time or specific fuel consumption of the helicopter).
Preferably, the predefined operating point is defined in terms of the power rating that encompasses the predominant span of the flight operating profile so that lhe split path transmission system 100 is operated with equal torque splitting between the forward and aft split load paths of each gear train 1 û6L / 106R for most of the flight operating profile. This ensures that unequal gear train tooth loading, i.e., deleterious mechanical effects due to meshing interaction between gears and pinions, in the split path transmission system 100 is mlnlmlzed.
By way of example, the predefined operating or design point for the described embodiment of the split path transmission system 100 was quantified at the 100% power rating at the left and right helical bull pinions 118L, 118R, i.e., Tin = 17,636 in-lb (1,993 joule) . Once the predefined operating point is specified, the forward and aft torque distribution curves 50, 52/ 60, 62 of each gear train 106L / 106R are analytically manipulated by transposing the forward and aft torque distribution curves 50, 52 / 60, 62 of each gear train 106L/ 106R so that. each torque distribution curve 50, 52/ 60, 62 passes through the predefined operating point, i.e., Tin = 1 7,636 in-lb (1,993 joule) wherein TRFwd = TRAft = 8,8 1 8 in-lb - ~996 joule), i.e., the equal load sharing point,. The transposition of the forward W O 97/22817 PCT~B96/01390 and aft torque distribution curves 50, ~2 / 60, 62 of each gear train 106L / 106R is effected so that the slopes of the respective torque distribution curves 50, 52 / 60, 62 remains constant, i.e., the net torsional spring rates of the forward and aft split load paths of each gear train 1 06L / 106R are unchanged.
Figure 7 illustrates the analytically-manipulated forward and aft torque distribution curves for the right gear train 106R described hereinabove (the analytically-manipulated forward and aft torque distribution curves are identified by reference numerals 60' and 62', respectively, and correspond to the forward and aft torque distribution curves 60 and 62, respectively, illustrated in Figure 4). Mathematically, the analytically-manipulated forward and aft torque distribution curves 60', 62' are defined by the equations TRFwd = .5902 lTin) -1591 (Equation 5), and TRA~ = .4098 (Tin) + 1 591 I Equation 6) .
The constants of equations 5, 6 define the intersection of the respective analytically-manipulated torque distribution curve 60', 62' with the ordinate axis and are proportional to the magnitude of the change in the gear train timing dimension required in the forward or aft split load path of the right gear train 106R to achieve equal torque distribution at the predefined operating point of Tin ~ 17,636 in-lb (1,993 joule).
An examination of Figure 7 shows that the analytically-manipulated forward torque distribution curve 60', which has the steeper slope, i.e., .5902 versus .4098 for the analytically-manipulated aft torque distribution curve 62',intercepts the abscissa at about Tin = 2,696 in-lb (305 joule) (see reference character 64). The input torque defined by interception point 64 is proportionalto the degree of "relief" that must be effected in the timing of the forward split load path of the right gear train 106R. Once the relief has been effectuated, all of the torque coupled into the right gear train 1 û6R is transmitted through the aft split load path of the right gear train 106R over an input torque ran~e 65 of Tin =
û in-lb (0 joule) to Tin = 2,696 in-lb (305 joule). This phenomenon is indicated by reference character 66 in Fig,ure 7 which identifies the total-torque segment ofthe anaiytically-manipulated aft torque distribution curve 62' wherein all of the WO g7J2~817 PCTnB96~013gO

input ~orque Tin coupled into the right gear train 106R is transmitted through the aft split load path, i.e., the slope of the segment 66 of the aft torque distribution curve 62 is 1.0 (Tin = TRAfl).
Once the input torque Tin exceeds the value of Tin = 2,696 in-lb ~305 joule~, 5 torque is transmitted through the forward split load path at a progressively increasing rate lsee analytically-modified forward torque distribution curve 60') such that equal torque splitting between the forward and aft split load paths isachieved at the predefined operating point, i.e., Tin = 17,636 in-lb (1,993 joule) wherein TRFwd = TRAfl = 8,818 in-lb (996 joule). Equal torque splitting occurs since t0 the analytically-modified forward and aft torque distribution curves 6û', 62' of the right gear train 1û6R are converging instead of diverging (contrast with theforward and aft torque distribution curves 6û, 62 of Figure 4 which are diverging at the 1ûû% power rating). While the right gear train 1û6R will be subjected to deleterious tooth loading as the input torque Tin is increased beyond that of the predefined operating point, i.e., Tin = 17,636 in-lb ( 1,993 joule), due to asymmetrical loading in the forward and aft split load paths, it will be appreciated that proper selection of the predefined operating point ensures that such time of operation, and hence, the effects of deleterious tooth loading, is minimized, i.e., the predefined operation or design point by definition 20 encompasses the predominant span of the flight operating profile.
The foregoing procedure is repeated for the forward and aft torque distribution curves 50, 52 of the left gear train 106L to define the analytically-modified forward and aft torque distribution curves therefor so that the timing change required in the forward or aft split load paths of the left gear train 106L
25 can be quantified. It will be appreciated that, due to the slopes of the forward and alFt torque distribution curves50, 52, the timing change in the left gear train 106L will be dissimilar from the timing change required in the right gear train 106R. In point of fact, since the aft torque distribution curve 52 has thesteeper slope (0.5379 versus 0.4621 for the forward torque distribution curve 5û~, 30 the aft split load path of the left gear train 106L must be "relieved" such that SUBSTITUTE SHEET (RULE ~6~

CA 02240361 1998-06-ll W O 97/22817 PCT~B96/01390 initially all input torque Tin will be transmitted through the forward split load path of the left gear train 1 06L.
To mechanically effectuate the timing changes identified by the foregoing procedure, the timing relationship of the index teeth of one split load path of each gear train 106L / 106R is modified during fabrication of the gears and pinions of the respective gear train 106L/ 106R to achieve equal torque splitting at the predefined operating point. With respect to the example regarding the forward and aft split load paths of the right gear train 106R
described in the preceding paragraphs, the timing relationship of the index teeth in the forward split load path, i.e., the index teeth ITIl6, ITl18 of the right forward spur gear 116RFwd, right forward double helical bull pinion 118RFwd combination, is modified (relieved in the described example) so that torque is not transmitted through the forward split load path until input torque Tin exceeds 2,696 in-lb (3û5 joule).
This is exemplarily illustrated in Figure 8 which depicts the right forward spur gear 11 6RFwd and the right forward helical bull pinion 1 18RFwd as fabricated with the modified timing relationship. The first timing line TLl, which passes through the pitch point of the index tooth ITlls of the right forward helical bull pinion 118RFwd, does not pass through the pitch point of the index tooth ITIl6 of the right forward spur gear 116RFwd. A second timing lineTL2 is extended outwardly from the coaxial center of rotation CCR to pass through the pitch point of the index tooth ITll6 of the right forward spur gear 116RFwd. The perpendicuiar distance (at the pitch point of the index tooth ITll6) between thefirst and second timing lines TLl, TL2 defines a gap 70 that quantifies the timing relationship between the index tooth ITIl6 of the right forward spur gear 11 6RFwd and the index tooth ITIls of the right forward helical bull pinion 113RFwd that provides equal torque splitting between the forward and aft split load paths of the right gear train 106R at the predefined operating point. The magnitude of the gap 70 is the ratio of the interception point between ~he abscissa and the analytically-modified torque distribution curve having the steeper slope divided W O 97~228t7 PCTnB96~139~

by the net torsional spring rate of the split load path having a torque distribution curve with the lesser slope.
Once the timing adjustment has been introduced, then at the no-load condilion the gap 70 results in a spatial separation, i.e., no mechanical contact, between the gear teeth of the right forward helical bull pinion 113RFwd and the r gear teeth of the central bull gear 1 û8. Concomitantly, however, the gear teeth of the right aft helical bull pinion 118RAft and the teeth of the central bull gear 1 U8 are in mechanical contact at the no-load condihcn.
Input torque Tin initially coupled into the right gear train 106R causes 1û simultaneous rotation of the forward and aft spur gears 116RFwd, 116RAft (via meshing interactions between the right bevel pinion 1 1 OR, right bevel gear 1 1 2R, and the right spur pinion 1 1 4R). The rotation of the aft spur gear 1 1 6RAft results in all of the inpuf torque Tin being transmitted to the main rotor shaft 102 through the afi split load path as a result of the meshing interaction between the gear teeth of the right aft helical bull pinion 1 18RAft and the gear teeth of the central bull gear 108.
The rotation of the forward spur gear 1 1 6RFwd, in contrast, does not cause meshing interaction between the forward helical bull pinion 118RFwd and the central bull gear 108, but rather results in freewheeling therebetween. However,as the input torque is increased there is a continuing reduction in the spatial separation between the gear teeth of the forward helical bull pinion 1 1 8RF~d and the gear teeth of the central bull gear 108 due elastic deformations of the loaded components of the aft split load path. As the input torque Tin exceeds 2,696 in-ib (305 joule), meshing interaction between the aear teeth of the forward helical bull pinion 113RFwd and the gear teeth of the central bull gear 1U8 occurs such that torque is coupled through both the forward and aft split load paths in the manner described hereinabove.
A method 1û according to the present invention for fabricating a split path transmission system that provides equal torque splitting between the forward and aft split load paths of each gear train thereof at a predefined operating point by providing an intentionai timing difference between the - 2~ -W O 97/22817 PCTnB96/~1390 forward and aft split load pafhs of each gear train has been described in detailin the preceding paragraphs with respect to a particular embodiment of the split path transrnission system 100. To summarize, with reference to Figure 9, the method lOincludes:
a step 12 of identifying the torque distribution curves for the forward and aft split load paths of one gear train of the split path transmission system by either empirical or analytical means;
a step 14 of depicting the forward and aft torque distribution curves identified in step 12 in a graphical format wherein the abscissa of the graph represents input torque Tin coupled into the one gear train from the associated engine and the ordinate of the graph represents the torque being transmitted through the forward and aft split load paths, respectively;
a step 16 of selecting a predefined operating poir,t for the split path transmission system wherein equal torque splitting between the forward and aft split load paths of the one gear train is desired;
A step 18 of analytically modifying the forward and aft torque distribution curves as required, including the substeps of ~ a substep 18a of analytically modifying the forward torque distribution curve such that the forward torque distribution curves passes 2û through the predefined operating point; and/or a sub step 1 8b of analytically modifying the aft torque distribution curve such that the aft torque distribution curves passes through the predefined operating point;
a step 20 of identifying an interception point with the abscissa of the analytically-modified forward or aft torque distribution curve having the steeper slope to quantify an input torque range; and a step 22 of altering the timing relationship of the gears and pinions of the split load path represented by the analytically-modified torque distribution curve having the steeper slope to provide an intentional timing difference between the gears and pinions of the forward and aft split load paths of the one gear train so that during operation of the split path transmission system the other split wo s7n2sl7 PCTnB96/01:~90 load path transrnits all torque inputTin to the main rotor shaft over the input torque range.
Step 18 may be broken down into two substeps, 18a, 18b, as described hereinabove. It will be appreciated that once the forward and aft split load paths have been identified and depicted in graphical format in steps 12, 14 described hereinabove, the split load path having the steeper slope is readily identifiable. At a minimum, only the torque distribution curve having the steeper slope needs to be analytic~lly modified to provide identification and quantification of the input torque range Isee step 20). For completeness, both lû the forward and aft torque distribution curves may be analytically modified by means of steps 1 ~a and 18b.
Step 22 involves the alteration, either during initial fabrication or by subsequent modification, e.g., grinding, of prefabricated gears and pinions, of the intermeshing gears and pinions of the forward or aft split load path 15 represented by the analytically-modified torque distribution curve having thesteeper slope so that the timing relationship of such intermeshing gears and pinions is dissimilar to the timing relationship of the corresponding gears and pinions of the other split load path, i.e., an intentional timing difference is effected between the forward and aft split load paths of each gear train of the 2û split path transmission system 100. Once the alteration step has been accomplished, the split path transmission system should be subjected to empirical testing to verify that an equal torque splitting between the forward and aft split load paths of the one gear train has been achieved at the predefined operating point. If required, steps 12-22 may be repeated as 2~ necessary for the one gear train until equal torque splitting between the forward and afl split load paths of the one gear train is achieved. Steps 12-22 are further implemented for each remaining gear train of the split path transmission system le.g., for the embodiment of the split path transmission system 1ûû described hereinabove, the left gear train 1 û6L) to achieve an intentional timing difference 3û between the forward and aft split load paths of such gear train(s).
.

CA 02240361 1998-06-ll W O 97/22817 PCT~B96/01390 While the foregoing disclosure of the method according to the present invention has been presented in terms of a split path transmission system havingtwo independent gear trains, it will be appreciated that the method of the present invention is applicable to split path transmission systems composed of asingle gear train or more than two independent gear trains, e.g., three independent gear trains.
Therefore, although the method according to the present invention has been shown and described herein with respect to a certain detailed embodiment of a split path transmission system, it will be understood by those 1 û skilled in the art that a variety of modifications and variations of the method are possibie in light of the above teachings. It is therefore to be understood that,within the scope of the appended claims, the present invention may be practiced otherwise than as specifically described hereinabove.

What is cluil-.Ed is:

- 2$ -

Claims (5)

Claims

1. A method of fabricating a split path transmission system including at least one gear train having forward and aft split load paths, comprising the steps of:(a) identifying the torque distribution curves of the forward and aft split load paths of the one gear train of the split path transmission system by eitherempirical or analytical means;
(b) depicting the forward and aft torque distribution curves of step (a) in a graphical format wherein the abscissa of the graph represents torque input Tin coupled into the one gear train and the ordinate of the graph represent the torque transmitted through the forward and aft split load paths, respectively;
(c) selecting a predefined operating point for the split path transmission system wherein equal torque splitting between the forward and aft split load paths of the one gear train is desired;
(d) analytically modifying the forward and aft torque distribution curves as required to pass through the predefined operating point;
(e) identifying an interception point with the abscissa of the analytically-modified forward or aft torque distribution curve having the steeper slope to quantify an input torque range; and (f) altering the timing relationship of the gears and pinions of the split load path represented by the analytically-modified torque distribution curve having the steeper slope to provide an intentional timing difference between the gears and pinions of the forward and aft split load paths, respectively, wherein during operation of the split path transmission system the other split load path transmits all torque input Tin to the main rotor shaft over the input torque range.

2. The method of claim 1 further comprising a step of:
(g) empirically testing the split path transmission system to verify that equal load splitting between the forward and aft split load paths is occurring at the predefined operating point.

3. The method of claim 2 further comprising the step of:
(h) repeating steps (a) - (g) as required until equal torque splitting between the forward and aft split load paths of the split path transmission system is achieved.

4. The method of claim 1 further comprising the step of:
(i) implementing steps (a) - (f), as required, for each remaining gear train of the split path transmission system to provide an intentional timing difference between the gears and pinions of the forward and aft split load paths, respectively, of each remaining gear train such that equal torque splitting in the forward and aft split load paths of each remaining gear train is achievedat the predefined operating point.

5. A split path transmission system as fabricated by the method of claim 1.