USAAVLABS TECHNICAL REPORT 65-5604

TRANSMISSION STUDY FOR TANDEM-ROTORSHAFT-DRIVEN HEAVY-LIFT HELICOPTERS

R-37913 9- - :.j 1

By

Mack oL1 o c" •Y

September 1965

i. S. ARMY AVIATION MATERIEL LABORATORIES

FORT EUSTIS, VIRGINIA

CONTRACT DA 44-177-AMC-241(T)

VERTOL DIVISIONTHE BOEING COMPANY

DEPARTMENT OF THE ARMYU S ARMY AVIATION MATERIEL LABORATORIES

FORT EUSTIS VIRGINIA 23604

This report represents a part of the USAAVLABS program tvinvestigate mechanical transmission system concepts for a

shaft-driven heavy-lift helicopter of the 75,000- to 95.0o0-

pound gross weight class. The purpose of this invest! .%.)onwas to determine the high risk or problem areas that couid

be expected in the development of a drive train for a mechan-

ically driven heavy-lift helicopter.

This report presents a comparative analysis of various powertrain configurations for use in a Tandem-Rotor, Shaft-Driven,Heavy-Lift Helicopter. The report also presents discussionsof development areas wherein successful accomplishments wouldrender an improved mechanical drive system.

This command concurs with the contractor's conclusions reportedherein.

Task IM121401D14414Contract DA44-177-AMC-241 (T)

USAAVLABS Technical Report 65-56

September 1965

TRANSMISSION STUDY FOR TANDEM-ROTOR

SHAFT-DRIVEN HEAVY-LIFT HELICOPTERS

R-379

byJ. Mack

Prepared by

VERTOL DIVISIONTHE BCING COMPANY

Morton, Pennsylvania

forU. S. ARMY AVIATION MATERIEL LABORATORIES

FORT EUSTIS, VIRGINIA

SUMMARY

This report presents the results of a study of mechanicaldrive systems for a tandem-rotor, heavy-lift helicopter (HLH).This work was performed for the U. S. Army Aviation MaterielLaboratories (USAAVLABS) by the Vertol Division of Boeing,under Contract DA 44-177-AMC-241(T), between 27 June 1964 and27 December 1964.

Design characteristics for the helicopter and description ofthe mission were given by USAAVLABS. Drive system configura-tions were selected by the contractor.

The objectives of the study were as follows:

1. To define problem areas peculiar to a helicoptercapable of fulfilling the mission requirements

2. To indicate development programs necessary to elimi-nate high-risk problem areas

3. To provide sufficient weights information to permitcomparison of mechanical drive systems with otherdrive systems

The primary conclusions drawn from this investigation are asfollows:

1. The confidence level is high for the satisfactorysolution of the tandem heavy-lift-helicopter mechani-cal drive system. A comparison has been made of theHLH and current technology, using certain factorswhich generally indicate state of the art. The HLEdoes not exceed present levels in the more signifi-cant factors.

2. Programs of drive system component development arerequired in three areas to ensure high confidence.These are: gear surface durability, engine control,and overrunning clutches. Other investigative areasare defined in this study as highly recommended, butnot essential to the HLH.

3. Confidence in the tandem-rotor drive system is in-creased because total power is shared between two

iii

rotor transmissions. This minimizes the requiredextension to present day drive system technology.

4. Specific weight of the study drive system approxi-mates 0.58 pound per horsepower.

5. Advanced gear, bearing, and shafting technologiescontribute to a 10-percent reduction in drive systemweight.

6. Supercritical speed shafting provides a 160-poundweight decrease compared to subcritical designs. Aconcurrent improvement in system reliability is ex-pected by reduction in dynamic components.

7. Final reduction ratios required can be accommodatedby conventional gearing arrangements. However, high-ratio devices competitive in efficiency to presentstandards are particularly applicable to the HLH.

8. Bevel gearing capable of transmitting HLH power doesnot exceee current experience levels, except thatsurface durability (scoring) requires investigationand possible corrective active.

Investigations cf (1) gear tooth surface capacity, (2) multi-engine control, and (3) overrunning clutches are recommendedto achieve confidence in the reliability of the HLH drivesystem:

1. Gear tocth scoring is an important surface failurephenomenon. The requirements of the HLH indicatemain-power gearing which is within the presentlyaccepted scoring region. To avoid this problem, andmaintain minimum-weight transmissions, it is necessaryto improve scoring resistance by analysis and test ofpromising solutions.

2. A requirement for more than two engines will increasepilot responsibility for power management and match-ing. To determine the magnitude of the problem, it isrecommended that a multiengine simulator be used.Further, if this problem proves serious, automaticcontrol methods must be investigated for the enginesapplicable to the HLH requirement.

iv

3. Increased torque requirements and placement of theclutch at the transmission input combine to raiserubbing velocity over the level of present experi-ence. Clutch life and reliability must be evaluatedat contemplated overrun velocities. Clutch position-ing at the transmission-engine interface presentsdynamic unknowns which also require study,

In addition, the study indicates that investigations of (1)gear tooth strength, (2) supercritical-speed sbafting, (3)bearing analysis, and (4) materials improvement are particu-larly worthy of effort, and that they will provide maximumreturn to HLH and other drive systems:

1. To assist in obtaining the required improvement insurface durability and to obtain higher load-carryingcapacity for the same gearing weight, a program toimprove gear tooth bending fatigue strength isrecommended.

2. The po,,!er requirements and length of the HLH combineto increase the significance of interconnect shafting.Supercritical speed operation, by eliminating shaftsupports, reduces weight. By reducing potential fail-ure points, reliability is increased. Continuation ofpresent efforts through a flight demonstration isrecommended.

3. Larger bearings operating, in some cases, at higherspeeds require rigorous analytical approaches toensure reliability. Confident use of available in-creases in capacity requires more complete knowledgeof loads and velocities within the bearing.

4. The use of titanium is being widely explored fordrive system application. Successful application inmany areas which can take advantage of its uniquecharacteristics will provide substantial weight re-duction to the drive system. Materials and processesin the area of ferrous metals promise greaterstrength-weight ratios, increased predictability, anddimensional stability in large sizes. Continuingprograms to evaluate promising approaches should beencouraged. The results of these programs can be ex-pected to bene3fit current helicopter drive systems,

v

as well as the heavy-lift helicopter of the future.

This etudy has investigated several tandem-rotor HLH concepts.A high level of confidence has been demonstrated for thesuccessful application of the mechanical drive systems in-vestigated and in all areas the design solutions are withincurrent technology. However, the drive systems are alsosubject to improvement. Programs to advance technology aresuggested.

vi

CONTENTS

Page

SUMMARY . . . . . . . . . . . . . . . . . . . . . iii

LIST OF ILLUSTRATIONS ....... .............. ix

LIST OF TABLES ............ ................. xiii

INTRODUCTION ................ .................. 1

DATA SUPPLIED ................. .................. 8

Design Characteristics .......... .......... 8Missions ................ ................... 8

ANALYSIS OF PROBLEM ............. ............... 10

Requirements of HLH ........... ............. 10Objectives ................ .................. 10Mechanical Design Criteria .... .......... .. 13Design Stress Levels ...... ............. .. 16Powerplant Selection ...... ............. .. 29Transmission System Configuration Analysis . . 35Final Reduction Subsystem Analysis ........ .. 40Bevel Gearing Considerations ... ......... .. 51

RESULTS ............... ..................... 55

Configuration I ......... ............... 55Configuration II ........ ............... .. 67Configuration III ......... .............. 69

EVALUATION OF THE STUDY ....... ............. 75

Commitments .......... ................. 75Recommended Problem-Area Programs ........ .. 76Recommended Development-Area Programs . 78Weight Evaluation ........... ............... 82

vii

Page

BIBLIOGRAPHY ............ ................... .. 159

DISTRIBUTION ............ ................... .... 160

APPENDIXES .............. ................... .. 161

I. Equations for Equivalent Number of Meshes(Me) for Power Loss Comparisons.. . . . 161

1I. Summary of Bearing Loads and Gear StressLevels for Configuration I ... ........ .. 163

III. Table XI, Summary of Lubrication SystemAnalysis for Configuration IA .. ...... 169

Viii.

ILLUSTRATIONS

Figure Page

1. Comparison of HIM Drive System withCurrent System .......... ............ 3

2. Study Configurations I and IA . . .. 4

3. Study Configuration II ...... ........ 5

4. Study Configuration III ..... ....... 6

5. Effect of Torque on Gear Diameter(For a Constant Face/Diameter Ratio) . 7

6. Specification Missions ...... ........ 9

7. Advanced Stress Levels - Effect on

Drive System Characteristics ...... 18

8. Weight Effect of Stress Levels -?orward Transmission ... ......... .. 19

9. Comparison of Bevel Gear Sizes . . . . 20

10. Bearing Life Improvement byImproved Methods .... ........... .. 24

11. Bending Strength Test Results . . . . 25

12. Specimen - Gear Bending Test ..... 27

13. Spiral Bevel Gear Bending StressVersus Probability of Survival . . .. 28

14. Power Requirements . . .......... 31

15. Engine Availability and GrowthPlanning (Based on Anticipated Funding) 34

16. Comparison of Drive System Config-urations ........ ............... .. 39

ix

Figure P-age

17. Effect of Parallel Meshes on Weight . . 43

18. Final Reduction Methods - SchematicComparison ......... .............. 44

19. Weight Comparison of ReductionMethods .......... ............... 45

20. Equivalent Meshes of Final ReductionMethods .......... ................ 49

21. Engine Housing Planform. ........... .. 57

22. Exhaust Studies ...... ............ .. 59

23. Maintainability - Plan View ........ .. 60

24. Maintainability - Side View ........ .. 61

25. Maintainability - View Looking Forward 62

26. Comparison of Supercritical and Sub-critical Shafting .... ........... .. 64

27. Configuration I - Weight Effect ofAlternate Approaches ... ......... .. 65

28. Reliability Through Redundancy . . .. 71

29. Dual Load-Path Gear .... .......... .. 73

30. Weight Comparison of Study DriveSystems .......... ................ 84

31. Weight Trends Comparison .. ....... .. 88

32. Effect of Design Horsepower onWeight ........... ................ 90

33. Weight Effect of Stress Level andMaterial ........... ............... 93

x

Figure Page

34. Canc idate Configurations .. ....... .. 95

35. Forward Rotor Transmission - Conven-tional Stress Level Ge° s and Bearings 99

36. Forward Rotor Transmission - AdvancedStress Level Gears and Bearings . . .. 101

37. Drive System - Configuration I -SK 14216 ........ ............... 103

38. Airframe Arrangement - ConfigurationI - SK 14217 ......... ............. 107

39. Forward Rotor Transmission - Config-uration I - SK 14218 ..... ......... .. 111

40. Aft Rotor Transmission - ConfigurationI - SK 14219 ....... .... . 113

41. Aft Rotor Shaft - Configuration I -SK 14221 ............. 119

42. High-Mounted Reduction - AlternateConfiguration IA - SK 14222 ......... .. 121

43. High-Mounted Reduction AirframeArrangement - Alternate ConfigurationIA - SK 14225 ..................... 123

44. Drive System - Configuration II -

SK 14226 ......... ............... .. 125

45. Airframe Arrangement - ConfigurationII - SK 14227 ...... ............. .. 129

46. Forward Rotor Transmission - Config-uration II - SK 14228 .... ......... .. 131

47. Aft Rotor Transmission - ConfigurationII - SK 14229 ........ ............. .. 135

xA

Figure Page

48. Airframe Arrangement - ConfigurationIII- SK 14242 ..... ........... ... 137

49. Aft Rotor Transmission - Configura-tion III - SK 14243 ........ ........ .. 139

50. Study Configuration IV - SK 14245 . . 141

51. Drive System - Configuration V -SK 14236 ......... ............... .. 143

52. Combining Transmission - Configura-tion V - SK 14237 ...... ........... .. 145

53. Study Configuration VI - SK 14244 . . . 149

54. Compound Reducer Study - SK 14232 . . . 151

55. Three-Stage Reducer Study - SK 14231 . 153

56. Split-Torque Reducer Study -

SK 14230 ......... ............... .. 155

57. Roller Gear Reducer Study - SK 14233 . 157

58. Bearing Summary - HLH Configuration I . 163

xii

TABLES

Table Page

I Miss~on Design Criteria ... ......... .. 15

LI Bearing Design Criteria ... ......... .. 15

III Gear Stress Level Tabulation ...... 28

IV Powerplant Tabulation .... ......... .. 30

V Configuration IA Helicopter WeightSummary (Transport Fuselage) ....... ... 85

VI Configuration IA Propulsion GroupWeight Summary ....... ........... 86

VII Configuration IA Lubrication SystemWeight Summary ..... ............. 86

VIII Component Weight Summary .. ........ 87

IX Spiral Bevel Gear Summary ... ........ .. 165

X Spur Gear Summary .... ............ . .67

XI Summary of Configuration IA LubricationSystem Analysis ...... ............ .. 169

xiii

INTRODUCTION

The helicopter and rotor system on which this transmissionstudy was based is, in part, the result of the mission re-quirements and general characteristics supplied by USAAVLABSand, in part, the result of past Vertol Division study. Acontinuing program of study may modify certain characteris-tics of the heavy-lift helicopter (HLH), as study conclusionsand recommendations are assimilated. It is therefore desir-able to indicate the sensitivity of the drive system studyconclusions to such design characteristics. Rotor disc load-ing and rotor thrust are selected for certain conditions. Ifthe conditions change, what will be the effect on the drivesystem?

Figure 1 illustrates a comparison of the HLH and currenttechnology. Major study configurations I, IA, II, and IIIare shown in Figures 2, 3, and 4.

Drive system torque is effectively independent of rotor discloading for a constant gross weight and rotor tip speed.While the power required increases with smaller diameters(higher disc load), the rotor rpm increases to maintain aconstant tip speed. The relationship between induced rotortorque, thrust (gross weight), rotor radius, and rotor speedis shown in equation (1). Total torque increases slightlywith disc load, because total power includes profile power.

3

Torque Thrust 2

S(2 P 7)2( OR)

The torque-power relationship is:

Torque = Pia (2)

Rotor power requirement is based upon:

P. = (Thrust)Vi (3)

/ ThrustVi 2 2P r a2 (4)

•-• . - • .1

wherePi = induced power P = density of air

Vi = induced velocity R = rotor radius

2 = radians/second

Study areas (such as gear sizes) which depend upon torque are

not greatly influenced by variations of possibly ± 20 percent

in rotor disc loading, at constant thrust. This indicates

that rotor diameter alone is not a powerful factor in the

study conclusions.

A change in thrust (or gross weight) affects the torque re-

quirement, assuming constant rotor disc loading, constant

tip speed, and aerodynamically equivalent rotors. A 20-per-cent increase in mission weight would, by referenze to the

foregoing equation, indicate that torque is increased by 32percent. The effect of aircraft gross weight change on drive

system conclusions is therefore more significant than is theeffect of rotor disc load. While considering this, it is

well to remember that gear diameter is responsive to the cuberoot of torque (Figure 5). Therefore, a one-third increase

in torque will increase gear diameter and velocity by 10 per-

cent. Since this is exactly the amount rotor rpm decreased

in the example, conclusions relating to gear velocity are

still relevant.

The effect of rotor diameter influences the transmissionsystem arrangement in a less obvious manner. In the early

study configurations it became apparent that system simplifi-

cations could be made by directing the engine inputs to the

rotor gearbox, thereby eliminating engine nose boxes and bevel

gearing. To do this it was necessary to house the engines in

an airfoil section protruding from the fuselage. The effect

of rotor downwash on this section is dependent upon rotordiameter. The rotor diameter used in the study vehicles is

large enough to ensure that downwash is not significant.

With a lesser diameter, it would become increasingly signifi-

cant. Therefore, the final transmission arrangement would be

the result of trade-off studies between rotor diameter, drag

area, fuselage length, and transmission complexity.

2

MAXIMUM VELOCITIES

40 - ...+ 20 <- CURRENT

% <D VERTOL

C:- - 120- i [1- EXPERIENCE

BEVEL BEARINGS CLUTCH PLANETSGEAR DN FPM FPMFPM

MAXIMUM FORCES401

20--CURRENT'

- - - -V R O

20 EXPERIENCE

40

MAXIMUM DIMENSIONS CURRENT HELICOPTER

INDUSTRY EXPERIENCE

40-

+ 20- CURRENT

% VERTOL

20- EXPERI ENCE

40

BEVEL PLNT LIGHT LIGHTGEAR RING ALLOY ALLOY

,DIAMETER D±AMETER 1FORGING( )ASTING(2)

(1)MAX. (2)MAX.DIMENS ION DIAMETER

FIGURE 1. COMPARISON OF HLH DRIVE SYSTEM WITH CURRENTSYSTEM.

C

En

0

a ro0

00

0

z

H

I-4

00

00P 4 C

10

00H

>~. 0

-. -00

oc

El)C:

r)0z

0z

200

100

100 200 300 400 500

TORQUE %

FIGURE 5. EFFECT OF TORQUE ON GEAR DIAMETER (FOR ACONSTANT FACE/DIAMETER RATIO).

7

DATA SUPPLIED

The following helicopter design characteristics and missiondescription were furnished by the study contract document:

DESIGN CHARACTERISTICS

1. Gross weight: 75,000 to 85,000 pounds2. Turbine-powered3. Safe autorotation at design gross weight4. Design load factor: 2.5 at design gross weight5. Crew minimum of 1 pilot, 1 copilot, and I crew chief6. All components to be designed for 1200 hours between

major overhaul, and 3600 hours' service life

MISSIONS

Transport Mission (Fiwrej 6)

1. Payload: 12 tons (outbound only)2. Radius: 100 nautical miles3. Vcruise, 12-ton payload: 110 knots4. Vcruise, no payload: 130 knots5. Hovering time: 3 minutes at takeoff, 2 minutes

at midpoint (with payload)6. Reserve fuel: 10 percent of initial fuel7. Hover capability: 6000 feet, 950 F. (OGE)8. Mission altitude: sea-level standard atmosphere9. Fuel allowance for start, warm-up, and

takeoff: MIL-C-5011A

Heavy-Lift Mission (Figure 6)

1. Payload: 20 tons (outbound only)2. Radius: 20 nautical miles3. Vcruise, 20-ton payload: 95 knots4. Vcruise, no payload: 95 kniots5. Hovering time: 5 minutes at takeoff, 10 minutes

at destination (with paylaad)6. Reserve fuel: 10 percent: of initial fuel7. Hover capability: sea Level, 59 0 F. (OGE)8. Missiop altit-ade sea-level standard atmosphxre9. Fuel allowance for start, warm-up, and

takeoff: MIL-C-5011A

a

Ferry Mission (Figure 6)

1. Ferry range (no payload, STOL takeoff):1500 nautical miles

2. Reserve fuel: 10 percent of initial fuel

3. Fuel allowance for start, warm-up, and takeoff:

MIL-C-5011A

4. Minimum design load tactor of 2.05. Mission altitude: sea-level standard atmosphere

6. Best speed for range

TRANSPORT MISSION

6000 3-MIN HOVER950 %ZP OUTBOUND

12 TONS AT 110 KN SL STD

2-MIN HOVERWITH LOAD

-- INBOUND

0 TONS AT 130 KN

HEAVY-LIFT MISSION

"20 NMOUTBOUND ---

5 MIN 20 TONS AT 95 K1

• SL STD10 MIN WITH LOAD

- INBOUND0 TONS AT 130 KN

FERRY MISSION

1500 NM

SBEST SPEED FOR RANGE SL STD

FIGURE 6. SPECIFICATION MISSIONS.

9

ANALYSIS OF PROBLEM

REQUIREMENTS OF HLH

The transmission study configurations are based upon thefollowing requirements:

1. Maximum design horsepower: 15,2002. Rotor configuration: Tandem3. Rotor diameter: 96 feet4. Rotor rpm: 134

The maximum design horsepower (15,200) is based upon the poweravailable at standard conditions trom advanced versions ofcurrent engines. The installed horsepower requirement isderived from the specified 6000-foot, 95-degree, transporthover. The maximum sea-level horsepower required by themission characteristics is 11,500. The rotor system will bedesigned to have an eventual capability of absorbing theavailable 15,200 shaft horsepower (slip). This additionalpower can be utilized to increase aircraft capability by pro-viding higher speed or greater payload-carrying capacity thanspecified in the mission definition. Designing the drivesystem to 15,200 shp provides assurance that the study con-clusions reflect the most arduous condition expected. Systemweights must be factored for comparison with systems desig-tedto lesser power requirements.

Rotor diameter was based upon a conservative disc loading forthe gross weight supplied. This loading, 5.55 pounds persquare foe: at 80,000 pounds3 gross weight, is directly com-parable to Vertol Division experience for the CH-47A Chinook.As in the Chinook, it allows margin for efficient growthversions at higher qros; weights. Rotor rpm is also deter-mined from Vertol Division experience, and results in a tipspeed of 700 feet per second.

OBJECTIVES

A necessery preliminary to the design of the system is theestablishment of the design objectives. These are used asS9uide during the conceptual. study phases, and as a checklist for evaluation in the final design stage. The overallc-bjectives for the HLH drive system are that it shall perform

10

its functions with maximum reliability, minimum total effec-

tive weight, and minimum maintenance.

The functions of the drive system are to:

1. Distribute power from prime movers to rotors2. Make the necessary speed reduction3. Synchronize rto. rs4. React rotor force3 to airframe5. Slow and stop rotors6. Power aircraft accessories

System

The means by which objectives will be attained in the systemare as follow_:

1. Reduce gear mesh loss to a minimum by arrangement ofthe syster.. and choice ot intermediate ratios.

2. Reduce number of gearcase assemblies to a minimum.3. Increase reliability by approaches such as inspec-

tion systems for incipient failure, use of materialwith inherently slow crack propagation, and re-dundancy.

4. Maintain system rotational speeds at a maximum upto the final reduction, consistent with final driveratio available.

5. Eliminate power-mixing functions on interconnectshaft.

Subsystem

The means by which objectives will be attained in the sub-systems are as follows:

Transmission Assemblies

1. Use multiple load paths wherever possible.The failure of any single component shall notbe catastrophic in effect.

2. In addition to, or in lieu of, the precedingparagraph, ensure that the fatigue failurepropagation rate of the component shall beslower than the normal inspection period.

11

3. Provide failure warning of nonvisiblecomponents.

4. Increase material reliability by advancedprocessing techniques.

5. Eliminate or damp resonant frequencies withinthe operating range.

6. Consider emergency nonlubricated operation,especially for bearings.

7. Increase power/weight ratio by equalized loadingof planet gears.

8. Develop high-ratio, high-efficiency finalreduction devices.

9. Minimize weight of rotor support and reactionstructure.

10. Consider the following maintenance requirements:(1) increased TBO - 1200 hours between over-hauls, 3600 hours to retirement, and capabilityof reliable operation between 300-hour peri-odics; (2) field equipment and skill limitation;and (3) accessibility of components.

Shafting

1. Utilize supercritical speed operation for lessweight and increased reliability. However, allshaft systems shall also have provision forsubcritical operation by addition of supportingmembers.

2. Minimize number of shaft joints and flexiblecouplings, consistent with deflection andhandling requirements.

3. Consider the effect of end supports when crit-ical speed is calculated, especially for high-speed application.

4. Incorporate redundant design wherever practi-cable, especially in the attachment areas.

5. Make shaft couplings and attachments visuallyinspectable.

6. Increase misalignment capacity in shaftcouplings and spline joints.

7. Improve lubrication of splined joints wheremisalignment is to be accommodated.

8. Reduce balance requirements of shafting bysupercritical speed operation.

12

9. Reduce or eliminate lubrication maintenanceof shaft assemblies.

Lubrication System

1. Jets shall be multiple or shall be removableand inspectable from the outside of thegearcase.

2. Filters shall have indication of condition,preferably in cockpit.

3. Oil temperatures shall not exceed the presentmaximum.

4. Internal lubrication passages shall bemachined and inspectable.

5. Situation of the oil cooler and blower shallbe as close as possible to each transmission.

Rotor Brake

Rotor brake design objectives shall include thefollowing:

1. The rotor brake shall not be mounted on amain drive shaft.

2. Brake actuation and control shall besimplified and failsafe.

3. Consider shieldinq of brake disc.

MECHANICAL DESIGN CRITERIA

Design criteria for the HLH drive system was obtained from apower--required analysis for each phase of the flight regime.Gears, splines, and shafts were designed for unlimitedfatigue life, and 3600-hour wear life. Bearings were designedaccording to a cubic mean power taking into account all ex-pected powers shown in Tables I and II.

The cubic mean power is derived from Palmgrert (Reference 5)and from others as follows:

Fm = N1 + F23 N2+ (5)

N

13

where

Fm = cubic mean powerFl = power acting N1 % of timeF 2 = power acting N2 % of timeN = 100% of time

The minimum bearing design life was 1200-hour B10 (B10 is thelife that 90 percent of the bearings of an apparently identi-cal lot can be expected to meet or exceed). This design lifeis consistent with the contractual requirements and currentdesign objectives. The design criteria "Percent of Time" wasderived from analysis of the expected overall mission profileof the H4LH. This analysis was used to estimate the amount oftime which the HLH would spend upon each general task duringits total service lile. For example, percent of time per-forming the heavy-lift mission, percent performing the trans-port mission, and percent in ferry to and from the theatreof operation. It will be noted that a high percentage (31percent) of the total life was estimated for the heavy lifthover which requires 11,500 shp. This decision indicatesthe tendency to provide a conservative bearing design cri-teria.

The maximum available sea-level power of 15,200 shp is pre-dicted for only 5 percent of the time. Its effect on thecubic mean power is arithmetically negligible. Bearingsassociated with gear support functions were not, therefore,affected by the maximum power criteria. Rotor shafts androtor shaft support bearings were sized by preliminary aero-dynamic analysis of rotor loads, as well as by the maximumtorque requirement. A rotor shaft bearing load schedule,consistent with the expected flight conditions, was set upand used to obtain required bearing capacities. The powerdistribution between rotors was determined after a review offlight test data :rom current tandem rotor helicopters andafter consideration of any special requirements of the HLH.A distribution fat:tor consistent with prrx,:at experiencewas decided upon.

Gears, shafts and splines are designei for continuous opera-tion at that propo.-tion of maximum takeoff power (15,200 shp)dictated by their position in the system and by the applica-tion of the distribution factor (in the rotot transmissions).

14

TABLE I. MISSION DESIGN CRITERIA

SHP forFlight Condition Percent of Time Rotor Tip Std Day

Speed (fps) SL

Climb 5 700 11,500

Vmax 5 700 15,200

Vcriiise 40 700 7,100

HoverTransport mission 14 700 9,800HL mission 31 700 11,500

Autorotation 5 875 0

Remarks:

1. The power distribution to either rotor is to bebased upon current data.

2. Alternating torque is to be considered as 15% of meanfor fatigue life.

3. Rotor shaft included angles (measured from aircraftwaterline) to be:

forward: 100-1/2°aft: 85-1/20

4. Rotor rotation shall be determined by the particulartransmission configuration.

TABLE II. BEARING DESIGN CRITERIA

Condition Cubic Mean Power Tip Speed Life B10

Normal 10,000 709 ]1200

Engine(s) out Single-engine power 700 300

15

: :7:!m . m ! , , m13 l ll I- -m -l I -l l l l - I i l i l i i l l l l

They are not affected by cubic mean load schedules.

DESIGN STRESS LEVELS

A safe stress level is based upon experience with a particulartype of design and with a particular method of calculation.It is considered safe if final performance meets standardsof reliability and maintenance-free operation predicated assatisfactory for the product. Since considerations of re-liability and maintainability are paramount in aircraft, thestrong tendencies for the designer are to formulate conserv-ative stress levels and use conservative analysis techniques.This is especially true beczuse of the immediate pressuresupon the aircraft manufacturer to produce a safe and econom-ically feasible mechanism with a minimum of time availablefor development and uprating of component stress levels.However, in the course of programs such as those proposed inthe present HLH study, an opportunity will be presented toevaluate improvements in materials and technology realized

in the past decade and to further develop proven concepts inorder to yield lower-weight, more reliable mechanical drivesystems.

To this end the Vertol Division has used the results of itsown and others' experience to propose attainable goals instress levels for the time period of the HLH. This is under-stood to be the 1970 era, or approximately 8 to 10 years be-yond thie basic design of current medium-capacity helicopters.The goals established have been used in the design studies ofthis report, and have been compared to 1960-era levels interms of weight, and of other significant criteria (Figure 7).

These goals are predicated upon the following:

1. Developments of the past few years have indicatedthat the improvement potential exists.

2. No technical breakthroughs will be required for

their attainment; rather, an evolutlonary-type im-provement will be required to substantiate thesepresent indications.

3. An aggressive development program will be supportedto realize this evolutionary improvement within the

16

time permitted. To implement this program, testand development facilities in existence, and planned,should be utilized to provide a level of confidencewhich will permit advanced stress levels to beapplied to the HLH.

It should be emphasized that the full use of advanced stresslevels represents weight and size advantages to the HLH.These stress levels are not, in the main, prerequisites tothe successful solution of the HLH drive system; a solutionmaking use of 1960-era stress levels is practicable. Thisis illustrated by Figure 35, which delineates an HLH rotortransmission designed to current stress levels. This maybe compared to Figure 36 which shows the same transmissionusing advanced levels. Figure 8 compares the weights ofthese two transmissions. The reduction in planetary ringgear size from 38 inches to 33 inches by a combination ofimprovements in bearings and gears is especially noteworthy.The effect of advanced stress levels on the spiral bevelgears is shown in Figure 9.

Confidence in the feasibility of obtaining more from amaterial presupposes increased knowledge of the material,and of the environment in which it operates. If thisconfidence is borne out by testing, the increased capa-bility can be used to provide greater reliability at astress level comparable to that now used. Two optionscan therefore be derived from a successful developmentprogram: improved power to weight, or improved reliability.The ultimate trade-off between these two will be influencedby field experience accumulated before the final designdecisions are made, as well as by the success of the devel-opment effort in meeting and exceeding the objectives.

Areas where increased stress levels were investigated are-he following:

1. Spiral bevel gearing2. Spur and helical involute-form gearing3. Rolling-element bearings

17

~ -- 7

CHARACTERISTIC REDUCTION OF CHARACTERISTIC VALUE

-20% - 10 % 0I I

BEVEL GEARPITCH LINE VELOCITY

PLANET GEARPITCH LINE VELOCITY

BEARINGDN VALUE *

PLANET GEARCENTRIFUGAL FORCEIBEVEL GEARCENTRIFUGAL FORCE

BEVEL GEAR DIAMETER

PLANET RING DIAMETER

SD = ID of bearing in MMN = RPM

CONVENTIONAL

STRESS LEVEL

FIGURE 7. ADVANCED STRESS LEVELS - EFFECT ON DRIVESYSTEM CHARACTERISTICS.

18

4000-

3800"

3600-

3400

3200

3000

WEIGHT(POuNDs)

CONVENTIONAL ADVANCEDSTRESS LEVEL STRESS LEVE•L

(SK 14220) (SK 14223)

FIGURE 8. WEIGHT EFFECT OF STRESS LEVELS - FORWARDTRANSMI.SSION.

19

............

ADVANCED DESIGN

CON VENTIONALDESIGN

\I OUTPUT - -

INPUT

II

0 1 2 3

SCALE - INCHES

FIGURE 9. COMPARISON OF BEVEL GEAR SIZES.

20

Areas where increased stress levels were not used are:

1. Shafts under combined loading with possiblefretting at connections, and with overridingrigidity and spring rate design criteria.

2. Gear carriers and webs under combined loading.3. Light alloy forgings and castings.

It is believed that successful application of new materialcombinations and techniques to the first group couldinvluence the acceptable limits placed upon the second.However, there is generally in the second group a lessdirect and less provable (particularly by componenttesting) relationship between stress level and servicereliability. There is also a possible degradation offatigue endurance by reason of the size factor in the largeforgings and castings required for the HLH drive system.It was, therefore, assumed for this study that suchelements would be developed and extended during futuregrowth of an HLH vehicle in being.

Information on the progress being made in development ofadvanced stress levels was obtained front the followingsources: major bearing manufacturers, the Gleason GearWorks (Reference 8), Vertol Division's experience ingrowth design and in development testing. In consultingthe outside sources, the following questions were asked:How does your present day experience compare with publisheddesign allowables (or lives); what is your estimate forgrowth in the next 5 years; upon what technical advance-ments is your growth prediction based: what does existingtest data indicate for growth potential?

Bearings

On the question of gre'-th in bearing capacity, the consensusis that a life expectancy of 3 to 10 times present catalogratings could be realized by 1970. New m-terials and tech-niques have shown increases of 20 to 30 times rated life, intest lots. Life improvements have been twofold or threefoldin the past few years, but these h6;ve been hidden by moredifficult operating conditions, such aa the use of MIL-L-7808oil, and at times, by unforeseen dynamic loads and misalign-ments. Qualified in terms of load for the same design life,

21

future bearings may be expected to ca-.y 1.4 to 2.2 timespresent loading.

Bearing life-load relationship is derived from the followqingequations (Reference 5):

1. For ball bearings: L (6)

2. For roller bearings: L =(9) 3.33 (7)

where

L = lifeC = load ratingP = load

The methods of attaining oearing life and load improvementwill include the following:

1. Reduction of unwanted inclusions by melting steelin vacuum.

2. The use of new types of bearing steels such as M-50.3. The systematic improvement of existing (52100)

bearing steel by evaluation of the effect on fatiguelife of trace elements, and by modification of themajor chemical constituents (Reference 3). Traceelements have been present in the chemical analysisbecause past methods of steel making could noteliminate them. New methods, including vacuummelting, are now able to control or exclude them.

4. Elimination of end grain and segregated material,by new forging techniques for raceways. This isparticularly pertinent to ball thrust bearings,wherein the ball elements run against the side ofthe race, and where end grain exists when races aremachined from tube.

5. Manufacturing techniques to improve bearing elementfinish and roundness.

6. Peening of bearing surfaces to improve residualstress conditions and possibly disperse carbidestending to form fatigue nuclei, and to improve endgrain effects.

22

• .- -. •

These, and other approaches, typify the manufacturing tech-niques which have already indicated significant bearing life-load capacity improvement. Some are presently in use in theBoeing-Vertcl CH-47A. Figure 10 illustrates the degree ofimprovewe"nt expected from various approaches by one bearing,manufacturear (Reference 5).

The question of bearing predictability, or the spread in testlife experienced in an apparently identical bearing lot wasdiscussed. The majority opinion was that little decrease inthis spread has been noted, despite the life increases ex-perienced as a result of material and deaign improvement.A significant improvement in predictability apparently aw~itsa basic explanation of bearing fatigue causality; this hasyet to be made.

Apart from the material. and processing improvements described,analytical techniques must keep pace to ensure utilizationof bearings to maximum advantage. The theory- is in largepart known, and requires systematizing and prograiming tomake it available for design use. With existing and futureprograms, the external forces upon the bearing, and internalforces velocities, and directions of the rolling elementsmust be quantified with exactitude. The method of analysiswill include the rigidity cf the races, centrifugal and gyro-scopic forces upon the elements, and other effects whichdifferentiate a real bearing from the si-.,plifying assumptionsof normal calculation methods. T1his difference becomes ofmajor importance with high-speed, highly loaded bearingsoperating with combined loading and with varying degrees ofhousing rigidity.

Spur and Helical Gearing

The power-carrying capacity of a given gear depends upon geartooth strength, and gear tooth surface contact capacity.Testing conducted by Vertol and others has given strongindication that aircraft gear tooth•Sending strength ishigher than generally assumed. For example. testing recentlyconcluded under Contract DA-23-204-AMC-02693(T) by VertolDivision has shown a mean failure leve! approximately fivetimes higher than norm.El design levels (Figure 11). Thisprogram was designed to evaluate virious gear tooth grindingeffects in co-mparative testing of basically simrilar gears.

23

7:7-- - --n• ,- -m -- C m

(INFORMATION COURTESY OF SKF INDUSTRIES,

INC., KING OF PRUSSIA, PA.)

CALCULATED LIFE LI 0

A. 52100 STEEL 0-1500 RPMAfRMELT X-9700 RPM

I

E. 52100 STEEL I(CONSUMABLE IVACUUM MELT) I

II

C. M50 STEEL I X(COSUMABLE IVACUUM MELT) I

I1-2-5

D. 52100 STEEL I XXXMULTIPLE REMELTSlst,2nd, and5th INGOTS I

IE. MODIFIED 52100 j

STEELS1. VACUUM MELT X XXX X X

2. AIRMELT 0(X 'GD 0 0

5 10.8 20 50 100 200LI 0 IIFE (MILLIONS OF REVOLUTIONS)

POINTS REPRESENT ESTIMATED LIFE OF INNERRINGS FROM SAME HEAT OF STEEL (MINIMUMLOT TESTED WAS 30 BEARINGS), OR THEAVERAGE OF SEVERAL LOTS.

FIGURE 10. BEARING LIFE IMPROVEMENT BY IMPROVED METAODS.

24

170

140 ~~L&L

H~~- ----~j -.. ILOWER LIMIT:::...

130__ I : :~

U) 90 __ _ - - -FAILURES~ .:.I .: ASSOCIATED z: 4 . .... .

80 GERTOT......SCORING70- -

50

40

10 5 106 10 7

N -CYCLES

SYM4BOL DESCRIPTION

0 -

(RUNOUT IS INDICATED

BY ARROW -

FIGURE, 11. BENDING STRENGTH TEST RESULTS.

25

The gearing used was representative of the material and pro-cessing used in helicopter drive systems. (See Figure 12 fortest specimen characteristics.) Because further improvementcan be expected by development, it is believed that safeworking bending stress can be realistically increased for the1970 era.

In this program, failure data were obtained economically andexpeditiously by using a gear research test stand. Most air-craft gear testing is not purposely continued to bendingfatigue failure, especially in a gear case designed to acceptthe loading without distortion. This type of investigationis recommended for continuation, as are means of achievingimproved strength-weight ratios.

Gear contact capacity is another consideration for gearsizing, and is of the same importance as bending strength.Both must advance by equal amounts to provide weight savingsin gearing. It is believed that present design contact stressis nearer the permissible limit than is present design bend-ing stress. This is evidenced by the frequency of contactdistress phenomena occurrences as compared to the incidenceof tooth failures. However, it is believed that severalpromising means of increasing contact capacity will lead togains in this area commensurate with the gains to be expectedin bending capacity. In consequence, the design of the HLHspur and helical gearing uses an index of bending capacity50 percent in excess of that now currently used, and a sim-ilar index for contact capacity 20 percent in excess of thatnow used. These increases are mutually compatible in pro-ducing a balanced improvement in gear size.

Bevel Gears

The allowable stress for a given gear type, as pointed outabove, is an index number for that type, within a certainenvironment. It is not, therefore, necessary that the bevelgear stress number improvement duplicate that for parallelshaft gearing. After consultation with the Gleason GearWorks, and review of Vertol Division experience, it was de-cided to predict a 12 percent increase in bending strengthcapability, and a similar increase in surface wear capabilityas goais attainable with the aid of a bevel gear advancementprogram. Figure 13 illustrates a statistical analysis of

26

FILLET DETAIL

0.376 -0.000 DIA.

ON 4.000 BOLT!\CIRCLE DIA.

10.000

0 3.3705 +.0005PITCH DIA 3.0000

GROUND FLANK & ROOT GROUND FLANK ONLY

PROTUBERA14CE HOBBEDFILLET

FILLET DETAIL FILLET DETAILFOR - 2 & -3 FOR - 1 ONLY

GEAR DATA: TYPE - SPURNUMBER TEETH -70DIAMETRAL PITCH -7PRESSURE ANGLE -250

MAXIMUM LEAD ERROR .0003 IN/INMAXIMMA TOOTH TO TOOTH

SPACING ERROR -. 0003

FIGURE 12. SPECIMEN - GEAR BENDING TEST.

27.

LINE 2 ADVANCED LEVEL

MMEN: P - . 50------ +--€---€- "•5b = 60,000 ••

7z S1 .b S35,500

R .9999

S Sb 5 54 000 a

L, I *EI, -COVNTOA LEVL-

S•'----- •Sb - 30,000

0 .0 • o PROBABILITY OF SURVIVAL a '

FIGURE 13. SPIRAL BEVEL GEAR BENDING STRESSVERSUS PROBABILITY OF SURVIVAL

TABLE III. GEAR STRESS LEVEL TABULATION

Item Corn entional Advanced

Bevel Gear*

Bending Allowable - (psi) 30,000 35,000Contact Allowable - (psi) 225,000 250,000Load Distribution Factor 1.1 1.1

Spur Gear**

Bending Allowable - (psi) 30,000 45,000Contact Allowable - (psi) 150,000 185,000

*Calculation by Gleason tooth geometry and gear rating

program.**Calcuilations:

Wt- Pd3ending (Sb) F Z dF Yk ________

Contact (Sc) 3180 ! Wt R + r4 FSin20 Rr

Scoring (Tf) AGMA Flash Temperature Scoring Index217.01

28

bevel gear bending fatigue stress related to reliability.The basic data is drawn from Reference 8. Line 1 includes

all data points, and is the basis for the conventional stresslevel of 30,000 psi at R = 0.9996. Line 2 was drawn from the

same data after discarding the lowest failure points which,

by reason of material and processing, are not representativeof HLH gearing technology. Approximately 10 percent of thefailures were thus discarded to obtain a 35,500 psi stresslevel at R = 0.9999, which was the design assumption for theadvanced stress level.

POWERPLANT SELECTION

The primary emphasis in this study was on the solution of the

HLH transmission system, beginning at the engine shaft. The

engine output connection was considered as the interface be-tween engine and transmission.

Investigation of engines was therefore concentrated on thefollowing:

1. A review of available engines, in the horsepowerrange applicable to two-, three-, or four-engine

installations in the HLH, as presented by majormanufacturers: General Electric, Lycoming, Allison,and Pratt and Whitney

2. Determination of numbers of engines required toperform the defined HLH mission

3. Preliminary design studies for each major engineconsidered, to determine the effect of its charac-teristics on the transmission arrangement and onthe total reduction ratio requirement

4. Liaison with the engine manufacturers during thelayout design stage, and discussion with them of

major design problems

A tabulation of the available engines is given in Table IM.

The design criteria for engine power was determined by the95-degree 6000-foot altitude hover with payload, required for

the transport mission (Figure 14). The effects of one-engine-

out on the transport mission at standard conditions repre-

29

~~• •- - ==

0. 0 0 Mfg.

3~r "€ 5'

00 j- o

00

S0

41 "

0 to Deaigaat.> 0 x 0

I C '0 M

lb.- . - ye00VO m (n _q Vn

0- -

0 ,,00"*3 I n 0

-" o"o 0 0o

00

rfo

0 0 0 0 00O 0 w. 0oE 0 10 C' w0£- ~ ~ a -A*~ 0 - ' ,' ~

<0is. ( 2)

C', '0 PI0

00 wz*

-$ at 1 1t - I

0 ~ ~ ~ 1 0l 0 rv

0 0

0'

-t 0.1 00

m 0

0~ 0(1

0 Mis.l m t I

0 r- I ~

0~~ RPMeg-u

n.- -t -q -- -

REQUIRED POWER AT 6,000 FT - 95 0 F!

13,000 AIRFOIL = NACA 001213,000 W= TIST 90

17,000. ..

SHAFT HP .9REQ'D 10,0000

8,000

30 40 50 60

ROTOR RADIUS - FT

BASIC ACCESSORY POWER = 50 HPDOWNLOAD FACTOR = 1.04XMSN EFFICIENCY = 96%NUMBER OF BLADES = 3OVERLAP = 31%TIP SPEED - 670 FPSCT/T .1

FIGURE 14. POWER REQUIREMENTS.

31

sented less severe criteria, it: the case of three- or four-engine installations. After examination of possible combina-tions, i.; waq decided to limit the study to consideration ofthree- or four-engine installations, in the 5000-shp and 3800-shp classes, respectively. Two engines fall within eachclassification: the T-18 and T-56 in the 5000-shp class; theT-64 and T-55 in the 3800-!.hp class.

The performance of engines within a class was not sufficientlyanalyzed to permit a choice of one particular engine as opti-mum for the HLH. Tqhe delineation of a particular engine inthe drive system design studies should not be construed asindicating preference for that engine, because, in fact, nosuch choice was attempted.

In the investiqation of & three- versus four-engine conficu--ation, however, some general characteristics were notedwhich influenced the final design toward four engines. HLHpower requirements limited the three-engine design choice tothe T-78 and T-56 engine families. These are presently con-

fi'sured as fixed-shaft, not free-turbine, engines. A rotorengagement clutch is therefore necessLry for helicopter use.A review of engagement clutch design with Allison, and con-sideration of Vertol expe.'ience with the Allison-poweredH-16A and H-16B helicopters indicated to Vertol that such amechanism, wliile feasible, would represent a mandatory devel-

opment item. It would also present the transmission systemwith certain requirer~ients, notably, the provision of oil flowand oil cooling during the clutch encagement cycle. Theserequirements would impose penalties upon the transmissionlubrication system.

A review of the fixed-shaft engine power-speed characteris-tics =, -mea to indicate a less favorable response to powerdemari then qas available in free-turbine engines. The com-bined Droblems of clutch and pceer characteristizs tended toplace the fixed-shaft engine in an unfavorable position,relative to tne free turbine.

The availability of free-turbine versions of these engineswas discussed. It was understcod that if these were offered,they would he in a zear-3riv2 configuration only. Thispossibility was examired, anC it was concluded that a three-engine, rear-drive requirement woule add to transmission

32

- -- " - -

complexity to the extent of requiring at least one additionalbevel gear transmission. For this reason, the three-engine,T-56 family version, while able to meet the mission powerrequirements, was not ccntinued into the detail study area.The T-78 family, while showing attractive weights and spec-ifics, was not able to meet the mission power zequirement;for this reason, it was dropped from the study.

The preliminary examination of transmission systems compat-ible with three engines did indicate that such a transmissionsystem would meet problems similar in descripcion and magni-tude to the four-engine systems, with the added requirementof the engagement clutch.

The four-engine configuration was confined, as stated, toconsidezation of the T-64 and T-55. It will be noted thatthe planned development of both follows similar lines, andthat the standard-cond.tions output horsepower predicted is3800 in the 1970 perioo. The transmission system was de-signed to accept this power, although it is beyond the maxi-mum requirements of the mission specification for the air-craft used as the study vehicle. The T-55B-1 and theT-64-S4A enginec coulO provide 10-miaute-rating power suffi-cient to maintain the HLH transport mission hover. (Thishover requirement is for three minutes at 6000-foot, 950Fconditions.) The required power could be provided withinthe 3400-shp design. It is understood that both manufac-turers coulc provide the 10-n.inute rating in advance of the3800-shp military-rated design. Figure 15 indicates possibleengine growth timing.

The major problem which must be explored before a four-engine(or three-engine) HLH is designed is in the engine controlarea. As explained 4n Recommendations, iL is felt that thiscan be done by simulation, and that various control modescan be investigated sufficiently to provide confidence inthe HLH configuration. Improvements in engine control onpresent two-engine production aircraft are currently underinvestigation by Vertol Division. The results of this ii-vestigation may provide assistance to the HLH program.

33

LEGEND4600"- LYCOMING in T-55

GENERAL Illlllll T-64ELECTRIC _

C4 z

z H

-Z __

ZE- -

4200 -- 0

195 166 16 198 16 90 17

>4H E-1 H~E-N A4CPT E-4NE-.

U 4 E-4

0 H H 0 o 0Z

H E' ~~ E-~ E-

3800J

P Z Z Z ZE-4z 0 :z 0 0 =0

3400--~

3000--

2600--

195 196 197 168 169 170 17

YEAR

HLH ERIO

FIGURE ~ 15mNIEAALBLT NDGOT LNIG(AEON ATICIATEDFUNDNG)

3000 34

TRANSMISSION SYSTEM CONFIGURATION ANALYSIS

In this section, the factors governing selection of a trans-mission system are outlined, the method of selection iE de-s-ribed, and comments are made on six configurations selectedfor preliminary study. Vertol Division experience in thedesign and production of helicopters makes for full recogni-tion of the interdependence existing between the airframe andthe dynamic systems. This experience, as well as the experi-ence gained in previous investigations and reviews of drivearrangements and components, has been used in the selectionof the drive systems for preliminary study.

Methodology

To systematize the possible approaches to be considered duringthis study, a comprehensive study Z drive system arrange-ments was made. This included al.. the realistic approachesthat could be visualized toward distributing power from two,three, or four engines to two rotors. These approaches werezeviewed and grouped by such major drive system criteria as:number of gears, number of transmissions, and number of powermeshes, to provide a ranking order. The ziumber of powermeshes reflects upon transmission weight as well as reliabil-ity. The power lost in a gear mesh is turned to heat, in-stead of lifting payload. At normal power loading, a singleextra mesh in the HLH may be equivalent to nearly nine hun-dred pounds, and will contribute this amount to the totaleffective weight of the drive system. Total effective weightis derived as follows:

Total effective weight = (dead weight) +(mesh loss equivalent weight) (8)

Mesh loss equivalent weight = number of meshes xhorsepower per mesh x power loss x power loading (9)Example: 1 x 15,200 x 0.0075 x 7.5 850 (10)

Power loss input power - output power (11)input power

Power loading = installed poweraircraft gross weight (12)

35

-%

Other criteria used to indicate the final study corfigura-tions, but not considered amenable to a simple number valueare:

1. State-of-the-art advances requireJ in such areas asvelocities, sizes, and loads

2. Effect of engine position on inlet ducting, exhaustposition, reingestion, and accessibility

3. Comparative reliability, as indicated by overallcomplexity and by position of combining gears withinthe rotor synchronizing loop

4. Structural requirements necessary to mount enginesand transmissions

5. Aerodynamic effect of necessary protrusions andhousings

6. Maintenance of aircraft center-of-gravity withinacceptable limits

Within the comprehensive group of configurations, it wasevident that many mechanical problems were similar; typicalcomponent arrangements reasserted themselves as solutions tothese problems. System studies were chosen to include asmany of these arrangements as possible. This was done sothat a number of solutions could be synthesized for problemsencountered in the continuous process of refining the HLHaircraft and drive system.

Discussion of Configurations

The configurations discussed in this section are shown inFigure 34. In addition to the schemes discussed, this illus-trLtion also presents some alternate versions which wereconsidered prior to selection of study configurations.

Major divisions of the schemes considered were establishedby the number of engines and by engine location.

Two-engine versions may be derived from arrangements whichgroup two engines forward and two aft. Alterr.ate arrange-ments to many of the versions ir.. 'de various positions ofcombining gears, an aft planetary -icated a'. the top of thepylon, and minor changes in gear ratios and intermediatespeeds. These are not shown.

36

The configurations are ranked by mesh loss although, as ex-plained, this is only one of a number of judgement criteria.Increased mesh loss is generally indicative, however, ofincreased weight and _omplexity and decreased reliability.Figure 16 indicates a partial list of the criteria used toselect the study systems.

Four-Engine Configurations

Configuration I was chosen on the basis of low losses anddesirable simplicity (Figure 2). This configuration and itsvariants is discussed in the RESULTS section, and illustratedby Figures 37 through 43. A version of this basic configura-tion (referred to as IA) divides the aft transmission into alower bevel drive, and a final reduction atop the pylon. Itis otherwise identical to the parent configuration.

Configuration IV was selected for limited study on the basisof being directly representative of a CH-47A transmissionsystem, enlarged, and with two added engines. The advantagescenter about engine location; they are ideally positioned fordirect inflow, ducts are not required, and the possibilityof exhaust reingestion can be inferred as minimal. The poddedinstallation would provide maximum accessibility. To conferthese advantages, the transmission system increases in com-plexity. The number of gearboxes is increased by five; ofthis number, four may be identical engine nose boxes which,however, present particular problems in remote oil supplyand scavenging. The rotor gearboxes are identical to thoseof Configuration I. This configuration has not been examinedin detail; Figure 50 delineates the arrangement and inter-mediate speeds. Reference 7 includes a study of a systemsimilar except for horsepower.

Configuration V was selected for limited study, particularlyin the area of the combining transmission (Figures 51 and 52).This is a particular problem, in that, with a parallel enginearrangement, the center distances of the input pinions areestablished at such a distance thac idler gears are requiredexcept at large reduction ratios (6 to 1 or more). Withoutidlers, gear pitch line velocity becomes untenably high atlower ratios. The insertion of idlers results in increasedweight, mesh losses, and complexity. The combiner trans-mission shown demonstrates both approaches by showing one

37

design with idlers, and one without. In this application,the reduction ratio is such that it represents a borderline

case for the use of idlers. With this configuration, a two-

stage planetary can be used above the combining gears. The

arrangement does not represent a simplification over Config-uration I, nor does it reduce the number of bevel qears.Further objections are the difficulty of access to theengines, and to the vertical shaft, which runs through the

center of the engine pack. With a two-engine input at eachpylon, the problem of accessibility would be eased. Thecombiner transmission would be basically the same but with

two inputs. The transmission system comments would pertainequally to this minor variation. The four-engine-aft config-

uration is depicted in Figure 51 with appropriate shaftspeeds and gear ratios.

Configuration VI was selected for limited study, a majorvariation on the parallel-engine concept. In this case,

the engines have been staggered to reduce the required com-biner box input spans. A skewed engine arrangement was in-vestigated at the same time, also with the objective of re-

ducing combiner size, and eliminating idler gears. The

skewed arrangement, in which engine inputs converged on acommon point, was discarded when the necessary size of theengine compartment was established. Configuration VI, shown

in Figure 53, represents a more attractive solution tc thetransmission system and to engine compartmentation. There

is a major problem apparent, however, in the provision of

satisfactory ai: intakes to the engines. The extreme aftposition of the engines indicates that a reasonable aircraftCG location may be difficult to attain; in addition, penal-ties may be incurred in providing adequate mounting struc-ture.

Configuration II was selected as a major study configuration.This arrangement is discussed in the RESULTS section. It isillustrated by Figures 44 through 47.

Configuration III was selected as a major study configurationin order to investigate the effect of a dual-shaft system onsize and weight. It is shown in Figures 48 and 49, and isdiscussed in the RESULTS section.

Configuration X is another arrangement of dual shaft and rear

38

SYSTEM CONPIGOUATIONFACTOR

I II III IV V VI

TRANSMISSION COUNT o o o o

GEAR COUNT 0+ @0@ 0 +

POWER LOSS 0oe~

COMPONENT VELOCITY (,AX.) 0 0 0 o - -o-

USE OF PROVEN CONCEPTS 0 o ÷ +÷TOTAL EFFECTTVE WEIGHTo

ACCESSIBILITY oF DRIVE o o o o@ •SYSTEM

ACCESSIBILITY OF POWERPLANTS

EXHAUST RECIRCULATION0

ENGINE AIR INTAKES ± AN- o o

AIRFRAME COMPLEXITY -- 4

ADDITIONS TO MINIMUM0@00

OUTBOARD PROFILE

COMPARATI VE RATINGS

O SUIpERIOR

€- AVERAGE

O INFERIOR

FIGURE 16. COMPARISON OF DRIVE SYSTEM CONFIGURATIONS.

39

1

mounted engines and represents a dualized version of Config-uration I. It illustrates the penalties of direct transla-tion from single shaft to dual shaft, in that the mesh lossand number of gears are increased. This is avoided in Config-uration III by arranging two engines with a direct drive intothe interconnect shaft and combining the other two into therotor transmission.

Three-Engine Configurations

Three-engine transmission systems were given limited studyto determine whether specific problems, and specific advan-tages, accrue to the use of three engines. As noted inPOWERPLANT SELECTION, the major study effort was concentratedon : four-engine vehicle.

Configuration XII has three engines arranged in a "T" forma-tion, combining upon one gear. The low power loss, and num-ber of bevel gears (6) make this an attractive s:hematicarrangement. Limited study of this system revealed thatvelocity of components directly related to the input sectionincreases by approximately 10 percent as compared to thefour-engined Configuration I. These components include bevelgears, overrunning clutch, and bearings. Such an i~-reasedoes not necessarily represent a crossover into problem areasfor these components, although, as noted previously, theoverrunning clutch requires development effort.

Aside from the input section, the rem-ainder of the three-engine transmission systems investigated show no unique prob-lems. The engagement clutch has been reviewed under POWER-PLANT SELECTION and represents a major development item, ifthe use of fixed-shaft engines is contemplated. The rotortransmission final reductions shown in Configuration XII, forexample, ire identical in arrangement and size to those usedin Study Configuration II.

FINAL REDUCTION SUBSYSTEM ANAl YSIS

In this section, various methods of obtaining the final re-duction of speed to rotor rpm are considered. The factorsleading to a choice of method are outlined, and the differentapproaches are compared.

40

Subsy;stem Requirements

The final reduction is required to deliver power to the rotorshaft efficiently, reliably, and within a compact outline.The proportion of the total system ratio reduction that itaccepts is dependent upon matching its own characteristicswith those of the drive system ahead of it. Because no methodof obtaining the entire reduction in one package (withoutmultiplication of gear stages) is now available to the heli-copter designer, the practice is generally to take ratio re-ductions wherever a direction-changing gear mesh must existin the system. The final reduction is left with the residualratio which for various reasons was not extracted in the pre-ceeding meshes. This residuum is typically 15 to 30 percentof the total reduction, in the medium transport helicopter.

The HLU configurations under study have total system -atiorequirements of 100 : 1 to 150 : 1, depending upon engineselection. This is roughly twice the CH-47A ratio. Thegearing upstream of the final reduction was arranged to giveeither one or two meshes in the different study configura-tions. By using the ratio reduction potential of thesemeshes to the fullest extent, the requirement for final re-duction ratio was set at approximately 45 to 1 (with onepreceeding mesh), and 20 to 1 (with two preceeding meshes).This selection took into account present state of the art infinal reduction methods. Other methods now under developmentwill, if successful, influence a choice to a higher finalratio in order to keep intermediate speeds higher, and torqueand weight lcwer.

To conform to the state-of-the-art requirement, all powertransmission was by involute-tooth-form gcaring, althoughrequirements exist to which the non-invol-.ite tooth form(conformal) could be applied with advantage.

It is more economical, from a subsystem weight standpoint,to reduce any given ratio by load-sharing meshes, than bytransmitting all load through one mesh. As an example ofthe weight characteristics of single-mesh versus parallel-mesh subsystems, reference is made to Figure 17. This indi-cates the weight saving effect of using a number of parallelmeshes to replace one. It is evident that while three orfour parallel meshes decrease weight significantly, a

41

diminishing improvement is obtained from more than this num-ber. Using assumptions other than shown, the optimum numberwill change, but thb change will not be substantial. Inaddition to the accepted benefits of multiple meshing, whichmostly relate to the desirability of packing in the mostworking metal within a given envelope, these systems lendthemselves to epicyclic (planetary) mction, with the addedadvantage of increase in ratio, and decrease in powei loss.T1erefore, the final reduction investigations concentratedon parallel mesh and epicyclic motions.

The ratio requirement of under 20 to 1 was well suited tothe proven ability of the two-stage epicyclic approach. Theassembly for Configuration I, shown in Figure 39, drawsheavily upon the developed CH-47A planetary system. Gearmounting, ratio division between first and second stages,and component sizing are the result of this successful ex-perience. This solution, therefc.e, has a high confidencelevel. It is worthy of note that the HLH under study doesnot require a new concept of final reduction in order toobtain an effective drive system.

Systems which combine engines on a gear which is immediatelyadjacent to the final reduction input (Configuration II, andothers) require a higher numerical ratio than the foregoing.To achieve the required 45 to 1 ratio with parallel-mesh,epicyclic methods, an investigation of arrangements was con-ducted. The results are given in the following four studies:

1. Compound first stage and single second stage(Figure 54)

2. Three stages (Figure 55)3. Split-torque first stage and single second stage

(Figure 56)4. Poller gear, three stages (Figure 57)

Schematic representations of the four types are shown inFigure 18.

The results. of the study are shown in Figure 19. Based upontotal effective weight, the chosen method, which is entirelywithin the state of the art, is the compound and single con-figuration SK 14232. This is shown in Configuration II.The roller gear represents an attractive alternative, if

42

7t

100

90

WEIGHT% 80

70-

60

1 4 8NUMBER OF MESHES

ASSUMED: FACE WIDTHS ARE EQUAL

FIGURE 17. EFFECT OF PAPALLEL MESHES ON WEIGHT.

43

mLm

COMPOUND AND 3-STAGE PLANETARY(1SINGLE STAGE (D SK 14231

SK 14232

SPLIT TORQUE (4 ROLLER DRIVE SK 14233 4SK 14230 '/(EXTENDED FOR CLARITY)

FIGURE 18. Y-INAL REDUCTION METHODS - SCHEMATIC COMPARISON.

44

3600-

3500-

3400-

3300-

3200-

TOTAL 3100 1EFFECTIVE -'*

WEIGHT 3000-PER

ASSEMBLY II(POUNDS)

Liffi- ?JSK 14230 ISK 14231 SK 14232

SPLIT-TORQUE I3-STAGE COMPOUND

20001

1900]

1800-

DEAD 10WEIGHT 10PER

ASSEMBLY

(OU'NDS)I

SK 14230 SK 14231 SK 14232SPLIT-TORQUE 3-STAGE COMPOUND

FIGURE 19. WEIGHT COMPARISON OF REDUCTION METHODS.

45

further testing continues to uphold early model-test indica-tions. Comments on each arrangement are grouped below underthe main criteria of choice:

Power Loss

The "equivalent number of meshes" were calculated ac-cording to the method of Buckingham (Reference 1) andothers, to obtain the power losses. Because an epicyclicsystem entails relative motion of gear certers, the ve-locity of engagement is the sum of tangential and gear-center (orbital) velocities. Mesh loss is the productof load and velocity; therefore, a lowered resultantgear mesh velocity gives a smaller power loss than afixed (nonorbiting) system. This is generally true ofthe epicyclic systems used in helicopter drives. Onthe other hand, it is possible in a differential systemto increase velocity, so that the resultant power lossis far above that which would be obtained if the gears werefixed.

The effect of relative velocity of the gear elementswas taken into account in the analysis of the methods.An apparently small change in mesh lo:s can signifi-cantly affect the total effective weight of the method,and therefore, its competitive position. For example,the three-stage planetary, with a low dead weight buta comparatively high power loss, is relegated to a lowposition in overall weight standing.

To reduce the dead weight of the rcller drive, a furtherreduction stage was added. This is shown in Figure 18as Arrangement 4. This arrangement was investigatedas an epicyclic, with fixed ring gear. The mesh lossis slightly less than that of a fixed-center system,because of the obital motion in the first and finalstages. However, the orbital speed, equivalent to out-put rpm, of the gears contributes little to reductionof first-stage resultant velocity and mesh loss. Inthe same way, the split-torque methcd does not gainsignificantly in mesh loss because the carrier is ro-tating at output rpm (at this ratio, 2 percent of inputrpm). This is practically a fixed condition when com-pared to a simple planetary, where orbital rpm may be

46

30 percent of input rpm, with the effective mesh losssubstantially reduced in consequence.

It is significant that the lowest power losses for a 45to 1 method are not greatly in excess of those for the20 to 1 reduction. Therefore, the higher ratio ispraccical from a total effective weight standpoint.It is, however, close to the practical upper limit forthe methods investigated, unless an added stage is used.Figure 20 shows equivalent meshes and range of ratio forthe methods studied. Refer also to equations in Appen-dix I.

Vo..lume and Weight

The maximum diameter of the final reductions was deter-mined _by the ring gear required for the last planetstage. The load carried by this stage is equivalent torotor torque. In the arrangements shown (Figure 18), itis distributed over six planet gears.

The split-torque design (Arrangement 4) necessitates arotary ring gear. Enclosing this gear will necessitatea gear case from four to six inches larger than wouldbe required for designs using a fixed ring. To reactload, the split-torque design shown uses the planetcarrier. There is an inherent weight penalty in thatthe split-torque arrangement requires all the componentsof the fixed-ring designs plus a ]arge disc, to make thetransition from ring gear to rotor shaft. Similarly,in the first stage the rotating ring gear requires asubstantial disc to carry out load.

Arrangements 1 and 3 use essentially similar firststages. Clearance diameter for the rotating clustergears is less than the final ring gear because of theratio split chosen. However, because only three planetgears can be geometrically accommodated, the first stageis heavy and deep, by comparison with the six-gear finalstage.

To reduce first-stage bulk, it was separated into twostages; this resulted in Arrangement 2. A lower weightis possible because more planet gears can be fitted in,

47

to share load, when the ratio is iower. Arrangement 2,the three-stage planetary is lowest in volume and indead weight.

In order to V.t in more parallel-mesh gears at a givenratio, the roller gear drive uses offset planet gears.The primary weight advantages inherent in this approachare: first, that the second and any succeeding stagesbenefit by two meshes per gear; and second, that thesecond--stage pinion is driven from both ends, thus re-during torsional deflection in this small-diameter, wide-face gear. Because of the small diameter, it is possibleto obtain ratios of 35 to 1 (as shown in Figure 57) withthree meshes, in a reasonably compact design.

A roller gear ratio of 45 to 1, consistent with theother designs, was investigated. It is desirable tolimit the diameter of the final reduction unit. Toaccomplish this in two stages, pinions of small diameterand correspondingly large-face width were required. Afurther effort was then made to compact the 45:1 assemblyby adding another stage (Figure 18). The diameter ofthe ring gear was reduced, as would be expected, at thecost of additional mesh loss. Because of the relativelysmall amount of operational experience with this drive,and because time did not permit optimization of thedesign, no final weight estimate was made. Tf resultsof planned full-scale tests are favorable, the deviceshould be at least competitive in weight with more con-ventional systems. To demonstrate superiority in totaleffective weight, it would be necessary to establish,from full-scale testing, that the method of positioningresults in higher efficiency or greater gear stressallowables than are obtained with conventional planetar-ies at the same ratio.

Load Shazing

To gain from parallel meshes, all paths must share inthe load. If all paths cou]d be induced to carry ex-actly equal loads, under all conditions, a significantimprovement in load-carrying ability of the assemblywould be realized. The magnitude of inequality isapproximated from calculation and experimental evidence.

49

S3-STAGEi•

4.0

ROLLER 4-STAGE0

En•n •SPLIT TORQUEcnn

0~ 3.0

H ROLLER 3-STAGEC2 (SK 14233)

(SK 14218)

2.0 9 i i I10 20 30 40 50

FINAL REDUCTION RATIO

00 0 0 REFER TO FIGURE 18 AND TO TEXT

FIGURE 20. EQUIVALENT MESHES OF FINAL REDUCTION METHODS.

49

Load sharing is highly dependent upon component accuracy.It may be improved in some cases by the fcllowingapproaches:

1. Controlled elasticity of the supporting members,as for example, the planet carrier posts

2. Mechanical freedom in members; for example,floating (spline-supported) sun gears, and/orring gears which are allowed to shift radiallyunder load

3. Equalization of axial forces among a group of(helical) gears, which will automaticallyequalize tangential forces. (This is partic-ularly applicable to a compound planetary suchas SK 14232 (Figure 54); various methods, in-cluding use of hydraulic devices, have beenproposed.)

4. Mechanical equalizing levers or cams to allowplanet gears to adjust position to inequalitiesin tangential load

Approaches 1, 2 and 3 are either in use, or appearespecially suitable for helicopter final drives.Approach 1 has been shown in the drawings.

To share load in a different sense, that is, betweenstages, the split-torque method has been suggested.Analysis of the split-torque design shown in Figure 56(SK 14230) disclosed that although 25 percent of thetorque was in fact transmitted from the first-stagecarrier direct to the output shaft, the benefit to thesecond stage was minimal. This was due to the non-orbiting design of the second stage, whereby its re-duction ratio was numerically less than a geometricallysimilar orbiting arrangement. This meant that the inputrpm to this stage was necessarily less, to accomplishthe same overall ratio. Therefoie, the torque, for thesame power, was greater. This effectively negated thetorque split in the first stage, and resulted in second-stage gear sizes which are the same as those used inthe other designs.

The primary advantage seen in the split-torque deviceis that it reduces orbital velocity of the first-stage

50

planets. This is an important consideration when theinput rpm approaches power turbine speed. The centri-fugal effect is not, however, significant in the speedrange shown here, where centrifugal force representsonly 10 percent of the load upon the first-stage planetbearing.

Conclusion of Design Comments

The designs shown in Figures 54, 55, 56, and 57 uniformlyreflect the choice of stress levels proposed for study forthe HLH time period. If such designs were to be constructedbefore implementing a complete evaluation program, a moreconservative approach would necessarily be taken. The changesin size and weight would be of the same order as that shownby comparison of Figures 9 and 10. The single most importantsize effect would be growth of the final stage planetary dia-meter by 15 percent. This would place the ring gear in asize range above that noW accommodated in heat treatmentpresses and gear grinding machinery. The ring gear designedto advanced stress levels could be handled with existing pro-duction equipment. Based on present experience, the heattreatment distortion expected in a carburized ring gear ofeither diameter will create manufacturing problems. To reducethese, an investigation of materials and processes should beconducted. One suggested approach is the use of through-hardening steels, to eliminate the carburizing cycle.

BEVEL GEARING CONSIDERATIONS

The HLH drive system, by reason of the horsepower transmitted,requires that more-tlhn-normal consideration be given to cer-tain aspects of the beval gearing design. These are: (1)mounting accuracy, (2) vibratory loads, and (3) scoringhazard.

Mounting Accuracy

Bevel gear design calcilations for stress and load carryingcapacity involve certain assumptions regarding the toothbearing which will be obtained in the gear mountings underload. Unless these assumptions are closely fulfilled, actualstresses may be considerably different from the calculatedvalues, and early failures may result. If the gear mountings

51

could be made very rigid, the behavior of the tooth bearingunder load would be predictable and a suitable tooth bearingcould be developed based on experience. In helicopter drivesystems, however, mounting designs differ greatly, and rigid-ity is sacrificed to some extent in favor of weight.

The physical dimensions of the HLH transmission cases, andthe magnitude of the gear loads imposed upon them, indicatethat adequate mounting rigidity of the gears is more diffi-cult to obtain than in smaller, less highly loaded transmis-sions. The larger gear teeth used in larger transmissionsare not, unfortunately, tolerant of proportionately higherdeflections. Therefore, emphasis must be placed on designapproaches which are inherently rigid, and which thereforeminimize weight.

The design approaches used throughout the HLH study haveincluded:

1. Large-diameter, rigid, gear shafts2. Straddle-mounting of the pinion member3. Preloaded thrust bearings to reduce axial motion4. Double-web support for large-diameter gears

Vibratory Loads

A gearing design may produce natural resonant frequenciesin the gear and its supporting web and shaft which are withinthe spectrum of the vibratory inputs. A minimum-weight de-sign, operating at higher meshing frequencies, is particu-larly susceptible to this problem, which, if not checked, islikely to cause catastrophic failure. The primary causationfor a given gear set is forcing frequency; tooth stresslevel, rim velocity, and gear diameter do not, taken singly,cause a resonance problem.

In the past, the technique for obtaining a solution to bevelgear rim and web resonance had not been well defined. Pre-vious investigations of parallel shaft gearing (Reference 4)led, however, to a general understanding of the experimentaltechniques which could be used to indicate the natural fre-quencies of the gear on its support. With these known, areliable appraisal of the existence and magnitude of theproblem can be made by comparing the natural frequencies to

52

the range of operating frequencies. If the natural frequen-cies fall within this range, corrective action is required.One corrective action consists of attempting to stiffen thegear support and raise the natural frequency by adding irrtal.It has been shown that this is not particularly effective,at least within the restrictions of a flight-weight trans-mission. A more powerful approach is to absorb the resonantenergy by damping. This has been successfully accomplishedin a number of instances. It represents a reliable andminimum-weight solution to the problem.

As previously noted, the 'evel of stress in gearing is not,of itself, likely to predispose the gear set towards aresonance problem. It is not foreseen that the advancedstress levels proposed will make the problem more likely tooccur, or more difficult to solve. In fact, by reducing geartooth size, they may reduce the magnitude of the forcingfunction. However, it may be difficult to stiffen large-diameter gears sufficiently to avoid the entire range offorcing frequencies, not withstanding the inherently rigiddesigns employed. Therefore, the HLH gearing will be testeefor dynamic response i, the design stage. This can be donereadily by using semifinished or "boiler-plate" models. Ifdamping is desirable, this too can be evaluated by the samerelatively quick and inexpensive method. There is, therefore,a proven approach to the elimination of rim and web resonanceproblems; this approach will be applied to the HLH gearing.

Scoring Hazard

Scoring is defined here as the welding, and subsequenttearing out, of gear tooth asperities. It is a function ofgear tooth load, lubricant film strength, velocity, and otherfactors which may raise local surface temperatures above themelting point. A severe scoring situation is rapidly evi-denced, and is self-perpetuating. It is one of the boundaryconditions which determine gear load-carrying capacity, andas such, is of the highest importance to the HLH drive system.

The load-carrying requirements of the HLH drive system neces-sitate relatively large tooth sizes, for bending strength.The size of the tooth is indicative of the sliding velocity,which is a factor in heat generation. The combination ofcoarse diametral pitch (large tooth sizes) and high unit

53

tooth load is con6ucive to the development of scoring. How-ever, Vertol's past experience and that of others indicatethat the scoring hazard can be overcome through the develop-ment of suitable surface treatments or deposited coatingssuch as silver plate, tuftriding, alphatizing, etc. Addi-tional possibilities for resisting the scoring hazard existthrough the development of improved surface finish and topog-raphy and the proportioning of tooth geometry to reduce theinstantaneous temperature rise.

Designs using advanced tooth bending stresses will resultin smaller tooth sizes. This will be conducive to lesseningof scoring tendencies.

Indications, using the Gleason scoring index, are that theHLH bevel gearing, without modification, is within the scoringproblem area. As noted, there are solutions which have beenapplied to other gearing, and which have been found effective.It is now necessary to establish the approach for the HLHrequirements of load level and tooth geometry and size.

54

RESULTS

In the previous section, TRANSMISSION SYSTEM CONFIGURATIONANALYSIS gave the background for the configurations thatwould be further studied. These were selected from among alarge number of possible arrangements. This section is de-voted to analysis and discussion of those systems chosen forfurther study.

By the requirements for this study, two configurations wereto be pursued up to the point where a selection could bemade; one configuration was to be completed.

As the study developed, it became apparent that severalwidely different arrangements should be considered to indi-cate direction for both transmission and airframe develop-ment. The minimum requirements of the study were thereforeexceeded, in that two configurations were completed, and athird configuration was continued into a study of one gearbox.

CONFIGURATION I

The transmission system which has been chosen as the one inwhich most confidence can be placed at this time, is StudyConfiguration I. This configuration was chosen because:

1. Engine arrangement in aft fuselage minimizes theproblems expected from recirculation and properinflow.

2. There is a minimum number of gearboxes (two).

3. Selection and placement of speed reduction mech-anism results in final drive and bevel gearingarrangements which are extensions of CH-47A experi-ence.

Inherent Reliability

The combining of multiple engines by means of a single gearcase forming part of the aft rotor transmission representsa substantial reduction in number of transmissions, com-plexity of lubrication system, and number of meshes, whencompared to the configuration shown on Figure 50. The

55

arrangement also removes the combining function from theinterconnect shaft, with an expected improvement in systemreliability. Further study may prove the practicability ofa breakaway section, or clutch, between combiner and rotorcases, to free the rotor system from the combining gears incase of a catastrophic failure of the latter. This would per-mit a safe autorotational descent with rotors synchronized.

Further, the rotor brake has been located on a shaft exten-sion of the combining box, away from safety-of-flight shaft-ing. The plane of the disc is clear of the accessoriesmounted above and forward, on top of the combining box.

Aerodynamic Consideration

The position of the engines, required to accomplish the trans-mission system simplifications enumerated, requires a stubwing extension from the base of the pylon. The structuraland aerodynamic effects must -eceive more detailed analysisthan is possible in a transmi-sion study. However, to answerthe most immediate question, preliminary analysis was madeof hover download, intake efficiency of engine ducts, andexhaust recirculation probabilities. This investigation pro-vided a satisfactory level of confidence in those areas.

Hover Download

The mean span of the housing (stub wing extension) forthe four aft-mounted engines is about 28 percent ofblade radius, based on z 48-foot total radius. Withthe present trend of 20-percent root cutout and slip-stream contraction. hover download is significant onlyfor the parts of the fuselage which are outside 25-per-cent radius from the rotor center. The download on thepresent arrangement is about 150 pounds for the twoengine housings (Figure 21).

For the forward and aft arrangements, the engine hous-ings are both less than 25-percent rotor radius fromthe centerline of rotation and do not, therefore, con-tribute significantly to the hover download.

For lower rotor radii, corresponding to higher discloadings, the engine arrangements shown in Figures 2

56

!/25% OF ROTOR

RADIUS

FIGURE 21. ENGINE HOUSING PLANFORM.

57

and 3 may be moved inboard. The air inlets for thisconfiguration would be deepened to meet the mass flowrequirements.

Inlet Ducting

Inlet losses are held to a minimum by the large intakeducts. Velocities in the ducts are of the order of 100feet per second, thus presenting a low dynamic pressureto obstacles such as the drive shafts and to 90-degreeturns in the ducting. At the engine inlets, the airvelocity increases to approximately 250 feet per second,eliminating adverse pressure gradients due to diffusion.

In forward flight, the airflow external to the inletducting traverses an NACA Series I cowling lip on theleading edge of the airfoil-shaped wing. This lip isdesigned to prevent airflow separation on the wing dur-ing all flight conditions. Steep descents will providethe greatest angle of attack to the cowling lip. Thelip geometry can be optimized by wind tunnel tests.

Exhaust Recirculation

In a configuration study, exhaust recirculation couldbe examined by the test technique described in Reference6. The approach has been successfully used to determineoptimum engine location to prevent possible impingementof exhaust gases on the aft pylon. Figure 22 illustratesone step in the investigation using smoke flow in a 1/8-scale helicopter model. An extension of this approachis believed applicable to determination of flow patternsfrom engine exhausts ove:" various flight regimes.Engine exhaust recirculatioi, would necessarily be pre-dicted and corrected at an early stage in the final de-sign.

Maintainability and Size of Components

To illustrate the accessibility features of Configuration I,Figures 23, 24, and 25 are included. These show the use ofthe engine housing as a work platform for engines and afttransmission. The engines are removable by lowering straightdown from the installed position, as indicated on Figure 21.

58

Tr-2

FIGURE 22. EXHAUST STUDIES.

59

T9

II

... .. ..0/ ...

.. .

1-3

H~~~ ............... .... ...

l• K>"

\H

S' !H

. . . . . . . .

H . . o. . . . . . . . . . . . . . .

L. . . ..

09

I -\*I j

I \

1�

o 1

'I.Lii .11.0

z

zwH I.HI \J�i.\. 'I:.

HLJ

I.../

9-49

to

z... .. ..

I -4

Maintainability, although not covered in this report, is con-sidered an important factor in the comparison and selectionof a transmission design.

Variations of Configuration I

A number of alternates have been investigated within thegeneral framework of Configuration I. These were:

Supercritical Versus Subcritical Speed Shafting

The comparison is shown in Figures 26 and 27. At thesame shaft rpm, and using similar diameters, the super-critical shaft shows advantage by weight reduction (160pounds) and by reduction in number of sections (2 versus10). This is despite the fact that a steadyrest supportis placed in the center of the supercritical shaft torestrain shaft deflection under flight maneuver condi-tions. The work currently in progress at the VertolDivision under USAAVLABS sponsorship will, at its conclu-sion, provide additional information and an increasedconfidence level in the use of supercritical shaft de-signs in this application.

High-Speed Interconnect Shaft

An interconnect shaft system operating at engine outputrpm was investigated for possible advantage over amedium-speed (6000 rpm) system. In both cases, thesupercritical speed approach was used. The medium speedwas selected for this configuration because of thefollowing:

1. There was no change in total 2ffective weight(Figure 27). A 9-percent increase in power loss,combined with a heavier final reduction, servedto cancel the weight advantage of higher-speed,lower-torque shaft and bevel gears.

2. The rotor transmission becomes more complex andlarger in volume to provide the additional ratioreduction necessary to the high-speed system.

3. The bevel gear pitch line velocity was substan-tially increased by the higher shaft speed.

63

600-

537

500-

40037WEIGHT(POUNDS)

300-

200

SUBCRITICAL SUPERCRITICALSHAFT SHAFT

3030-

NUMBER20-

MAJORCOMPONENTS

10 8[ ........

SUBCRITICAL SUPERCRITICALSHAFT SHAFT

SCHEMATICS{SUBCRIT~ICAL e II

SUPPORT ANDCOUPLINGDAMPER

FIGURE 26. COMPARISON OF SUPERCRITICAL AND SUBCRITICALSHAFTING.

64

120%-

110%-

100% ::..

90% ISYSTEM

WEIGHT80

I SPE MONEI HF

SHAFT PLANETARYSYSTEM%

FIGURE 27. CONFIGURATION I -WEIGHT EFFECT OF ALTERNATEAPPROALijES.

65

Maximum velocity is 180 percent of that selectedfor Configuration I.

Final Reduction Position

The final speed reduction gearing for the aft rotor maybe mounted with the bevel gears at the pylon base(Figure 41), or separated from the bevel gears at thepylon top (Figures 42 and 43). The first approach pro-vides the minimum number of gear cases (two). It reactsrotor torque at the pylon base, with a direct load pathto the fuselage longitudinal structure. However, itrequires a shaft above the final reduction gearing capa-ble of transmitting rotor torque at rotor speed, andalso capable of resisting the bending imposed by rotorloading. The alternative zpproach, to place the finalreduction gearing at the top of the pylon, reduces thetorque requirement and eliminates the bending from therotor shaft. There is, in consequence,a substantialreduction in shaft size and weight. The reduction inweight chargeable to the tray mission system is esti-mated at 7 percent of the system by this approach. Toset against this, an additional transmission is, ineffect, created by the separation of bevel gearing andplanetary gearing. Torque reaction must be carrieddown the aft pylon with a possible increase in struc-tural weight. The structu:al weight increase has notbeen included in the above 7 percent. It will be re-called that this alternative approach was used in theVertol H-16 helicopters. Further study of the com-parative advantages is indicated, and will be conductedbefore a final configuration is selected.

Effects of Engine Choice

,he major effects upon the transmission system of achoice between the two engines, T-64 or T-55, are asfollows:

1. A ratio change would be required because of theJifferent engine output speeds. The T-64 wouldrequire an overall ratio of 13,600 1 134 =

101.5 to 1. The T-_55 would require an overallratio of 16,000 1 134 = 119.5 t3 1. The

66

difference could be taken in the first reductionratio (Configuration I), giving a 14-percentlarger pini,n for the T-64 than the T-55, be-cause of higher input torque and lower ratio.Gear pitch line velocity would be approximatelythe same. The remaining syste sizes and speedswould be unchanged.

2. The engine torquemeter, now a part of the T-64output shaft casing, would be revised to agreewith a new and longer output shaft casing.

3. To preserve the same direction of rotor rotation,the combining section of the aft transmissionwould be changed to place combiner gears on theopposite sides of their pinions. This wouldaccommodate the opposite shaft rotations of theT-64 and T-55.

CONFIGURATION II

Any study of tandem-rotor helicopter engine installationsmust inevitably consider the merits of mounting engines onboth the forward and aft pylons. Such an arrangement hasbeen used in the past by Vertol Divisioi on the H-16 seriesof helicopters. The installation of engines close-coupledto the rotors offers certain apparent and actual advantages;however, recent Vertol Division practice (as instanced by the107 family) has been to group the engines at the rear pylon.The reasons for this are:

1. Problems of exhaust reingestion are minimized. Themultitude of flight conditions under which the heli-copter operates makes even the most straight-forwardinlet-exhaust relationship a matter of concern andextensive study. Positioning of exhausts and in-takes along the fuselage increases the complexityof the problem, and decreases the probability of asolution satisfactory for all flight regimes.

2. Forward location of the engines may increase noise,beat, and vibration levels in the cabin area. Thepenalty is increased shielding or damping materialto counteract these.

67

3. Engine controls, fuel distribution, and servicingrequirements are separated as fax as is possiblewithin tie confines of the fuselage, thereby in-creasing weight and complexity. From a vulner-ability viewpoint, this penalty might be admissible.

There are some possible advantages from forward and aftengines. Ithe HLH concept on which this study is based hasa particular requirement for a fuselage length greater thanthat of the 107 family, but comparable to that of the H-16.It would appear that the length of fuselage might increasethe significance of such considerations as:

1. A normally unloaded interconnect shaft, with (underthe worst conditions) a requirement for one-half theshaft torque capacity normally required of the all-aft-engine configuration.

2. Dispersion or dilution of the exhaust stream in itspassage from forward to aft pylons.

In addition, at least one of the forward and aft engine con-figurations initially analyzed demonstrated that 6implifica-tion could be achieved by this arrangement; it also showedthe least number of gear meshes, and consequently the besteffective weight, of all transmission system arrangementsconsidered. Because of these benefits, it was decided toevaluate and compare this configuration (Configuration II)with Configuration I. T'his comparison could indicate whetherthe drive system advantage was sufficient to outweigh the ex-pected additional complexity required in other systems by thisarrangement. Comments on the system are as follows:

Component Arranqement

Engines are combined upon a bevel gear, which is (schemat-ically) mounted upon the final drive input. This is a mostsimple and satisfactory arrangement which does, however, re-quire a high-zatio, efficient, and compact final drive. Theanalysis of various forms of final drive is .o be found inthe FINAL REDUCTION SUBSYSTEM ANALYSIS section. Briefly, acompound planetary gearing arrangement was developed for thisapplication. It is considered that no exorbitant weight orefficiency penalties were incurred, as compared to the two-

68

stage final reduction shown in Configuration I.

The interconnect shaft system, as was noted, operates un-loadeC when power supply and demand are equalized betweenrotors. The design case for this shaft takes into accountan engines-out condition, so that shaft and bevels are de-signed with sufficient fatigue strength to accept the powerof one engine. The important consideration is that, comparedto Configuration I, the advantages of this unloaded shaftsystem are not substantial for the following reasons:

1. Supercritical speed shafting, used in both config-urations, has reduced the weight of both shaft sys-tems. For this reason, the difference in designloading between the two systems results in a reduceddelta weight for Configuration II of approximately1 percent of the transmission system weight.

2. The indeterminate load direction and unloaded cycleswhich the shafting, couplings, bevel gears, andsplines would experience would not necessarily re-sult in higher reliability for an unloaded system.

Inherent Reliability

This is, as indicated, not necessarily improved over Config-uration I. It may be less, because of the greater complexityof the rotor transmission. The rotor transmissions incorpo-rate the combining function; in Configuration I, this isdelegated to a separate combining gear case which is notwithin the synchronizing loop.

CONFIGURATION III

It is necessary to continually increase helicopter drivesystem reliability. There are at least two general approachesto the achievement of this; they are:

I. Arrangement of the drive system to provide maximumsimplicity, and reduction in the number of locationswhere the failure of one element can contribute tothe loss of an aircraft

2. Detail design and processing of the transmission

69

components to ensure predictable and uniform re-sponse to the environmental conditions to whichthey are subjected (Consideration must be givento the occurrence of environmental conditions whichlie outside the expected range.)

It is to be expected that approach 1 will be met insofar asthe separate, and sometimes conflicting, demands of trans-mission and airframe allow. Approach 2 is increasingly at-tainable as material technology and design experience advance.However, a condition may be foreseen wherein increasinglysophisticated manufacturing controls will produce decreasingimprovements in reliability.

A third approach is therefore indicated. This is the use ofparallel load paths, which yield theoretical improvements inreliability of the magnitude shown in Figure 28. The extentto which the actual improvement matches the mathematical pre-diction depends largely upon the independence of each load-carrying path. The load paths must be so designed and con-strained that failure of one does not cause the failure ofthe other. If this is not done successfully, then the over-all system reliability is degraded. The actual reliabilityimprovement depends also upon inspectability of each loadpath.

Many instances already exist of multiple load paths purpose-fully applied to modern helicopter drive systems, and moreapplications are continually under study and development.While recognizing this and taking advantage of it in allstudy designs, it was thought proper to extend the approachand to investigate an HLH drive system which dualizes theinterconnect shafting and spiral bevel gearing. This isdelineated in Study Configuration III. The objective is to

maintain complete interconnection of the rotors, despite thefailure of a shaft, coupling, or bevel gear. The followingcomments can be made on this configuration:

1. The 32-inch separation of the two interconnect shaftsallows inspection of each, and permits structure

between them to act as a barrier in case of failureof one.

2. Spiral bevel gears and mountings are located in

70

mmm-/mm -mmm- o m m

- .99

0.8

E-4

001.0.6 .90

0 W

FOrnELMNTEI ARLLL

0.4

1 -(FILUEPOABILITY OF SINGE ELEMENTn

FIGURE 28. RELIABILITY THROUGH REDJNDANCY.

71

opposite sides of the gearbaxes, spaced and sepa-rated as in the case of the shafts. Any set ofgears must, in case of a jamming-type failure, beable to separate without causing more than a localfailurr to the support structure. This presupposesa deta.l design of the case which would permit this,and would also contain the debris. Such a designis believed practicable.

3. The combining gears would be constructed as a pir-allel dual-face gear set. A tosted example of tnisis shown in Figure 29. The gear is a double helicalunit made of two single gears bolted -co a commonflange which is inserted between them. Th- testresulted in the nearly complete destruction of theteeth in one gear. The other gear successfullycarried load for some time after failure of thefirst, and maintained complete synchronization ofthe transmission drive. A development of this prin-ciple is believed to be particularly well suited tothe dual drive system combiner gear set.

4. The final drive is a planetary gear set comparableto that of Configuration I. The reasons for notduplicating the final drive are as follows:

a. Planetary gearing is inherently partially re-dundant by virtue of multiple meshing. Thisattribute may be developed, to increase thechances of survival after a failure, by variousapproaches now under consideration.

b. Historically, the higher-speed drive systemelements are more prone to catastrophic-typefailure. These are duplicated.

C. The mathematics of failure probability showthat duplication of some elements in a con-nected series contributes to an overall im-provement in the reliability of the series.Therefore, those elements that can be dupli-cated without incurring weight penalties areduplicated.

72

EL

L7J

5. The estimated dead weight for Configuration III isclose to that for single-shaft systems (I and II).This is because advantage is taken of the weight-saving possibilities of dual shafting in thisarrangement. For example, the straight-throughdrive from the two rear engines permits the use ofengine-speed dual shafts and gearing. The gearsthat collect the dual inputs are also used for speedreduction. The total number of meshes and the finalreduction ratio are equivalent to those of a lower

speed system (Configuration I) while the advantagesof a high speed system are also obtained. A dualizedconversion of a -:ingle system (Figure 34) will beheavier than the single system.

6. The rearmost engines, with air intakes as shown inFigure 48, have no ram air recovery in forwardflight. The consequent power loss, as compared tofull ram recovery, is estimated at 5.2 percent.The transverse engines are able to recover part ofthe ram effect, giving an estimated loss of 1.8percent. Loss of ram recovery is not directly com-parable to transmission power loss, since heat re-jection devices are not involved. The effect is toincrease specific fuel consumption, generally by0.6 percent for each 1.0-percent power loss. Atthe critical hover condition the effect of reduced

power must of course be considered along withpropulsion system efficiency in the final design.

In hover, power losses are estimated at 1.0 to 1.2percent.

7. Dual shaft systems are advantageous to the HLH be-cause power carried per bevel mesh is reduced to one-half that of the single-shaft system. The resultis that the size and peripheral speed of the bevelgearing are only slightly greater than those em-ployed in current systems.

74

EVALUATION OF THE STUDY

COMMITMENTS

The commitments of this study were several in number: (1) todefine the high risk development areas of a tandem-rotor HLFmechanical drive system; (2) to indicate the exploratorydevelopment programs necessary prior to the final configurationdesign; (3) to estimate weight, size, and efficiency of thedrive system.

To fulfill these commitments, the study investigated a numberof mechanical transmission systems. All contain componentswhich are possible candidates for inclusion in an HLH drivesystem; these systems were analyzed so that the study conclu-sions could be drawn from the broadest base. It was believedthat concentration on one configuration only could obscure thefact that much configuration analysis and appraisal of the HLHitself remains to be done. This further and continuing effortmay initiate requirements that would enhance the value of acurrently less attractive transmission system.

A review of this rather wide selection of configurations showedthat, in general, the transmissions do not require radical de-partures from present usage, either in principle or in physicalsize, to perform the HLH mission requirements. This is true,even though the horsepower transmitted is 3 to 4 times greaterthan that of present tandem-rotor machines. The reasons canbe summarized as follows:

1. The tandem-rotor helicopter, by virtue of powersharing between two rotors, does not require in anyone transmission the capacity required in that of asingle-rotor helicopter of equivalent gross weight.Therefore, a tandem-rotor HLH does not require amajor step in transmission size beyond present daysingle-rotor machines of the largest capacity.

2. Generally, transmission component %2iameters follow acubic rule. Therefore, an increase of perhaps 200percent in rotor power from a medium transport typesingle-rotor machine to an HLH tandem will increasesizes on the order of 25 percent.

75

3. Perhaps more important than size, component valocityand concomitant &ynamic considerations can be heldto a smaller increase over present practice than even25 percent. This can be done by selection of gearratios at the high-speed end, and by the lower rotorspeed required by the rotor diameter necessary forthe maximum gross weight specified.

4. Finally, techniques of dynamic control and analysishave progressed beyond those used in design of thecurrent generation of helicopters. It is hoped thatan aggressive program of still further basic improve-ment will be stimulated by customer interest. Withthe tools on hand, only evolutionary development incertain areas is necessary to assure high confidencein the HLH drive system. with successful improvementprograms, it is believed that gains in weight and re-liability will accrue to the benefit of the HLH.

The exploratory development programs found necessary to provideconfidence in the HLH drive system are grouped below under theheading of PROBLEM-AREA PROGRAMS. In addition, under theheading of DEVELOPMENT-AREA PROGRAMS are grouped programswhich, although applicable to helicopter drive systems gener-ally, are expected to result in measurable benefit to the pro-jected HLH. As such, they may be considered outside the pri-mary obligation of this study. However, it is through theseprograms that the fullest realization of the HLH payload po-tential and overall effectiveness will be obtained.

RECOMMENDED PROBLEM-AREA PROGRAMS

The criteria used for defining a problem area are as follows:

1. The characteristics imposed by the requirements ofthe HLH are significantly more severe than thoseexisting in current practice.

2. Analysis of the problem indicates dynamic effectsthat prohibit: a confident extension of presentpractice without trial.

3. Solution is not necessarily possible in a short timespan.

76

4. Alternate approaches impose a penalty.

Items to which all of these criteria apply are consideredmandatory development items. They apply specifically to theHLH.

Overrunning Clutches

Location of the overrunning clutch on the transmission inputshaft (Figure 40) results in a clutch rubbing velocity50-percent higher than current experience. It also places theclutch in close proximity to turbine-frequency vibrationsources. Power transmittal requirements also contribute to theincreased velocity. The recommended approaches to increasingconfidence in overrunning clutches applied to the HLH would in-clude an investigation of the factors which influence theability of sprag-type clutches to survive a high rubbing ve-locity. Material hardness and finish, sprag geometry, andlubrication would be considered. An investigation of thehydrodynamic lifting of sprags from the rotating race shouldbe made. This might be conducted by methods similar to thoseused in determining gear tooth oil-film thickness.

A first-phase effort should include overrunning tests ofpresent-day-technology sprag clutches, with variaticns in lu-brication as to quantity, points of admission to, and egressfrom, the working surface. It is understood that such testingwill be conducted in the near future by a major clutch manu-facturer. An investigation must also be made of the vibratoryimpulses which may be imposed by location of the clutch at theengine-transmission interface; additional inputs from thedamped, supercritical speed shafting should be included.

It is also recommended that a study be made of the overrunningclutch function. This would determine whether fundamentalchanges in design approach can reduce sensitivity to lub:.i-cation, improve reengagement control, and increase overall re-liability. This study should also consider elimination of themechanical overrunning clutch, possibly by tranbferring some ofits functions to the power turbine.

Scoring of Gear Teeth

The unusually larce loads imposed upon the HLH gearing indicate

77

the need for coarse gear tooth pitches, in order to realize aminimum-weight gear set with a satisfactory balance betweenstrength and surface durability. Coarse pitches, combinedwith other factors which may be present, are conducive to earlyscoring-type failuzre. it is therefore recommended that furtherinvestigation be conducted to increase the confidence in thesegear applications. This would include the following:

1. Detail examination of typical HL11 pitches for scoring,using advanced computer techniques for optimization oft]-e form.

2. Tests of sample gears using optimized tooth forms,various surface finishes, and various materials andsurface treatments.

3. Verification that the analytical technique providesdesign information sufficient to eliminate scoring inHLH-type gears.

Engine Control System for a Four-Engined Helicopter

Current twin-engined helicopters present a number of powermanagement problems to the pilot. Maintaining uniform torqueoutput from each engine may divert the pilot's attention fromprimary flight functions. For this reason, automatic powermanagement is being developed for turboshaft engines to be usedin twin-engined helicopters. The extension of this principleto four-engined helicopters should be investigated.

Since the engines presently being considered for the H•{ do nothave an automatic power management system, the problem of powermanagement for a four-engined helicopter must be investigatedto define its scope.

A program for setting up and operating a four-engine simulatorwith engine instrumentation is recommended. The results ofengine and flight-contiol inputs would be presented through theinstrumentation, thereby permitting evaluation of the problemof pilot power management.

RECOMMENDED DEVELOPMENT-AREA PROGRaM

It is highly desirable that certain technical areas be given

78

i-. m --- - - _ _ _ _ _

the stimulus of development, in order to provide advantagesfor all classes of helicopters equipped with mechanical drivesystems. To illustrate the effect of advanced stress levelsupon the HLH transmission, desi7n studies are shown and aresummarized in Figure 7 and Figure 8. Attainment of thesestress levels is believed practicable in the 1970 time period;however, this will necessitate pursuit of active developmentprograms in the immediate future.

The recommended development areas which have shown improvementpotential in the design study are as follows:

Improvement in Gear Tooth Strength

A demonstrated improvement in bending fatigue strength of gearteeth will reduce tooth size and tend to reduce both scoringand dynamic problems. The improvement is also desirable toprovide reductions in weight and size.

The recommended program would include the following:

1. Analysis of the geometry factors influencing geartooth strength

2. A continuation of the gear fatigue testing conductedby the Vertol Division, with the objective of improvingprocessing variables such as heat treatment, and ofinvestigating improved materials and gear tooth forms

3. Sufficient testing to provide a high level of confi-dence in the reliability of gearing at the increasedstress levels, with optimized combinations of form,material, and process technique

Supercritical-Speed Shaftinq

The supercritical speed shaft has proved of particular advan-tage to the HLH configurations under study. With subcriticalspeed shafting, the combination of increased shaft length andincreased horsepower required for the RLH would make for in-creases in shaft weight and complexity. However, by sub3ti-tuting supercritizal speed shafting, major gains are realizedin several areas. These are: (1) an estimated 30-percent re-duction in shaft weight, and (2) a 75-percent reduction in thenumber of major component assemblies incliding supports. Each

79

- m -• •••m mm m m • m m -i m n m m mmm m m

extra support assembly represents additional mechanicalcomplexity and is, therefore, a potential failure origin, aswell as being another maintenance item.

A recommended program for the further development of the super-critical speed shaft would be directed as follows:

1. Continuation at Vertol Division of present bench testsof a full-scale shaft, culminating in testing of theshaft installed in an aircraft

2. Development of an aircraft-type damper suitable foruse on the HLH shaft system

Bearings

In addition to the general development effort being exerted bythe bearing manufacturers to improve materials and manufac-turing processes, it is necessary that specialized studies beconducted by the transmission designer. These will lead tofurther understanding of specific helicopter transmission prob-lems. It is therefore recommended that the following investi-gations be pursued:

Planet Support Bearing

To achieve the reduced planetary diameter made possible byfull use of improved bearing capacity, it is desirable toextend existing analytical techniques. The type of load-ing, outer race deflection, and centrifugal effect are allunique to a planet bearing. The planet gear surrounds thebearing as a rotating outer race. Loads are applied tothe flexible (outer) race at opposite points.

The resultant deflections have been shown to have substan-tial effect on the position of the load zone, and upon peakloads. Preliminary analysis (Reference 2) has indicatedthe benefits of permanently deforming the inner race tocompensate for elastic deformation of the outer. Thiswould be designed to extend the load zune and average peakloads.

A computer program should be formulated to further analyzethe effect on roller loads. A follow-on test program is

80

necessary to evaluate the analysis, using an existingplanetary system.

HiL-Speed Bearings

Again, the confident use of improved bearing capacitymakes necessary a reduction in simplifying assumptions.Presently available programs are progressing toward arigorous analysis of the rolling-element forces andmotions which determine actual bearing life. It is neces-sary to continue and systematize this analysis, to definemore exactly these body forces. It is also important toinvestigate the circumistances that cause bearing elementsto skid instead of roll, and to eliminate this problem.

The extension of present analytical effort to include bodyforces arising from high speed operation is required. Theeffective spring rate of the rotating bearing should alsobe qualified, to better solve for bearing external loads.The effect of misalignment on roller bearings, with result-ing edge loading of the rollers, should be analyzed, andverified by test.

Bearing Cage Design

In recent years, aircraft bearing capacity has been raisedby increasing the size and number of rolling elements.This has, within a given envelope, reduced the space avail-able for a cage to locate and control the elements. So faras is known, this increasingly-critical component has notreceived analytical treatment to determine the forcesacting upon it, and the consequent stress levels. Bearingcage design should be reviewed, following analyticaleffort, and various designs, materials, dimensional clear-ances, and lubricant coatings evaluated by test programs.This work will be directly applicable to increasing theemergency nonlubricated capability of bearings, since thecage is considered the most critical bearing element whenlubrication is denied.

Low-Speed Bearings

A bearing, such as a rotor shaft support bearing, operatingat low speed and under high load is subject to unique prob-lems. It is postulated that certain speed-load combin-

81

ations prevent formation of a sufficient oil film to sep-arate the race and rolling elements, thereby leading topremature failure. Advanced bearing materials may reducethis problem, but the effect is substantially to deratethe bearing, by using up potential capacity improveme,.t tocompensate for a problem. Successful investigation andsolution will allow significant weight saving in thesenormally large and heavy bearings.

A program to investigate lubrication should be initiated.This would include application of lubricating oil to dif-ferent bearing areas, and also the combination of fluidand solid-film lubrication. An extended program would alsoinvestigate specialized types of bearings fundamentally de-signed to withstand th• combined radial and thrust loadconditions typical of these applications.

Planet Load Equalization

The final reduction planetary system accounts for two-thirdsof the gearing weight of the drive system. it is, therefore,the area where substantial weight gain can be realized fromsuccessful improvement techniques. The maldistribution ofplanet loading increases planet gear and bearing stresses,which in turn increase the planet system weight. Equalizationof planet load can reduce the cyclic load peaks, enablinghigher power to be transmitted through equivalent weight gears.New approaches are required, to provide load equalization with-out increasing complexity and weight. A'study of design pa-rameters is recommended, together with a comparison of existingtest information with calculation. If a satisfactory solutionis found, it would be recommended that the program be continuedto prototype evaluation, using ;:xisting transmission and standsas test hardware.

WEIGHT EVALUATION

To obtain a realistic weight estimate of a drive system fromconceptual drawings requires the application of experiencefactors obtained from designing and building helicopter drivesystems. Without this background, the estimation is almostcertain to neglect hidden factors which significantly influencesystem weight. Trend curves were used to cross-check cal-culation in the transmission weight estimates of this report.

82

In reviewing the drive system weights of the six systemsstudied in this report, several important points warrant dis-cussion. These points are:

1. Dead weight comparison between the six systems

2. Total effective weight (including power loss) andits effect on comparison of systems

3. The effect of designing for available power

4. The correlation in major subsystem weights betweenthe HLH and current experience

5. -e weight effect of transmission improvements

Dead Weight Considerations

The six study drive systems, compared on a dead-weight basis(neglecting power loss), show a difference of 7 percent betweenthe lightest and the heaviest (Figure 30). This is not en-tirely unexpected: since these configurations were selectedfrom a large number of possible candidates as being likely toproduce minimum-weight systems. Compensating features amongthe systems tend to equalize weight. The additional weight ofthe engine nose gearboxes (Configuration IV), for example, wascounterbalanced by a somewhat lighter combining transmissionthan is found in Configuration I. The weight advantage ofhigher shaft rpm was cancelled by extra weight in the reductiongearing (Configuration I alternate). The weight of additionalbevel gearing was compensated for by the higher speed and lowertorque made possible, without additional meshes, by the posi-tion of the combining gears (Configuration III, Dual ShaftSystem). Tables V through VmI show subsystem weights. Figure31 compares historical trends with HLH system weights.

Total Effective Weight Considerations

The effect of power loss, when added to dead weight, makes thecomparison between systems more telling. There is a spread ofapproximately 13 percent between the highest and lowest totaleffective weight. Total effective weight is a better selectioncriterion than dead weight since it accounts for the power thatis not available to lift, but which instead is turned to heat

83

DESIGN CONDITIONS15,200 TOTAL SHP134 ROTOR RPM

ADVANCE STRESS LEVELS

10,000-

DEADWEIGHT 9,000(POUNDS)

8,000 rFFI fIA I II III IV V VI

CONFIGURATION

42 36 43 47 49 50 52

FIGURE NUMBER

15,000

14,000TOTAL

EFFECTIVE.......WEIGHT..

(Pu~) 13,000- [1..."!m.m.mgi.j.Im... hh...L.....L..m...L.L....L~~ 1 1 I

FIGURE~~~~~..... 3..EIH.CMARSNOF... RIESYT

8 4. . ...

TABLE V. CONFIGURATION IA HELICOPTER WEIGHT SUMMARY(TRANSPORT FUSELAGE)

Mission Weight(Ib)

Item Heavy-

Transport Lift Ferry

(1800 nm)

Weight empty: 44,576 44,576 44,576Rotor group (Note 1) 9,520Body group 10,380Alighting gear 3,940Flight controls 2,706Propulsion group (Note 2) 12,213Electrical and electronics 1,275Furnishings and equipment

(Note 3) 824Miscellaneous (Note 4) 3,718

44,576

Fixed useful load 890 890 890

Fuel 8,550 3,850 45,034

Auxiliary tankage --.--- 5,500

Cargo 24,000 40,000 --

Takeoff gross weight 78,016 89,016 96,000

Design gross weight 80,000 ......

Notes

1. Three rotor blades with 48-foot radius and 3.17-foot chord.2. Transmission design limit is 15,200 shaft horsepower.3. Cabin interior includes cargo provision for troops.4. Includes 5-winch cargo system.

85

TABLE VI. CONFIGURATION IA PROPULSION GROUP WEIGHT SUMMARY

Item Weight(lb)

Engines - LTC 4K-2 (4 req'd) 2,500Air induction (4 req'd) 20Exhaust (4 req'd) 64Lube system - engine (4 req'd) 30Fuel system: 480

Tanks, fittings 345Engine lines 60Boost pumps 12

Systems, controls, supports 63480

Engine controls (4 req'd) 92Starting - hydraulic - engine 178

(4 req'd)Drive system: 8,849

Aft transmission 4,080Forward transmission 3,306Aft rotor shaft 320Interconnect shaft 327

Engine shafts 50Lube system 625Rotor brake i11Miscellaneous 30

8,849Total 12,213

TABLE VII. CONFIGURATION IA LUBRICATION SYSTEM WEIGHT SUM-MARY

Item Weight(lb)

Oil cooler - fwd transmission 54Oil cooler - aft transmission 87Blower and drive shaft 55Plumbing 181Supports 11Cooler baffles 5Oil (at 7.75 lb per gallon) 232

Total 625

86

TABLE VIII. COMPONENT WEIGHT SUMMARY

Component Weight (ib) for ConfigurationsI II III IV V VI

Aft transmission (Note 1) 2886 2511 2734 2371 4220 2844Fwd transmission (Note 2) 3306 3732 3539 3304 3304 3377Engine transmission -- -- -- 472 -- --

Combining transmission -- -- -- 352 -- --

Aft rotor shaft (Note 3) 1946 1946 1946 1946 320 1946Interconnect and

engine shaft 377 269 284 377 367 377Lube system (Note 4) 595 575 625 655 625 595Rotor brake 11 i1 111 111 111 111Miscellaneous 30 30 30 30 30 30

Total 9251 9174 9269 9507 8866 9169

Notes:

1. Aft transmission includes accessory drive and blower drive;combining section where applicable.

2. Fwd transmission includes fwd rotor shaft.3. Aft rotor shaft includes thrust bearing.4. Lube system includes oil.

87

HLH CONFIG 1}tiii ~j4Ifl{T I (CONVENTIONAL)

0) 1PRE-1960-DESIGN

E4 TRED.......

*CURRENT TREND

loooI

E40)7

>4

7rjEn~~ VA

.o ...... ... L .L .....i. 1_ _ _

1 ~ ~~~ ~ ~ ~ 2 01 6 7 a 100lo

FUNCTION OF DESIGN POWER

*AT 15,200 HP

FIGURE 31. WEIGHT TRENDS COMPARISON.

88

and dissipated with resulting additional weight for coolersand blower. Configuration II, because of the elimination ofpower carry-through in the interconnect shaft bevel gearing,has a low number of meshes, and therefore, a low total effec-tive weight. Configurations IA, I, III, V, and VI carryslightly highe. total effective weight; Configuration IV issubstantially higher.

The conclusion to be drawn is that selected best transmissionsystems can vary considerably in arrangement, and yet beidentical, within the limits of accuracy of trend calculation,in dead weight.

The selection of the optimum arrangement was, in this case,based upon factors other than dead weight.

Effect of Power Criteria

The drive system power criteria used in this study were baEedupon the transmittal of full available power from advancedversions of current engines, at a rotor rpm corresponding tca 48-foot blade radius. The system weights (Figure 30)uniformly reflect these criteria. Less severe criteria couldbe cons 4dered--for example, the power required to meet exactlythe mission specification (11,500 shp). The effect on trans-mission system weight -s shown in Figure 32. This shows that,by limiting power to 11,500 shp, a weight reduction of 20 per-cent should be realized, resulting in a Configuration I systemweight of 7100 pounds.

It was believed that the commitments of this study were bestmet by using transmission design criteria which would be con-sistent with potential growth horsepower. It is historicallydemonstrable that powerplant growth leads tranrmission capa-bility and that, as increased power becomes available, improvedhelicnpter performance is readily accepted. Transmissionproblem areas have therefore been defined for a system of suf-ficient capacity to accept this growth power, with attendantcomponent size, velocity, and weight. It is emphasized thatfinal design of the HLH drive system will reconsider alternatepower criteria and that, in consequence, the weight of thesystem may be reduced.

89

= -. - -MT"

CONFIGURATION IA SYSTEMADVANCED STRESS LEVELS EMPLOYED

ROTOR RPM IS CONSTANT (134 RPM)ROTOR DIA IS CONSTANT ( 96 FT)

9,000

-DES IGNS~POINT

8,000

S"7,U000

0 E-AMISSION SPECIFICATION POWER

11,000 12,000 13,000 14,000 15,000 16,000

DESIGN HORSEPOWER

FIGURE 32. EFFECT OF DESIGN HORSEPOWER ON WEIGHT.

90

Correlation of HLH to Current Tandem-4e.licoptp_ Transmission

Subsystem weight percentages showed close agre-ament with thosecurrently experienced in tandem-rotor drive systems. Themajor difference was in the percent system weight attributedto the aft rotor shaft. Rotor blade radius is an influencingfactor on pylon height; pylon height detemnines the necessarylength of this shaft. The blade radius selected for thisstudy was sufficient to increase the aft shaft portion of thetotal weight. An approach to reducing shaft weight is to placethe final reduction immediately beneath the aft rotor. A ver-tical interconnect shaft replaces the rotor shaft from bevelcase to final reduction; weight is reduced as a consequence ofthe reduced torque. Bending forces and bending stiffness re-quirements are eliminated by a moment-carrying support for therotor shaft above the final reduction gearing.

Effect of Transrmission Improvements

The transmission improvement areas which contribute to weightreduction and which are embodied in the estimates (Figure 30)are advanced gear stress levels, increased bearing capacity,and supercritical speed shafting. The combined effect of theseimprovements is to give an estimated 1200-pound weight re-duction in Configuration I drive system, as compared to thesame configuration using 1960-era design practice. This re-presents a 13-percent weight reduction, at no sacrifice int. ansmission capacity.

The macnitude of stress-level increase and the effect of super-critical shafting is detailed in foregoing sections. The at-tainment of these improvements is possible within the timeperiod allotted for Hill development. To meet overall reli-ability a:id maintainability goals, it will be necessary topursue development programs such as those outlined in the pre-ceding section. Existing programs leading to the realizationof operational supercritical shafting and improved bearingsmust be continued.

Decreased face width and diameter of planetary and bevel gear-ing account for the largest part of the reduction in systemweight. Increased bearing capacity makes possible a furtherreduction, especially in the planetary section. Reduced gear

91

diameters impose greater reaction loads because of highertangential forces. Without a commensurate improvement inbearing capacity, it is found that bearing diameter often be-comes the decisive sizing factor for housings, mountings, andplanet ring gears.

An additional approach to weight reduction is the substitotionof titanium for ferrous components other than gears and bear-ings. These could include rotor shafts, gear shafts andcarriers, shaft adapters and couplings, bolting hardware,spacers, and shims. The decision to use titanium for a spec-ific application is predicated upon continuing research on thematerial cbaracteristics and allowable stress levels. Such aprogram is presently under way at Vertol Division. It is alsopredicated upon stiffness requirements, especially in shaftapplications, where the lowered modulus of titanium may requireincreased diameter to maintain stiffness while preserving theweight advantage.

Assuming that maximum use can be made of titanium in the abovementioned components, the weight advantage to the HLH drivesystem represents an estimated 800 pounds. This latter ap-proach can be applied separately from t1p" )f advanced gearand bearing stress levels, since the two approaches are in-dependent. Added together, they represent a possible drivesystem weight saving of 16 percent of a conventionally designed(1960-era) system. System weights shown in Figure 30 do notinclude the effects of using titanium. Figure 33 compares theweight effect of conventional design, first with that of ad-vanced stress levels and supercritical shafting, and then withthe effect of these plus titanium.

92

a • •* - -• - - • • • • •-

100 .........

90 . .

WEIGHT %......

80.......

90 ....A..... ... C

FIUE3.WEIGHT EFETOTESLEE N AEIL

893

HR = HIGH REDUCTION

PLANETARY TRANSM 1

INTERMEDIATE SPEED ¶)ENGINE S3PEED

CONFIGURATION I CONFIGURATION VII1 42

38 5.29 2 (40 S,5.03 2 (10 S/B) I

NO. NO NO. NO. NO. NOSEQUIV. GEARS EQUIV o GEAF

MESHES XMSN AND KSH•S BOXESN

ER ENG. 6O j (S/B) PER ENG. E(s/

43 5 .03 26.03 6 (15 S/B) I (4 S,

ILR

LR

INTERMEDIATE

Ilt.TERM-EDIATE SPEED &

CONFIGURATION VIII CONFIGURATION V

FIGURE 34. CAN4DIDATE CONFIGURI

A9

HR =HIGH REDUCTION LR =LOW REDUCTION

PLANETARY TRANSMISSION PLANETARYTRANSMISSION

ENGINE SPEED INTERMEDIATE SPEED

CONFIGURATION VII CONFIGURATION IV

42 6.03 I 6 43L. 5.29 (10 S/B) 6 (15 S/B)

NO. NO. NO. NO. NO.EQUIV NO* GEARS EQUIV N GEARSMESHES XMSN AND MESHES XES AND

PER ENG. BOXES (S/B) PER ENG. BOXES (S/B)f•'i 37 5.3 37

5.03 2 (4 S/B) 1 5.03 2 (4 S/B)

LR IALRINTERMEDIATE SPEED INTERMEDIATE SPEED

S, ,LR LR

SCONFIGURATION V CONFIGURATION VI

Ot FIGURE 34. CM.DIDATE CONFIGURATIONS. (Sheet 1 of 2)

____

HR HIGH REDUCTIONPLANETARY

TRANSMISS ION

INTERMEDIATE' SPEED

CONFIGURATION II CONFIGURAT:42

4.29 2 (10 S/B) 4.79 2

NO. N NO.NO. NO.EQUIV. GEARS EQUIV. NO.MESHES BOXES AND I MESHFiS XMSNBOE MEHE BOXES-PER ENG. (S/B) PER ENG.B

S6.03 (225/B) 4.29 2

~ LRR

ENGINE SPEED

ENGINE SPEED

CONFIGURATION X CONFIGURA

FIGURE 34. CANDIDA 1

97 Am, •

- , , .... ... . ..

HnRLR

V INTERMEDIATE SPEED HRI LR

ENGINE SPEED

AT' CONFIGURATION IX CONFIGURATION III44 5. 3244

4.79 2 (4 S/B) 2 (10 S/B)

NO. NO. NO. NO.EQUIV. NO. GEARS EQUIV. NO. GEARSEI XMSN MESHES XMSN ARS

ES MESHES BOXES BOXESPER ENGS (S/B) PER ENG. (S/B)

4.9 246 394.29 2 4S/B) 4.62 2 (7 S/B)

HR HR

, ENGINE SPEED HR " •

HR ENGINE SPEED

k, CONFIGURATION XI CONFIGURATION XII

ATURE 34. CANDIDATE CONFIGURATIONS. (Sheet 2 of 2)

718

11I

FIGLPZ 3. FRWAR ROTR TANSMSSIO COVENTONA

STRESS~~ ~ ~ ~ LEE EASADBERNS

99N

) III

I' A

rK� �I -j -'I

'7-

�

/

-- E,5O 4' PM _______ �.d IP7�C�WNfCT

o , I � * f

¶t�*�f.

8

134 ePM

23oR Pwl

IV ,'-,-- --,-

FIGURE 36. FORWARD ROTOR TRANSMISSION - ADVANCED STRESSLEVEL GEARS AND BEARINGS.

101 A

J-7

5 3'ZE, " C.IF S vAk$,

-A_______6'0*A

av=I

AWEMN

2* L _

FIUR 37 DR-,r -YTM-C~FGIRTO K126

(Sheet1 of 2

103iŽ D5~~

-Sr 4'~ A~FxAM

131

'^R B RA K

~ 5...........

RIF-

:E::N4EC. SHAFT

FIGURE 37 DRIVE SYSTEM - CONFIGURATION I - SK 14216.(Sheet 2 of 2)

105 A

ROTOR, S-4wA-T TnlktiS- 5ý,PPO~r. I SKI14 219.

-POWER PLANT.-

SHAF T DAMPEIR..5 14W

Ej~'41z SWT-IGjloo~ 'ICP T6 ALTERNATE

FLEXIBLE JLI-.K

0 50 100

SCALE Ov IhNtEs.

8B

F'-GURW 38. AIRFRAME ARRANGF2ENT CONFIGURATION I

SK 1.4217. (Sheet 1 of 2)

107A

- -WOr

qVA)ww 5gwr^%%m ofMCmmfl

AP7 "ý- Al

-a tot --- 4 A

4-41,- -"-

4m * 100w-a

J--

7F w~ArF ~

.rr XAMB AM __-

AA~r J4-VJ T7~ 4-

-r _Ao As~-

A W -~ 40a IW- 4?AM ivz%-,a ,A Amw AWW A/4O A j~qvOzw i

C ac-N- Aats AJ AV'Ai

-W .- Z M - Wr-~rCO

"v7-

FIGtL'*E 38. AIRFRAME ARRANGEMENT -CONFIGURATION I-SK 14217. (Sheet 2 of 2)

109 A

AIVA

J-J, ftw

Ile0~~ A- AIX

JWo-4ýAlI~tAv iL AA AFAIC0

Amw~~APeA -mf-o mvrW~

-AlfT 11FA',VPJ

:7z7

41A? -&-ff-

V ID AW7AW

AuwB

/

1341PPM

rs "

ow 0

FIGURE 39. FORWARD ROTOR TRANSMISSION - CONFIGURATION I -

SK 14218.

Mae

#AOUNTINr. FACE NCW4C

1 ci, , "&p

K/17

ICE5$OMY ~LUSE PUMP

Aii

ROTORJ, SHAFT

IB4 RPM

S\ . -_ --- ----- - ---

4B P- P .

RATIO -7.jT

BLOWERP DRPI',E-\

FWD _NTRONC5HAFT 6i5olF?'.i

FIGURE 40. AFT ROTOR TRANSMISSION - CONFIGURATION I -

SK 14219. (Sheet 1 of 3)

113 A

- I - - .. ... .._ - - - - - - . -"' "'' • -' • •-•---- - -' .. . . .. . .. . .

ý_H AFT

-A ------ •

12L~

tTt

PM.G, 0 _RPM

. / --•• . ,/, j-..R _• • .OTO R • •-'-

SA-IE-CE

I- ,

S. . .. 0 I • • •

SCALF.- INC• liFS

8!

FWD }I~'i j i .

T " ,L . HAV rTG_50_RP 5YN4CH

i t 11

OVER-PtiN1~ING

'-EIJGIE N'PJTS.i6,000 RPM, 08425345S6

1• , . I I 1 1

SECT. 33- 5ChLE- INCHES

FIGURE 40. AFT ROTOIR TRANSMISSION - C0NFIZ11RAkCION I -

SK 14219. (Sheet 2 of 3)

115

I-4YDPUMP -~ . M NN, MOTOR

4-1-

Hyn PUMP ------ HYD Pump

0~ -

SCLE INCHE

LjURBE PUMP -p OO RNMISIb-CNFGR1 I

slK~~- 1419 (See13of3w117

I

,.~Pj _______/

ROTOr .;#AF, r

FIGURE 41. AFT ROTOR SHAFT - CONFIGURATION I - SK 14221.

119

A

V4840

RomeR s#AFr

wlrl Or 4 0 40

T*OVSAV3310'V

8PTMqAC

_ Iiiw.C.ik

rA".t5 r £7f,,-m4;

FIGURE 42. HIGH-MOUNTED REDUCTION - ALTERNATE CONFIGURATIONIA - SK 14222.

121 A

,,- .0&

rA",s~ rraw

SC, 16 lf- iM Af 3

ALI

'WERNATE CONFIGU~RATION

kNOt sift AN

ZI Z- AWIM

-r i-lowesNw *S

ALTE OMMT COFIGAO IA -i S I425

123if

SK14228. FORWARD QOTORTPAANSMISSION ASS E W,13Ly.

:FK414?l AIRFRAME/ APRANCMENT

\\--RIGID COUPLING =IPOINTINTERCONNECT SHAFT DAMPER.

SFLEXIBLE COUPUm.

FIGURE 44. DRIVE SYSTEM - CONFIGURATION II - SK 14226.(Sheet 1 of 2)

125

A

THRUST SUPPORT.

AFT ROTOR $AiTT

k134

R

/\

SLOWER - LUBE- OIL

O LUNB. OIL AFT ROTOR SKAFT.M

It ". .

A ACCESSORY CEARBOX.

I L MILFOINT SuPPolrT. RIGID ~ CO PLN ~Rrom BRAKE .

ThTECOWNcT IAFTDAMPR.\SK14229. AFT ROoreINTECONECT MFTDAMPR.ýTRANSMISSION-

FLEX IBLE COUPLING -

0 so loo.

SCALE OF INC'3wES.

POWER PLANT.

/ (G4 ALTE•NAT'E

FLV~lBLE COj?L'N46.

I ENGINE S"AFT-16,000 RPM ,_TERCONNECT SHAFT

SHAFT DAMP-R. 6100 R PM

SSK 14229..

II

KI

FIGURE 44. DRIVE SYSTEM - CONFIGURATION II - SK 14226.

(Sheet 2 of 2)

127 A

AECT SHAFT ROTOR SHAFT 14UST SUPPORT. i SK*t4M,

I7

0 5D too.

SCALE OF INCHES.

B+

.p k.-

Z- AW ft~

ASW-SVWALW

zm Am '-r 5?t~nlW A*9W&W~r WAl

S,~Af ASi4F Jde whr~

PThD ~

V"? ASE~-

I=~ A~L- -

Am Avo-

FIGURE 45. AIRFRAME ARRANGEMENT - CONFIGURATION IISKC 14227.

129 A

~-A-

A~Aw~~ AKW 0

W47j'v7'ftW SWd~.P~t - V0m

AWVto W * W ----- I- -

~Wpr

mo -I---?- ~4" e~mr lIw

IB

X17'~

11~~ -#- SK122. (b t1 f2

413

A

7"'ýf FPLANIETARY SE-MeOW.

/((REFER TO SKC. 14229., A. SCNOJE T .

LO PM

Ll

71::

OVERRUNNING CLUTCH

SECTION A.A. HENGINE INPUT *6PO R PM

FIGURE 46. FORWARD ROTOR TRANS4I SSION - CONFIGURATION11 - SKC 14228. (Sheet 2 of 2)

133

,T

482

~~II* BLCW-NE.R DRIVE - y~

FW D- roOO RPMSYN~CH 5H-AFT ~'

ENIJCIW~E INPUT-

FOR. 5ECTIOMl VIEW

IFI!!

FIGURE 47. AFT ROTOR TRANSMISSION - CONFIGURATION II -

SK 14229.

135 A

Iý4~ ýVPiz

:1ACCE55-Q. D)RVE

"A~ I

RPM

__TO BRAKE

A-

5CALE- iNCHES

-IGURATION II-

sl~oo~V A IR~ RIAN DO AfN ,w*- AV.*

11ýWW M AAW~rAA

ZALW MAW ~"c

mm~~L Arp e

AW4qI AkINJ

FIGUR 48.AIRFAM ARRANVIEk ACOFGUAI~ I

137A

mAM AV-A AWA c-

_-.fXM OAWV iN7NA~

WA, wet,

.%Vvwf 949 PW AM~h A~W9

Z"W&Y ~AwvEW.

AW.FIC? AU

B;WMJ~fWW7~

I

II

ro A.1 W~_-

1 ,

FIGURE 49. AFT ROTOR TRANSMISSION - CONFIGURATION III -

SK 14243.

139 A

~r'o 49's roi

I ~ T O Ak r4S'O

mvkve Idwr ZAOybVq TOIA, e7Are*,~0 sem A) 4dS -4

A-A~

-lCVM S 1UCO4

1414

GA4 R.et4.

ifMM't"&JNIM PA

Glo M'Im ft"ft*I~~T5 I.AFT w sO~1~bic55i@T. uabgI

. .......

L. ...

a-0" t _____.

- -j.~2, .~ SCE SK.14237.

4E KANUU

143<

[ - ~ REAR eorok ozwE s.-,,tj.-4--

---REMUTIONGEABX

FOUR T.64EZGNS

VIEW wc±t4 FOtC.DRD.

A,- JOR T.64E4GINES.

ILW.

FIGURE 52. COMBINING TRANSM4ISSION - CONFIGURATION V-SIC 14237. (Sheet 1 of 2)

145A

-ANA

4AAM71,O~ A*?C~,C

jrojA-A

A/

A/

F j CO -C'

/ /

I /

FIGURE 52. COMBINING TRANSMISSION - CONFIGURAT'ION V -

SK 14237. (Sheet 2 of 2)

147 A

ADIN-/OAf

-7b*

EMISS ION - CONFIGURATION V-teet 2 of 2)

147

BLOWER LUSEOIL CCOLING_

•-7F~ti ' RTO I E•

FIUR 5. TUEyCONFIGURTIO VI SK 14244.__

149 - ; "

VIEW LOOKIN,. F•-•wARD

FI'G'URE 53. STUDY CON;FIGURATION V]" - 5K< 14244.

149 A

~ __- - .. H Ji| -

THRUST 5!;PMCT

FAFT ROTOW 5-AVT-

- ~134 R.P.M.

6LWER -LUISE

~OIL CCOLING-.

L....,AE. POWR PLANT

I T5! 51%WN

-(G ALc?NFATj.

-TIN

I. ___ I I IAFT R07DRTRANS¶I5S~iO'.

SCAL.E OF INICHES.

SECTIOHUM

P07roj /L4d RPM4

I EI

.- W-I

FIGURE 54. COMPOUND REDUCER STUDY -SK 14232.

I If

ILL,

,oil

O 3 4 5~ S(cT/OA'

A !

- -r• I I '

I i i t__ _ _ _ _ _ _ _ _ _ _ _ _ _

aldo RPM ___

-V -

rI

FIGURE 55. THREE-STAGE REDUCER STUDY -SK 14231.

1534

,r --- I•N-/ } -

'-1

A'

II

.1. / *1 t

I // , _

. ,�

N K'N

K'N .1 4.'

I---- N _ _ _

- '--------------.----- ,' N.

- - - J I -

_____ _________ I V-- '--------'-t-----' K;' \ I, /

____________________ \\'�� � ii

N I �

'I \ �NN

t I---� '

/ N

// � 'i�'Ks��-- 1-

// �

I NI I! �I \j\ �K'- A

SEC 7/ON AA

8

ROTOR 13RP

jA

I -• -

I I

t- i

j i • aoloi RPM '

FIGURE 56. SPLIT-TORQUE REDUCER STUfDY -SK 14230. ,

155

S- - •

\-- 'I \"

,"I .

S "I • -:

I,/

-' ICJA V4L fT iiN "a 7. , ii

/ #:f

,:I, •B ,

"I f I

*,~ 44flCO 0

-J

/ 139 85PD/ _ RINr4 (.L.AR

-N,--"

\\

1& Pi - -

k"/ " .P,,NF--75-62)

--\ , \•

';SL RUT-/ R\ !Rt./

FGR 57 ROLE GEA REDCE "D SK 1433

157

\!~~~~ \ iO\ -TLIN4• •)"

_- • i .-- "/ '--. .. I

'-SRLC T R•L •,bG• -/ETOx

157

S" I - ! ! ! I II I II-- S

4400 0 D;r ~IkCA$E

"INC

~~iit

IT -jINPT -~- IiROTORIPT3 80PD _______ ___ 'F

5HF SNS (Z) 154 RPM-~ 5MAFT 1 S0) -

1! 75 PD'*4ANETS 16

15-82 PD

PLANETS ( 2)

\ 10,75 PD REACTION~.ANE.T('~ 41 5"4AFT

~~-~---STRUCTURM.L RWlts

SECTION .,

SK 14233.

IB

BIBLIOGRAPHY

1. Buckingham, E., Manual of Gear Design, The IndustrialPress, New York, New York.

2. Harris, T.. and Broschard J., Analysis of an ImprovedPlanetary Gear Transmission Bearing, SKF Industries,Inc.

3. Morrison, L.D., et al, "Tie Effect of Material Varia-bles on the Fatigue Life of AISI 52100 Steel BallBearings", ASLE Transactions, Volume 5, No. 2, November1962.

4. McIntyre, W.L., How to Reduce Gear Vibration Failure,paper given before AGMA, February 1964.

5. PaJmcren, A., Ball and Roller Bearing Engineering,SKF industries, Inc.

6. Stalter, J.L., Wind Tunnel Test of a 1/8-Scale Modelof the YHC-IB Helicopter, Engineering Research ReportNo. 344, University of Wichita, Wichita, Kansas,August 1959.

7. The Vertol Division of Boeing, Heavy Lift HelicopterStudy, Report R-283, prepared for USAAVLABS, FortEustis, Virginia, June 1962.

8. Gleason Gear Works, "Improved Method for EstimatingFatigue Life of Bevel and Hypoid Gears", SAE QuarterlyTransactions, Vol. 6, No. 2, April 1952.

159

mm • •a 5t

APPENDIX I

EQUATIONS FOR EQUIVALENT NUMBER OF MESHES (Me)FOR POWER LOSS COMPARISONS

ma = actual number of meshesR = total ratio

For Single-Stage Epicyclic aypes:

Simple planetary: ma = 2; Me = ma (1 - i)R

Compound planetary: ma = 2; Me = ma (I - 1)R

Split-torque planetary: ma = 4; Me = ma (1 - 1 D3R 2RD1

Where

Dl = diameter first sun gearD3 = diameter first ring gear

For Study Arrangements (Figure 18):

Split-torque compound: ma = 4; Me = ma (1 - 1 D4D2R 2RD3Dl

Where

Dl = diameter of first sun gearD2 = diameter of gear of compound planetaryD3 = diameter of pinion of compound planetaryD4 = diameter of first ring gear

Compound-simple planetary: ma = 4; Me ma (I 1 1 )2Rc 2Rs

Where

Rc = compound planetary ratioRs = simple planetary ratio

161

Three-stage planetary: ma = 6; Me = ma (1 - _1I_ _3R 1

1 1-

3R2 3R3

Where

R1 = first-stage ratioR2 = recond-stage ratioR3 = third-stage ratio

Epicyclic Roller Gear:

Odd number planet rows: Me = ma (1 - )R

Even nuaaher planet rows: Me = ma (1 + 1 )R

Where

R = total ratio

Roller gear (Figure 57): ma = 3 = Me

162

APPENDIX II

SUMMARY OF BEARING LOADS AND GEAR STRESS LEVELSFOR CONFIGURATION I

FWD. XMSN AFT. XMSN

POWER RATING:CUBIC MEAN POWER = 10,000 HP @ 16,000 RPM COMB. XMSN

(ROTATED 900)

NOTES: 1. All B1 lives shown are based on four-e ine operationwith the exception of bearings o & a which arebased on a two-engine-out condition.

2. All Dearing capacities were increased by a 70%growth factor.

FIGURE 58. BEARING SUMMA

163 A

LOADS3 LEVELS 0 u

H 0Z E-

1 Roller 300x380x38 49,000 -- 134

2 Ball 320x480x74 39,000 51,000 134

3 Spher-Rol 95x200x67 57,000 -- 443

4 Spher-Ro! l10x200x53 27.,.000 -- 1270

S5 Ball 160x220x28 -- . 340

6 Roller 220x270x24 9,400 -- 2350

S7 Ball 130x200x33 6,100 3,000 2350

8 Roller 90x60x30 9,000 -- 6150

9 Ball l0x20ox38 o7,000 6150

10 Ball 105x160x26 -- 1,750 6150

- 11 Roller 105x160x26 5,500 -- 615012 Roller 105xl60x26 6,000 -- 6150

13 Ball 105x160x26 -- 1,950 6150

"Z 14 Ball 110x200x38 -- 7,800 6150

15 Roller 130x200x33 9,500 -- 6150

S16 Roller 5i0x210x28 11,500 -- 2350

17 Ball 240x309x30 5,200 3.100 2350

is Ball 410x490x38 5,500 -- 13419 Ball 320x480x74 14,000 |51.000 134

20 Roller 130x200x37 6,000 -- 6150

COMB. XMSN 21 Ball 130x200x33 -- 4,150 6150900) 22 Rollcr 120x189x2S 11,000 -- 6150

z 23 Roller 75x160x37 7,1i00 -- 16000

24 Roller 70x125x24 2,500 -- 16000

25 Ba1l 70xl10x20 -- 690 16000

• 26 Ball 70xl50x35 -- 2,750 16000-2_ýnne operation0e operhtiona0 27 Rollzr 73x160x37 7,000 -- 16000

(~22 which are ____

28 Rol ler 70x125x24 2,500 -- 16000

)y a 70Xo 29 Bal 70x]10x20 -- 690 16000

30 Ball == 70xl50x35 -- 2,750 16000

FIGURE 58. BEARING SUMMARY - HLH CONFIGURATION I.

IMAB

LOADS

4 A a) • " < ,4 H

.34 300x380x38 4900,0 -- 134 4 97,000 121,500 2560

.34 320x480x74 39,000 51,000 134 43,000 152,000 198,000 2700

43 95x200x67 57,000 -- 443 163,000 165,000 1300

!70 110x200x53 27,000 -- 1270 105,500 131,000 2500

40 160x220x28 -- . 340 54,300 -- --.

r 220x270x24 9,400 -- 2350 517,000 43,750 52,750 2200

.50 130x200x33 6, 100 3. 000 2350 306,000 45,000 46,800 1420

.50 90x160x30 9,000 -- 6150 553,000 56,000 61,000 1570

l50 110x200x38 -- 7,000 6150 677,000 52,500 63,000 1970

50 105x160x26 -- 1,750 6150 646,000 13,100 28.000 11i00

50 105x160x26 5,50:) -- 6150 646,000 34,000 48,500 3800

50 105x160x26 6,000 -- 6150 646,000 37,000 48,500 29-50

50 105xl6Ox26 -- 1,950 6150 646,000 14,800 28,000 8130

50 110x200x38 -- 7,800 6150 677,000 59,000 63,000 1440

50 130x20ix33 9,500 -- 6150 800,000 59,000 75,000 2640

50 150x210'ý28 11,500 -- 2350 353,000 53,750 64,000 2160

50 240x309x30 5,200 3,100 2350 565,000 29,000 34,000 1990

34 410x490x38 5,500 -- 134 55,000 11,000 40,000 47200

34 3 14.000 51,000 134 43,000 124.000 198.000 48

50 130x200x33 6,000 -- 6150 800,000 21,000 75,000 12200

50 130x200x33 -- 4,150 6150 800,000 31,500 46,800 3700

50 120xl8Ox28 11,000 -- 6150 738,000 39,000 60,000 7000

00 1 75x160x37 7,100 -- 16000 1,200,000 59,000 71,000 2190

Q0 1 70x125x24 2,500 -- 16000 1,120,000 21,000 39,000 9780

00 70x110x20 -- 690 16000 1,120,000 7,250 13,400 7620

1 70x150x35 -- 2,750 16000 1,120,000 29,000 43,000 3960

' 17'5x160x37 7,000 -- 16000 1,200,000 58,000 71,000 2340

0 70x1 25x24 2,500 -- 16000 1,120,000 21,000 39,000 9580

1 70xllOx20 -- 690 16000 1,120,000 7,250 13,400 7620

I 70).150x35 -- 2,750 ,,16000 1, 120,,000 29,000, 43,000 3960C

• - mmmmu -

TABLE IX. SPIRAL BE 1

E 0

z E- 0z

0 Hoo P 2 22

o z PINION 2.600 10.000 26 20 25 220 100030' R.H. CCW 3.780 3.7-

GEAR 2.600 26.154 68 20 25 78030' 100030' R.H. CW 3.780 3.7

0

O Z PINION 2.600 10.000 26 20 25 200 85030' L.H. CW 3.780 3.7

GEAR 2.000 26.4 ,568 20 25 65030' 85930' R.H. CCW 3.780 3.7

z PINION 5.000 5.000 25 20 25 2100° 92030 L.H. CW 2.250 .

AFT

PINION 5.000 5.000 25 20 25 2030 82030 L.H. CW 2.250 .2

0W

AFT

GEAR 5.000 13.00') 65 20 25 62030' 82'030_IR. 1. CCW 2.20 .2

J--

165

BEVEL GEAR SUMMARY

Model.: HLH Configuration i

Power Rating: Four Engine -

_15,200 HP at 16 000 RPM

zo uO 4 I

3.780 6150 93,600 34,400 196,000 1580 1.732 16,000

3.780 2350 245,000 34,600 196,000 975 1.732 16,000

3.780 6150 93,600 34,700 198,000 1580 1.732 16,000

3.780 2350 245,000 34,900 198,000 975 1.732 16,000

1220 16000 15,000 34,000 210,000 1195 2.000 2,0

I l

2.250 16000 15,000 34,000 210,000 1195 2.000 21,000

.250 6150 39,000 34,000 210,000 7410 2.000 21.0002.250 16000 15,000 34,000 210,000 1195 2.000 21,000

!.250 6150 39,000 34,000 210,000 740 J2.000 21,000

B- J

TABLE X. SPUR G'UR

SSUN 3.000 8.333 25 2.360 .164 .516 250 2350

E4 (4)PLANET 3.000 12.333 37 2.300 .199 .516 250 1270

~----------------

RING 3.000 33.000 99 2.030 -- .516 250 --

SSUN 3.000 13.000 39 4.340 .157 .. ,15 250 474

- E4 (6)< '* PLANET 3.000 10.000 30 4.280 .160 .515 250 443

-V

_ _ _ _ _ _ _ __ _ __ _ KRING 3.000 33.000 99 3.76U .515 25 --

167 A

'UR GEAR SUMMARY

Model: HLH - Configuration I

Power Rating: Four Engines -15,200 HP at 16.000 PY

m z to zn

0 E-4 W E-4;z 9 >U 0z C-4

0 E. W E# 0 9C

61,250245,000 36,700 184,000 2,050,000 452 .509 1.474 6220

90,750 35,600 184,000 2;570,000 452 .537 1.474 6375

243,000971,000 35,000 100,000 -- .620 --- 7220

202:0001,210,000 39,400 184,500 615,000 340 .544 1.491 7160

155,000 41,200 184,500 538,000 340 .527 1.491 7250

512,0003,070.000 40,000 125,500 .620 8280

,8

APPENDIX III.

TABLE XI. SUMMARY OF LUBRICATION SYSTEM ANALYSISFOR CONFIGURATION IA

Horse- (BTU/Min) (gpm)Unit Location Efficiency power Heat Flow

Forward Bevel 99.35 7600 2090 6.35Rotor XMSN 1st planet 99.03 7600 3120 9.50

2nd planet 99.03 7600 3120 9.50Rotor Shaft . ...

8330 25.35

Aft Bevel 99.35 7600 2090 6.35Rotor XMSN Comb.bevels 99.35 3800x4 4180 12.70

1st planet 99.03 7600 2090 6.352nd planet 99.03 7600 2080 6.35Accessory 98.00 250 210 0.64

10660 32.39

Total Flow (gpm) = 25.3532.39

57.74

Notes

1. Horsepower total = 15,200, equally divided.2. BTU/Minute = hp x 42.43. Flow (gpm) = BTU

7.75 x .53 x A TWhere: 7.75 = lb per gallon MIL-L-7808

.53 = specific heat MIL-L-7808A T = temperature rise = 80*F

4. The calculated 80OF temperature rise will resultin an actual expected 45°F rise according to VertolDivision experience factor.

169

Securizy Classification

DOCUMENT CONTROL DATA- R&D(Security classilfication of title, body of abstract end indering annotation must be entered .. hen the overall report ix c I. sIed)

IORIG.INATING ACTIVITY (Corporate a•uthor) 26Z REPO RqT SECURiTY C ;.ASSlFICATIONVertol Division• UnclassifiedThe Boeing Company 2b GOUP

Morton, Pennsylvania N/A3 REPORT TITLE

TRANSMISSION STUDY FOR TANDEM-ROTOR SHAFT-DRIVENHEAVY-LIFT HELICOPTERS

4 DESCRIPTIVE NOTES (Type of report and inctuaive d"!as)

Final; 27 June to 27 December 1964S AUTHOR(S) (Last name. firt name, initial)

Mack, John

6 REPORT DATE 7 7- TOTAL NO OF PAGEs 7b NO OF REFS

September 1965 169 Eight11a CONTRACT OR GRANT NO. DA 44-177-AMKC- to ORIGINATOR.'S REPORT NUSERS)

PRJC NO 241(T) USAAVLABS Technical Report 65-56

c TASK IM121401D14414 flb OTHER RtEPORT NOS) (Arty ohrnu.bers S.ho may be e..e.,edthis report)

d. R-37910 A VAILAFILITY/LIMITATION NOTICES

alified requesters may obtain copies of this report from DDC.This report has been furnished to the Department of Commerce for sale to thepblic-

I t. SUPPLEMENTARY NOTES 1it. SPONSORING MILITARY ACTIVCAY

I U.S. Army Aviation MaterielILaboratoriesFort MiaSttis. Virginia 23604

I1. AGSTRACT

Mechanical drive systems for heavy-lift tandem-rotor helicopterswere studied.

Three- and four-engine configurations were analyzed. Configurationsincluded aft-mounted and fore- amd aft-mounted engines, single anddual drive systems, and high-mounted and low-mounted aft planetariesWeights were estimated.

Problems of gear surface durability, multiengine control, and over-running clutches were defined. The effects of increased gear toothbending fatigue strength, higher bearing capacitj supercritical-speed shafting, and the use of titanium and improved ferrous metalswere evaluated. In the more significant factors, the satisfactorysolution of the heavy-lift helicopter drive system lies within thecurrent state-of-the-art.

DD !M1473 UNCLASSIFIEDSecurity Classification

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I

UNCLASSIFIEDSecu rtit CldssiIfi(cat ion

LINK A L:NK B LINK CK E Y WO RDS R R LROE t~ I ROLE ý

Heavy-lift helicopterTransmission ITandem-rotor helicopterPlanetary reduction gearing '

Drive shaftingMultiengine controlSupercritical-speed shafting 1Gear-surface durabilityGear tooth bending fatigue strengthBearing capacityTitanium

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UNCLASSIFIED"-Security Classification

• • • a~ 579245a